JP2002242715A - Compression ignition type engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To enlarge a compression ignition range toward its high-load side. SOLUTION: During a medium-load operation, the phase of rotation of an intake cam and an exhaust cam is varied by a variable valve timing mechanism in the direction to reduce the period of a negative valve overlap and to set the closing timing IVC of the intake valve in the vicinity of an intake bottom dead center(BDC) (b). By reducing the period of the negative valve overlap at a medium load, the amount of residual gas confined in a combustion chamber is reduced and intake heating is also reduced. By setting the closing timing IVC of the intake valve in the vicinity of the intake bottom dead center(BDC), volumetric efficiency is decreased and thus self-ignition is restrained whereby stable compression ignition performance can be achieved even if the operating region of the engine is in a region (high-load side) where a air-fuel mixture is richer.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a compression ignition type engine for igniting a mixture in a combustion chamber at multiple points by adiabatic compression.

[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] However, leaning of the air-fuel mixture is limited due to prolonged combustion period and unstable combustion. To overcome this problem, in-cylinder injection is used to increase the limit by stratified charge combustion, in which the mixture is stratified around the spark plug while maintaining a stratified state, to ensure ignitability. Since the air-fuel mixture is concentrated, there is a problem that the combustion temperature becomes high and NOx tends to increase.

On the other hand, a diesel engine has a high compression ratio and a substantially lean air-fuel ratio, thereby making it possible to almost eliminate a pump loss. Therefore, the diesel engine has a high thermal efficiency. , Soot may be emitted, and the exhaust gas characteristics are inferior.

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 order to realize compression ignition combustion, it is necessary to activate fresh air using high-temperature residual gas heat.
One method is to advance the closing timing of the exhaust valve and delay the opening timing of the intake valve to form a negative overlap period in which both valves close before and after the exhaust top dead center. There is known a technique for trapping residual gas in a combustion chamber in the first half of the stroke.

[0006] For example, Japanese Patent Application Laid-Open No. 2000-64863 provides a negative valve overlap period in which both the exhaust valve and the intake valve are closed before and after the top dead center of the exhaust gas, and the pre-heating of the residual gas trapped in the combustion chamber is increased. Discloses a technique for promoting compression ignition.

In this prior art, during low-load operation, the residual gas amount is increased by advancing the closing timing of the exhaust valve, and the residual gas is compressed by retarding the opening timing of the intake valve. The work required to do so is recovered to prevent a drop in thermal efficiency.

[0008]

By the way, in the above-mentioned prior art, the thermal efficiency is improved by controlling the negative valve overlap period before and after the top dead center of the exhaust gas. Thermal energy fluctuates according to the engine load. That is, the residual gas temperature during the low-load operation is low, and gradually increases as the operation shifts to the medium-load operation.

In the prior art described above, the ignition is performed during the compression stroke by utilizing the heat energy of the residual gas confined in the combustion chamber during the negative valve overlap period. In some cases, the gas does not reach the ignition temperature, causing poor ignition, or, conversely, early ignition, which makes it difficult to obtain stable compression ignition performance. As a result, there is a problem that the compression ignition region becomes narrow.

SUMMARY OF THE INVENTION In view of the above circumstances, an object of the present invention is to provide a compression ignition engine capable of obtaining a stable compression ignition performance over a wide range without being affected by load fluctuations.

[0011]

SUMMARY OF THE INVENTION In order to achieve the above object, the present invention makes it possible to form a negative valve overlap period in which both the exhaust valve and the intake valve are closed before and after the top dead center of the exhaust gas. In a compression ignition type engine provided with a variable valve mechanism, the negative valve overlap period is narrowed and the closing timing of the intake valve is advanced as the engine load increases.

In such a configuration, the thermal energy of the residual gas trapped during the negative valve overlap period is
During low load operation, it is low, and gradually increases as it shifts to medium load operation. As the engine load increases, the valve closing timing of the intake valve is advanced to reduce the volumetric efficiency and negative valve overlap period , The auto-ignition is suppressed by reducing the intake air heating, and compression ignition can be performed in a region where the air-fuel ratio is deeper.

In this case, preferably, the closing timing of the intake valve at the time of medium load operation is set near the intake bottom dead center.

