JP6744765B2 - Control method and control device for in-cylinder direct injection internal combustion engine - Google Patents

Description

The present invention relates to control of a direct injection type internal combustion engine.

As a control of an internal combustion engine that directly injects fuel into the cylinder, control that injects fuel during the compression stroke and uses the tumble flow as the gas flow in the cylinder to unevenly distribute the air-fuel mixture around the ignition plug to perform stratified combustion. It has been known. By performing such stratified combustion, the cylinder as a whole can be made leaner than the stoichiometric air-fuel ratio, so that fuel efficiency can be improved. Further, Patent Document 1 describes a control for significantly retarding the ignition timing (for example, after compression top dead center) when performing the above-mentioned stratified combustion for the purpose of promoting warm-up of a catalyst for purifying exhaust gas. Has been done.

Japanese Patent No. 3963088

However, in the configuration in which the air-fuel mixture is carried around the spark plug by utilizing the gas flow in the cylinder, the air-fuel mixture carried around the spark plug is easily flowed by the gas flow in the cylinder. In particular, when retarding the ignition timing as in the control described in the above document, it is difficult to hold the air-fuel mixture around the ignition plug until the ignition timing. That is, it is difficult to perform stratified charge combustion by the control described in the above document, and combustion stability is deteriorated.

Therefore, an object of the present invention is to enable stable combustion even when the ignition timing is significantly retarded.

According to one aspect of the present invention, the first fuel injection that forms a gas flow in the cylinder and forms a homogeneous air-fuel mixture leaner than stoichiometry in the entire cylinder in the first half of the intake stroke is performed, and the first fuel injection is performed around the spark plug . A direct injection internal combustion engine for a direct injection type internal combustion engine, in which a second fuel injection that forms a mixture richer than a stoichiometry is performed after a compression stroke and spark ignition is performed on the mixture formed by the fuel injected in the second fuel injection. A control method is provided. This control method is a fuel injection that is executed before the second fuel injection during the compression stroke, and increases the momentum of the injected fuel spray as the intensity of the gas flow increases, thereby burning the gas flow. The combustion pre-injection to weaken the fuel to a stable strength and the fuel pressure so low that the momentum of the fuel spray corresponding to the strength of the gas flow cannot be secured after the engine is started. to execute if unable Rukoto weaken the gas flow until it can be ensured, including the compression ratio control for increasing the effective compression ratio in the cylinder direct injection internal combustion engine.

According to the above aspect, it becomes possible to hold the air-fuel mixture around the spark plug until the ignition timing, and as a result, stable combustion is possible even when the ignition timing is significantly retarded.

FIG. 1 is a schematic configuration diagram in the vicinity of a combustion chamber of a direct injection type internal combustion engine. FIG. 2 is a diagram showing the relationship between the fuel injection timing and the tumble flow intensity. FIG. 3 is a diagram showing an example of the relationship between fuel spray and tumble flow. FIG. 4 is a diagram showing another example of the relationship between the fuel spray and the tumble flow. FIG. 5 is a timing chart of the tumble flow strength. FIG. 6 is a timing chart showing an example of the implementation timing of the first compression injection. FIG. 7 is a map showing the momentum of the fuel spray. FIG. 8 is a diagram showing the behavior of the hydraulic pressure when the engine is started. FIG. 9 is a diagram showing the relationship between fuel pressure and combustion stability. FIG. 10 is a diagram showing the difference in tumble flow strength depending on the compression ratio. FIG. 11 is a diagram showing the relationship between the compression ratio and the combustion stability. FIG. 12 is a diagram showing the relationship between the fuel pressure and the combustion stability for each compression ratio. FIG. 13 is a flowchart showing a control routine according to this embodiment. FIG. 14 is a timing chart when the control of the first embodiment is executed. FIG. 15 is a diagram showing the relationship between the compression ratio and the HC emission amount. FIG. 16 is a diagram showing the relationship between the oil water temperature and the combustion stability. FIG. 17 is a timing chart showing a modified example of the control routine of the first embodiment. FIG. 18 is a diagram for explaining the control routine according to the second embodiment. FIG. 19 is a diagram showing the relationship between the closing timing of the intake valve and the effective compression ratio. FIG. 20 is a diagram showing the relationship between the valve overlap period and the combustion stability. FIG. 21 is a diagram for explaining a modified example of the control routine according to the second embodiment.

Hereinafter, embodiments of the present invention will be described with reference to the accompanying drawings.

(First embodiment)
FIG. 1 is a schematic configuration diagram around a combustion chamber of a cylinder direct injection internal combustion engine (hereinafter, also referred to as “engine”) 1 to which the present embodiment is applied. Although FIG. 1 shows only one cylinder, this embodiment can be applied to a multi-cylinder engine.

The cylinder block 1B of the engine 1 includes a cylinder 2. A piston 3 is housed in the cylinder 2. The piston 3 is driven by a variable compression ratio mechanism VCR to reciprocate in the cylinder 2.

The variable compression ratio mechanism VCR is a mechanism known by the applicant. The lower link 13 is rotatably fixed to the crank pin 14A of the crank shaft 14. The lower link 13 and the piston 3 are connected via the upper link 12. Further, the lower link 13 and the control shaft 16 are connected via the control link 15. The control link 15 is connected to a position deviated from the rotation axis of the control shaft 16. The control shaft 16 is rotationally driven by an electric motor 17 via, for example, a rack and pinion mechanism.

With the above configuration, when the control shaft 16 rotates and the control link 15 is pulled down, the lower link 13 rotates about the crank pin 14A as an axis, and the upper link 12 is pushed up. As a result, the top dead center position of the piston 3 rises. On the contrary, when the control link 15 is pushed up, the upper link 12 is pulled down and the top dead center position of the piston 3 is lowered. As described above, the variable compression ratio mechanism VCR can change the so-called mechanical compression ratio by changing the position of the top dead center of the piston 3.

