JP7251287B2 - Homogeneous mixture compression ignition engine controller - Google Patents

Description

The present invention relates to an apparatus for controlling an engine capable of premixed compression ignition combustion in which fuel injected into a combustion chamber is mixed with air and burned by self-ignition.

In recent years, attention has been paid to premixed compression ignition combustion (HCCI combustion) in which gasoline fuel mixed with air is burned by self-ignition in a combustion chamber. Homogeneous charge compression ignition combustion is a form in which the air-fuel mixture burns multiple times at the same time. It is said to be very advantageous in terms of (thermal efficiency).

As one type of the above-described premixed compression ignition combustion, a combustion type that combines combustion by self-ignition of an air-fuel mixture and forced combustion using a spark plug has been proposed. That is, spark ignition is used as a trigger to forcibly burn part of the air-fuel mixture by flame propagation (SI combustion), and to burn the other air-fuel mixture by self-ignition (CI combustion). Such combustion is hereinafter referred to as partial compression ignition combustion.

As an example of an engine that employs the partial compression ignition combustion, the one disclosed in Patent Document 1 below is known. Specifically, in the engine of Patent Document 1, when performing partial compression ignition combustion (referred to as SI-CI combustion in the document), the amount of heat release due to SI combustion with respect to the total amount of heat release in one cycle The EGR rate in the combustion chamber, the timing of spark ignition by the spark plug, etc. are controlled so that the SI rate, which is the ratio of , coincides with a target value (target SI rate) determined according to the engine load (target torque).

JP 2018-084183 A

Here, in the engine of Patent Document 1, the geometric compression ratio is set to a relatively high value (for example, 17 or more and 20 or less) in order to enable partial compression ignition combustion of the air-fuel mixture. Such a high compression ratio tends to induce abnormal combustion such as pre-ignition and knocking especially in a high load range of the engine. In this regard, in Patent Document 1, the combustion type in the high load range is switched from partial compression ignition combustion to SI combustion, that is, to a combustion type in which all of the air-fuel mixture is forcedly combusted by spark ignition. However, if the partial compression ignition combustion is prohibited in the high-load region of the engine (the entire air-fuel mixture is SI-burned) in this way, even if abnormal combustion can be avoided, sufficient fuel efficiency can be improved. no effect will be obtained.

On the other hand, in order to enable partial compression ignition combustion in a high load range, it is proposed to inject fuel not only in the intake stroke but also in the compression stroke. This is because the fuel injection during the compression stroke suppresses the temperature rise in the combustion chamber, which is considered to work to suppress abnormal combustion. However, if the fuel injection amount during the compression stroke is increased indiscriminately in order to suppress abnormal combustion, a locally rich air-fuel mixture, which leads to the generation of smoke and an increase in NOx, is likely to be formed in the combustion chamber, resulting in engine failure. Emission performance may deteriorate.

The present invention has been made in view of the above circumstances, and provides a control device for a premixed compression ignition engine capable of suppressing abnormal combustion in a high load range and improving emission performance. intended to provide

In order to solve the above problems, the inventors of the present application sufficiently enhance the scavenging action of the combustion chamber (the action of discharging high-temperature burned gas from the combustion chamber) in the high load range, A study was made to reduce dependence on fuel injection during the compression stroke (hereinafter referred to as post-injection), which is required for normalization of partial compression ignition combustion. As a result of investigation, it was found that if the valve overlap period in which both the intake and exhaust valves are open is appropriately set, sufficient scavenging action will be exhibited, reducing the dependence on post-injection in the high-load range and reducing excessive fuel consumption. It has been found that a good emission performance is ensured by eliminating the need for an extra post-injection. This is not limited to the case of adopting partial compression ignition combustion (that is, the combustion mode in which a part of the air-fuel mixture is subjected to premixed compression ignition combustion), but also to the case of adopting the combustion mode in which all of the air-fuel mixture is subjected to premixed compression ignition combustion. But it's the same.

The present invention is made based on the findings as described above. That is, a control device according to a first aspect of the present invention includes a combustion chamber, an injector that injects fuel into the combustion chamber, an intake valve that opens and closes an intake port for introducing intake air into the combustion chamber, and An exhaust valve that opens and closes an exhaust port for discharging exhaust gas, and controls an engine capable of premixed compression ignition combustion in which the fuel injected from the injector is mixed with air and burned by self-ignition. a variable valve mechanism capable of changing a valve overlap period in which the open period of the intake valve and the open period of the exhaust valve overlap; post-injection for injecting fuel during a compression stroke while driving the variable valve mechanism to form the valve overlap period of a predetermined amount or more so that the mixture of fuel and air undergoes premixed compression ignition combustion; and a combustion control unit that causes the injector to perform a pre-injection that is included in the intake stroke or the first half of the compression stroke and that injects fuel at a timing earlier than the post-injection, wherein the variable valve mechanism controls the opening timing and the opening timing of the intake valve. A phase-type intake valve variable mechanism that changes the closing timing simultaneously and by the same amount, wherein a portion of the high load region on the low speed side is defined as a first region, and a portion of the high load region on the high speed side is defined as a second region. Then, the combustion control unit controls the injector so that the start timing of the post-stage injection is later in the first region than in the second region, and the valve overlap period is set to be later than in the second region. The combustion control unit controls the variable valve mechanism so that it becomes longer in the first region, and the combustion control unit causes the intake valve to open before the exhaust top dead center in both the first region and the second region, and The intake valve is closed after the bottom dead center of the intake, and the operating phase of the intake valve in the second region is shifted to the retard side from the operating phase in the first region. It is characterized by controlling a variable mechanism (Claim 1).

According to the control device according to the first aspect , in the high-load region of the engine where the amount of heat release is large, in addition to pre-injection that injects fuel at a relatively early timing such as the intake stroke, fuel is injected during the compression stroke. is executed, the temperature rise in the combustion chamber during the compression stroke can be reduced by the latent heat of vaporization of the post-injection, and the occurrence of abnormal combustion, which is a concern in the high load range, can be effectively suppressed. can do.

In particular, in the first region, which is a part of the low speed side in the high load range, the start timing of the post-stage injection is delayed and the valve overlap period is extended compared to the second region on the high speed side, so abnormal combustion is prevented. It is possible to obtain a sufficient temperature control effect commensurate with the risk.

That is, by setting the start timing of the post-injection in the first region to a more retarded timing in the compression stroke, the latent heat of vaporization of the fuel by the post-injection is applied at the timing when the temperature of the combustion chamber rises due to compression. It is possible to enhance the cooling effect of the combustion chamber by the latent heat of vaporization. Similarly, by expanding the valve overlap period in the first region, a flow of intake air blowing through from the intake port to the exhaust port is formed during the valve overlap period, and as a result, the burned gas remaining in the combustion chamber is is promoted to the exhaust port, and the scavenging action can be enhanced. This, together with the effect of the latent heat of vaporization of the fuel due to the post-injection described above, leads to a sufficient suppression of temperature rise in the combustion chamber during the compression stroke. Oxidation reaction can be suppressed, and pre-ignition, which is abnormal combustion in which the air-fuel mixture ignites prematurely, can be suppressed.

The low-temperature oxidation reaction is a slow oxidation reaction that occurs before the high-temperature oxidation reaction (substantial combustion reaction) that generates high thermal energy with a flame, and is the latter half of the compression stroke when the combustion chamber becomes hot. is known to occur in According to the findings of the inventors of the present application, this low-temperature oxidation reaction is likely to occur remarkably under operating conditions such as those in the first region, that is, under conditions of low rotational speed and high load. Moreover, when the low-temperature oxidation reaction appears prominently, the temperature of the air-fuel mixture in the combustion chamber (air-fuel mixture before combustion) rises due to the reaction heat, so the air-fuel mixture self-ignites earlier than the regular combustion start time. , the probability of occurrence of pre-ignition increases. On the other hand, in the present control device , the valve overlap period is extended and the start timing of the post-stage injection is retarded during operation in the first region. Due to the synergistic effect, the combustion chamber can be cooled to the extent that the reaction level of the low-temperature oxidation reaction is sufficiently lowered, and the induction of abnormal combustion such as pre-ignition due to the low-temperature oxidation reaction can be effectively prevented. can be suppressed to

Moreover, according to the configuration of the control device , which can be expected to have a high scavenging action, there is no need to excessively increase or retard the post-stage injection to suppress preignition, so the deterioration of emission performance due to the post-stage injection can be prevented. can be minimized. That is, if the valve overlap period expansion control described above is not executed in the first region, the amount of the post-stage injection is greatly increased to suppress preignition, or the injection timing is greatly increased within the compression stroke. Retarding (closer to compression top dead center) may be necessary. However, if this is done, a locally rich air-fuel mixture is likely to be formed in the combustion chamber, which may cause smoke or an increase in the amount of NOx generated. On the other hand, according to the present control device in which the valve overlap period is extended in the first region, a sufficient scavenging action is exhibited during the valve overlap period, so that the post-injection is excessively increased or retarded. becomes unnecessary, and the above-mentioned problems (generation of smoke, etc.) can be avoided, and relatively good emission performance can be ensured.

On the other hand, in the second region, which is on the higher speed side than the first region, the fluidity in the combustion chamber is relatively high, so there is a possibility that a sufficient scavenging action can be obtained even if the valve overlap period is shortened. In addition, the actual time from the start of compression of the air-fuel mixture by the piston to the end of compression (heat-receiving time of the air-fuel mixture) is shortened, and the reaction level of the low-temperature oxidation reaction naturally decreases, so serious abnormal combustion such as pre-ignition occurs. risk is lower. On the other hand, in the present control device , the valve overlap period is shortened and the start timing of the post-stage injection is advanced in the second region, so that the required level of cooling effect corresponding to the risk situation in this region is ensured. Thus, while suppressing abnormal combustion, formation of a locally rich air-fuel mixture can be avoided, and emission performance can be improved.

Further, in this control device , when the valve overlap period is reduced in the second region, the phase-type variable intake valve mechanism is driven to retard the closing timing of the intake valve with respect to the intake bottom dead center. Therefore, the effective compression ratio of the engine, that is, the volume of the combustion chamber when the piston is at top dead center and the volume of the combustion chamber at the closing timing of the intake valve (in other words, the timing at which the piston substantially starts compression) As a result, the pumping loss can be reduced while suppressing knocking, which is a concern in the second region.

That is, in the second region where the rotational speed is relatively high and low-temperature oxidation reaction does not easily occur, the risk of preignition occurrence decreases as described above. Knocking, which is abnormal combustion in which the combustion gas is rapidly burned by local self-ignition, is likely to occur. Since knocking mainly depends on the environment of the combustion chamber at the time when the combustion of the air-fuel mixture starts, it is considered that the higher the temperature of the combustion chamber at the time of the start of combustion, the more likely it is to occur. On the other hand, in the present control device , the operating phase of the intake valve is shifted to the retarded side in the second region to lower the effective compression ratio. can be reduced, and the occurrence of knocking can be effectively suppressed. In addition, since the effective compression ratio is lowered in the second region where the intake inertia effect is high due to the relatively high rotation speed, the necessary and sufficient compression ratio can be obtained while enjoying the reduction in pumping loss due to the substantial reduction in compression allowance due to the piston. A charged air amount can be obtained, and the fuel efficiency of the engine can be improved.