2) The valve closing timing of the intake valve during low load operation is retarded to a position beyond the intake bottom dead center.

[0015]

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.

An injection hole of an in-cylinder injector 11 as a fuel injection means faces the center of the top surface of the combustion chamber 3, and a piston 2 provided in the injection direction of the in-cylinder injector 11 is provided. A curved concave-shaped piston cavity 2a is formed on the top surface. Further, the ignition portion of the ignition plug 12 faces one side of the combustion chamber 3 (a squish area in the present embodiment).

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.

Further, a compression ignition intake cam and a compression ignition exhaust cam are connected to a variable valve timing (VVT) mechanism. This VVT mechanism variably sets the valve timing according to the engine load Lo by changing the rotation phase of a compression ignition intake cam and a compression ignition exhaust cam by an actuator such as a hydraulic solenoid (or an electromagnetic solenoid). is there. Reference numeral 16 denotes a knock sensor, 17 denotes a water temperature sensor, and 18 denotes an O2 sensor.

The 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 sensors, and generates an output voltage proportional to the amount of depression of an accelerator pedal 32. A 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.
3b. Furthermore, this output port 2
5, an in-cylinder injector 11 and a spark plug 12 are connected via a drive circuit (not shown).

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 the engine is in the compression ignition region or the spark ignition region. When the engine is in the compression ignition region, the throttle valve 9 is fully opened (FIG. 5).
(Refer to (d)), the fuel injection amount, the injection timing, and the valve timings of the intake and exhaust valves 6 and 7 that can obtain the optimal compression ignition combustion are set. When the operation region is in the spark ignition region, normal spark ignition control is executed.

In order to obtain the optimum compression ignition combustion in the compression ignition region, the temperature of the fresh air sucked in the intake stroke is increased by the heat energy of the residual gas, and the residual gas and the fresh air are combined with each other in the compression stroke. It is necessary to raise the temperature to a temperature at which compression ignition is possible by adiabatic compression in the above. But,
The thermal energy of the residual gas fluctuates according to the engine load, is low during low load operation, and increases as the operation shifts to medium load operation.

Therefore, when performing the compression ignition control, the valve timing is continuously changed in accordance with the engine load, and as shown in FIGS. 4 (a), (b) and 5 (c), the engine load is reduced. As the vehicle shifts from the low load operation to the medium load operation, the valve closing timing EVC of the exhaust valve 7 is gradually retarded and the valve opening timing IVO of the intake valve 5 is gradually advanced, so that the negative valve overlap period Is controlled to be wider on the low load side and narrower as it shifts toward the middle load direction.

As a result, the amount of residual gas confined in the combustion chamber 3 increases on the low-load operation side and decreases as the engine shifts to the medium-load operation direction. By increasing the amount of residual gas on the low load operation side, the temperature of the air-fuel mixture is increased by the heat energy of the residual gas, and compression ignition at a leaner air-fuel ratio becomes possible. On the other hand, as the vehicle shifts to the medium load operation direction, the self-ignition is suppressed by reducing the residual gas amount,
Avoid knocking.

In this case, if the closing timing of the intake valve 6 is constant, the region where compression ignition is possible becomes narrow. For example, FIG.
As shown in FIG. 3A, when the valve closing timing IVC of the intake valve 6 is fixed at a position beyond the intake bottom dead center (BDC), FIG.
Compression ignition combustion is possible only in a narrow range of a low load region surrounded by hatching shown in FIG. This is because when controlling the negative valve overlap period and performing compression ignition combustion using the heat energy of the residual gas, the thermal energy of the residual gas increases as the shift to the medium load direction occurs, This is because knocking and the like easily occur.

Therefore, in the present embodiment, in the compression ignition region, the negative valve overlap period is variably set according to the engine load Lo to control not only the air-fuel mixture temperature but also the closing of the intake valve 6. The volume efficiency is controlled by variably setting the valve timing IVC so as to obtain appropriate compression ignition combustion.