Further, the piston 3 has a cavity 10 described later on a crown surface 3A (hereinafter, also referred to as a piston crown surface 3A).

A cylinder head 1A of the engine 1 includes a concave combustion chamber 11, and an intake passage 4 and an exhaust passage 5 that connect the combustion chamber 11 and the outside of the engine. The combustion chamber 11 is of a so-called pent roof type, in which a pair of intake valves 6 are arranged in the opening of the intake passage 4 and a pair of exhaust valves 7 are arranged in the opening of the exhaust passage 5. A spark plug 8 is arranged along the axis of the cylinder 2 at a substantially central position of the combustion chamber 11 surrounded by the pair of intake valves 6 and the pair of exhaust valves 7.

The intake passage 4 is shaped so that the intake air flowing into the combustion chamber 11 forms a tumble flow. That is, the intake passage 4 functions as a gas flow generation device.

A fuel injection valve 9 is arranged so as to face the combustion chamber 11 at a position between the pair of intake valves 6 of the cylinder head 1A. The directivity of the fuel spray injected from the fuel injection valve 9 will be described later.

The fuel injection device of this embodiment is a so-called common rail type. The high-pressure pump 41 boosts the fuel sucked up by the low-pressure pump 42 from the fuel tank 43 and supplies it to the common rail 40. As a result, the common rail 40 is brought into a predetermined high pressure state, so that the fuel injection valve 9 can inject fuel at a high pressure. The low-pressure pump 42 is driven by, for example, an electric motor. The high-pressure pump 41 is driven by a camshaft shaft on the intake side or the exhaust side of the engine 1.

The intake valve 6 and the exhaust valve 7 are driven by a variable valve mechanism 20 as a valve overlap period adjusting mechanism. The variable valve mechanism 20 controls the valve timings of the intake valve 6 and the exhaust valve 7, that is, the valve opening timing and the valve closing timing so that the valve overlap period in which both the intake valve 6 and the exhaust valve 7 are opened occurs. Anything that can be changed is sufficient. The valve opening timing is the timing at which the valve opening operation is started, and the valve closing timing is the timing at which the valve closing operation is ended. In the present embodiment, a known variable valve mechanism 20 that changes the rotational phase of the camshaft that drives the intake valve 6 and the camshaft that drives the exhaust valve 7 with respect to the crankshaft 14 is used. A known variable valve mechanism that can change not only the rotation phase but also the operating angles of the intake valve 6 and the exhaust valve 7 may be used. Further, the variable valve mechanism 20 is not limited to one that can adjust both the opening and closing timings of the intake valve 6 and the exhaust valve 7, and may be one that can adjust only one of them. For example, even if only the opening/closing timing of the intake valve 6 can be adjusted, another mechanism is adopted if the valve overlap period between the opening period of the intake valve 6 and the opening period of the exhaust valve 7 can be lengthened or shortened. You may.

An exhaust gas purification catalyst for purifying the exhaust gas of the engine 1 is provided downstream of the exhaust passage 5 in the exhaust flow. The exhaust purification catalyst is, for example, a three-way catalyst.

The piston 3 includes the cavity 10 on the piston crown surface 3A as described above. The cavity 10 is provided on the piston crown surface 3A at a position biased toward the intake side. The fuel injection valve 9 is arranged so that, when the fuel is injected when the piston 3 is near the compression top dead center, the fuel spray is directed to the cavity 10. The cavity 10 is shaped so that the fuel spray (B in the drawing) that collides and bounces back toward the spark plug 8.

The cavity 10 may be located at the center of the piston crown surface 3A or at any other position as long as the condition that the fuel spray collides is satisfied.

The fuel injection amount, fuel injection timing, ignition timing, etc. of the engine 1 are controlled by the controller 100 according to the operating state of the engine 1. The controller 100 is composed of a microcomputer provided with a central processing unit (CPU), a read only memory (ROM), a random access memory (RAM) and an input/output interface (I/O interface). It is also possible to configure the controller 100 with a plurality of microcomputers.

The fuel injection timing referred to here is the timing at which fuel injection is started. Further, in order to execute these controls, the engine 1 includes a crankshaft angle sensor, a cooling water temperature sensor 32, an air flow meter for detecting an intake air amount, an accelerator opening sensor 31 for detecting an accelerator pedal depression amount, an exhaust gas purification catalyst. Various detection devices such as a catalyst temperature sensor 33 for directly or indirectly detecting the temperature of the are provided. The accelerator opening sensor 31 functions as an acceleration request sensor that detects an acceleration request from the driver, but the acceleration request sensor is not limited to this. For example, a manual accelerator operation can be applied, and the form of the operator is not limited as long as the requested acceleration amount can be detected.

In the present embodiment, the configuration in which the stratified mixture is formed around the spark plug 8 using the cavity 10 will be described, but the present invention is not limited to this. For example, the fuel injection valve 9 may be arranged adjacent to the ignition plug 8 and the fuel injected from the fuel injection valve 9 with a short drive pulse may stay around the ignition plug 8 to form a stratified mixture. Good.

Next, control of the engine 1 having the above configuration will be described.

First, a conventionally known control at the time of cold start will be described.

As a control during cold start, super retarded stratified combustion for promoting activation of an exhaust purification catalyst is known (for example, Japanese Patent Laid-Open No. 2008-25535). At this time, during cold starting, combustion is performed under severe conditions from the viewpoint of low temperature combustion stability, so control that can ensure combustion stability is desired.

