JP2002256924A - Combustion control device for compression ignition type engine - Google Patents

Abstract

(57) [Problem] To enable compression ignition combustion over a wide range without extremely increasing the compression ratio. When an operation region is in a compression ignition region, a negative valve overlap period is formed in which both an intake valve and an exhaust valve close before and after top dead center of exhaust, and during this period, the first injection is performed. (S16) and stratified spark ignition (S1).
8) The mixture of the residual gas and the fuel is stratified and the intake gas flowing in the intake stroke is heated and heated by the combustion gas. As a result, even if the compression stroke start temperature rises and the compression ratio is about 10 to 15, the in-cylinder gas can be reliably raised to the compression ignition temperature by adiabatic compression during the compression stroke.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a combustion control system for a compression ignition engine capable of obtaining stable compression ignition combustion without extremely increasing a compression ratio.

[0002]

2. Description of the Related Art As means for improving the thermal efficiency of a four-cycle engine, it is known to increase the specific heat ratio of the working gas to thereby increase the theoretical thermal efficiency by making the air-fuel mixture lean. Further, by making the air-fuel mixture lean, even when the engine is operated with the same torque, more air is sucked into the engine, so that the pumping loss can be reduced.

[0003] Generally, in such lean burn, stratified charge combustion in which the mixture is collected around the spark plug in a stratified state to ensure ignitability is often employed. However, in stratified charge combustion, since a rich air-fuel mixture is concentrated around the spark plug, there is a problem that the combustion temperature becomes high and NOx tends to increase. Further, in a region where the air-fuel mixture is dense, the specific heat ratio is small, and the theoretical thermal efficiency tends to decrease. Furthermore,
In the stratified combustion, the combustion period is prolonged and the combustion is likely to be unstable, so there is a limit to lean air-fuel ratio.

On the other hand, a diesel engine has a high compression ratio and can substantially eliminate pumping loss by making the air-fuel ratio substantially lean, so that it has high thermal efficiency. However, since it is diffusion combustion, it has a low air utilization rate and a low output. , Soot may be emitted, and the exhaust gas characteristics are inferior. In a region where the air-fuel mixture is dense, the specific heat ratio is small, and the theoretical thermal efficiency tends to decrease.

Therefore, as a means for solving such a problem, there has been proposed a compression ignition type engine in which a gasoline mixture is ignited at multiple points by adiabatic compression without using a spark plug. In the compression ignition type engine, rapid combustion with short flame propagation is realized by multipoint ignition without using spark ignition, so that a local high-temperature portion is hardly formed in the combustion chamber, and NOx emission is greatly reduced. be able to.

[0006]

However, if the compression ratio is set to a high value of about 15 to 18 in order to obtain compression ignition combustion, the combustion pressure increases rapidly during high load operation due to an increase in the fuel injection amount. As a result, knocking is likely to occur.

To cope with this, for example, Japanese Patent Application Laid-Open No. 9-28
Japanese Patent No. 7528 proposes a technique for reducing the temperature of air-fuel mixture in a combustion chamber by external EGR or controlling the temperature of intake air by a cooling device provided in an EGR passage.

However, the combustion control based on the external EGR or the intake air temperature has a problem that the response is slow, and it is not possible to obtain a good follow-up property with respect to a torque change during running.

In order to avoid the occurrence of knocking,
It is possible to adopt a so-called Miller cycle in which the actual compression stroke is set shorter than the expansion stroke so that the expansion ratio becomes larger than the compression ratio. There is a problem that a sufficient amount of air is not supplied to the room, which causes a decrease in torque.

SUMMARY OF THE INVENTION In view of the above circumstances, the present invention provides a compression ignition engine capable of performing compression ignition combustion over a wide range while suppressing the occurrence of knocking and capable of greatly reducing exhaust emissions in all operation regions. It is an object to provide a combustion control device.

[0011]

In order to achieve the above object, the present invention provides an in-cylinder injector for directly injecting fuel into a combustion chamber with a spark plug and an exhaust valve and an intake valve both before and after exhaust top dead center. A variable valve mechanism capable of forming a negative valve overlap period in which the valve is closed, wherein the first fuel is injected during the negative valve overlap period, and An injection timing setting means for injecting a second fuel after opening the valve;
Ignition timing setting means for burning a mixture formed by the first fuel injected during the negative valve overlap period by spark ignition during the negative valve overlap period.