A control device according to a second aspect of the present invention includes a combustion chamber, an injector that injects fuel into the combustion chamber, an intake valve that opens and closes an intake port for introducing intake air into the combustion chamber, and exhaust gas from the combustion chamber. An exhaust valve that opens and closes an exhaust port for discharging the fuel, and controls an engine capable of premixed compression ignition combustion in which the fuel injected from the injector is mixed with air and burned by self-ignition. a variable valve mechanism capable of changing a valve overlap period in which the open period of the intake valve and the open period of the exhaust valve overlap; and fuel injected from the injector in a high engine load range. post-injection for injecting fuel during the compression stroke while driving the variable valve mechanism to form the valve overlap period of a predetermined amount or more so that the mixture with air undergoes premixed compression ignition combustion; a combustion control unit that causes the injector to perform a pre-injection that is included in the stroke or the first half of the compression stroke and that injects fuel at a timing earlier than the post-injection, and a part of the low speed side in the high load region is a first region. , when a part of the high-speed side in the high-load region is set as the second region, the combustion control unit controls the injector so that the start timing of the post-stage injection is later in the first region than in the second region. and controlling the valve variable mechanism so that the valve overlap period is longer in the first region than in the second region, and the combustion control unit starts the post-injection in the first region. The injector is controlled so that the timing is maintained at a predetermined timing in the middle of the compression stroke regardless of the rotational speed ( Claim 2 ).

According to the control device according to the second aspect , as in the first aspect, it is possible to both suppress abnormal combustion in the high load range and improve the emission performance. Moreover, since the start timing of the post-stage injection in the first region is held at the predetermined timing in the middle of the compression stroke regardless of the rotational speed, the fuel injection control in accordance with the rotational speed is simplified, and the pre-injection in the first region is performed. The combustion chamber can be cooled to a level at which abnormal combustion such as ignition is suppressed. In addition, even under conditions where the risk of pre-ignition occurrence is the highest, the start timing of the post-injection can be set in the middle of the compression stroke (in other words, there is no need to retard it until the end of the compression stroke), so excessive local enrichment of the air-fuel mixture can be prevented. can be avoided, and deterioration of emission performance can be effectively suppressed.

In the control device according to the second aspect, preferably, the combustion control unit advances the post-injection start timing as the rotational speed increases in a portion of the low speed side in the second region, and The injector is controlled so that the start timing of the post-stage injection is held at a predetermined timing at the beginning of the compression stroke regardless of the rotational speed in a part of the high speed side within the second region ( Claim 3 ).

In this way, when the post-injection start timing is advanced as the rotational speed increases in a part of the lower speed side of the second region, pre-ignition occurs as the rotational speed increases from the lower limit speed of the second region. The start timing of the post-injection can be set to an appropriate timing (a more advanced timing) in accordance with the rapid decrease in the risk of occurrence of , and the emission performance can be improved. In addition, in a part of the second region on the high speed side, the start timing of the post-injection is held at the predetermined timing at the beginning of the compression stroke, thereby improving emission performance and suppressing knocking, which is a concern on the high speed side. The combustion chamber can be cooled at the level obtained.

In the above configuration, more preferably, the combustion control unit controls the injector so that the injection amount ratio of the post-injection in the first region and the second region is constant regardless of the rotational speed . Item 4 ).

According to this configuration, it is possible to obtain a required level of cooling effect while simplifying the fuel injection control according to the rotational speed.

A control device according to a third aspect of the present invention includes a combustion chamber, an injector that injects fuel into the combustion chamber, an intake valve that opens and closes an intake port for introducing intake air into the combustion chamber, and exhaust gas from the combustion chamber. An exhaust valve that opens and closes an exhaust port for discharging the fuel, and controls an engine capable of premixed compression ignition combustion in which the fuel injected from the injector is mixed with air and burned by self-ignition. a variable valve mechanism capable of changing a valve overlap period in which the open period of the intake valve and the open period of the exhaust valve overlap; and fuel injected from the injector in a high engine load range. post-injection for injecting fuel during the compression stroke while driving the variable valve mechanism to form the valve overlap period of a predetermined amount or more so that the mixture with air undergoes premixed compression ignition combustion; a combustion control unit that causes the injector to perform a pre-injection that is included in the stroke or the first half of the compression stroke and that injects fuel at a timing earlier than the post-injection, and a part of the low speed side in the high load region is a first region. , when a part of the high-speed side in the high-load region is set as the second region, the combustion control unit controls the injector so that the start timing of the post-stage injection is later in the first region than in the second region. and controlling the valve variable mechanism so that the valve overlap period is longer in the first region than in the second region, and the combustion control unit controls a part of the low speed side in the first region , the valve overlap period is maintained at the first period regardless of the rotation speed, and the valve overlap period is changed to the first period as the rotation speed increases in a part of the high speed side in the first region. , wherein the variable valve mechanism is controlled so as to be shorter than ( claim 5 ).

According to the control device according to the third aspect , similarly to the first aspect, it is possible to both suppress abnormal combustion in the high load range and improve the emission performance. Moreover, since the valve overlap period is uniformly set to a large value in a part of the low speed side in the first region, a sufficient scavenging action can be exhibited under conditions where the risk of occurrence of preignition is the highest. Ignition can be effectively suppressed. Also, in a part of the high speed side in the first region, the higher the rotation speed, the shorter the valve overlap period. The overlap period can be ensured, and pre-ignition, which occurs at a lower risk on the higher rotation side, can be suppressed to a necessary level by the scavenging action during the valve overlap period.

In the control device according to the first aspect, the combustion control unit advances the start timing of the post-injection and increases the injection amount ratio of the post-injection as the rotational speed increases during operation in the first region. The injector may be controlled so as to increase ( Claim 6 ).

According to this configuration, it is possible to set an appropriate post-injection start timing and injection amount ratio in accordance with an increase or decrease in the risk of preignition occurring due to a change in rotational speed, thereby further enhancing the effect of suppressing preignition. be able to.

In the control device according to the first to third aspects, preferably, the engine includes an intake passage and an exhaust passage that communicate with the combustion chamber, and an EGR device that recirculates exhaust gas discharged to the exhaust passage to the intake passage. and the combustion control unit burns the air-fuel mixture by premixed compression ignition combustion in a medium load range in which the load is lower than that in the high load range, and controls exhaust gas recirculated to the combustion chamber through the EGR device. The EGR device is controlled so that the external EGR rate, which is a ratio, is lower in the high load range than in the medium load range ( claim 7 ).

In this way, when the external EGR rate is made relatively low in the high load region of the engine, a sufficient amount of air can be introduced into the combustion chamber in the high load region, and a sufficient amount of air corresponding to the load can be obtained. can generate high torque. On the other hand, when the external EGR rate is lowered in the high load range, the risk of abnormal combustion (pre-ignition or knocking) increases, but this abnormal combustion is sufficiently suppressed by the effects of the above-described expansion of the valve overlap period and post-injection. be done. That is, according to the above configuration, it is possible to suppress abnormal combustion while ensuring high output torque.

In the control device according to the first to third aspects, preferably, the engine includes a spark plug that ignites the air-fuel mixture in the combustion chamber, and the combustion control unit, during operation in the high load range, Part of the air-fuel mixture is burned by flame propagation from the ignition point of the spark plug, and the rest of the air-fuel mixture is burned by self-ignition, so that partial compression ignition combustion is performed near compression top dead center. The ignition plug is caused to perform spark ignition at the timing ( claim 8 ).

In this way, when part of the air-fuel mixture is burned by flame propagation triggered by spark ignition by the spark plug, the ignition timing of the air-fuel mixture can be accurately adjusted according to the timing of spark ignition. , it is possible to realize an engine with excellent practicality that exhibits stable output regardless of fluctuations in the external environment, load, and the like.

As described above, according to the control device for a premixed charge compression ignition engine of the present invention, it is possible to both suppress abnormal combustion in a high load range and improve emission performance.

1 is a system diagram schematically showing the overall configuration of a compression ignition engine according to one embodiment of the present invention; FIG. It is the figure which showed collectively the sectional view of an engine main body, and the top view of a piston. FIG. 4 is a diagram showing lift curves of an intake valve and an exhaust valve; 3 is a block diagram showing a control system of the engine; FIG. 1 is an operation map in which engine operation regions are classified according to differences in combustion modes; 4 is a graph showing a waveform of heat release rate during SPCCI combustion (partial compression ignition combustion). FIG. 6 is an enlarged operation map in which the first operation region (A1) in the map of FIG. 5 is divided according to differences in fuel injection control, etc. FIG. 4 is a flow chart for explaining control operations that are executed while the engine is running; FIG. 8 is a time chart showing a fuel injection pattern executed in a high load region (D1) in the map of FIG. 7, and (a) to (c) show injection patterns under different engine speed conditions. . FIG. 8 is a time chart showing a fuel injection pattern executed in a medium load range (D2) in the map of FIG. 7; FIG. 4 is a graph showing changes in the external EGR rate set in the high load range and the medium load range according to the engine load. FIG. 4 is a diagram for explaining the injection amount/injection timing of the pre-stage injection and the post-stage injection executed in the high load range, and FIG. , (b) is a graph showing the relationship between each injection amount ratio of the front injection and the rear injection and the engine rotation speed. FIG. 4 is a diagram for explaining the valve timing in the above high load range, where (a) is a graph showing the relationship between the valve overlap period and the engine speed, and (b) is the opening/closing timing of the intake/exhaust valves and the engine speed. It is a graph which shows the relationship with speed. FIG. 4 is a diagram for explaining a modification of the above embodiment, where (a) is a graph showing the relationship between the start timing of the post-injection and the engine speed, and (b) is a graph showing the relationship between the injection amount ratio of the post-injection and the engine speed. is a graph showing the relationship of

(1) Entire Configuration of Engine FIGS. 1 and 2 are diagrams showing a preferred embodiment of a compression ignition engine (hereinafter simply referred to as engine) to which the control device of the present invention is applied. The engine shown in this figure is a 4-cycle gasoline direct injection engine mounted on a vehicle as a power source for running. It has an exhaust passage 40 through which exhaust gas discharged from the engine body 1 flows, and an external EGR device 50 that recirculates part of the exhaust gas flowing through the exhaust passage 40 to the intake passage 30 .

The engine body 1 includes a cylinder block 3 in which a cylinder 2 is formed, a cylinder head 4 attached to the upper surface of the cylinder block 3 so as to block the cylinder 2 from above, and a reciprocatingly slidable insertion into the cylinder 2. and a piston 5 that is The engine main body 1 is typically of a multi-cylinder type having a plurality of (for example, four) cylinders, but for the sake of simplification, only one cylinder 2 will be described here.

A combustion chamber 6 is defined above the piston 5, and fuel containing gasoline as a main component is supplied to the combustion chamber 6 by injection from an injector 15, which will be described later. The supplied fuel burns while being mixed with air in the combustion chamber 6, and the piston 5 reciprocates vertically by receiving the expansion force due to the combustion.