In super retarded stratified charge combustion, the controller 100 sets the ignition timing to the first half of the expansion stroke, for example, 10 to 30 deg after the compression top dead center. Further, the controller 100 executes so-called multi-stage injection, in which the required fuel amount per one cycle is injected plural times. In the case of the two-stage injection, the controller 100 sets the first fuel injection timing in the first half of the intake stroke, and sets the second fuel injection timing in the latter half of the compression stroke by the fuel spray by the ignition timing. Set the timing so that you can reach the area around.

Here, the first fuel injection amount and the second fuel injection amount (stratified injection amount) in the case of the two-stage injection will be described.

The air-fuel ratio of the exhaust gas discharged by the above-mentioned super retarded stratified charge combustion is stoichiometric (theoretical air-fuel ratio). The controller calculates the amount of fuel that can be completely combusted with the amount of intake air per cycle (hereinafter, also referred to as total fuel amount), as in a general fuel injection amount setting method. A part of the total fuel amount, for example, 20 to 90% by weight is set as the first injection amount, and the rest is set as the second injection amount.

In the super retarded stratified charge combustion, the air-fuel ratio of the exhaust gas may be leaner than stoichiometric.

When the fuel injection amount is set as described above, the fuel spray injected in the first fuel injection diffuses into the cylinder 2 without colliding with the cavity 10, mixes with air, and is stoichiometric throughout the combustion chamber 11. A leaner homogeneous mixture (A in the figure) is formed. Then, the fuel spray (B in the figure) injected in the second fuel injection (stratified injection) collides with the cavity 10 and reaches the vicinity of the spark plug 8 by being wound up, and around the spark plug 8. Concentrate to form a richer mixture than stoichiki. As a result, the air-fuel mixture in the combustion chamber 11 becomes stratified. If spark ignition is performed by the spark plug 8 in this state, combustion that is resistant to disturbance and suppressed in misfire is performed. By the way, although the above-mentioned combustion is stratified combustion, it is referred to as super-retarded stratified combustion in order to distinguish it from general stratified combustion in which the ignition timing is before compression top dead.

It should be noted that the above-described first fuel injection is divided into two, and the amount of fuel required per cycle is divided into three injections, two in the intake stroke and one in the compression stroke, as a three-stage injection. Good. In this case, the third injection is stratified injection.

Here, the fuel injection timing in super retarded stratified charge combustion will be described with reference to FIG. FIG. 2 is a timing chart in which the horizontal axis is the crank angle. In the figure, IT1 shows the fuel injection timing of the first time, IT2 shows the fuel injection timing of stratified injection, the solid line A shows the strength of the tumble flow, and the solid line B shows the fuel adhesion characteristic to the piston 3. In the figure, TDC means top dead center and BDC means bottom dead center.

It is known that tumble flow is effective in improving fuel efficiency and reducing exhaust emissions. Since the tumble flow is formed by the intake air that flows into the combustion chamber 11 after the intake valve 6 is opened, the strength of the tumble flow gradually increases during the intake stroke. However, when the volume of the combustion chamber 11 increases as the piston 3 descends, the flow velocity of the tumble flow decreases. Therefore, the strength of the tumble flow reaches the first peak (1st peak) during the intake stroke and starts to decrease.

In the compression stroke, the volume of the combustion chamber 11 decreases as the piston 3 rises, causing the flow velocity of the tumble flow to increase, and in response to this, the strength of the tumble flow also increases and the second peak ( 2nd peak). After that, when the piston 3 further rises, the tumble flow is crushed, so that the strength of the tumble flow gradually decreases, and eventually the tumble flow disappears.

The fuel adhesion property to the piston 3 indicates how much fuel adheres to the crown surface 3A of the piston 3 when the injected fuel collides with the piston 3. As shown in the figure, the closer the piston 3 is to the top dead center, the larger the fuel adhesion amount. This is because the distance between the piston 3 and the fuel injection valve 9 decreases as the piston 3 approaches the top dead center, and more fuel collides with the piston 3.

If the fuel attached to the piston 3 is carried over to the next cycle without vaporizing or burning during the cycle, the liquefied fuel is accumulated in the piston 3. In this state, if the combustion mode is switched to the normal combustion mode in response to the completion of warming up of the exhaust purification catalyst or an acceleration request, the accumulated fuel is burned by the combustion flame propagating to the piston 3, and the PM emission amount increases. End up. The normal combustion mode referred to here is a combustion mode in which fuel injection is performed in the intake stroke or compression stroke, and spark ignition is performed at MBT (optimal ignition timing) or at an ignition timing close to this.

From the viewpoint of suppressing the PM emission amount, the crank angle range in which the fuel injection timing can be set is limited based on the amount of fuel adhering to the piston 3. The NG range in FIG. 2 is an example of a crank angle range in which setting of fuel injection timing is prohibited. It should be noted that the allowable amount of adhering fuel is larger on the intake top dead center side than on the compression top dead center side. This is because the time to the ignition timing is longer in the case of adhering to the piston 3 at the intake top dead center side than to the case of adhering to the compression top dead center side, and the amount of evaporation/vaporization from the adhering to the ignition Because there are many.

As described above, the fuel injection timing of the stratified charge injection is within a range in which the fuel spray collides with the cavity 10 and the collided fuel spray forms a stratified mixture around the spark plug 8 by the ignition timing. It is set in consideration of the amount of fuel adhered to the piston 3.

By the way, as shown in FIG. 2, when the tumble flow is still strong at the fuel injection timing of the stratified charge injection, the fuel spray injected by the stratified charge injection is made to flow by the tumble flow, and is stratified around the ignition plug 8. It becomes difficult to form a mixture. Therefore, it is difficult to form the stratified mixture by forming the intake passage 4 so as to enhance the tumble flow in order to improve the fuel efficiency and reduce the exhaust gas, and the combustion stability in the super-retarded stratified charge combustion decreases. End up.

