JP2016044671A - Control device of compression ignition type engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To improve exhaust emission performance and to prevent combustion noise while securing ignitability in compression ignition, in a compression ignition type engine.SOLUTION: A control device of a compression ignition type engine discharges a part of a combustion gas by internal EGR means to a specific intake passage (intake port 161) as a part of a plurality of intake passages connected to one cylinder 18, when an operating state of an engine body (engine 1) is in a high load region in a compression ignition combustion region, injects a predetermined amount of fuel into the specific intake port by a port injector 68, introduces a part of the combustion gas to the intake passage (intake port 162) excluding the specific intake passage by external EGR means (EGR passage 50), and supplies the remaining amount of fuel into the cylinder by a second injector (direct-injection injector 67).SELECTED DRAWING: Figure 8

Description

  The technology disclosed herein relates to a control device for a compression ignition engine.

  For example, Patent Document 1 describes an engine that performs compression ignition combustion of an air-fuel mixture in a cylinder in a compression ignition combustion region set on a low load side. In the region where the compression ignition combustion is performed, the engine introduces internal EGR gas into the cylinder to increase the temperature state in the cylinder, thereby improving the ignitability and combustion stability, The fuel efficiency is improved by making the fuel ratio leaner than the stoichiometric air fuel ratio. On the other hand, in a region where compression ignition combustion is performed in a high load region, the external EGR gas cooled in addition to the internal EGR gas is introduced into the cylinder, thereby optimizing the temperature state in the cylinder and ensuring ignitability. In addition, while avoiding rapid combustion, the exhaust gas emission ratio using the three-way catalyst is improved by setting the air-fuel ratio of the air-fuel mixture to the stoichiometric air-fuel ratio.

  Patent Document 2 describes an engine including a direct injection injector and a port injector that injects fuel into an intake port. In this engine, in the low load operation region, fuel is injected into the intake port only from the port injector, while in the high load operation region, fuel is injected from both the port injector and the direct injection injector.

JP 2014-51928 A JP 2012-163028 A

  The engine described in Patent Document 1 includes a so-called direct injection injector that directly injects fuel into a cylinder, and compresses ignition (CI) or controls an air-fuel mixture formed in the cylinder. Combustion is performed by self-ignition (Controlled Auto Ignition: CAI), but the so-called single-stage injection, in which fuel is injected into the cylinder by the direct injection injector in a single injection, has a low local equivalent ratio in the cylinder (equivalent There is a problem that a large amount of air-fuel mixture (with a ratio φ of less than 1) is formed, the ignitability of compression ignition is lowered, the combustion temperature is lowered, and HC and CO are increased. In contrast, so-called multi-stage injection, in which fuel injection into a cylinder is divided into multiple times, has a mixture distribution to a region with a high local equivalent ratio such that the equivalent ratio φ exceeds 1, compared to single-stage injection. There is an advantage that the ignitability of compression ignition can be increased. However, due to the formation of an air-fuel mixture that is distributed from a high local equivalence ratio distribution to a low distribution, soot is generated from the overconcentrated part, or the combustion temperature is high in the lean region near the stoichiometric and excessive oxygen. This causes a new problem that NOx is generated.

  In particular, when the engine operating state is in a high load region, the amount of fuel increases, so the above-described problems caused by the mixture distribution in the cylinder become more prominent, and the internal EGR gas and the external EGR gas Even if both are introduced into the cylinder and the temperature state in the cylinder is optimized, rapid combustion may occur, leading to an increase in NOx and an increase in combustion noise.

  The technology disclosed herein has been made in view of the above points, and the purpose of the technology is to improve the exhaust emission performance and reduce the combustion noise in the compression ignition type engine while ensuring the ignitability of the compression ignition. To avoid it.

  The technology disclosed herein relates to a control device for a compression ignition engine, and the device includes a cylinder to which each of a plurality of intake passages is connected, and in a predetermined compression ignition combustion region, An engine main body configured to compress and ignite the air-fuel mixture, an intake valve provided in each of the plurality of intake passages and configured to open and close the intake passage, and connected to the one cylinder A port injector disposed in a specific intake passage that is a part of the plurality of intake passages and configured to inject fuel into the specific intake passage, and configured to supply fuel into the cylinder By performing the second injector, the preceding valve that opens the intake valve during the period from the expansion stroke to the exhaust stroke, and the main valve that opens at least during the intake stroke, a part of the combustion gas is Part of the combustion gas through an internal EGR means configured to be introduced into the cylinder together with fresh air through a constant intake passage and an EGR passage connected at least to an intake passage other than the specific intake passage And an external EGR means configured to be introduced into the cylinder, and a control unit configured to operate the engine body through injection control of the port injector and the second injector.