A crankshaft 7 that is an output shaft of the engine body 1 is provided below the piston 5 . The crankshaft 7 is connected to the piston 5 via a connecting rod 8 and is rotationally driven around the central axis according to the reciprocating motion (vertical motion) of the piston 5 .

The geometric compression ratio of the cylinder 2, that is, the ratio between the volume of the combustion chamber 6 when the piston 5 is at top dead center and the volume of the combustion chamber 6 when the piston 5 is at bottom dead center, is determined by SPCCI combustion, which will be described later. As a value suitable for (partial compression ignition combustion), it is set to 14 or more and 20 or less, preferably 16 or more and 18 or less.

In the cylinder block 3, a crank angle sensor SN1 for detecting the rotation angle (crank angle) of the crankshaft 7 and the rotation speed (engine speed) of the crankshaft 7, and a cooling device that flows through the cylinder block 3 and the cylinder head 4 are provided. A water temperature sensor SN2 is provided to detect the temperature of water (engine water temperature).

The cylinder head 4 has an intake port 9 for introducing air supplied from an intake passage 30 into the combustion chamber 6, and an exhaust port 10 for introducing exhaust gas generated in the combustion chamber 6 to an exhaust passage 40. , an intake valve 11 for opening and closing the opening of the intake port 9 on the combustion chamber 6 side, and an exhaust valve 12 for opening and closing the opening of the exhaust port 10 on the combustion chamber 6 side. As shown in FIG. 2, the valve type of the engine of this embodiment is a 4-valve type consisting of 2 intake valves and 2 exhaust valves. That is, in this embodiment, two intake ports 9 and two exhaust ports 10 are opened to the combustion chamber 6 of one cylinder 2, and the intake valves 11 and the exhaust valves 12 correspond to the number of the ports 9 and 10. are provided two by one cylinder 2 .

The intake valve 11 and the exhaust valve 12 are driven to open and close in conjunction with the rotation of the crankshaft 7 by a valve mechanism including a pair of camshafts arranged in the cylinder head 4 .

A valve mechanism for the intake valve 11 incorporates an intake VVT 13 capable of changing the opening/closing timing of the intake valve 11 . Similarly, the valve mechanism for the exhaust valve 12 incorporates an exhaust VVT 14 capable of changing the opening/closing timing of the exhaust valve 12 . The intake VVT 13 (exhaust VVT 14) is a so-called phase-type variable mechanism, and changes the opening timing and closing timing of the intake valve 11 (exhaust valve 12) simultaneously and by the same amount. The combination of the intake VVT 13 and the exhaust VVT 14 corresponds to the "variable valve mechanism" of the present invention, and the intake VVT 13 corresponds to the "variable intake valve mechanism" of the present invention.

FIG. 3 is a diagram showing lift curves of the intake valve 11 and the exhaust valve 12 (IN indicates the lift curve of the intake valve 11 and EX indicates the lift curve of the exhaust valve 12). As shown in this figure, the intake valve 11 and the exhaust valve 12 may be driven so that the valve opening periods overlap across exhaust top dead center (TDC in FIG. 3). This overlapping period, that is, the period in which both the intake valve 11 and the exhaust valve 12 are open, is called a valve overlap period. The valve overlap period can be adjusted by controlling the intake VVT 13 and the exhaust VVT 14 described above.

In FIG. 3, the solid-line waveform indicates the lift curve when the valve overlap period is set relatively long, and the broken-line waveform indicates the lift curve when the valve overlap period is set relatively short. For example, in a high load region where the pressure of intake air (supercharging pressure) pressurized by the supercharger 33 (to be described later) increases, when the valve overlap period is expanded as shown by the waveform of the solid line, the overlap During the period, the flow of intake air blowing through from the intake port 9 to the exhaust port 10 is formed. As a result, the burned gas remaining in the combustion chamber 6 is promoted to be discharged to the exhaust port 10, and the so-called scavenging performance is improved. . On the other hand, when the valve overlap period is extended in a low load region where the intake pressure is low, burned gas (exhaust gas) flows from the exhaust port 10, which is still open, into the combustion chamber 6 in the first half of the intake stroke. As a result of being pulled back, the internal EGR rate, which is the proportion of the exhaust gas remaining in the combustion chamber 6, increases. As described above, in the engine of this embodiment having the intake VVT 13 and the exhaust VVT 14, the scavenging performance and the internal EGR rate can be adjusted by expanding or contracting the valve overlap period in the high load range or low load range of the engine. It has become.

As shown in FIGS. 1 and 2, the cylinder head 4 has an injector 15 for injecting fuel (gasoline) into the combustion chamber 6, and the fuel injected from the injector 15 into the combustion chamber 6 is mixed with intake air. A spark plug 16 is provided for igniting the air-fuel mixture.

As shown in FIG. 2 , a cavity 20 is formed in the crown surface of the piston 5 by recessing a relatively wide area including the central portion thereof toward the opposite side (downward) of the cylinder head 4 . A squish portion 21 made of an annular flat surface is formed radially outside the cavity 20 on the crown surface of the piston 5 .

The injector 15 is a multiple injection hole type injector having a plurality of injection holes at its tip, and is capable of radially injecting fuel from the plurality of injection holes (F in FIG. 2 indicates each injection hole). represents the spray of fuel injected from the holes). The injector 15 is arranged at the center of the ceiling surface of the combustion chamber 6 so that its tip faces the center of the crown surface of the piston 5 (the center of the bottom of the cavity 20).

The spark plug 16 is arranged at a position slightly shifted toward the intake side with respect to the injector 15 . The position of the tip portion (electrode portion) of the spark plug 16 is set so as to overlap with the cavity 20 in plan view.

As shown in FIG. 1 , the intake passage 30 is connected to one side surface of the cylinder head 4 so as to communicate with the intake port 9 . Air (fresh air) taken from the upstream end of the intake passage 30 is introduced into the combustion chamber 6 through the intake passage 30 and the intake port 9 .

The intake passage 30 includes, in order from the upstream side thereof, an air cleaner 31 that removes foreign matter from the intake air, a throttle valve 32 that can be opened and closed to adjust the flow rate of the intake air, a supercharger 33 that compresses and delivers the intake air, and a supercharger. An intercooler 35 for cooling the intake air compressed by the feeder 33 and a surge tank 36 are provided.

Each portion of the intake passage 30 is provided with an airflow sensor SN3 that detects the flow rate of intake air, an intake air temperature sensor SN4 that detects the temperature of the intake air, and an intake pressure sensor SN5 that detects the pressure of the intake air. An airflow sensor SN3 and an intake air temperature sensor SN4 are provided at a portion between the air cleaner 31 and the throttle valve 32 in the intake passage 30, and detect the flow rate and temperature of intake air passing through these portions. The intake pressure sensor SN5 is provided in the surge tank 36 and detects the pressure of intake air in the surge tank 36 .

The supercharger 33 is a mechanical supercharger (supercharger) mechanically linked to the engine body 1 . The specific type of the supercharger 33 is not particularly limited, but any known supercharger such as Lysholm type, Roots type, or centrifugal type can be used as the supercharger 33 .

Between the supercharger 33 and the engine body 1, an electromagnetic clutch 34 is interposed which can be electrically switched between engagement and disengagement. When the electromagnetic clutch 34 is engaged, driving force is transmitted from the engine body 1 to the supercharger 33, and supercharging by the supercharger 33 is performed. On the other hand, when the electromagnetic clutch 34 is released, the transmission of the driving force is interrupted and the supercharging by the supercharger 33 is stopped.

A bypass passage 38 for bypassing the supercharger 33 is provided in the intake passage 30 . The bypass passage 38 connects the surge tank 36 and an EGR passage 51 (to be described later) to each other. A bypass valve 39 that can be opened and closed is provided in the bypass passage 38 .

The exhaust passage 40 is connected to the other side surface of the cylinder head 4 (the surface opposite to the intake passage 30) so as to communicate with the exhaust port 10. As shown in FIG. Burned gas generated in the combustion chamber 6 is discharged to the outside through the exhaust port 10 and the exhaust passage 40 .

A catalytic converter 41 is provided in the exhaust passage 40 . The catalytic converter 41 includes a three-way catalyst 41a for purifying harmful components (HC, CO, NOx) contained in the exhaust gas flowing through the exhaust passage 40, and particulate matter (PM) contained in the exhaust gas. A GPF (gasoline particulate filter) 41b for collecting is incorporated.

An A/F sensor SN6 that detects the oxygen concentration in the exhaust gas is provided upstream of the catalytic converter 41 in the exhaust passage 40 .

The external EGR device 50 has an EGR passage 51 connecting the exhaust passage 40 and the intake passage 30 , and an EGR cooler 52 and an EGR valve 53 provided in the EGR passage 51 . The EGR passage 51 connects a portion of the exhaust passage 40 downstream of the catalytic converter 41 and a portion of the intake passage 30 between the throttle valve 32 and the supercharger 33 . The EGR cooler 52 cools the exhaust gas (external EGR gas) recirculated from the exhaust passage 40 to the intake passage 30 through the EGR passage 51 by heat exchange. The EGR valve 53 is provided in the EGR passage 51 downstream (closer to the intake passage 30 ) than the EGR cooler 52 so as to be openable and closable, and adjusts the flow rate of the exhaust gas flowing through the EGR passage 51 .

(2) Control System FIG. 4 is a block diagram showing the engine control system. A PCM 100 shown in this figure is a microprocessor for overall control of the engine and the like, and is composed of a well-known CPU, ROM, RAM and the like.

Detection signals from various sensors are input to the PCM 100 . For example, the PCM 100 is electrically connected to the crank angle sensor SN1, the water temperature sensor SN2, the airflow sensor SN3, the intake air temperature sensor SN4, the intake pressure sensor SN5, and the A/F sensor SN6. Information (that is, crank angle, engine rotation speed, engine water temperature, intake flow rate, intake air temperature, intake pressure, exhaust oxygen concentration) is sequentially input to the PCM 100 .

Further, the vehicle is equipped with an accelerator sensor SN7 for detecting the opening of an accelerator pedal operated by the driver of the vehicle (hereinafter referred to as accelerator opening), and a traveling speed of the vehicle (hereinafter referred to as vehicle speed). A vehicle speed sensor SN8 is provided, and detection signals from these sensors SN7 and SN8 are also sequentially input to the PCM 100. FIG.

The PCM 100 controls each part of the engine while executing various determinations and calculations based on the input information from each sensor. That is, the PCM 100 is electrically connected to the intake/exhaust VVT 13, 14, the injector 15, the spark plug 16, the throttle valve 32, the electromagnetic clutch 34, the bypass valve 39, the EGR valve 53, etc. Control signals are output to each of these devices based on the above.

Specifically, PCM 100 functionally includes determination unit 101 , combustion control unit 102 , and storage unit 103 .

The combustion control unit 102 is a control module that controls the combustion of the air-fuel mixture in the combustion chamber 6, and controls each part of the engine so that the output torque of the engine and the like are appropriate values according to the driver's request. The determination unit 101 is a control module for making various determinations necessary to determine the details of control by the combustion control unit 102 . The storage unit 103 stores various data necessary for processing in the determination unit 101 and the combustion control unit 102 .