Since the strength of the tumble flow decreases as it approaches the compression top dead center as described above, it becomes easier to form the stratified mixture as the fuel injection timing of the stratified injection is delayed. However, as the fuel injection timing is delayed, the amount of fuel adhering to the piston 3 increases. Therefore, in the generally known general fuel injection control, the fuel injection timing of the stratified injection takes into consideration the balance between the ease of forming the stratified mixture and the amount of fuel adhered to the piston 3. Was there.

On the other hand, in the present embodiment, by executing the fuel injection control described below, the combustion stability of the ultra-retarded stratified charge combustion is ensured even when the tumble flow is strengthened in order to improve the fuel efficiency performance and reduce the exhaust gas. To do. Further, according to the fuel injection control of the present embodiment, the amount of fuel adhering to the piston 3 does not increase in order to secure combustion stability.

3 and 4 are diagrams for explaining the outline of the fuel injection control according to the present embodiment.

When fuel is injected during the compression stroke, as shown in FIG. 3, the vortex center of the main flow of the tumble flow (hereinafter, also simply referred to as “vortex center of tumble flow” or “vortex center”) is positioned lower than the fuel spray. In some cases, as shown in FIG. 4, the vortex center of the tumble flow may be located higher than the fuel spray. Note that FIG. 4 shows a state in which the piston 3 is lifted from the state of FIG. Since the vortex of the tumble flow is crushed as the piston 3 rises, the vortex center of the tumble flow is higher in FIG. 4 than in FIG.

In the case of FIG. 3, at the collision position between the fuel spray and the tumble flow vortex, the tumble flow strength is increased by the combination of the vector in the traveling direction of the fuel spray and the vector in the rotation direction of the vortex of the tumble flow.

On the other hand, in the case of FIG. 4, at the collision position of the fuel spray and the tumble flow vortex, the vector in the traveling direction of the fuel spray opposes the vector in the rotation direction of the tumbling vortex, so the tumble flow strength decreases. To do. The term "opposing" means that both vectors form an angle that hinders the rotation of the vortex of the tumble flow, and is not limited to the case where the angle formed by both vectors is 180°. ..

FIG. 5 is an enlarged view from the vicinity of the 2nd peak in FIG. 2 until the tumble flow disappears. The broken line in the figure shows the tumble flow intensity of FIG. 2, and the solid line shows the tumble flow intensity when fuel is injected at the fuel injection timing of FIG. As shown in the figure, when the fuel is injected at the fuel injection timing shown in FIG. 4, the height of the 2nd peak is suppressed and the timing at which the tumble flow intensity decreases is advanced.

Therefore, in the present embodiment, before the stratified injection, fuel injection is performed to weaken the strength of the tumble flow to the extent that a stratified mixture can be formed around the spark plug 8. That is, in the state as shown in FIG. 4, the fuel injection for changing the tumble flow in the direction in which the combustion by the spark ignition is stabilized (hereinafter, also referred to as “compression early injection”) is performed, and then the stratified injection is performed. ..

The total amount of fuel is not changed even when the compression early injection is performed. For example, the injection amount of the first fuel injection performed during the intake stroke is reduced by the amount of injection in the compression previous injection. Since the fuel spray injected in the pre-compression injection collides with the tumble flow to promote diffusion and mixing, the amount of fuel adhering to the piston 3 does not increase by performing the pre-compression injection. It should be noted that the injection amount of the stratified injection may be reduced and the injection may be performed in the first compression period injection. Further, if the emission performance falls within the allowable range, the total fuel amount may be increased by the amount of the first compression injection without changing the injection amounts of the first fuel injection and the stratified injection.

Here, the fuel injection timing of the compression early injection will be described.

FIG. 6 is a timing chart showing an example of the timing of the compression early injection. FIG. 6 shows two examples (a) and (b).

In the tumble flow intensity chart, the solid line shows the case where the compression early injection is not executed, the broken line shows the case where the fuel injection timing of the compression early injection is (a), and the chain line shows the case where the fuel injection timing of the compression early injection is (b). Showing. Further, in this chart, the injection timing of the stratified injection is the timing on the retard side limit. The retardation side limit of the stratified charge injection is determined based on the ignition timing in consideration of the period required for the fuel injected in the stratified charge injection to form the stratified charge mixture.

Timing t1 is a timing before the 2nd peak of the tumble flow intensity. If the compression early injection is executed at the timing t1, the 2nd peak of the tumble flow decreases as shown in the figure, and the strength of the tumble flow at the start of the stratified injection decreases, so the combustion stability of the stratified combustion improves.

Timing t2 is a timing after the 2nd peak of the tumble flow intensity. Even when the compression pre-injection is executed at the timing t2 when the second peak has passed, the rate of decrease in the strength of the tumble flow is accelerated as shown in the figure, so that the combustion stability of the stratified charge combustion is improved.

Therefore, the combustion stability of the stratified charge combustion can be improved regardless of whether the fuel injection timing of the early compression period is (a) or (b).

It should be noted that the pre-compression injection advances at a position where the vector in the traveling direction of the fuel spray and the vector in the rotating direction of the tumble flow vortex are opposed at the collision position between the fuel spray and the tumble flow vortex. be able to. Further, the pre-compression pre-injection can be retarded within a range in which the pre-compression pre-injection can be completed by the fuel injection timing of the stratified injection.

Next, the fuel injection amount and the fuel injection pressure (hereinafter, also referred to as “fuel pressure”) of the pre-compression injection will be described.

The combustion stability of the stratified charge combustion improves as the tumble flow strength at the time of executing the stratified charge injection decreases. In view of this, in the present embodiment, the controller 100 controls the pre-compression injection so that the higher the tumble flow intensity, the greater the decrease in the tumble flow intensity.