  The control unit discharges a part of the combustion gas to the specific intake passage by the internal EGR means when the operating state of the engine body is in a high load region that is equal to or higher than a predetermined load in the compression ignition combustion region. In addition, a predetermined amount of fuel is injected into the specific intake passage by the port injector, and a part of the combustion gas is introduced into the cylinder by the external EGR means through the intake passage other than the specific intake passage. At the same time, the remaining fuel amount is supplied into the cylinder by the second injector.

  Here, a plurality of intake passages are passages connected to one cylinder, and include an intake port provided in the cylinder head and an independent passage of the intake manifold continuous to the intake port.

  The specific intake path is a partial intake path among a plurality of intake paths connected to one cylinder, and the specific intake path includes one or a plurality of intake paths. Not all intake passages connected to one cylinder become specific intake passages. For example, when there are two intake passages connected to one cylinder, there is one specific intake passage. In addition, when there are three intake passages connected to one cylinder, there are one or two specific intake passages.

  The port injector may be disposed only in a specific intake path among the plurality of intake paths. Moreover, you may arrange | position to all of several intake passages. In this case, the port injector disposed in the intake passage other than the specific intake passage corresponds to the second injector. As will be described later, the second injector may be a direct injection injector that directly injects fuel into the cylinder.

  According to this configuration, when the operating state of the engine body is in a high load region that is equal to or higher than a predetermined load in the compression ignition combustion region, a part of the combustion gas is discharged into the specific intake passage by the internal EGR means. Specifically, the intake valve is opened during the period from the expansion stroke to the exhaust stroke (that is, the preceding valve opening). The combustion gas discharged to the specific intake passage by the internal EGR means is relatively high temperature.

  The control unit also injects a predetermined amount of fuel from the port injector into the specific intake passage from which the combustion gas is discharged. Since the operating state of the engine body is equal to or greater than the predetermined load and the required fuel amount is relatively large, the fuel amount injected from the port injector is a part of the required fuel amount. The fuel injection amount is set according to the operating state of the engine body. As will be described later, since the fuel injected from the port injector is mainly used to ensure the ignitability of compression ignition combustion, the predetermined fuel amount is an amount that can ensure the ignitability and combustion. The amount is set so as not to increase the noise. By injecting fuel into the intake passage by the port injector, the air-fuel mixture introduced into the cylinder is homogenized.

  Thus, the combustion valve, fuel, and fresh air are introduced from the specific intake passage into the cylinder by main opening of the intake valve during the intake stroke. At this time, the lift amount of the intake valve may be adjusted so that the equivalence ratio of the air-fuel mixture introduced from the specific intake passage into the cylinder is in the range of 1.0 to 1.4. The lift amount of the intake valve can be realized by adopting a known mechanical or hydraulic variable drive mechanism as a drive mechanism of the intake valve. In this way, it is possible to form a homogeneous and high-temperature mixture layer in the cylinder.

  On the other hand, a part of the combustion gas is introduced into the intake passages other than the specific intake passage by the external EGR means. Since this combustion gas is introduced into the intake passage via the EGR passage, the temperature is lower than that of the combustion gas produced by the internal EGR means. Note that an air-cooled or water-cooled cooler may be provided in the EGR passage to positively lower the temperature of the combustion gas introduced into the intake passage other than the specific intake passage. Thus, relatively low-temperature combustion gas and fresh air are introduced into the cylinder from the intake passage other than the specific intake passage.

  Further, the remaining combustion injected from the second injector is introduced into the cylinder. If the second injector is a port injector, fuel is introduced into the cylinder from an intake passage other than the specific intake passage. If the second injector is a direct injection injector, fuel is injected into the cylinder. In any configuration, it is possible to form an air-fuel mixture layer containing relatively low-temperature combustion gas in the cylinder. This air-fuel mixture layer is also preferably homogeneous in order to ensure exhaust emission performance.

  Thus, in order to realize temperature stratification of the air-fuel mixture in the cylinder, the homogeneous and high-temperature air-fuel mixture formed in the cylinder first self-ignites near the compression top dead center, receives the self-ignition, In particular, the air-fuel mixture layer containing low-temperature combustion gas is sequentially ignited. As a result, combustion slows down, suppressing an increase in combustion temperature, suppressing generation of RawNOx, and avoiding an increase in combustion noise.

  The controller may reduce the amount of fuel injected by the port injector as the load on the engine body increases.