(3) Control according to operating state FIG. 5 is an operation map for explaining differences in control according to engine speed/load. As shown in the figure, the operating range of the engine is roughly divided into four operating ranges A1 to A4 depending on the difference in combustion mode. Assuming a first operating area A1, a second operating area A2, a third operating area A3, and a fourth operating area A4, the third operating area A3 is an extremely low speed area in which the engine speed is less than the first speed N1. , the fourth operating region A4 is a high speed region where the engine speed is equal to or higher than the second speed N2, and the first operating region A1 is a speed region other than the third and fourth operating regions A3 and A4 (low/medium speed area), the second operating area A2 is the remaining area other than the first, third, and fourth operating areas A1, A3, and A4. (low-to-medium speed/low load area). A reference load L1 positioned at the boundary between the first operating region A1 and the second operating region A2 corresponds to the lower limit load (supercharging line) at which the supercharger 33 is driven.

An outline of combustion control in the first to fourth operating regions A1 to A4 will be described below.

(3-1) First Operating Region In the first operating region A1 of low to medium speed and high load, partial compression ignition combustion (hereinafter referred to as SPCCI combustion) combining SI combustion and CI combustion is performed. SI combustion is a combustion mode in which the air-fuel mixture is ignited by a spark generated from the spark plug 16, and the air-fuel mixture is forcibly burned by flame propagation that spreads the combustion area from the ignition point to the surroundings. CI combustion is a combustion mode in which an air-fuel mixture is combusted by self-ignition under an environment of sufficiently high temperature and high pressure due to compression of the piston 5 or the like. SPCCI combustion, which is a combination of SI combustion and CI combustion, involves SI combustion of part of the air-fuel mixture in the combustion chamber 6 by spark ignition performed in an environment just before the air-fuel mixture self-ignites, and the SI It is a combustion mode in which, after combustion, other air-fuel mixture in the combustion chamber 6 is subjected to CI combustion by self-ignition (due to further increase in temperature and pressure accompanying SI combustion). “SPCCI” is an abbreviation for “Spark Controlled Compression Ignition”.

FIG. 6 is a graph showing changes in the heat release rate (J/deg) depending on the combustion waveform, that is, the crank angle, when the SPCCI combustion is performed as described above. As shown in this figure, in SPCCI combustion, heat generation due to SI combustion and heat generation due to CI combustion occur consecutively in this order. At this time, due to the fact that the combustion speed is faster in CI combustion, the rise of heat release is steeper in CI combustion than in SI combustion. Therefore, the waveform of the heat release rate in SPCCI combustion has an inflection point X that appears at the timing (θci described later) at which SI combustion is switched to CI combustion.

As a specific form of the SPCCI combustion as described above, in the first operating region A1, control is performed to perform SPCCI combustion of the air-fuel mixture while performing supercharging by the turbocharger 33 . In order to realize such SPCCI combustion accompanied by supercharging, in the first operating region A1, the PCM 100 controls each part of the engine as follows.

The opening degree of the throttle valve 32 is adjusted to a value such that the amount of air corresponding to the theoretical air-fuel ratio is introduced into the combustion chamber 6 through the intake passage 30 . That is, in the first operating region A1, the air-fuel ratio (A/F ) is set in the vicinity of the stoichiometric air-fuel ratio (14.7), more specifically, to a value slightly richer than the stoichiometric air-fuel ratio (eg, about 12 to 14). Then, the air-fuel ratio in the combustion chamber 6 is matched with the target air-fuel ratio based on the target value of the air-fuel ratio (target air-fuel ratio) and the oxygen concentration in the exhaust gas detected by the A/F sensor SN6. The obtained opening degree of the throttle valve 32 is determined, and the throttle valve 32 is controlled according to this determination.

The supercharger 33 is turned on. That is, the first operating region A1 belongs to a region of the reference load L1 or higher, which is the lower limit of the supercharging region, and supercharging by the supercharger 33 is required. Therefore, in the first operating region A1, supercharging by the supercharger 33 is performed by engaging the electromagnetic clutch 34 and connecting the supercharger 33 and the engine body 1 . At this time, the bypass valve 39 is opened so that the pressure in the surge tank 36 (supercharging pressure) detected by the intake pressure sensor SN5 coincides with a predetermined target pressure for each engine load/rotational speed condition. degree is controlled.

Here, in SPCCI combustion in which SI combustion and CI combustion are combined, it is important to control the ratio of SI combustion and CI combustion according to operating conditions. Therefore, in this embodiment, attention is paid to the SI rate, which is the ratio of the amount of heat release due to SI combustion to the total amount of heat release due to SPCCI combustion (SI combustion and CI combustion), and the engine is adjusted so that this SI rate becomes an appropriate value. control each part of

The SI rate will be described with reference to FIG. In FIG. 6, the crank angle θci corresponding to the inflection point X at which the combustion mode switches from SI combustion to CI combustion is defined as the start timing of CI combustion. In this case, the amount of heat release due to SI combustion corresponds to the area R1 of the waveform of the heat release rate on the advanced side of θci (the start timing of CI combustion), and the amount of heat release due to CI combustion is greater than θci. It can be regarded as equivalent to the area R2 of the waveform of the heat release rate located on the retarded side. Using these areas R1 and R2, the SI rate can be defined as R1/(R1+R2).

In the first operating region A1 where SPCCI combustion is performed, each part of the engine is controlled so that the SI rate and θci described above match predetermined target values. That is, in the first operating region A1, a target SI rate, which is a target value of the SI rate, and a target θci, which is a target value of θci, are set for various conditions of different engine loads/rotational speeds. A plurality of control variables such as the injection amount/injection timing of fuel from the injector 15, the spark ignition timing (ignition timing) by the spark plug 16, and the EGR rate (external EGR rate and internal EGR rate) are the target SI rate. and the target θci are controlled to be a combination that can be realized. The external EGR rate is the weight ratio of the external EGR gas (exhaust gas recirculated to the combustion chamber 6 through the EGR passage 51) in the total gas in the combustion chamber 6. The internal EGR rate is It is the weight ratio of internal EGR gas (burned gas remaining in the combustion chamber 6 due to internal EGR) to the total gas in the combustion chamber 6 .

For example, the fuel injection amount/injection timing is determined by a predetermined map in consideration of the target SI rate and target θci. Regarding the external EGR rate and the internal EGR rate, the opening/closing timing of the intake/exhaust valves 11 and 12 and the degree of opening of the EGR valve 53, which are the main influencing factors of both EGR rates, are also dependent on the target SI rate and the target SI rate. It is determined by a map determined by considering θci.

On the other hand, the timing (ignition timing) of spark ignition by the spark plug 16 is determined by calculation using a predetermined model formula so as to obtain the target SI rate and the target θci.

As described above, in the first operating region A1, the ignition timing, the fuel injection amount/injection timing, and the intake/exhaust valves 11 and 12 are determined by a method that combines a predetermined map and calculation using a model formula. The opening/closing timing and the degree of opening of the EGR valve 53 are controlled so as to obtain a combination that provides an appropriate SI rate and θci (target SI rate and target θci) predetermined for each operating condition.

(3-2) Second operating region In the second operating region A2 of low to medium speed and low load, control is performed to cause SPCCI combustion of the air-fuel mixture in a state in which supercharging by the turbocharger 33 is stopped (state of natural intake). executed. In order to realize such SPCCI combustion under natural intake, the PCM 100 controls each part of the engine in the first operating region A1 as follows.

The opening degree of the throttle valve 32 is controlled so that the air-fuel ratio is variable depending on whether it is inside or outside a substantially rectangular boundary Q set within the second operating region A2. That is, in the region outside the boundary Q in the second operating region A2, the throttle valve 32 is adjusted so that the air-fuel ratio (A/F) of the air-fuel mixture in the combustion chamber 6 substantially matches the stoichiometric air-fuel ratio (14.7). opening is adjusted. On the other hand, in the area inside the boundary Q in the second operating area A2, the opening of the throttle valve 32 is adjusted so that the air-fuel ratio is greater than the stoichiometric air-fuel ratio, for example, about 20-35. In other words, in the region inside the boundary Q, control is executed to form an A/F-lean air-fuel mixture having an air-fuel ratio greater than the stoichiometric air-fuel ratio in the combustion chamber 6 and to subject the air-fuel mixture to SPCCI combustion. In the outer region of , control is executed to form a stoichiometric air-fuel mixture close to the stoichiometric air-fuel ratio in the combustion chamber 6 and perform SPCCI combustion of the air-fuel mixture. The boundary Q is defined by the upper limit speed (second speed N2) and the upper limit load (reference load L1), and the lower limit load (lowest load of the engine).

The supercharger 33 is turned off. That is, the electromagnetic clutch 34 is released to disconnect the supercharger 33 from the engine body 1, and the bypass valve 39 is fully opened to stop supercharging by the supercharger 33.

In the second operating range A2 as well, the target SI rate and the target θci are determined for each load/rotational speed condition, as in the case of the first operating range A1. The fuel injection amount/injection timing, the opening/closing timing of the intake/exhaust valves 11 and 12, and the opening degree of the EGR valve 53 are determined using a predetermined map to achieve the target SI rate and target θci. is determined to be the same value as As for the ignition timing by the ignition plug 16, the ignition timing that can achieve the target SI rate and the target θci is determined by calculation using a predetermined model formula.

(3-3) Third operating region and fourth operating region Third operating region A1 (very low speed region) in which the engine speed is lower than the first speed N1, and fourth operation in which the engine speed is equal to or higher than the second speed N2 In region A4 (high speed region), control is executed to burn the air-fuel mixture by SI combustion. For example, the entire amount of fuel to be injected in one cycle is injected from the injector 15 during the intake stroke, and spark ignition is performed by the spark plug 16 near the compression top dead center. This spark ignition initiates SI combustion, and all of the air-fuel mixture in the combustion chamber 6 is combusted by flame propagation.

(4) Injection control in the first operating region Next, a more specific example of control in the first operating region A1, particularly differences in fuel injection patterns according to operating points, will be described. The fuel injection pattern in the first operating region A1 where the load is relatively high among the SPCCI combustion execution regions can be broadly divided depending on in which of the two divided regions D1 and D2 shown in FIG. 7 the engine is operated. . Further, even in the divided region D1, the tendency of change in the fuel injection timing according to the rotational speed differs between the low speed side divided region D11 and the high speed side divided region D12. Hereinafter, the divided region D1 in the first operating region A1 is referred to as a high load region, the divided region D2 is referred to as a medium load region, the low speed side divided region D11 in the high load region D1 is referred to as a low speed/high load region, and the high speed side divided region. D12 is called a medium speed/high load range. In the example of FIG. 7, the medium load range D2 occupies the load range from the reference load L1 (supercharging line) which is the lower limit load of the first operating range A1 to the higher load L2, and the high load range D1 (low speed/ The high load range D11 and the medium speed/high load range D12) occupy the load range from the load L2 to the maximum load L3 in the first operating range A1. The low-speed/high-load region D11 and the medium-speed/high-load region D12 are adjacent to each other with the boundary speed Nx as the boundary. The low speed/high load region D11 corresponds to the "first region" of the present invention, and the medium speed/high load region D12 corresponds to the "second region" of the present invention.