The reason why the tumble flow intensity is lowered by the pre-compression injection is that the vortex of the tumble flow and the fuel spray injected by the pre-compression injection collide with each other, and the angular momentum of the vortex of the tumble flow is reduced. Therefore, as the tumble flow intensity is higher, the controller 100 increases the momentum of the fuel spray injected in the pre-compression injection.

FIG. 7 is a map showing the relationship between the momentum of the fuel spray, the fuel injection amount and the fuel pressure. As shown, as the fuel pressure increases and the fuel injection amount increases, the momentum of the fuel spray increases. That is, as a method of increasing the momentum of the fuel spray, for example, the fuel injection amount may be increased or the fuel pressure may be increased. Of course, the fuel injection amount may be increased and the fuel pressure may be increased.

The tumble flow strength increases as the intake amount increases, that is, as the throttle valve opening increases. Therefore, as the throttle valve opening increases, the controller 100 increases at least one of the fuel pressure and the fuel injection amount.

Further, when the throttle valve opening is constant, the higher the engine speed, the higher the tumble flow intensity. Therefore, the controller 100 may increase at least one of the fuel pressure and the fuel injection amount as the engine speed increases.

As described above, in the present embodiment, super retarded stratified charge combustion is executed in order to early activate the exhaust purification catalyst when the engine is started. Then, in order to secure the combustion stability of the super-retarded stratified combustion, the tumble flow strength is reduced by performing the pre-compression injection.

By the way, since the high-pressure pump 41 is driven by the engine 1, it takes time from the start of the engine until the fuel pressure rises.

FIG. 8 is a timing chart showing the behavior of fuel pressure with the vertical axis representing fuel pressure and the horizontal axis representing time. The broken line in FIG. 8 indicates the engine rotation speed.

When the cranking is started at the timing T0, the high pressure pump 41 also starts operating. Timing T1 is the timing when the high pressure pump 41 performs the pressurization once, and timing T2 is the timing when the high pressure pump 41 performs the pressurization twice. The fuel pressure gradually increases at timing T1 and timing T2. It is assumed that the fuel pressure has reached a level at which fuel injection for engine start is possible at timing T2.

When fuel injection is started after timing T2, the fuel pressure rises and falls. This is because the fuel pressure decreases while one of the fuel injection valves 9 of the multi-cylinder is open, and the fuel pressure rises by the high-pressure pump 41 while all the fuel injection valves 9 are closed. Is.

When the engine 1 starts operating immediately before the timing T3 and the engine rotation speed increases, the fuel pressure also increases accordingly.

In this way, the fuel pressure starts increasing after the engine 1 is started, so the fuel pressure is low immediately after the engine is started. Further, since the momentum of the fuel spray is small when the fuel pressure is low, the effect of lowering the tumble flow strength is small even if the pre-compression injection is performed.

Further, as the fuel pressure becomes lower, the fuel injected to form the stratified mixture becomes more difficult to move around the spark plug 8 after colliding with the cavity 10 of the piston 3. As a result, if the super-retarded stratified charge combustion is performed in a state where the fuel pressure is low, there is a possibility that the combustion stability cannot be secured. On the other hand, if the start of super retarded stratified charge combustion is waited for until the fuel pressure rises sufficiently, activation of the exhaust purification catalyst is delayed and exhaust performance deteriorates. FIG. 9 summarizes the relationship between the combustion stability and the fuel pressure described above.

FIG. 9 is a diagram showing the relationship between the combustion stability and the fuel pressure when performing the above-described super-retarded stratified charge combustion with a constant compression ratio. As described above, the lower the fuel pressure, the lower the combustion stability. The vertical axis of FIG. 9 is the combustion stability, and the horizontal axis is the fuel pressure. Here, the fuel pressure when the combustion stability is at the allowable limit is P1. That is, after the timing T3 in FIG. 8, if the fuel pressure reaches P1, super retarded stratified charge combustion can be stably performed.

Next, the relationship between the combustion stability and the compression ratio will be described.

FIG. 10 is a diagram showing the relationship between the strength of the tumble flow and the compression ratio when the fuel pressure is constant. The vertical axis of FIG. 10 is the tumble ratio, and the horizontal axis is the crank angle. The “tumble ratio” here is an index showing the strength of the tumble flow. The broken line in FIG. 10 shows the case where the compression ratio (ε) is 10.5, and the solid line shows the case where the compression ratio (ε) is 8.

The lower the compression ratio, the more easily tumble flow is generated, and the more easily the strength is maintained. Therefore, as shown in FIG. 10, both the 1st peak and the 2nd peak have a higher compression ratio. Also, the lower the compression ratio, the more the strength of the tumble flow is maintained up to the retard side. That is, in order to secure the combustion stability, it is necessary to increase the momentum of the fuel spray injected in the early compression injection as the compression ratio is lower. In other words, the higher the compression ratio, the higher the combustion stability.

FIG. 11 is a diagram showing the relationship between the combustion stability and the compression ratio when the above-described super-retarded stratified charge combustion is performed with the fuel pressure kept constant. The vertical axis of FIG. 11 is the combustion stability, and the horizontal axis is the compression ratio. As shown in FIG. 11, the higher the compression ratio, the higher the combustion stability. Here, the compression ratio when the combustion stability is at the allowable limit is ε1. That is, if the compression ratio is higher than ε1, combustion stability can be secured.

FIG. 12 is a diagram summarizing the above-mentioned relationship between the combustion stability, the fuel pressure, and the compression ratio, with the vertical axis representing combustion stability and the horizontal axis representing effective fuel pressure. In FIG. 12, ε1<ε2<ε3.

As shown in FIG. 12, the higher the compression ratio, the lower the fuel pressure when the combustion stability reaches the allowable limit. That is, even if the fuel pressure is low, it becomes easier to secure the combustion stability by increasing the compression ratio. Therefore, in the present embodiment, when the fuel pressure is low, such as immediately after the engine is started, the variable retardation ratio mechanism is used to increase the compression ratio, thereby rapidly starting the super retarded stratified charge combustion.