In such a configuration, in the compression ignition region,
Before and after the exhaust top dead center, a negative valve overlap period is formed in which both the intake valve and the exhaust valve close, and the first fuel is injected into the residual gas trapped in the combustion chamber, and this fuel and By burning the mixture with the residual gas by spark ignition during the negative valve overlap period, the gas temperature in the combustion chamber is increased, and the intake air sucked during the intake stroke is heated by the heat energy of the combustion gas. As a result, the gas temperature at the start of the compression stroke rises, and the gas temperature in the combustion chamber can be raised to a temperature at which compression ignition combustion is possible by adiabatic compression during the compression stroke.

In this case, preferably, the amount of fuel injected at the first time is reduced as the engine load increases.

[0014]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS An embodiment of the present invention will be described below with reference to the drawings. FIG. 1 shows an overall configuration diagram of a compression ignition type engine.

1 is an engine body, 2 is a piston, 3 is a combustion chamber, 4 is an intake port, 5 is an exhaust port, 6
Denotes an intake valve, 7 denotes an exhaust valve, and a throttle valve 9 is interposed in an intake passage 8 communicating with the intake port 4. The throttle valve 9 is connected to an electronic control throttle device (not shown) for electronically controlling the throttle opening.
Further, an in-cylinder injector 11 is provided at the center of the top surface of the combustion chamber 3.
A curved concave piston cavity 2a is formed on the top surface of the piston 2 opposed to the injection direction of the in-cylinder injector 11. Furthermore,
One side of the combustion chamber 3 (a squish area in the present embodiment)
The ignition portion of the ignition plug 12 is exposed. Note that reference numeral 16 denotes a knock sensor, and 17 denotes a water temperature sensor.

Further, an exhaust passage 2 communicating with the exhaust port 5 is provided.
8, an O2 sensor 18 as an example of an air-fuel ratio detecting means is provided.
A three-way catalyst 29 for purifying by oxidizing O and HC and reducing NOx is interposed. The air-fuel ratio detecting means may be a wide-range air-fuel ratio sensor.

The compression ratio of the engine employed in this embodiment is set at 10 to 10 so that high-load operation by ordinary spark ignition can be realized while suppressing abnormal combustion such as knocking.
It is set to about 15.

The intake valve 6 and the exhaust valve 7 are connected to the variable valve mechanisms 13a and 13b, respectively. In the present embodiment, each of the variable valve mechanisms 13a and 13b includes a dual cam including a spark ignition intake cam and a compression ignition intake cam, and a spark ignition exhaust cam and a compression ignition exhaust cam. These cams are switched according to the operation area.

When the compression ignition intake cam and the compression ignition exhaust cam are selected, as shown in FIG.
Is advanced from the exhaust top dead center (TDC), and the valve opening timing IVO of the intake valve 6 is set at the exhaust top dead center (TD).
C), a negative valve overlap period is formed in which both the valves 6 and 7 are closed before and after the exhaust top dead center (TDC). Due to this negative valve overlap period,
The intake air supplied to the combustion chamber 3 during the intake stroke is heated and heated by the thermal energy of the residual gas confined in the combustion chamber 3. The intake cam for compression ignition and the exhaust cam for compression ignition may be connected to a variable valve timing (VVT) mechanism capable of variably setting the rotation phase according to the engine load Lo.

On the other hand, as shown in FIG. 5B, when the spark ignition intake cam and the spark ignition exhaust cam are selected, the two valves 6, 7 are placed before and after the exhaust top dead center (TDC). Is returned to the normal valve timing in which a positive valve overlap period in which both valves are opened is formed.

Signals detected by these sensors are input to an electronic control unit (ECU) 20. An electronic control unit (ECU) 20 includes a CPU 21, a ROM 22, a RAM
23, an input port 24, an output port 25, etc., and are connected to each other by a bidirectional bus 26.

The input port 24 is connected to a crank angle sensor 31 for generating a crank pulse for each set crank angle in addition to the above-described sensors, and generates an output voltage proportional to the amount of depression of an accelerator pedal 32. The load sensor 33 is connected via an A / D converter 34. The output port 25 is connected to each of the variable valve mechanisms 13a, 13a, via an intake valve drive circuit 36a and an exhaust valve drive circuit 36b.
13b, is connected to the ignition plug 12 via an ignition drive circuit 36c, and is further connected to the in-cylinder injector 11 via an injector drive circuit 36d.