FIG. 8 is a flowchart showing a specific control procedure performed during operation in the first operating area A1. As a premise for applying this flow chart, the engine is assumed to be in a semi-warm state or a warm state. Whether the engine is in the semi-warm/warm state is determined based on the engine water temperature detected by the water temperature sensor SN2. For example, when the detected engine water temperature is 70° C. or higher, it is determined that the engine is in the semi-warm/warm state, and the flowchart of FIG. 8 is applied.

When the control shown in this flowchart starts, the determination unit 101 of the PCM 100 determines in step S1 whether or not the current operating point of the engine is included in the high load region D1 shown in FIG. That is, the determination unit 101 determines the engine rotation speed detected by the crank angle sensor SN1 and the engine load specified from the detection value (accelerator opening) of the accelerator sensor SN7, the detection value (vehicle speed) of the vehicle speed sensor SN8, and the like. Based on this, the current operating point of the engine is specified on the operating map of FIG. 7, and it is determined whether or not the current operating point is included in the high load region D1 in the map.

If the determination in step S1 is YES and it is confirmed that the current operating point of the engine is included in the high load region D1, the combustion control unit 102 of the PCM 100 proceeds to step S2 to drive the supercharger 33. to supercharge the intake air. That is, the combustion control unit 102 engages the electromagnetic clutch 34 to connect the turbocharger 33 and the engine body 1, and adjusts the opening of the bypass valve 39 to control the boost pressure.

Next, the combustion control unit 102 proceeds to step S3 and selects the first injection pattern shown in FIG. 9 as the fuel injection pattern by the injectors 15 . FIGS. 9A to 9C show injection patterns at a plurality of operating points P1 to P3 (FIG. 7) with different rotational speeds, representing the injection patterns in the high load range D1. As shown in the figure, when the first injection pattern is selected, one fuel injection Fa is performed during the intake stroke and one fuel injection Fb is performed during the compression stroke. Hereinafter, Fa will be referred to as pre-injection and Fb will be referred to as post-injection.

More specifically, in the first injection pattern, the start timing of the pre-injection Fa is set somewhere in the intake stroke (more specifically, the middle of the intake stroke), and the start timing of the post-injection Fb is set in the early or middle of the compression stroke. is set to either Although the details will be described later in the next section (5), each start timing of the front injection Fa and the rear injection Fb is changed mainly according to the engine rotation speed. In other words, when the first injection pattern is selected, the combustion control unit 102 starts the pre-injection Fa during the intake stroke and the post-injection Fb during the early or middle period of the compression stroke, and controls the injections Fa and Fb. The injector 15 is controlled so that the start timing changes according to the engine speed. For example, the start timing of the post-injection Fb is set so as to be later on the low speed side (low speed/high load region D11) than on the high speed side (medium speed/high load region D12) (see FIG. 12A, which will be described later). ).

In this specification, the terms "early stage", "middle stage", and "late stage" of a process mean the following. That is, in this specification, each period when an arbitrary stroke such as an intake stroke or a compression stroke is divided into three equal periods is defined as "early period", "middle period", and "late period" in order from the front. For this reason, for example, the (i) early stage, (ii) middle stage, and (iii) late stage of the compression stroke are respectively defined as (i) 180 to 120° CA before top dead center of compression (BTDC) and (ii) 120 to 60° BTDC. CA, (iii) each range from BTDC60 to 0°CA.

Next, the combustion control unit 102 proceeds to step S4, and sets the injection amount and injection timing of each fuel injection (pre-injection Fa and post-injection Fb) in the first injection pattern to engine load (required torque) and rotation speed. determined based on A map M1 is referred to for determination of the injection amount/injection timing. The map M1 defines the injection amount/injection timing of each injection Fa, Fb in the first injection pattern for each engine load/rotational speed condition, and is stored in the storage unit 103 in advance. The injection amount/injection timing according to this map M1 is determined in consideration of the target SI rate and target θci explained in (3-1) above. In other words, in step S4, the injection amounts and injection timings of the injections Fa and Fb in the first injection pattern are determined to be values suitable for achieving the target SI rate and target θci. Details of the injection amounts/injection timings of the pre-injection Fa and the post-injection Fb determined in this manner are as described in (5) below.

Next, a description will be given of the control when it is determined NO in step S1, that is, when it is confirmed that the current operating point of the engine is not included in the high load region D1. In this case, the determination unit 101 proceeds to step S5 and determines whether or not the current operating point of the engine is included in the middle load range D2.

If the determination in step S4 is NO and it is confirmed that the current operating point of the engine is not included in the medium load range D2, that is, if the engine operates in an operating range (second, If it is confirmed that the engine is operating in any one of the third, fourth operating regions A2, A3, A4), the combustion control unit 102 proceeds to step S9 to determine which operating region A2, A3, A4. Check if the region includes the operating point, and execute combustion control according to the result. For example, when it is confirmed that the operating point is included in the second operating region A2, the combustion control unit 102 performs control to burn the air-fuel mixture by SPCCI combustion, as in the case of operating in the first operating region A1. . However, in the second operating region A2, unlike in the first operating region A1, the supercharger 33 is stopped (engagement of the electromagnetic clutch 34 is released). Further, when it is confirmed that the operating point is included in the third operating area A3 or the fourth operating area A4, the combustion control unit 102 performs control to burn the air-fuel mixture by SI combustion instead of SPCCI combustion.

On the other hand, when the determination in step S5 is YES and it is confirmed that the current operating point of the engine is included in the middle load range D2, the combustion control unit 102 proceeds to step S6 to drive the supercharger 33. to supercharge the intake air.

Next, the combustion control unit 102 proceeds to step S7 and selects the second injection pattern shown in FIG. 10 as the fuel injection pattern by the injectors 15 . FIG. 10 shows an injection pattern at a typical operating point P4 (FIG. 7) in the medium load range D2. As shown in this figure, when the second injection pattern is selected, one fuel injection Fc is performed during the intake stroke. The start timing of this fuel injection Fc is set, for example, in the first half of the intake stroke. Unlike the first injection pattern (FIG. 9) described above, fuel injection during the compression stroke (injection corresponding to post-injection Fb) is not executed.

Next, the combustion control unit 102 proceeds to step S8 to determine the injection amount and injection timing of the fuel injection Fc in the second injection pattern based on the load and rotation speed of the engine. A map M2 is referred to for determination of the injection amount/injection timing. The map M2 defines the injection amount/injection timing of the fuel injection Fc in the second injection pattern for each engine load/rotational speed condition, and is stored in the storage unit 103 in advance. The injection amount/injection timing according to this map M2 is determined in consideration of the target SI rate and target θci described above. In other words, in step S8, the injection amount and injection timing of fuel injection Fc in the second injection pattern are determined to be values suitable for realizing the target SI rate and target θci.

When the fuel injection pattern (injection amount/injection timing) and injection pressure are set as described above, the combustion control unit 102 proceeds to step S10 to control the intake and exhaust valves based on the load and rotation speed of the engine. The opening/closing timings (valve timings) of 11 and 12 are determined, and intake/exhaust VVTs 13 and 14 are controlled with the determined valve timings as targets. The map M3 is referred to for determining the valve timing here. The map M3 defines the valve timing for each engine load/rotational speed condition, and is stored in the storage unit 103 in advance. By using this map M3, the valve timing and overlap period (FIG. 3) are set to values suitable for achieving the target SI rate and target θci.

For example, the valve overlap period in the high load region D1 is changed mainly according to the rotation speed, and all of the values are set to a predetermined value greater than zero (for example, approximately 35° CA in crank angle) or more. Details will be explained in the next section (6), but the valve overlap period in the high load range D1 is designed to be longer in the low speed side (low speed/high load range D11) than in the high speed side (medium speed/high load range D12). is set to The valve overlap period in the medium load range D2 is also changed according to the rotation speed, but its value is generally shorter than the valve overlap period in the high load range D1.

Next, the combustion control unit 102 proceeds to step S11, determines the opening degree of the EGR valve 53 (EGR opening degree) based on the engine load and rotation speed, and operates the EGR valve 53 with the determined EGR opening degree as a target. Control. A map M4 is referred to for determining the EGR opening. The map M4 defines the EGR opening for each engine load/rotational speed condition, and is stored in the storage unit 103 in advance. By using this map M4, the EGR opening is set to a value suitable for achieving the target SI rate and target θci.

The EGR opening degree determined through the map M4 in step S11 is generally made lower as the load increases within the first operating region A1. FIG. 11 is a graph showing changes in the external EGR rate (ratio of exhaust gas recirculated to the combustion chamber 6 through the EGR passage 51) realized by such control of the opening of the EGR valve 53 according to the engine load. . As shown in the figure, in the first operating region A1, the opening degree of the EGR valve 53 is controlled so that the external EGR rate decreases as the load increases. In other words, during operation in the first operating region A1, the degree of opening of the EGR valve 53 is higher in the high load region D1 (loads L2 to L3) than in the middle load region D2 (loads L1 to L2). The external EGR rate is controlled to be low.

Next, the combustion control unit 102 proceeds to step S12, and the actual EGR rate (external EGR rate and internal EGR rate) in the combustion chamber 6 and the temperature in the combustion chamber 6 near the compression top dead center ( In-cylinder temperature) is estimated. As described above, in this embodiment, the opening/closing timing (valve timing) of the intake/exhaust valves 11 and 12 and the opening degree of the EGR valve 53 (EGR opening) are determined by the map. Even if the timing and EGR opening are controlled, the EGR rate may fluctuate due to various factors such as response delay. In addition, fluctuations in the EGR rate, together with other factors such as the outside air temperature, lead to fluctuations in the in-cylinder temperature near the compression top dead center. Therefore, the combustion control unit 102 detects values (intake flow rate, intake temperature, intake pressure, etc.) detected by various sensors such as the airflow sensor SN3, the intake air temperature sensor SN4, and the intake air pressure sensor SN5, and the valve timing and the EGR opening. Based on the set value and a predetermined model formula, the actual EGR rate (external EGR rate and internal EGR rate) in the combustion chamber 6 at the closing timing (IVC time point) of the intake valve 11 and the IVC and the in-cylinder temperature in the vicinity of the compression top dead center (compression top dead center or its vicinity) arriving immediately after . The above model formula is a model formula that takes as input elements the recent history of each parameter such as intake air flow rate, intake air temperature, intake pressure, valve timing, EGR opening, etc. are set so that the external EGR rate and the internal EGR rate of , and the in-cylinder temperature in the vicinity of the top dead center of the compression can be estimated, respectively.

Next, the combustion control unit 102 proceeds to step S13 and determines the spark ignition timing (ignition timing) by the spark plug 16 based on the EGR rate and in-cylinder temperature estimated in step S12. Specifically, the combustion control unit 102 uses a predetermined model formula to determine the ignition timing by the spark plug 16 to achieve the above-described target SI rate and target θci. The model formula is a model formula having multiple parameters as input elements including the estimated EGR rate (external EGR rate and internal EGR rate) and the in-cylinder temperature. .theta.ci is set so as to obtain the ignition timing that matches the target SI rate and the target .theta.ci as much as possible. According to this model formula, the ignition timing is determined by the combination of the estimated EGR rate and the in-cylinder temperature within a predetermined crank angle range near the top dead center of the compression stroke. The more retarded the timing is, the more retarded the timing is. Conversely, the more difficult it is to ignite the air-fuel mixture, the more advanced the timing is calculated.