The closing timing of the intake valve 6 is, specifically,
It is set in the combustion control routine shown in FIG. In this routine, first, in step S1, based on the engine speed Ne and the engine load Lo, with reference to the operation region map shown in FIG. 3, it is determined whether the operation region is in the compression ignition region or the spark ignition region. Find out. Note that the compression ignition region in the present embodiment is set to a region surrounded by a solid line including a hatched region in the figure, that is, a low-medium rotation, low-medium load region. Further, the compression ignition region is divided into a stratified compression ignition region, a medium load region is divided into a uniform compression ignition region, and a fuel injection timing is relatively reduced in the latter half of the compression stroke in the stratified compression ignition region according to the engine load Lo. Set later. Further, in the uniform compression ignition region, the fuel injection timing is set from a time after the start of the negative valve overlap period to a relatively early time during the intake stroke (see FIG. 6).

When the operating region 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 S6.

In step S2, the throttle valve 9 is fully opened, and then in step S3, a 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
And the exhaust valve 7 is operated at the valve timing at the time of compression ignition.

Then, the process proceeds to a step S4, wherein a negative valve overlap period is set. As shown in FIGS. 4A and 4B, the negative valve overlap period is maximized during low load operation, minimized during medium load operation, and variably set in that range according to the engine load Lo. Is done.

Specifically, based on the engine load Lo,
The valve overlap period setting table shown in FIG. 5C is referenced with interpolation calculation using the engine load Lo as a parameter, and a drive signal is output to each of the variable valve mechanisms 13a and 13b. Then, each of the variable valve mechanisms 13a, 13
b, the VVT mechanism changes the rotation phase of the compression ignition intake cam and the compression ignition exhaust cam, respectively, and delays the valve opening timing IVO of the intake valve 6 during low load operation;
The valve closing timing EVC of the exhaust valve 7 is advanced to extend the negative valve overlap period. On the other hand, during the medium load operation, the valve opening timing IVO of the intake valve 6 is advanced and the valve closing timing EVC of the exhaust valve 7 is retarded to narrow the negative valve overlap period.

As shown in FIG. 4B, the cam profiles of the intake cam for compression ignition and the exhaust cam for compression ignition are:
The valve closing timing IVC of the intake valve 6 during medium load operation is set to be near the intake bottom dead center (BDC), and the valve opening timing EVO of the exhaust valve 7 is set to the expansion bottom dead center (BDC). , The closing timing EVC of the exhaust valve 7 and the opening timing I of the intake valve 6
The negative valve overlap period formed with VO is set at a position symmetrical with respect to the exhaust top dead center (TDC). Then, as the engine load shifts from the medium load operation to the low load operation shown in FIG. 7A, the compression ignition intake cam and the compression ignition exhaust cam change the rotational phase symmetrically.

As a result, as shown in FIG. 6, as the engine load shifts from the medium load operation to the low load operation, the valve timing of the exhaust valve 7 is advanced, while the valve timing of the intake valve 6 is changed to the exhaust timing. The valve 7 is retarded in a symmetrical direction.

As described above, at the time of low load operation in which the heat energy of the residual gas is the lowest, the valve closing timing of the intake valve 6 is retarded, so that the volumetric efficiency is improved by inertial supercharging,
More fresh air can be inhaled. In addition, since the negative valve overlap period at this time is set to be wide (see FIG. 5C), the amount of the residual gas confined in the combustion chamber 3 is large, and the heating of the intake air by the heat energy of the residual gas is difficult. As a result, good compression ignition performance can be obtained even with a leaner air-fuel ratio.

On the other hand, during a medium load operation in which the residual gas confined in the combustion chamber 3 has relatively high thermal energy, the negative valve overlap period is short, so that intake air heating by the residual gas thermal energy is suppressed. Further, since the closing timing of the intake valve 6 is set near the intake bottom dead center (BDC), the volume efficiency is reduced, and the compression pressure is reduced, so that self-ignition is suppressed, and the occurrence of knocking is avoided. Stable compression ignition combustion can be obtained with a rich air-fuel mixture. as a result,
As shown in FIG. 3, the uniform compression ignition region can be expanded from the conventional region indicated by hatching to the high load side.

Next, the routine proceeds to step S5, in which compression ignition fuel injection control is executed, and the routine exits. In the compression ignition fuel injection control, processing for variably setting the fuel injection amount and the fuel injection timing is performed.