  As the engine load increases, the temperature state in the cylinder increases, and the ignitability of compression ignition combustion increases. Therefore, as the load on the engine body increases, the amount of fuel injected by the port injector is reduced, so that the ignitability is ensured, while the combustion of the air-fuel mixture layer that is first subjected to compression ignition becomes abrupt. It is avoided that the pressure rise in the cylinder becomes steep. That is, it is possible to avoid an increase in combustion noise when the load on the engine body is high.

  The plurality of intake passages including the specific intake passage may be configured not to generate a swirl in the cylinder by not giving angular momentum to the gas flowing into the cylinder.

  For example, the intake port may be a so-called straight port. By doing so, the gas flow in the cylinder, particularly the flow in the rotational direction around the cylinder axis (that is, swirl) is weakened. Thereby, as described above, the air-fuel mixture layer formed by the gas introduced into the cylinder from the specific intake passage, and the air-fuel mixture layer formed by the gas introduced into the cylinder from the intake passage other than the specific intake passage Therefore, the temperature stratification of the air-fuel mixture in the cylinder can be maintained without being stirred.

  The second injector may be a direct injection injector that injects fuel into the cylinder, and the control unit may start fuel injection by the direct injection injector in the latter half of the compression stroke. Here, the second half of the compression stroke may be the second half when the compression stroke is divided into the first half and the second half.

  By doing so, the timing at which fuel is injected into the cylinder from the direct injection injector becomes relatively late, so that the timing at which the air-fuel mixture containing low-temperature combustion gas ignites can be delayed. On the other hand, by injecting fuel into a cylinder in a state where the pressure in the cylinder is increased, the period from fuel injection to self-ignition becomes relatively short. As a result, after the homogeneous and high-temperature air-fuel mixture layer has ignited, the air-fuel mixture layer containing the low-temperature combustion gas can be reliably ignited sequentially. Thus, when the load on the engine body is high, the compression ignition combustion is slowed down.

  As described above, according to the control apparatus for a compression ignition type engine, when the operating state of the engine body is in the high load region in the compression ignition combustion region, the combustion gas is generated in the specific intake passage by the internal EGR means. And a predetermined amount of fuel is injected by the port injector, a part of the combustion gas is introduced into the intake passage other than the specific intake passage by the external EGR means, and the remaining fuel amount is injected by the second injector. Is introduced into the cylinder, so that a homogeneous and high-temperature mixture layer and a mixture layer containing low-temperature combustion gas can be formed in the cylinder, and the compression ignition combustion is slow while ensuring ignitability. Therefore, it is possible to improve exhaust emission performance and avoid an increase in combustion noise.

It is the schematic which shows the structure of an engine. It is a block diagram concerning control of an engine. It is sectional drawing which shows schematically the structure of the drive mechanism of an intake valve. It is a figure which illustrates the lift curve of an intake valve. It is a figure which illustrates the operation control map of an engine. It is explanatory drawing which illustrates the formation process of air-fuel | gap when an engine load is low. It is a figure which compares the distribution frequency of the local equivalence ratio of the air-fuel | gaseous mixture formed in a cylinder resulting from the difference in the control which concerns on fuel injection. It is explanatory drawing which illustrates the formation process of air-fuel | gap when an engine load is high. It is explanatory drawing which illustrates the formation process of the air-fuel | gaseous mixture when an engine load is high different from FIG.

  Hereinafter, an embodiment of a control device for a compression ignition engine will be described with reference to the drawings. The following description of preferred embodiments is exemplary.

(Entire engine configuration)
1 and 2 show a schematic configuration of an engine (engine body) 1. The engine 1 is a gasoline engine that is mounted on a vehicle and supplied with a fuel containing at least gasoline. The engine 1 includes a cylinder block 11 provided with a plurality of cylinders 18 (only one cylinder is shown in FIG. 1, but four cylinders are provided in series, for example), and the cylinder block 11 is arranged on the cylinder block 11. The cylinder head 12 is provided, and an oil pan 13 is provided below the cylinder block 11 and stores lubricating oil. A piston 14 connected to the crankshaft 15 via a connecting rod 142 is fitted in each cylinder 18 so as to be able to reciprocate. On the upper surface of the piston 14, a cavity 141 like a reentrant type in a diesel engine is formed. The cavity 141 faces a direct injection injector 67 described later when the piston 14 is positioned near the compression top dead center. The cylinder head 12, the cylinder 18 and the piston 14 having the cavity 141 define a combustion chamber. The shape of the combustion chamber is not limited to the shape illustrated. For example, the shape of the cavity 141, the upper surface shape of the piston 14, the shape of the ceiling portion of the combustion chamber, and the like can be changed as appropriate.