Next, the combustion control unit 102 proceeds to step S14, causes the injector 15 to inject fuel, and causes the spark plug 16 to perform spark ignition. That is, the combustion control unit 102 causes the fuel to be injected according to the fuel injection pattern determined in either step S3 or S7, and the injection amount of the fuel injection (Fa, Fb, Fc, etc.) included in the same injection pattern. / Injector 15 is controlled so that the injection timing coincides with the injection amount/injection timing determined in either step S4 or S8. Further, the ignition plug 16 is controlled so that spark ignition is performed at the timing determined in step S13.

(5) Setting Example of Injection Amount/Injection Timing of Pre-Injection and Post-Injection Next, a detailed setting example of the pre-injection Fa and post-injection Fb executed in the high load region D1 of the first operating region A1 will be described. As already explained, in the high load range D1, the first injection pattern shown in FIG. 9 is selected, whereby the front injection Fa is started during the intake stroke, and the rear injection Fb is started during the early or middle period of the compression stroke. be done. FIGS. 9A to 9C show injection patterns under different rotational speed conditions, corresponding to operating points P1 to P3 shown in FIG. 7, respectively.

9(a) to 9(c), the pulse shapes representing the fuel injections Fa and Fb respectively represent the injection start timing at the leftmost (advance side) position in each pulse shape, and the width of each pulse shape. The dimension (horizontal length) represents the injection quantity. In other words, the crank angle period corresponding to the width dimension of each pulse shape does not necessarily match the fuel injection period (the period during which the injector 15 is opened). That is, the higher the engine speed, the longer the injection period for obtaining the same injection amount. Quantities are not necessarily the same. On the other hand, the pulse shapes representing the fuel injections Fa and Fb in FIGS. 9(a) to 9(c) define the injection amount solely by their width dimensions. can be treated as proportional to the width dimension of the pulse shape.

As shown in FIG. 7, among the three operating points P1 to P3 corresponding to the respective injection patterns (FIGS. 9A to 9C), the operating point P1 on the low speed side is the low speed operating point even in the high load region D1. The two operating points P2 and P3 on the high speed side belong to the high speed side region, that is, the medium speed/high load region D12 in the high load region D1. The operating points P1, P2, and P3 are set so that the rotation speed increases in this order on an equal load line with the same load.

For example, the first speed N1 that is the lower limit speed of the low speed/high load region D11 (high load region D1) is 1000 rpm, and the boundary speed Nx that is the boundary speed between the low speed/high load region D11 and the medium speed/high load region D12 is 2000 rpm, and when the second speed N2, which is the upper limit speed of the medium speed/high load region D12 (high load region D1), is 4000 rpm, the rotational speeds of the operating points P1, P2, and P3 are 1500 rpm, 2250 rpm, and 3000 rpm, respectively. be able to.

As shown in FIGS. 9(a) to (c), the start timing of the pre-injection Fa is almost the same among the operation points P1 to P3, and all of them are set in the middle of the intake stroke.

The start timing of the post-injection Fb is set in the middle of the compression stroke at the operating point P1 belonging to the low speed/high load region D11 (Fig. 9(a)). On the other hand, at the two operating points P2 and P3 belonging to the medium-speed/high-load region D12, the start timing of the post-injection Fb is set more advanced than at the operating point P1 (FIGS. 9B and 9C). ). In the example shown in the figure, the start timing of the post-injection Fb is set in the first half of the compression stroke at both the operating points P2 and P3. The timing is set at a more advanced timing (in the vicinity of the intake bottom dead center) compared to that at point P2.

The injection amount ratio of the post-injection Fb (pre-injection Fa) is the same at any of the operating points P1 to P3, and is maintained at a constant value regardless of the rotational speed. The injection amount ratio of the post-injection Fb is set to a value significantly smaller than the injection amount ratio of the pre-injection Fa.

FIG. 12 is a diagram for explaining in more detail changes in the injection amounts/injection timings of the pre-injection Fa and the post-injection Fb according to the engine rotation speed. Specifically, FIG. 12(a) is a graph showing the relationship between the engine rotation speed and the injection start timing. BTDC). FIG. 12(b) is a graph showing the relationship between the engine rotation speed and the injection amount ratio (division ratio). The injection start timing and the injection amount ratio shown in each graph change only the rotation speed along the equal load line connecting the above-described operating points P1 to P3 (Fig. 7) (that is, while the load is fixed). This is the case when Therefore, plots corresponding to the operating points P1 to P3 are shown on the waveform of the polygonal line showing the change in injection start timing/injection amount ratio in each graph.

As shown in FIG. 12(a), the start timing of the post-injection Fb is a predetermined timing ( Here, it is set to BTDC90° CA). Further, in the speed range (from the boundary speed Nx to the second speed N2) belonging to the medium speed/high load region D12, the start timing of the post-injection Fb is set to be earlier as the rotational speed increases. More specifically, the start timing of the post-injection Fb is set from the middle of the compression stroke in a part of the low speed side in the medium speed/high load region D12 (the range from the boundary speed Nx to the speed corresponding to the operating point P3). It is shifted to the advance side as the rotational speed increases so that it becomes variable over the previous period. On the other hand, in a portion of the high speed side within the medium speed/high load region D12 (range from the speed corresponding to the operating point P3 to the second speed N2), the start timing of the post-injection Fb is close to the intake bottom dead center in the compression stroke. is fixedly set at the very beginning of (here, BTDC170° CA).

The start timing of the pre-injection Fa is set in the middle of the intake stroke in any of the speed ranges belonging to the high load region D1 (from the first speed N1 to the second speed N2). Specifically, the start timing of the pre-injection Fa is appropriately retarded or advanced in accordance with changes in the rotational speed within the high load region D1, but the values are all in the middle of the intake stroke (240 to 300° BTDC). CA).

In FIG. 12(a), the waveform representing the post-injection Fb is interrupted at the second speed N2, which is the upper limit speed of the high load region D1. In the 4th operating region A4, fuel injection (injection corresponding to the post-injection Fb) is not executed during the compression stroke. On the other hand, in FIG. 12(a), a thin two-dot chain line waveform continuous with the waveform of the pre-injection Fa is shown in the speed region corresponding to the fourth operating region A4. 4 represents the start timing of the fuel injection executed during the intake stroke in the operating region A4.

As shown in FIG. 12B, the injection amount ratio of the post-injection Fb (pre-injection Fa) depends on the rotation speed in the high load range D1 (the low speed/high load range D11 and the medium speed/high load range D12). It is set to take a constant value regardless of Here, the injection amount ratio of the post-injection Fb (pre-injection Fa) is the weight ratio of the fuel injected by the post-injection Fb (pre-injection Fa) in the total fuel injected during one cycle. In the case of this embodiment, all the required fuel in one cycle is injected by the pre-injection Fa and the post-injection Fb. be. Specifically, in the high load region D1, the injection amount ratio of the front injection Fa is uniformly set to 0.9, and the injection amount ratio of the rear injection Fb is uniformly set to 0.1.

As described above, in the fourth operating region A4 on the higher speed side than the high load region D1, fuel injection during the compression stroke (injection corresponding to the post-stage injection Fb) is prohibited, and only fuel injection during the intake stroke is permitted. be done. In FIG. 12(b), in the speed region (from the second speed N2) corresponding to the fourth operating region A4, the waveform of the thin two-dot chain line continuing to the post-injection Fb is at the position of 0 on the vertical axis, and the waveform of the pre-injection Fa The fact that a continuous thin two-dot chain line waveform is shown at the position of the vertical axis 1 represents this.

(6) Setting Example of Valve Timing Next, a specific setting example of the valve timing in the high load range D1 will be described. 13A and 13B are diagrams for explaining changes in valve timing according to rotation speed. FIG. Each opening/closing timing of the valve 11 and the exhaust valve 12 is shown. Specifically, in the graph of FIG. 13(b), the dashed-dotted line indicated by EVO indicates the opening timing of the exhaust valve 12, and the dashed-dotted line indicated by EVC indicates the closing timing of the exhaust valve 12. , IVO indicates the opening timing of the intake valve 11, and the solid line graph indicated by IVC indicates the closing timing of the intake valve 11, respectively. In addition, the numerical value on the vertical axis in the graph of FIG. 13(b) represents the crank angle after top dead center (deg. ATDC) with reference to exhaust top dead center. In each graph of FIGS. 13(a) and 13(b), the opening/closing timing and the valve overlap period set at a higher speed side (speed range corresponding to the fourth operating range A4) than the high load range D1 are narrowed. It is indicated by a two-dot chain line waveform.

As shown in FIG. 13(b), the opening timing of the exhaust valve 12 (hereinafter sometimes simply referred to as EVO) is in the expansion stroke slightly advanced from the bottom dead center of the expansion (ATDC-180°CA). It is set variably according to the rotation speed in the latter half. Specifically, EVO is maintained at a predetermined timing in the latter half of the expansion stroke regardless of the rotational speed in the speed range belonging to the low speed/high load region D11 (from the first speed N1 to the boundary speed Nx). On the other hand, in the speed range belonging to the middle speed/high load region D12 (from the boundary speed Nx to the second speed N2), the EVO is shifted to the advance side more than in the low speed/high load region D11. More specifically, the EVO in the medium speed/high load region D12 advances as the rotation speed increases in a small portion of the low speed side (near the boundary speed Nx), and takes a constant value in the other speed regions. is set to

The closing timing of the exhaust valve 12 (hereinafter sometimes simply referred to as EVC) is variably set according to the rotation speed in the first half of the intake stroke, which is slightly retarded from the exhaust top dead center (ATDC 0°CA). . Specifically, the EVC is maintained at a predetermined time in the first half of the intake stroke regardless of the rotational speed in the speed range belonging to the low speed/high load region D11 (from the first speed N1 to the boundary speed Nx). On the other hand, in the speed range belonging to the middle speed/high load region D12 (from the boundary speed Nx to the second speed N2), the EVC is shifted to the advance side more than in the low speed/high load region D11. More specifically, the EVO in the medium speed/high load region D12 advances as the rotation speed increases in a small portion of the low speed side (near the boundary speed Nx), and takes a constant value in the other speed regions. is set to

In other words, from the change trends of EVO and EVC as described above, the exhaust valve 12 is positioned above the expansion bottom dead center during operation in the high load range D1 (the low speed/high load range D11 and the medium speed/high load range D12). The valve is opened before and closed after exhaust top dead center, and the operating phase in the medium speed/high load region D12 is shifted to the advance side from the operating phase in the low speed/high load region D11. driven.