The electronic control unit (ECU) 20 determines the operating range based on the engine speed Ne calculated based on the signal from the crank angle sensor 31 and the engine load Lo detected based on the signal from the load sensor 33. It is checked whether it is in the compression ignition region or the spark ignition region. If it is in the compression ignition region, the throttle valve 9 is fully opened, and the fuel injection amount that can obtain the optimal compression ignition combustion is calculated. Perform ignition combustion control. When the operation region is in the spark ignition region, normal spark ignition combustion control is executed.

Since the compression ratio of the engine employed in the present embodiment is set at about 10 to 15, the engine is confined in the combustion chamber 3 during the negative valve overlap period in the compression ignition combustion control. Even if the intake air is heated and heated by the thermal energy of the residual gas, it is difficult to raise the temperature of the air-fuel mixture to the compression ignition temperature because the compression ratio is low. Therefore, in the present embodiment, the fuel is injected once to the residual gas confined in the combustion chamber 3 during the negative valve overlap period, and the fuel is stratified by spark ignition. By burning the residual gas, the combustion chamber 3 during the intake stroke
The heating of the intake air flowing into the cylinder is greatly enhanced, and the in-cylinder gas can be reliably heated to the compression ignition temperature by the adiabatic compression in the next stroke (compression stroke).

The combustion control, such as fuel injection control and ignition timing control, performed by the electronic control unit (ECU) 20 is specifically performed according to a combustion control routine shown in FIG.

In this routine, first, in step S1
Then, based on the engine speed Ne and the engine load Lo, referring to the operation region map shown in FIG. 4, it is determined whether the operation region is in the compression ignition region or the spark ignition region.
When it is in the compression ignition region, the process proceeds to step S2, and when it is in the spark ignition region, the process proceeds to step S5. Note that the compression ignition region in the present embodiment is set to a low-medium rotation and low-medium load region as shown in FIG.
It is set to other high rotation or high load regions.

In step S2, the throttle valve 9 is fully opened, and then in step S3, the signal for selecting the compression ignition intake cam and the compression ignition exhaust cam for the variable valve mechanisms 13a and 13b. And the intake valve 6
The exhaust valve 7 is operated at a valve timing that forms a negative valve overlap period before and after the exhaust top dead center (TDC) as shown in FIG.

When the compression ignition intake cam and the compression ignition exhaust cam are connected to the VVT mechanism, the rotation of the compression ignition intake cam and the compression ignition exhaust cam is performed based on the engine load Lo. During low load operation, the valve opening timing IVO of the intake valve 6 is retarded, and the exhaust valve 7
Is advanced, and the negative valve overlap period is extended almost symmetrically with respect to the exhaust top dead center (TDC). On the other hand, during the medium load operation, the valve opening timing IVO of the intake valve 6 is increased.
And the valve closing timing EVC of the exhaust valve 7 is retarded,
Exhaust top dead center (TDC) during negative valve overlap period
Is controlled to be substantially symmetrical with respect to.

As described above, the rotational phase of the compression-ignition intake cam and the compression-ignition exhaust cam is varied by the VVT mechanism so that the heat energy of the residual gas is the lowest. By extending the lap period, the amount of residual gas confined in the combustion chamber 3 is increased, thereby ensuring the ignitability of the fuel supplied by the first injection (hereinafter, abbreviated as “first injection”) described below. be able to. >> Also, by narrowing the negative valve overlap period on the medium load operation side where the heat energy of the residual gas is relatively high,
Intake heating by the heat energy of the residual gas is suppressed, and stable compression ignition combustion can be obtained even if the mixture is gradually enriched with an increase in engine load while avoiding knocking.

Next, the routine proceeds to step S4, in which compression ignition combustion control is executed, and the routine exits. This compression ignition combustion control is executed according to a compression ignition combustion control subroutine shown in FIG. In this routine, first, step S
At 11, the fuel injection amount is calculated by a map search or calculation based on the engine load Lo and the engine speed Ne. Next, the process proceeds to step S12, where the engine load Lo is determined.
<Calculation of injection ratio> indicating the ratio between the fuel amount at the time of the first injection and the fuel amount at the time of the second injection (hereinafter abbreviated as “second injection”) in accordance with the above. The injection ratio is set as << the ratio of the first injection, for example, at most 10% >>, and as the engine load Lo increases, the ratio of the fuel amount of the first injection (hereinafter abbreviated as "first injection fuel amount"). Until it reaches 0% ">>". Therefore, the ratio of the fuel amount of the second injection (hereinafter, abbreviated as “the fuel amount at the time of the second injection”) changes in the “range of 90 to 100%” as the engine load Lo increases.