  The gasoline engine 1 is set to a relatively high geometric compression ratio of 15 or more for the purpose of improving the theoretical thermal efficiency and stabilizing combustion by compression autoignition described later. The geometric compression ratio may be set as appropriate in the range of about 15 to 20, and may be 18, for example.

  The cylinder head 12 is formed with an intake port 16 and an exhaust port 17 for each cylinder 18. The intake port 16 and the exhaust port 17 include an intake valve 21 and an exhaust valve that open and close the opening on the combustion chamber side. 22 are arranged respectively. As schematically shown in FIG. 6 and the like, this engine 1 has two intake ports 161 and 162 and exhaust ports 171 and 172 connected to one cylinder 18. Accordingly, two intake valves 21 and two exhaust valves 22 are provided for each cylinder 18. Hereinafter, when the two intake ports are collectively referred to without being distinguished, reference numeral 16 is given. When the two intake ports are distinguished, reference numerals 161 and 162 are given.

  Among the valve systems that drive the intake valve 21 and the exhaust valve 22 respectively, a phase variable mechanism (hereinafter referred to as VVT (Variable ValveTiming) that can change the rotation phase of the intake camshaft with respect to the crankshaft 15 is provided on the intake side. 71) and a hydraulic variable drive mechanism 72 capable of changing the opening mode of the intake valve 21 are provided. The VVT 71 may adopt a known hydraulic or electric structure as appropriate, and the detailed structure is not shown. Details of the variable drive mechanism 72 will be described later. The drive mechanism of the intake valve 21 includes a VVT 71 and a variable drive mechanism 72.

  On the other hand, a VVT 73 capable of changing the rotation phase of the exhaust camshaft with respect to the crankshaft 15 is provided on the exhaust side. The VVT 73 on the exhaust side may adopt a hydraulic or electric known structure as appropriate, and the detailed structure is not shown. The drive mechanism of the exhaust valve 22 includes a VVT 73.

  FIG. 3 shows the configuration of the variable drive mechanism 72 of the intake valve 21. The variable drive mechanism 72 includes a camshaft 721, a cam 722 provided on the camshaft 721, a tappet 723 on which the cam 722 slides, a pump unit 724 coupled to the tappet 723, and a hydraulic pressure communicating with the pump unit 724. And a discharge path 725. The variable drive mechanism 72 is configured to be able to drive two intake valves 21 provided in one cylinder 18 independently of each other.

  The cam 722 for driving the intake valve 21 has two cam lobes 7221 and 7222 in this configuration example. The cam 722 pushes the tappet 723 down twice by two cam lobes 7221, 7222 during one revolution, in other words, during one combustion cycle of intake, compression, expansion and exhaust.

  The pump unit 724 includes a cylinder 7241 filled with hydraulic oil, and a plunger 7242 inserted in the cylinder 7241 and capable of reciprocating in the cylinder 7241. The plunger 7242 is also coupled to the tappet 723, and the plunger 7242 and the tappet 723 are biased toward the cam 722 by the spring 726. When the tappet 723 is pushed down by the cam lobes 7221 and 7222, the plunger 7242 moves down in the cylinder 7241, and the pressure of the hydraulic oil increases. The hydraulic oil pressure rises as the crank angle progresses and then falls to correspond to the profile of the cam 722. As described above, since the cam 722 has the two cam lobes 7221 and 7222, the hydraulic oil pressure rises twice during one combustion cycle.

  The hydraulic pressure discharge path 725 communicates with the cylinder 7241 of the pump unit 724, and a solenoid valve 7251 is interposed in the middle of the hydraulic pressure discharge path 725. The solenoid valve 7251 is a flow rate adjustment valve controlled by the PCM 10 to be described later, and its opening degree can be arbitrarily set between fully closed and fully open. When the solenoid valve 7251 is fully opened, the hydraulic oil pressure is not substantially increased because the hydraulic oil is discharged through the hydraulic pressure discharge passage 725 even when the plunger 7242 is lowered as described above. On the other hand, when the solenoid valve 7251 is fully closed, the hydraulic oil pressure is increased because the hydraulic oil is not discharged through the hydraulic pressure discharge path 725. Further, as will be described in detail later, the hydraulic oil pressure can be adjusted by adjusting the opening of the solenoid valve 7251 in accordance with the driving of the cam 722.