FIG. 13 is a flowchart showing a control routine for promptly starting the super retarded stratified charge combustion after the engine is started. This control routine is repeatedly executed, for example, in a cycle of several milliseconds after the engine is started. Hereinafter, description will be given according to the steps of the flowchart.

In step S100, the controller 100 determines whether or not execution of super retarded stratified charge combustion (FIR) is permitted. The conditions for permitting the super retarded stratified charge combustion are as follows, for example.

First, it is necessary that the engine speed be constant. That is, it is necessary that the engine rotation speed, which has risen from the cranking after the initial explosion, has converged to the idle rotation speed. Further, it is necessary that the oil water temperature is equal to or higher than a predetermined temperature. The predetermined temperature mentioned here is a temperature that becomes a threshold value of whether it is a so-called extremely low temperature or not. In an extremely low temperature environment, the friction of the engine 1 is large and the combustion stability is lowered, so that the ultra-retarded stratified combustion is not performed. In addition to this, it is necessary that the atmospheric pressure is equal to or lower than a predetermined pressure. As the atmospheric pressure becomes lower, the amount of oxygen taken into the engine 1 decreases, so the combustion stability decreases. Therefore, super retarded stratification combustion is not performed in places where atmospheric pressure is low, such as in highlands. In addition, in the determination of step S100, the above-described lack of fuel pressure is not determined.

The controller 100 executes the process of step S110 if the execution of the super retarded stratified charge combustion is permitted, and ends the routine if not.

In step S120, the controller 100 determines whether the differential pressure ΔP is greater than zero. The differential pressure ΔP is the difference between the actual fuel pressure and the target fuel pressure that is set as the fuel pressure that can ensure the combustion stability of the super retarded stratified combustion without increasing the compression ratio. It should be noted that this step only needs to determine whether or not the actual fuel pressure has a magnitude that can secure the combustion stability of the super-retarded stratified charge combustion with the compression ratio at the time of engine start. For example, if the target fuel pressure is set to a fuel pressure that is higher than the lower limit fuel pressure that can ensure the combustion stability of super retarded stratified combustion with the compression ratio at engine startup, even if the actual fuel pressure does not reach the target fuel pressure, Retardation stratified combustion may be permitted. The pressure difference between the target fuel pressure and the above lower limit fuel pressure may be set to a predetermined value α, and it may be determined whether or not the pressure difference ΔP is larger than the predetermined value α.

In step S120, the controller 100 calculates the energy (insufficient energy β) that is insufficient to reduce the tumble flow intensity. Here, the insufficient energy β will be described.

Since the tumble flow is the rotational movement of the gas in the cylinder, the kinetic energy Eair of the tumble flow is expressed by the equation (1) using the inertia moment I of the intake air and each acceleration ω.

Eair=1/2 Iω ² (1)
On the other hand, since the fuel spray makes a linear motion, the kinetic energy Espray of the fuel spray injected in the previous compression period is expressed by the equation (2) using the weight m and the speed v of the fuel spray.

Espray=1/2 mv ² (2)
If the kinetic energy of the fuel spray at the time of the pre-compression injection at the target fuel pressure is Espray_t, the kinetic energy Etum1 of the tumble flow whose strength has decreased to the extent that combustion stability can be secured by the pre-compression injection is given by the formula (3). To be done.

Etum1=Eair-Espray_t (3)
On the other hand, when the kinetic energy of the fuel spray when the pre-compression injection is performed at the actual fuel pressure is Espray_r, the kinetic energy Etum2 of the tumble flow whose strength is reduced by the pre-compression injection is expressed by the equation (4).

Etum2=Eair-Espray_r (4)
That is, the energy deficit β is the difference between Etum1 and Etum2.

After calculating the insufficient energy β as described above, the controller 100 calculates the compression ratio according to the insufficient energy β in step S130. Specifically, a table in which the relationship between the insufficient energy β and the compression ratio ε as shown in FIG. 13 is set is created in advance, and the compression ratio ε is obtained by searching this table. The compression ratio ε increases as the insufficient energy β increases.

In step S140, the controller 100 sets the compression ratio ε calculated in step S130 as the target compression ratio.

The variable compression ratio mechanism VCR is basically controlled based on the basic target compression ratio set by another control routine executed in parallel with this control routine. If the determination result in step S100 or S110 is no, the controller 100 controls the variable compression ratio mechanism VCR based on the basic target compression ratio. However, if the determination result in step S110 is yes, the controller 100 controls the variable compression ratio mechanism VLE based on the target compression ratio set in step S140 instead of the basic target compression ratio.

FIG. 14 is a timing chart when the above control is executed. The chart of the fuel pressure and the engine rotation speed is simplified as compared with FIG.

When cranking is started at timing T1 and spark ignition is performed at timing T2, the engine rotation speed increases, and the fuel pressure also increases accordingly. At timing T3, the engine rotation speed converges to the idle rotation speed, and super retarded stratified combustion is permitted. However, at timing T3, the fuel pressure has not reached the target fuel pressure, so the compression ratio is controlled to the target compression ratio set in step S140 of FIG.

When the fuel pressure reaches the target fuel pressure at timing T4, the determination result in step S110 of FIG. 13 becomes no, so the compression ratio is controlled to the basic target compression ratio.

The table showing the relationship between the insufficient energy β and the compression ratio ε shown in step S130 of FIG. 13 is set based on the combustion stability. However, as shown in FIG. 15, the HC emission amount [ppm] increases as the compression ratio increases. This is because the gas volume in the cylinder becomes smaller as the compression ratio becomes higher, so that the HC emission amount [ppm] evaluated by the HC amount per gas volume increases even if the HC production amount is the same. Therefore, the target compression ratio may be corrected so that the HC emission amount becomes smaller within a range where the combustion stability can be secured.