With the increase in the engine load Lo, the first injection fuel amount is reduced to reduce the heat energy due to the combustion of the residual gas, so that the air-fuel ratio is gradually enriched with the increase in the engine load Lo. Even in this case, stable compression ignition combustion can be obtained while avoiding abnormal combustion such as knocking, and high output can be achieved.

Next, the process proceeds to step S13, and in this step S13 and the next step S14, step S1 is executed.
The first injection fuel amount and the second injection fuel amount are calculated based on the fuel injection amount calculated in step 1 and the injection ratio calculated in step S12.

Thereafter, the process proceeds to step S15, and waits until the first injection timing is reached. This first injection timing is within a negative valve overlap period in which both the intake valve 6 and the exhaust valve 7 are closed, and a rich mixture around the ignition portion of the ignition plug 12 at the time of stratified spark ignition described later. This is a value indicating when the injection should be started in order to form the gas in a stratified state, and may be a fixed value or a variable value calculated back from the ignition timing based on the engine speed Ne.

When the first injection timing has been reached, the process proceeds from step S15 to step S16, where a first injection signal having a pulse width corresponding to the first injection fuel amount is output to the injector drive circuit 36d. Then, the first-injection fuel amount calculated in step S13 is injected into the combustion chamber 3 from the in-cylinder injector 11 (see FIG. 6A).

Next, the routine proceeds to step S17, and waits until the stratified spark ignition timing is reached. This stratified spark ignition timing is a fixed value, and the in-cylinder pressure at the time of combustion is determined by the exhaust top dead center (TD).
The timing is set such that the maximum is obtained just after C).

When the stratified spark ignition timing is reached,
The process proceeds from step S17 to step S18 to output an ignition signal to the ignition drive circuit 36c. Then, a stratified rich air-fuel mixture formed by the residual gas and the fuel formed around the ignition portion of the ignition plug 12 is forcibly ignited, and is subjected to flame propagation combustion (first combustion). Since the first combustion is also a positive work, the thermal efficiency does not decrease.

Thereafter, the process proceeds to step S19, and waits until the second injection timing is reached. The second injection timing is set to a relatively early timing after the intake valve 6 is opened during the intake stroke. By setting the second injection timing relatively early, it is possible to generate a uniform mixture before the gas temperature in the combustion chamber 3 reaches the compression ignition allowable temperature.

When the second injection timing has been reached, the process proceeds from step S19 to step S20, in which a second injection signal having a pulse width corresponding to the second injection fuel amount is output to the injector drive circuit 36d, and the routine exits. . Then, the second-injection fuel amount calculated in step S14 is injected into the combustion chamber 3 from the in-cylinder injector 11 (FIG. 6).
(b)).

At this time, during the intake stroke, the intake valve 6 opens and the intake air flowing into the combustion chamber 3 is heated by the combustion gas whose thermal energy is increased by the combustion during the negative valve overlap period. Is done. Then, the process proceeds to the next compression stroke while the in-cylinder gas and the fuel from the second injection are uniformly mixed.

In the compression stroke, the gas temperature at the start of the compression stroke is increased, so that even in the case of adiabatic compression with a compression ratio of about 10 to 15, the gas temperature in the combustion chamber 3 is reliably reduced to the compression ignition temperature. Can be raised. When the compression ignition temperature is reached, the air-fuel mixture in the combustion chamber 3 is ignited all at once and the flame does not propagate, that is, multipoint ignition combustion (inhomogeneous compression ignition combustion) in which an infinite number of spark plugs are arranged.
Is realized.

On the other hand, when it is determined in step S1 of the combustion control routine shown in FIG. 2 that the operation region is in the spark ignition region and the process proceeds to step S5, the normal combustion control by spark ignition is executed, and the routine exits. When the process shifts to the spark ignition combustion control, a signal for switching to the spark ignition intake cam and the spark ignition exhaust cam is output to the variable valve mechanisms 13a and 13b. As a result, the intake valve 6 and the exhaust valve 7 are operated at the valve timing at the time of normal spark ignition, that is, during the positive valve overlap period (see FIG. 5B) in which both the valves are opened from the end of the exhaust stroke to the beginning of the intake stroke. You. The cam profiles of the spark-ignition intake cam and the spark-ignition exhaust cam are set so as to maximize the volumetric efficiency.

At the same time, the throttle valve 9 is operated in conjunction with the accelerator pedal 32, and the fuel injection amount, fuel injection timing, ignition timing, etc. are returned to normal spark ignition combustion control (see FIG. 6 (b)). . Since these controls are publicly known, description thereof is omitted here.