  The cylinder 7241 of the pump unit 724 is also in communication with a chamber 727, and a piston 728 connected to the upper end of the stem of the intake valve 21 is disposed in the chamber 727. The piston 728 is disposed so as to be able to reciprocate in the chamber 727. When the hydraulic oil whose pressure has been increased by the pump unit 724 is supplied, the piston 728 descends in the chamber 727 and is urged in the valve closing direction by the spring 729. The intake valve 21 is pushed down to open the intake valve 21.

  FIG. 4 shows an example of a lift curve that the intake valve 21 can take. The solid lines L1 and L2 in FIG. 4 correspond to the lift curve of the intake valve 21 when the solenoid valve 7251 is normally fully closed. If the solenoid valve 7251 is fully closed, the hydraulic fluid boosted by the two cam lobes 7221 and 7222 is supplied to the chamber 727 as it is, so that the intake valve 21 opens twice as the crank angle advances. It will be.

  The fact that the intake valve 21 opens twice during one combustion cycle is utilized in the control of the internal EGR in the engine 1. In the engine 1, “internal EGR” means that the combustion gas discharged from the cylinder 18 to the intake port 16 side is reintroduced into the cylinder 18. That is, the intake valve 21 opens in the period from the expansion stroke to the exhaust stroke (that is, the preceding valve opening L1) and at least also in the intake stroke (that is, the main valve opening L2). A part of the combustion gas in the cylinder 18 is discharged to the intake port 16 side at the time of the prior valve opening, and the combustion gas and fresh air are introduced into the cylinder 18 at the time of the main valve opening.

  Further, if the solenoid valve 7251 is fully opened while the cam lobe 7221 pushes down the tappet 723 (in other words, while the cam lobe 7221 is operating) of the two cam lobes 7221 and 7222, the pressurized hydraulic oil is increased. Is not supplied to the chamber 727, the intake valve 21 does not open. Thereafter, when the solenoid valve 7251 is fully closed while the cam lobe 7222 pushes down the tappet 723 (in other words, while the cam lobe 7222 is operating), the intake valve 21 is opened as described above ( That is, the main valve opening L2). Thus, if the cam lobe 7221 is not substantially functioned and only the cam lobe 7222 is functioned, the intake valve 21 can be opened only once during one combustion cycle. This is used during normal engine operation without internal EGR control. Therefore, the variable drive mechanism 72 of the intake valve 21 can switch execution / non-execution of the internal EGR control.

  Further, by controlling the opening and closing of the solenoid valve 7251 at the time of the preceding valve opening, the lift amount and the valve opening timing of the preceding valve opening are changed, and the internal EGR amount introduced into the cylinder 18 can be adjusted. Become. Specifically, if the solenoid valve 7251 is fully opened at the beginning of the operation of the cam lobe 7221, the intake valve 21 can be maintained in the closed state, while the cam lobe 7221 is in operation (more precisely, By closing the solenoid valve 7251 (until the nose of the cam lobe 7221), the pressure of the hydraulic oil is started and the piston 728 is pushed, so the intake valve 21 is opened. Thus, the valve opening timing of the intake valve 21 can be delayed and the lift amount can be reduced (see L11 in FIG. 4). By further delaying the switching timing from opening to closing of the solenoid valve 7251, the opening timing of the intake valve 21 can be further delayed and the lift amount can be further reduced (L12 in FIG. 4). reference).

  Furthermore, when the operation of the cam lobe 7221 is started, if the solenoid valve 7251 is fully closed, the intake valve 21 opens as the hydraulic oil pressure increases, while the cam lobe 7221 is in operation. By opening the valve 7251, the pressure of the hydraulic oil is reduced, so that the intake valve 21 is closed by the urging force of the spring 729. Thus, the closing timing of the intake valve 21 can be advanced, and the lift amount and the operating angle of the intake valve 21 are reduced (see L13 in FIG. 4). The closing timing and the lift amount of the intake valve 21 are respectively changed according to the switching timing of the solenoid valve 7251 from closing to opening. For example, comparing L12 and L13 in FIG. 4 is equivalent to changing the valve opening timing while the lift amount of the intake valve 21 is substantially the same. This makes it possible to change only the phase of the preceding valve opening without changing the phase of the main valve opening L2.

  Further, if the rate of change of the opening of the solenoid valve 7251 from fully closed to fully open is increased, the valve closing operation of the intake valve 21 can be quickly performed. If the rate of change in opening is lowered, the intake valve 21 can be closed slowly. The variable drive mechanism 72 can also adjust the valve closing timing of the intake valve 21.