The opening timing of the intake valve 11 (hereinafter sometimes simply referred to as IVO) is variably set according to the rotation speed in the latter half of the exhaust stroke, which is slightly advanced from the exhaust top dead center (ATDC 0° CA). . Specifically, IVO is held at a predetermined timing in the latter half of the exhaust stroke regardless of the rotation speed in the speed range (from the boundary speed Nx to the second speed N2) belonging to the middle speed/high load region D12. On the other hand, in the speed range belonging to the low speed/high load region D11 (from the first speed N1 to the boundary speed Nx), the IVO is shifted to the advance angle side more than in the middle speed/high load region D12. More specifically, the IVO in the low-speed/high-load region D11 is set such that it advances as the rotational speed decreases in a portion of the high speed side and takes a constant value in a portion of the low speed side.

The closing timing of the intake valve 11 (hereinafter sometimes simply referred to as IVC) is variably set according to the rotation speed in the first half of the compression stroke, which is slightly retarded from the intake bottom dead center (ATDC 180° CA). . Specifically, the IVC is maintained at a predetermined time in the first half of the compression stroke regardless of the rotation speed in the speed range (from the boundary speed Nx to the second speed N2) belonging to the medium speed/high load region D12. On the other hand, in the speed range belonging to the low speed/high load region D11 (from the first speed N1 to the boundary speed Nx), the IVC is shifted to the advance side more than in the middle speed/high load region D12. More specifically, the IVC in the low-speed/high-load region D11 is set such that it advances as the rotational speed decreases in a portion of the high speed side and takes a constant value in a portion of the low speed side.

In other words, from the change trends of IVO and IVC as described above, the intake valve 11 is positioned above exhaust top dead center during operation in the high load range D1 (the low speed/high load range D11 and the medium speed/high load range D12). The valve is opened before the intake bottom dead center and closed after the intake bottom dead center, and the operating phase in the medium speed/high load region D12 is shifted to the lag side from the operating phase in the low speed/high load region D11. driven.

From the valve timing setting shown in FIG. 13(b), the valve overlap period during which both the intake and exhaust valves 11 and 12 are open changes in the high load region D1 with the tendency shown in FIG. 13(a). . According to FIG. 13(a), the valve overlap period in the low speed/high load range D11 is longer than the valve overlap period in the medium speed/high load range D12.

Specifically, the valve overlap period is constant regardless of the rotation speed in a part of the low speed side in the low speed/high load region D11 (the range from the first speed N1 to the speed corresponding to the operating point P1). It is held for a period (here, 70° CA), and its value is maximized within the high load region D1. The maximum value of the valve overlap period set here corresponds to the "first period" in the present invention. On the other hand, a part of the high speed side in the low speed/high load region D11 (the range from the speed corresponding to the operating point P1 to the boundary speed Nx) and a part of the low speed side in the medium speed/high load region D12 (the boundary speed Nx to a slightly higher speed), the higher the rotation speed, the shorter the valve overlap period. Furthermore, in the remaining speed range within the medium speed/high load range D12 (part of the high speed side within the same range D12), the valve overlap period is a constant period (here, 37°CA) regardless of the rotational speed. and its value is the lowest within the high load region D1.

(7) Effects As described above, in the present embodiment, during operation in the high load region D1 of the first operating region A1 where the engine load is high among the execution regions of SPCCI combustion, the intake and exhaust valves 11 and 12 The intake and exhaust VVTs 13 and 14 are controlled so as to form a valve overlap period in which both valves are open, and the front injection Fa that injects fuel during the intake stroke and the rear injection Fb that injects fuel during the compression stroke. The injector 15 is controlled so that and are executed. Specifically, in the high load region D1, the start timing of the post-injection Fb is set to be later in the low speed/high load region D11 in which the rotational speed is lower than in the middle speed/high load region D12 in which the rotational speed is high (Fig. 12(a)), and the valve overlap period is set to be longer in the low speed/high load region D11 than in the medium speed/high load region D12 (see FIG. 13(a)). According to such a configuration, there is an advantage that both suppression of abnormal combustion in the high load range D1 and improvement of emission performance can be achieved.

That is, in the above-described embodiment, in the high-load region D1 of the engine in which the amount of heat release increases, in addition to the pre-injection Fa during the intake stroke, the post-injection Fb that injects fuel during the compression stroke is executed. The rise in temperature of the combustion chamber 6 during the stroke can be reduced by the latent heat of vaporization of the post-injection Fb, and the occurrence of abnormal combustion, which is a concern in the high load range D1, can be effectively suppressed.

In particular, in the low-speed/high-load region D11, which is a part of the low-speed side of the high-load region D1, compared to the medium-speed/high-load region D12 on the high-speed side, the start timing of the post-injection Fb is delayed and the valve overlap period is expanded, it is possible to obtain a sufficient temperature suppression effect that is commensurate with the risk of abnormal combustion.

That is, by setting the start timing of the post-injection Fb in the low-speed/high-load region D11 to a timing on the more retarded side of the compression stroke, the fuel by the post-injection Fb is injected at the timing when the temperature of the combustion chamber 6 rises due to compression. can act on the latent heat of vaporization, and the cooling effect of the combustion chamber 6 by the latent heat of vaporization can be enhanced. In addition, similarly, by expanding the valve overlap period in the low speed/high load region D11, a flow of intake air blowing through from the intake port 9 to the exhaust port 10 is formed during this valve overlap period. The burned gas remaining inside is promoted to be discharged to the exhaust port 10, and the scavenging action can be enhanced. This, together with the effect of the latent heat of vaporization of the fuel by the post-injection Fb described above, sufficiently suppresses the temperature rise in the combustion chamber 6 during the compression stroke. The low-temperature oxidation reaction of the air-fuel mixture can be suppressed, and pre-ignition, which is abnormal combustion in which the air-fuel mixture ignites prematurely, can be suppressed.

The low-temperature oxidation reaction is a slow oxidation reaction that occurs before a high-temperature oxidation reaction (substantial combustion reaction) that generates high thermal energy with flame. above 650° C.) can occur later in the compression stroke. According to the findings of the inventors of the present application, this low-temperature oxidation reaction is likely to occur remarkably under operating conditions such as the low-speed/high-load region D11, that is, under conditions of low rotational speed and high load. Moreover, when the low-temperature oxidation reaction occurs remarkably, the temperature of the air-fuel mixture in the combustion chamber 6 (air-fuel mixture before combustion) rises due to the reaction heat. The probability of occurrence of pre-ignition, which is a phenomenon in which the air-fuel mixture self-ignites early, increases. In contrast, in the above-described embodiment, the valve overlap period is extended and the start timing of the post-injection Fb is retarded during operation in the low-speed/high-load region D11. Due to the synergistic effect with the cooling effect of latent heat, the combustion chamber 6 can be cooled to the extent that the reaction level of the low-temperature oxidation reaction is sufficiently lowered, and abnormal combustion such as pre-ignition is induced due to the low-temperature oxidation reaction. can be effectively suppressed.

Moreover, according to the above embodiment, in which a high scavenging action can be expected, there is no need to excessively increase or retard the post-injection Fb in order to suppress preignition. can be minimized. That is, in the low-speed/high-load region D11, if the valve overlap period expansion control described above is not executed, the post-injection Fb is largely increased in order to suppress preignition, or the injection timing is changed to the compression stroke. It may be necessary to significantly retard the engine (closer to compression top dead center). However, if this is done, a locally rich air-fuel mixture tends to be easily formed in the combustion chamber 6, which may cause smoke or an increase in the amount of NOx generated. On the other hand, according to the above-described embodiment in which the valve overlap period is extended in the low speed/high load region D11, a sufficient scavenging action is exhibited during the valve overlap period. It is possible to avoid the above-mentioned problems (generation of smoke, etc.) and ensure relatively good emission performance.

On the other hand, in the medium-speed/high-load region D12 on the high-speed side of the low-speed/high-load region D11, the fluidity in the combustion chamber 6 is relatively high, so that even if the valve overlap period is reduced, sufficient scavenging action is achieved. may be obtained. In addition, the actual time from the start of compression of the air-fuel mixture by the piston 5 to the end of compression (the heat receiving time of the air-fuel mixture) is shortened, and the reaction level of the low-temperature oxidation reaction naturally decreases, so serious abnormal combustion such as pre-ignition may occur. less risk of happening. In contrast, in the above-described embodiment, the valve overlap period is shortened and the start timing of the post-injection Fb is advanced in the medium-speed/high-load region D12. While ensuring the cooling effect to suppress abnormal combustion, it is possible to avoid the formation of a locally rich air-fuel mixture and improve the emission performance.

Further, in the above embodiment, when the valve overlap period is shortened in the middle speed/high load region D12, the phase-type intake VVT 14 is driven to make the closing timing of the intake valve 12 closer to the intake bottom dead center. Since it is shifted to the retarded side (see FIG. 13(b)), the effective compression ratio of the engine, that is, the volume of the combustion chamber 6 when the piston 5 is at the top dead center and the closing timing of the intake valve 12 (in other words, As a result, the ratio of the volume of the combustion chamber 6 to the volume of the combustion chamber 6 is relatively decreased at the time when the piston 5 substantially starts compression, and as a result, it is possible to reduce the pumping loss while suppressing the knocking that is a concern in the middle speed/high load range D12. can.

That is, in the medium-speed/high-load range D12 where the rotational speed is relatively high and low-temperature oxidation reaction is unlikely to occur, the risk of pre-ignition occurrence decreases as described above, but during the combustion of the air-fuel mixture, the Knocking, which is an abnormal combustion in which unburned gas located at . Since knocking mainly depends on the environment of the combustion chamber 6 when the combustion of the air-fuel mixture starts, it is considered that the higher the temperature of the combustion chamber 6 when the combustion starts, the more likely it is to occur. In contrast, in the above-described embodiment, the operating phase of the intake valve 12 is shifted to the retarded side in the middle speed/high load region D12 to lower the effective compression ratio. The temperature of the combustion chamber 6 can be lowered, and the occurrence of knocking can be effectively suppressed. In addition, since the effective compression ratio is lowered in the medium-speed/high-load range D12 where the intake inertia effect is high due to the relatively high rotation speed, the substantial reduction in the compression allowance due to the piston 5 reduces the pumping loss. A necessary and sufficient amount of charged air can be obtained while maintaining the fuel consumption performance of the engine.

In addition, in the above-described embodiment, the start timing of the post-injection Fb in the low speed/high load region D11 is held at a predetermined timing in the middle of the compression stroke regardless of the rotation speed. It is possible to cool the combustion chamber 6 at a level at which abnormal combustion such as pre-ignition in the low speed/high load region D11 is suppressed. In addition, even under conditions where the risk of occurrence of pre-ignition is the highest, the start timing of the post-injection Fb can be set in the middle of the compression stroke (in other words, there is no need to retard until the latter half of the compression stroke), so the air-fuel mixture is excessively locally enriched. can be avoided, and deterioration of emission performance can be effectively suppressed.

Further, in the above embodiment, the start timing of the post-injection Fb increases as the rotational speed increases in a portion of the low speed side within the medium speed/high load region D12 (the range from the boundary speed Nx to the speed corresponding to the operating point P3). Since the timing is advanced, the start timing of the post-injection Fb is set at an appropriate timing ( The timing can be set to a more advanced timing), and the emission performance can be improved. Further, in a portion of the high speed side within the medium speed/high load region D12 (range from the speed corresponding to the operating point P3 to the second speed N2), the start timing of the post-injection Fb is held at a predetermined timing in the early stages of the compression stroke. Therefore, the combustion chamber 6 can be cooled to a level capable of suppressing knocking, which is a concern on the high rotation side, while improving the emission performance.