As described above, according to the present embodiment, during the compression ignition combustion, before and after the exhaust top dead center (TDC), the negative valve overlap period in which both the intake valve 6 and the exhaust valve 7 close. Is formed, the residual gas is confined in the combustion chamber 3, the intake gas is heated and heated by the heat energy of the residual gas, and the fuel is injected during the negative valve overlap period. Since stratified combustion is performed by spark ignition, intake air heating can be greatly enhanced. As a result, stable compression ignition combustion can be obtained even when the compression ratio is about 10 to 15 without extremely increasing the compression ratio of the engine. Furthermore, by variably setting the first injection fuel amount in accordance with the engine load Lo, it is possible to ensure ignitability in a low load region and to suppress knocking in a medium load region. Compression ignition combustion can be performed over a wide range.

[0044]

As described above, according to the present invention,
In the compression ignition region, a negative valve overlap period is formed in which both the intake valve and the exhaust valve close before and after the exhaust top dead center, and fuel is supplied by the first injection to residual gas trapped in the combustion chamber. In addition, since the combustion is performed by spark ignition, the intake air heating is strengthened, and the gas temperature in the combustion chamber can be easily increased to a temperature at which compression ignition combustion is possible by adiabatic compression without extremely increasing the compression ratio. And stable compression ignition combustion can be obtained. As a result, it is possible to improve fuel efficiency and significantly reduce exhaust emissions.

In this case, by variably setting the fuel amount at the time of the first injection in accordance with the operation range, it is possible to ensure the ignitability in the low load range and to suppress the occurrence of knocking in the medium load range. Thus, compression ignition combustion can be performed over a wide range. As a result, in combination with the normal spark ignition combustion, it is possible to significantly reduce the exhaust emission in the entire operation range.

[Brief description of the drawings]

FIG. 1 is an overall configuration diagram of a compression ignition engine.

FIG. 2 is a flowchart showing a combustion control routine.

FIG. 3 is a flowchart showing a compression ignition combustion control subroutine.

FIG. 4 is an explanatory diagram showing an operation area map.

5A and 5B are explanatory diagrams showing valve timings of an intake valve and an exhaust valve for each operation region, wherein FIG. 5A shows a compression ignition region, and FIG. 5B shows a spark ignition region.

FIG. 6 is an explanatory diagram showing changes in in-cylinder pressure of compression ignition combustion and spark ignition combustion.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Engine main body 3 Combustion chamber 6 Intake valve 7 Exhaust valve 11 In-cylinder injector 12 Spark plug 13a, 13b Variable valve mechanism TDC Exhaust top dead center

──────────────────────────────────────────────────の Continued on the front page (51) Int.Cl. ⁷ Identification symbol FI Theme coat ゛ (Reference) F02D 41/04 385 F02D 41/04 385C 41/38 41/38 B 43/00 301 43/00 301J 301Z 301A F02P 5/15 F02P 5/15 B F term (reference) 3G022 AA07 EA02 EA07 GA01 GA05 GA08 3G023 AA03 AA04 AA05 AA06 AB01 AB05 AC04 AF01 AG01 AG03 3G084 AA00 AA03 BA13 BA15 BA16 BA23 DA10 DA11 FA10 FA29A33 FA AB02 BA08 BB01 BB06 BB13 DA01 DA02 DA04 DA09 DA12 DE03S EA01 EA02 EA03 EA04 EC10 FA05 FA16 FA17 FA18 FA21 GA05 GA06 HA12X HA13X HB01X HB02X HD05Z HE01Z HE03Z HF08Z 3G301 HA01 HA04 HA13 JA08 MA08 JA22 JA22 JA22 JA08 NC04 PD02A PD02Z PE03Z PF03Z

Claims (2)

[Claims] A negative valve overlap period for closing both an exhaust valve and an intake valve around an exhaust top dead center around an ignition plug, an in-cylinder injector for directly injecting fuel into a combustion chamber, and an exhaust top dead center. A compression-ignition engine having a variable valve mechanism capable of injecting a first fuel during the negative valve overlap period and a second fuel injection after the intake valve is opened. Setting means; and ignition timing setting means for burning a mixture formed by the first fuel injected during the negative valve overlap period by spark ignition during the negative valve overlap period. Combustion control device for compression ignition engine. 2. The combustion control device for a compression ignition engine according to claim 1, wherein the amount of fuel injected at the first time is decreased with an increase in engine load.