  Similarly to the preceding valve opening, it is possible to change the lift amount and valve opening timing of the main valve by controlling the opening and closing of the solenoid valve 7251 when the main valve is opened (for example, L21 and L22 in FIG. 4). , L23).

  As described above, since the variable drive mechanism 72 can independently drive the two intake valves 21 provided for one cylinder 18, one of the intake valves 21 is opened in advance, It is possible not to open the other intake valve 21 in advance. Even when both intake valves 21 are opened in advance, the lift amount and the valve opening period of the preceding valves can be made different from each other. Furthermore, the lift amount and valve opening period of the main valve opening of the two intake valves 21 can be made different from each other.

  Further, for each cylinder 18, a direct injection injector 67 that directly injects fuel into the cylinder 18 and a port injector 68 that injects fuel into the intake port 16 are attached to the cylinder head 12.

  The direct injection injector 67 is disposed so that its nozzle hole faces the combustion chamber from the central portion of the ceiling surface of the combustion chamber. The direct injection injector 67 directly injects an amount of fuel corresponding to the operation state of the engine 1 at an injection timing set according to the operation state of the engine 1 into the combustion chamber. In this example, the direct injection injector 67 is a multi-injector type injector having a plurality of injection holes, although detailed illustration is omitted. Thus, the direct injection injector 67 injects fuel so that the fuel spray spreads radially from the center position of the combustion chamber. For example, at the timing when the piston 14 is positioned near the compression top dead center, the fuel spray injected radially from the central portion of the combustion chamber flows along the wall surface of the cavity 141 formed on the upper surface of the piston. It can be paraphrased that the cavity 141 is formed so that the fuel spray injected at the timing when the piston 14 is located near the compression top dead center is contained therein. This combination of the multi-injection type direct injection injector 67 and the cavity 141 is an advantageous configuration for shortening the mixture formation period and the combustion period after fuel injection. In addition, the direct injection injector 67 is not limited to a multi-injection type injector, and may employ an external valve opening type injector.

  As shown in FIG. 1, the port injector 68 is disposed facing the intake port 16 or an independent passage communicating with the intake port 16 and injects fuel into the intake port 16. As conceptually shown in FIGS. 6 and 8, the port injector 68 is provided only in one of the two intake ports 161 and 162. The type of the port injector 68 is not limited to a specific type, and various types of injectors can be appropriately employed.

  The fuel tank (not shown) and the direct injection injector 67 are connected to each other by a fuel supply path. A high-pressure fuel supply system 62 including a fuel pump 63 and a common rail 64 and capable of supplying fuel to the direct injection injector 67 at a relatively high fuel pressure is interposed on the fuel supply path. . The fuel pump 63 pumps fuel from the fuel tank to the common rail 64, and the common rail 64 can store the pumped fuel at a relatively high fuel pressure. When the direct injection injector 67 is opened, the fuel stored in the common rail 64 is injected from the injection port of the direct injection injector 67. Here, although not shown, the fuel pump 63 is a plunger type pump and is driven by the engine 1. The high-pressure fuel supply system 62 including the engine-driven pump can supply a fuel with a high fuel pressure of 30 MPa or more to the direct injection injector 67. The fuel pressure may be set to about 120 MPa at the maximum. The pressure of the fuel supplied to the direct injection injector 67 is changed according to the operating state of the engine 1. The high pressure fuel supply system 62 is not limited to this configuration.

  Similarly, the fuel tank (not shown) and the port injector 68 are connected to each other by a low-pressure fuel supply path. On this low pressure fuel supply path, a low pressure fuel supply system 66 for supplying fuel with a relatively low fuel pressure to the port injector 68 is interposed. Although not shown in detail, the low-pressure fuel supply system 66 includes an electric or engine-driven low-pressure fuel pump and a regulator, and is configured to supply a predetermined pressure of fuel to each port injector 68. . Since the port injector 68 injects fuel into the intake port, the pressure of the fuel supplied by the low pressure fuel supply system 66 is set lower than the pressure of the fuel supplied by the high pressure fuel supply system 62.

  A spark plug 25 for forcibly igniting the air-fuel mixture in the combustion chamber is also attached to the cylinder head 12. In this example, the spark plug 25 is disposed through the cylinder head 12 so as to extend obliquely downward from the exhaust side of the engine 1. The tip of the spark plug 25 is disposed facing the cavity 141 of the piston 14 located at the compression top dead center.

  As shown in FIG. 1, an intake passage 30 is connected to one side of the engine 1 so as to communicate with the intake port 16 of each cylinder 18. On the other hand, an exhaust passage 40 for discharging burned gas (exhaust gas) from the combustion chamber of each cylinder 18 is connected to the other side of the engine 1.