Further, in the control routine of FIG. 13, the target compression ratio is maintained at a constant magnitude until the determination result of step S110 becomes no, that is, until the fuel pressure reaches the target fuel pressure. However, the target compression ratio set in step S140 may be corrected each time the calculation is performed, depending on the oil/water temperature. FIG. 16 is a diagram showing the relationship between the oil water temperature and the combustion stability. The oil water temperature mentioned here is the temperature of the lubricating oil or cooling water of the engine 1. As shown in FIG. 16, the combustion stability increases as the oil water temperature increases. This is because the higher the oil/water temperature, the lower the friction of the engine 1. When the engine 1 is cold-started, the oil-water temperature rises according to the elapsed time after the start, so that the combustion stability gradually improves according to the elapsed time after the engine starts. Therefore, it is possible to reduce the compression ratio by correcting the target compression ratio set in step S140 according to the oil/water temperature at each calculation. As a result, the amount of HC discharged can be reduced.

FIG. 17 is an example of a timing chart when the target compression ratio is corrected based on the oil/water temperature. Here, the target compression ratio 1 set at the timing T3 is corrected to the target compression ratio 2 lower than the target compression ratio 1 at the timing T3.5. It is not limited to the stepwise change as shown in FIG. 17, but may be changed so as to continuously decrease from the timing T3 to the timing T4.

As described above, the control of the present embodiment performs the fuel injection that forms the gas flow in the cylinder and injects the fuel to be burned around the spark plug after the compression stroke, and mixes the fuel injected by the fuel injection. It ignites a spark. In the control, the pre-compression injection for changing the gas flow toward the stable combustion, which is executed before the fuel injection during the compression stroke, is performed. Then, when it is not possible to change the gas flow to the point where combustion is stabilized by only the first compression injection, the compression ratio control for changing the effective compression ratio is executed. Specifically, as the compression ratio control, the effective compression ratio is increased according to the difference between the predetermined fuel pressure and the actual fuel pressure required to change the gas flow to the point where combustion is stabilized. As a result, even when the fuel pressure is low, such as immediately after the engine is started, it is possible to suppress a decrease in combustion stability.

In this embodiment, the engine 1 includes a variable compression ratio mechanism VCR that changes the mechanical compression ratio. Then, as the compression ratio control, the controller 100 raises the mechanical compression ratio by the variable compression ratio mechanism VCR. If the mechanical compression ratio rises, the effective compression ratio naturally rises, so that the decrease in combustion stability can be suppressed.

In the present embodiment, the increased effective compression ratio may be decreased when the actual fuel pressure increases. Since the exhaust temperature rises as the compression ratio decreases, the effect of promoting warm-up of the exhaust purification catalyst by super retarded stratified combustion is enhanced. It is also possible to reduce the amount of HC and PN emitted.

In the present embodiment, when the effective compression ratio that has been once increased is decreased, it may be decreased stepwise according to the increase in the actual fuel pressure. According to this, the effective compression ratio can be lowered while ensuring the combustion stability before the fuel pressure reaches the target fuel pressure, so that the effect of promoting warm-up of the exhaust purification catalyst and reducing the emission amount of HC and the like can be obtained. Can be increased.

(Second embodiment)
This embodiment is the same as the first embodiment in the technical idea of ensuring combustion stability by increasing the compression ratio in the state where the fuel pressure is low, but the method of changing the compression ratio is the same as that of the first embodiment. Is different. That is, the method of realizing the target compression ratio set in step S140 of FIG. 13 differs from that of the first embodiment. Hereinafter, differences from the first embodiment will be mainly described.

In the first embodiment, the mechanical compression ratio is changed using the variable compression ratio mechanism VCR, but in the present embodiment, the effective compression ratio is changed by changing the valve timing using the variable valve mechanism 20.

FIG. 18 is a valve timing chart. In the figure, IVO is the opening timing of the intake valve 6, IVC is the closing timing of the intake valve 6, EVO is the opening timing of the exhaust valve 7, and EVC is the closing timing of the exhaust valve 7. The left diagram of FIG. 18 shows the basic valve timing that is set based on the cooling water temperature and the like as in general engine control. The basic valve timing is set by another control routine executed in parallel with the control routine of FIG.

When the closing timing of the intake valve 6 is shifted from the basic valve timing to the bottom dead center side as shown in the right figure of FIG. 18, that is, when the intake valve 6 is advanced, the so-called effective compression ratio increases. As a result, the same effect as that of the first embodiment in which the mechanical compression ratio is increased can be obtained.

FIG. 19 is a diagram showing the relationship between the effective compression ratio and the closing timing of the intake valve 6. As shown in FIG. 19, the effective compression ratio increases as the closing timing of the intake valve 6 approaches the bottom dead center. Therefore, in the present embodiment, the controller 100 controls the closing timing of the intake valve 6 by using the variable valve mechanism 20 on the intake side so that the target compression ratio set in step S140 of FIG. 13 is achieved.

By the way, in the case where the variable valve mechanism 20 is configured to change the rotation phase of the camshaft with respect to the crankshaft 14, when the closing timing of the intake valve 6 is advanced, the opening timing of the intake valve 6 is increased by the advanced amount. Also advances. As a result, the valve overlap period increases. As the valve overlap period increases, the amount of gas (hereinafter, also referred to as “residual gas”) remaining in the cylinder after combustion increases. The combustion stability decreases as the residual gas increases. Therefore, as shown in FIG. 20, the combustion stability decreases as the valve overlap period increases.