Further, in the above embodiment, the injection amount ratio of the post-injection Fb in the high load region D1 (the low speed/high load region D11 and the medium speed/high load region D12) is constant regardless of the rotational speed (see FIG. 12). (b)), it is possible to obtain a required level of cooling effect while simplifying the fuel injection control according to the rotation speed.

In the above embodiment, the valve overlap period is uniformly set to the maximum value in a portion of the low speed side within the low speed/high load region D11 (the range from the first speed N1 to the speed corresponding to the operating point P1). Therefore, a sufficient scavenging action can be exerted under conditions where the risk of occurrence of preignition is highest, and preignition can be effectively suppressed. In addition, in a part of the high speed side within the low speed/high load region D11 (the range from the speed corresponding to the operating point P1 to the boundary speed Nx), the higher the rotation speed, the shorter the valve overlap period. It is possible to secure an appropriate valve overlap period that matches the high fluidity in 6 and the scavenging action is easily obtained, and the pre-ignition that occurs at a higher rotation side has a lower risk of occurrence due to the scavenging action during the valve overlap period. It can be suppressed at the required level.

Further, in the above embodiment, the opening of the EGR valve 53 is controlled so that the external EGR rate in the high load range D1 is lower than the external EGR rate in the medium load range D2 (see FIG. 11). A sufficient amount of air can be introduced into the combustion chamber 6 in the load range D1, and sufficiently high torque corresponding to the load can be generated. On the other hand, when the external EGR rate is lowered in the high load region D1, the risk of abnormal combustion (pre-ignition or knocking) occurring increases, but this abnormal combustion is sufficiently prevented by the effects of the expansion of the valve overlap period and the post-injection Fb described above. suppressed by That is, according to the above embodiment, abnormal combustion can be suppressed while ensuring high output torque.

(8) Modification In the above-described embodiment, during operation in the low speed/high load region D11, the start timing of the post-injection Fb is held at the predetermined timing (BTDC90°CA) in the middle of the compression stroke, and the injection amount of the post-injection Fb is Although the ratio is held at a constant value (0.1), at least one of these elements may be set variably according to the rotation speed in the same region D11. For example, as shown in FIG. 14, the higher the rotational speed in the low speed/high load region D11, the more retarded the start timing of the post-injection Fb and the greater the injection amount ratio of the post-injection Fb. According to this configuration, it is possible to appropriately set the start timing and the injection amount ratio of the post-injection Fb according to the increase or decrease in the risk of preignition occurring due to the change in the rotation speed, and the effect of suppressing the preignition is further improved. can be enhanced.

In the above embodiment, when the engine is operated in the high load range D1, the pre-injection Fa is performed once during the intake stroke and the post-injection Fb is performed once during the compression stroke. It is also possible to perform Fa and post-injection Fb in a plurality of times.

In the above embodiment, the pre-injection Fa is executed during the intake stroke, but the pre-injection Fa is completed by the middle of the compression stroke (BTDC90° CA) at the latest in the range on the advance side of the start timing of the post-injection Fb. do it. In other words, the injection timing of the pre-injection can be appropriately changed as long as the pre-injection is included in the intake stroke or the first half of the compression stroke and the fuel is injected earlier than the post-injection. The first half of the compression stroke is the first half when the compression stroke is divided into the first half and the second half, and is 180° to 90° CA before top dead center (BTDC) of compression.

In the above embodiment, part of the air-fuel mixture burns (SI combustion) due to flame propagation from the ignition point of the spark plug 16, and the rest of the air-fuel mixture burns due to self-ignition (CI combustion) (SPCCI combustion). ) has been described, but the engine to which the present invention can be applied may be any engine in which at least part of the air-fuel mixture performs homogeneous charge compression ignition combustion (HCCI combustion), For example, the present invention can also be applied to an engine in which all the air-fuel mixture in the combustion chamber undergoes premixed compression ignition combustion (an engine that does not require spark ignition by a spark plug).

6 combustion chamber 11 intake valve 12 exhaust valve 13 intake VVT (valve variable mechanism, intake valve variable mechanism)
14 Exhaust VVT (variable valve mechanism)
15 Injector 16 Spark plug 50 EGR device 102 Combustion control unit D1 High load region D11 Low speed/high load region (first region)
D12 Medium speed/high load area (second area)
D2 Medium load range Fa Front injection Fb Post injection

Claims (8)

A combustion chamber, an injector that injects fuel into the combustion chamber, an intake valve that opens and closes an intake port for introducing intake air into the combustion chamber, and an exhaust valve that opens and closes an exhaust port for discharging exhaust gas from the combustion chamber. A device for controlling an engine capable of premixed compression ignition combustion in which the fuel injected from the injector is mixed with air and burned by self-ignition,
a variable valve mechanism capable of changing a valve overlap period in which the open period of the intake valve and the open period of the exhaust valve overlap;
In a high load region where the engine load is high, the variable valve mechanism is driven so that the mixture of fuel and air injected from the injector undergoes premixed compression ignition combustion, and the valve overlap period is set to a predetermined amount or more. a combustion control unit that causes the injector to perform a post-injection that injects fuel during the compression stroke and a pre-injection that is included in the intake stroke or the first half of the compression stroke and injects fuel earlier than the post-injection; with
The valve variable mechanism includes a phase-type intake valve variable mechanism that simultaneously changes the opening timing and the closing timing of the intake valve by the same amount,
When a portion of the high load region on the low speed side is defined as a first region and a portion of the high load region on the high speed side is defined as a second region, the combustion control unit sets the start timing of the post-injection to the second region. controlling the injector to be slower in the first region than the valve overlap period, and controlling the valve variable mechanism so that the valve overlap period is longer in the first region than in the second region;
The combustion control unit opens the intake valve before the exhaust top dead center and closes the intake valve after the intake bottom dead center in both the first region and the second region. The variable intake valve mechanism is controlled such that the operating phase of the intake valve in the second region is shifted to the retard side from the operating phase in the first region. controller. A combustion chamber, an injector that injects fuel into the combustion chamber, an intake valve that opens and closes an intake port for introducing intake air into the combustion chamber, and an exhaust valve that opens and closes an exhaust port for discharging exhaust gas from the combustion chamber. A device for controlling an engine capable of premixed compression ignition combustion in which the fuel injected from the injector is mixed with air and burned by self-ignition,
a variable valve mechanism capable of changing a valve overlap period in which the open period of the intake valve and the open period of the exhaust valve overlap;
In a high load region where the engine load is high, the variable valve mechanism is driven so that the mixture of fuel and air injected from the injector undergoes premixed compression ignition combustion, and the valve overlap period is set to a predetermined amount or more. a combustion control unit that causes the injector to perform a post-injection that injects fuel during the compression stroke and a pre-injection that is included in the intake stroke or the first half of the compression stroke and injects fuel earlier than the post-injection; with
When a portion of the high load region on the low speed side is defined as a first region and a portion of the high load region on the high speed side is defined as a second region, the combustion control unit sets the start timing of the post-injection to the second region. controlling the injector to be slower in the first region than the valve overlap period, and controlling the valve variable mechanism so that the valve overlap period is longer in the first region than in the second region;
The combustion control unit controls the injector so that the start timing of the post-stage injection in the first region is held at a predetermined timing in the middle of the compression stroke regardless of the rotation speed. Control device for compression ignition engines. In the control device for the homogeneous charge compression ignition engine according to claim 2 ,
The combustion control unit advances the start timing of the post-injection as the rotation speed increases in a part of the low speed side in the second region, and advances the start timing of the post-injection in a part of the high speed side in the second region. A control device for a homogenous charge compression ignition engine, characterized in that the injector is controlled so that the start timing of is maintained at a predetermined timing at the beginning of the compression stroke regardless of the rotational speed. In the control device for the homogeneous charge compression ignition engine according to claim 3 ,
The combustion control unit controls the injector so that the injection amount ratio of the post-injection in the first region and the second region is constant regardless of the rotational speed. Formula engine controller. A combustion chamber, an injector that injects fuel into the combustion chamber, an intake valve that opens and closes an intake port for introducing intake air into the combustion chamber, and an exhaust valve that opens and closes an exhaust port for discharging exhaust gas from the combustion chamber. A device for controlling an engine capable of premixed compression ignition combustion in which the fuel injected from the injector is mixed with air and burned by self-ignition,
a variable valve mechanism capable of changing a valve overlap period in which the open period of the intake valve and the open period of the exhaust valve overlap;
In a high load region where the engine load is high, the variable valve mechanism is driven so that the mixture of fuel and air injected from the injector undergoes premixed compression ignition combustion, and the valve overlap period is set to a predetermined amount or more. a combustion control unit that causes the injector to perform a post-injection that injects fuel during the compression stroke and a pre-injection that is included in the intake stroke or the first half of the compression stroke and injects fuel earlier than the post-injection; with
When a portion of the high load region on the low speed side is defined as a first region and a portion of the high load region on the high speed side is defined as a second region, the combustion control unit sets the start timing of the post-injection to the second region. controlling the injector to be slower in the first region than the valve overlap period, and controlling the valve variable mechanism so that the valve overlap period is longer in the first region than in the second region;
The combustion control unit maintains the valve overlap period at a first period irrespective of the rotational speed in a portion of the low speed side within the first region, and maintains the valve overlap period in a portion of the high speed side within the first region. and controlling the valve variable mechanism so that the valve overlap period becomes shorter than the first period as the rotational speed increases in the control device of the premixed compression ignition engine. In the control device for the homogeneous charge compression ignition engine according to claim 1 ,
The combustion control unit controls the injectors so that the higher the rotation speed, the more advanced the start timing of the post-injection and the greater the injection amount ratio of the post-injection during operation in the first region. A control device for a homogenous charge compression ignition engine. In the control device for a homogeneous charge compression ignition engine according to any one of claims 1 to 6 ,
The engine includes an intake passage and an exhaust passage that communicate with the combustion chamber, and an EGR device that recirculates exhaust gas discharged to the exhaust passage to the intake passage,
The combustion control unit burns the air-fuel mixture by premixed compression ignition combustion in a medium load range in which the load is lower than the high load range, and controls the ratio of the exhaust gas recirculated to the combustion chamber through the EGR device. A control device for a premixed compression ignition engine, characterized in that it controls the EGR device so that the EGR rate is lower in the high load range than in the medium load range. In the control device for the homogeneous charge compression ignition engine according to any one of claims 1 to 7 ,
The engine comprises a spark plug that ignites the air-fuel mixture in the combustion chamber,
The combustion control unit performs partial compression ignition combustion in which part of the air-fuel mixture burns by flame propagation from the ignition point of the spark plug and the rest of the air-fuel mixture burns by self-ignition during operation in the high-load range. A control device for a premixed compression ignition engine, characterized in that the spark plug is caused to perform spark ignition at a predetermined timing near compression top dead center so that the above is performed.

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