  An air cleaner 31 that filters intake air is disposed at the upstream end of the intake passage 30. A surge tank 33 is disposed near the downstream end of the intake passage 30. The intake passage 30 on the downstream side of the surge tank 33 is an independent passage branched for each cylinder 18, and the downstream end of each independent passage is connected to the intake port 16 of each cylinder 18.

  Between the air cleaner 31 and the surge tank 33 in the intake passage 30, a water-cooled intercooler / warmer 34 that cools or heats the air and a throttle valve 36 that adjusts the amount of intake air to each cylinder 18 are arranged. It is installed. An intercooler bypass passage 35 that bypasses the intercooler / warmer 34 is also connected to the intake passage 30. The intercooler bypass passage 35 is connected to an intercooler for adjusting the flow rate of air passing through the passage 35. A bypass valve 351 is provided. Adjusting the temperature of fresh air introduced into the cylinder 18 by adjusting the ratio between the passage flow rate of the intercooler bypass passage 35 and the passage flow rate of the intercooler / warmer 34 through the opening degree adjustment of the intercooler bypass valve 351. Is possible. It should be noted that the intercooler / warmer 34 and its associated members can be omitted.

  The upstream portion of the exhaust passage 40 is constituted by an exhaust manifold having an independent passage branched for each cylinder 18 and connected to the outer end of the exhaust port 17 and a collecting portion where the independent passages gather. A direct catalyst 41 and an underfoot catalyst 42 are connected downstream of the exhaust manifold in the exhaust passage 40 as exhaust purification devices for purifying harmful components in the exhaust gas. Each of the direct catalyst 41 and the underfoot catalyst 42 includes a cylindrical case and, for example, a three-way catalyst disposed in a flow path in the case. The engine 1 does not include a NOx purification catalyst.

  The intake passage 30 and the exhaust passage 40 are connected via an EGR passage 50 for returning a part of the exhaust gas to the intake passage 30. The EGR passage 50 is connected to the upstream side of the direct catalyst 41 with respect to the exhaust passage 40 as shown in FIG. On the other hand, as shown in FIGS. 6 and 8, the intake passage 30 is connected only to the intake port 162 where the port injector 68 is not provided, of the two intake ports 161 and 162. As a result, the combustion gas recirculated to the intake side via the EGR passage 50 is introduced into the cylinder 18 only through the intake port 162.

  The EGR passage 50 includes a main passage 51 in which an EGR cooler 52 for cooling the exhaust gas with engine coolant is disposed, and an EGR cooler bypass passage 53 for bypassing the EGR cooler 52. Yes. The main passage 51 is provided with an EGR valve 511 for adjusting the amount of exhaust gas recirculated to the intake passage 30, and the EGR cooler bypass passage 53 has a flow rate of exhaust gas flowing through the EGR cooler bypass passage 53. An EGR cooler bypass valve 531 for adjustment is provided.

  The engine 1 is controlled by a powertrain control module (hereinafter referred to as PCM) 10. The PCM 10 includes a microprocessor having a CPU, a memory, a counter timer group, an interface, and a path connecting these units. The PCM 10 constitutes a control unit.

As shown in FIGS. 1 and 2, detection signals from various sensors SW <b> 1 to SW <b> 16 are input to the PCM 10. The various sensors include the following sensors. That is, the air flow sensor SW1 that detects the flow rate of fresh air, the intake air temperature sensor SW2 that detects the temperature of fresh air, the downstream side of the intercooler / warmer 34, and the intercooler / warmer 34 downstream of the air cleaner 31. A second intake air temperature sensor SW3 for detecting the temperature of fresh air after passing through the EGR gas temperature sensor SW4, which is disposed in the vicinity of the connection portion of the EGR passage 50 with the intake passage 30 and detects the temperature of the external EGR gas. An intake port temperature sensor SW5 that is attached to the intake port 16 and detects the temperature of the intake air just before flowing into the cylinder 18, and an in-cylinder pressure sensor SW6 that is attached to the cylinder head 12 and detects the pressure in the cylinder 18. The exhaust passage 40 is disposed in the vicinity of the connection portion of the EGR passage 50, and the exhaust temperature and the exhaust respectively. Exhaust temperature sensor SW7 and exhaust pressure sensor SW8 for detecting a force, and is disposed on the upstream side of the direct catalyst 41, the linear O ₂ sensor SW9, direct catalyst 41 and underfoot catalyst 42 for detecting the oxygen concentration in the exhaust gas The lambda O ₂ sensor SW10 that detects the oxygen concentration in the exhaust gas, the water temperature sensor SW11 that detects the temperature of the engine coolant, the crank angle sensor SW12 that detects the rotation angle of the crankshaft 15, and the vehicle An accelerator opening sensor SW13 for detecting an accelerator opening corresponding to an operation amount of an accelerator pedal (not shown), intake-side and exhaust-side cam angle sensors SW14, SW15, and a common rail 64 of the high-pressure fuel supply system 62 are attached. Further, a fuel pressure sensor SW for detecting the fuel pressure supplied to the direct injection injector 67 16.