The fuel injection amount is, as shown in FIG.
Control is performed to gradually increase the air-fuel ratio as the engine load Lo decreases. Note that reference numeral λ0 in the figure indicates a stoichiometric air-fuel ratio. The fuel injection timing is set based on the engine speed Ne and the engine load Lo. For example, when the operating region is in a low load region, the intake bottom dead center (BDC)
At a later time, that is, after the start of the compression stroke, the intake bottom dead center (BD) is set when the exhaust valve 7 is closed (at the start of the negative valve overlap period) at medium load and low / medium rotation.
Set relatively early, until C).

By injecting fuel from the in-cylinder injector 11 into the combustion chamber 3 at a time later than the intake bottom dead center (BDC), the combustion chamber 3 is reaching the gas temperature at which compression ignition combustion is possible. A stratified mixture is locally generated, and stratified compression ignition combustion can be performed with an extremely lean mixture. On the other hand, by setting the fuel injection timing relatively early after the exhaust valve 7 closes, a uniform mixture can be generated before the gas temperature in the combustion chamber 3 reaches the self-ignition possible temperature. When the temperature reaches the ignition temperature, this air-fuel mixture ignites all at once and combustion in which the flame does not propagate, that is, multipoint ignition combustion (uniform compression ignition combustion) in which an infinite number of spark plugs are arranged, is realized.

When it is determined in step S1 that the operation region is in the spark ignition region and the process proceeds to step S6, the normal combustion control by spark ignition is executed, and the routine exits. When shifting to spark ignition combustion control, the variable valve mechanism 13a, 1
A signal for switching to the spark ignition intake cam and the spark ignition exhaust cam is output to 3b. As a result, a positive valve overlap period in which the intake valve 6 and the exhaust valve 7 open together during normal spark ignition, that is, from the end of the exhaust stroke to the beginning of the intake stroke (FIGS. 4 (c) and 5 (c)) ))). 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 (see FIG. 5 (d)), and the fuel injection amount, fuel injection timing, ignition timing, etc. are returned to normal spark ignition control. Since these controls are known,
The description here is omitted.

[0042]

As described above, according to the present invention,
During compression ignition operation, as the engine load increases, the intake valve heating period is reduced by narrowing the negative valve overlap period, and the valve closing timing of the intake valve is advanced to reduce the volumetric efficiency. Self-ignition is suppressed, and stable compression ignition performance can be obtained in a region where the air-fuel ratio is deeper. As a result, the compression ignition region can be expanded to a higher load side.

[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 an explanatory diagram showing an operation area map.

4A and 4B are explanatory diagrams showing valve timings of an intake valve and an exhaust valve. FIG. 4A is a low-load operation, FIG.
(C) shows high load operation

FIG. 5 is an explanatory diagram showing a relationship between an engine load and each control characteristic.

FIG. 6 is an explanatory diagram showing a relationship between valve timing and in-cylinder pressure characteristics.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Engine main body 6 Intake valve 7 Exhaust valve BDC Intake bottom dead center IVC Closing timing Lo Engine load TDC Exhaust top dead center

 ──────────────────────────────────────────────────続 き Continuing on the front page (72) Inventor Hitoshi Ito 3-9-6 Osawa, Mitaka City, Tokyo Inside Subaru Research Institute, Inc. (72) Inventor Yohei Saiku 3-9-6 Osawa, Mitaka City, Tokyo F-term in Subaru Research Institute (reference) JA23 KA08 KA09 LA07 LB04 MA11 MA18 NA08 NC02 NE11 NE12 NE15 PC08Z PE03Z PE08Z PF03Z

Claims (3)

[Claims] 1. A compression ignition type engine having a variable valve operating mechanism capable of forming a negative valve overlap period in which both an exhaust valve and an intake valve are closed before and after a top dead center of an exhaust gas. A compression ignition type engine characterized in that the negative valve overlap period is narrowed and the intake valve closing timing is advanced as the pressure increases. 2. The compression ignition engine according to claim 1, wherein the closing timing of the intake valve during the middle load operation is set near the intake bottom dead center. 3. The compression ignition engine according to claim 1, wherein the closing timing of the intake valve during low load operation is retarded to a position beyond the intake bottom dead center.