That is, by advancing the closing timing of the intake valve 6 to raise the effective compression ratio, the effect of lowering the tumble flow intensity and increasing the combustion stability can be obtained, but the effect of increasing the residual gas is weakened. Factors also occur.

Therefore, when advancing the closing timing of the intake valve 6, the opening timing of the exhaust valve 7 is advanced by using the variable valve mechanism 20 on the exhaust side so that the valve overlap period does not change. Good. According to this, it is possible to suppress an increase in the residual gas due to the advance of the closing timing of the intake valve 6.

As described above, in the present embodiment, the variable valve mechanism 20 advances the closing timing of the intake valve 6 to raise the effective compression ratio. As a result, the same effect as when changing the mechanical compression ratio using the variable compression ratio mechanism VCR is obtained.

In the present embodiment, the opening timing of the exhaust valve 7 is advanced and the opening timing of the exhaust valve 7 is advanced so that the valve overlap period does not change even if the closing timing of the intake valve 6 is advanced. May be. According to this, the increase of the residual gas due to the increase of the valve overlap period can be suppressed.

In each of the above-described embodiments, the case where the super retarded stratified charge combustion is performed has been described, but the present invention is not limited to this. Strictly speaking, even if it is not in the stratified state, if it is a combustion mode in which the air-fuel mixture is concentratedly formed around the ignition plug and ignited, the same effect as the above-described embodiment can be obtained.

Further, it is needless to say that the present invention is not limited to the above-described embodiment, and various modifications can be made within the scope of the technical idea described in the claims.

1 In-cylinder direct injection internal combustion engine (engine)
3 Piston 6 Intake valve 7 Exhaust valve 8 Spark plug 9 Fuel injection valve 20 Variable valve mechanism 40 Common rail 100 Controller

Claims (8)

Form a gas flow in the cylinder,
In the first half of the intake stroke, the first fuel injection that forms a leaner homogeneous mixture than stoichiometric is performed in the entire cylinder,
The second fuel injection that forms a mixture richer than stoichiometric around the spark plug is performed after the compression stroke,
Spark ignition to a mixture formed by the fuel injected in the second fuel injection,
In the control method of the direct injection type internal combustion engine,
A fuel injection even during the compression stroke is executed before the second fuel injection, the combustion of the gas flow by increasing the momentum of the fuel spray intensity of the gas flow is higher injection stability Compression pre-injection to weaken the strength to
For even after the engine start fuel pressure is low enough not be ensured momentum of the fuel spray corresponding to the intensity of the gas flow, the only compression early injection is Ru weaken the gas flow to the point where that can ensure combustion stability When it is not possible to perform, compression ratio control for increasing the effective compression ratio of the direct injection type internal combustion engine,
A method of controlling a direct injection type internal combustion engine, comprising: The control method for a direct injection internal combustion engine, according to claim 1,
The compression ratio control, the higher the combustion gas flow is large difference between the actual fuel pressure to a predetermined fuel pressure required in order weakened until it stabilizes, the is control that greatly increase the effective compression ratio cylinder direct injection Method for controlling internal combustion engine. The control method for a direct injection type internal combustion engine according to claim 2,
The in-cylinder direct injection internal combustion engine further comprises a variable compression ratio mechanism for changing a mechanical compression ratio,
The compression ratio control is a control method for a direct injection type internal combustion engine, which is a control for increasing a mechanical compression ratio by the variable compression ratio mechanism. The control method for a direct injection type internal combustion engine according to claim 2,
The in-cylinder direct injection internal combustion engine further includes a variable valve mechanism that changes valve timing by changing rotational phases of an intake camshaft and an exhaust camshaft with respect to a crankshaft ,
The compression ratio control, the variable valve mechanism by direct in-cylinder is controlled to the closing timing after bottom dead center passes to advance to approach the bottom dead center raises the effective compression ratio of the intake valve injection Internal combustion engine control method. The control method for a direct injection type internal combustion engine, according to claim 4,
The compression ratio control is a control method for a direct injection type internal combustion engine, wherein the closing timing of the intake valve is advanced and the opening timing of the exhaust valve is advanced so that the valve overlap period does not change . The control method for a direct injection type internal combustion engine, according to any one of claims 2 to 5,
The control method for a cylinder direct injection internal combustion engine, wherein the compression ratio control is a control for decreasing the increased effective compression ratio to the effective compression ratio before being increased when the actual fuel pressure reaches a target fuel pressure . The control method for a direct injection internal combustion engine, according to claim 6,
The compression ratio control is a control method for a direct injection type internal combustion engine, which is a control in which, when the effective compression ratio is decreased , the oil water temperature is gradually increased while the actual fuel pressure is being increased . An intake passage arranged so that a tumble flow is formed in the cylinder;
A fuel injection valve for injecting fuel into the cylinder,
An igniter for spark igniting a mixture formed by the fuel injected from the fuel injection valve,
In the first half of the intake stroke, the first fuel injection that forms a homogeneous mixture that is leaner than stoichiometry is performed throughout the cylinder, and the second fuel injection that forms a mixture that is richer than stoichiometry around the spark plug is compressed. After that, the control unit
In a control device for a direct injection internal combustion engine, which comprises:
Wherein the control unit is a fuel injection to be executed even during the compression stroke before the second fuel injection, the tumble by increasing the momentum of the fuel spray strength of the tumble flow is higher injection The pre-compression pre-injection that changes the flow in a direction that stabilizes combustion and the pre-compression pre-injection only after the engine is started and the fuel pressure is so low that the momentum of the fuel spray corresponding to the strength of the tumble flow cannot be secured. Then, in the case where the tumble flow cannot be changed to a place where combustion becomes stable, a compression ratio control for increasing the effective compression ratio of the in-cylinder direct injection internal combustion engine,
A control device for a direct injection type internal combustion engine, comprising:

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