  The PCM 10 determines the state of the engine 1 and the vehicle by performing various calculations based on these detection signals, and according to this, the direct injection injector 67, the port injector 68, the spark plug 25, the intake-side VVT 71, and the variable Control signals are output to the actuators of the drive mechanism 72, the exhaust-side VVT 73, the high-pressure fuel supply system 62, and various valves (throttle valve 36, intercooler bypass valve 351, EGR valve 511, EGR cooler bypass valve 531). The PCM 10 estimates a combustion state including the temperature and pressure in the combustion chamber based on a preset model and the various detection signals described above, and outputs each control signal based on the combustion state. The engine 1 is operated.

(Engine operation control)
FIG. 5 shows an example of the operation control map of the engine 1. In order to improve fuel efficiency and exhaust emission performance, the engine 1 is controlled ignition (CAI) without performing ignition by the spark plug 25 in a low load region where the engine load is relatively low. To burn. In the example of FIG. 5, a region lower than the combustion switching load indicated by a solid line corresponds to a self-ignition region (CAI) in which CAI combustion is performed.

  As the load on the engine 1 increases, in CAI combustion, the combustion becomes too steep, causing problems such as combustion noise. Therefore, in the engine 1, in a high load region where the engine load is relatively high, the CAI combustion is stopped and the combustion is switched to the combustion by forced ignition (here, spark ignition (SI)) using the spark plug 25. In the example of FIG. 5, a region equal to or greater than the combustion switching load indicated by a solid line corresponds to a spark ignition region (SI) in which spark ignition combustion is performed. Thus, the engine 1 is configured to switch between the CAI mode and the SI mode in accordance with the operating state of the engine 1, in particular, the load of the engine 1. However, the boundary line for mode switching is not limited to the illustrated example.

  In the CAI mode, hot EGR gas having a relatively high temperature is introduced into the cylinder 18 in order to improve the ignitability and stability of the CAI combustion in the entire region from the low load region to the high load region. This is because the internal EGR gas is introduced into the cylinder 18. The introduction of hot EGR gas increases the compression end temperature in the cylinder 18 (that is, the temperature in the cylinder 18 when the piston 14 reaches compression top dead center), and the ignitability and stability of CAI combustion in a low load range. To increase.

  In the CAI mode, the internal EGR gas amount is adjusted while keeping the throttle valve 36 fully open, and the fresh air amount is also adjusted. This is advantageous for reducing pump loss. In the CAI mode, the internal EGR rate is changed according to the load level of the engine 1. Specifically, the internal EGR rate is increased when the load of the engine 1 is low, and the internal EGR rate is increased when the load of the engine 1 is high. Reduce EGR rate. In the low load region, the temperature in the cylinder 18 at the start of compression is increased by increasing the amount of hot EGR gas introduced, and the compression end temperature is increased accordingly. While the ignitability of compression self-ignition increases, the stability of compression self-ignition combustion increases. On the other hand, if the load on the engine 1 increases, the combustion gas temperature increases as the fuel injection amount increases, and the exhaust gas temperature increases together with the temperature state in the cylinder 18. Therefore, even if the amount of hot EGR gas introduced is reduced, the ignitability and stability of compression self-ignition can be ensured. If a large amount of exhaust gas is introduced into the cylinder 18, the specific heat ratio of the gas in the cylinder 18 becomes low, so that the compression end temperature becomes low even if the gas temperature at the start of compression is high. Therefore, the internal EGR rate may be limited to a preset maximum EGR rate.

  In this CAI region, the amount of hot EGR gas introduced corresponding to the load of the engine 1 is adjusted by adjusting the lift amount of the preceding valve opening of the intake valve 21. That is, when the EGR rate is changed from high to low as the load of the engine 1 increases, the lift amount of the preceding valve opening is reduced. Conversely, when the EGR rate is changed from low to high as the load on the engine 1 decreases, the lift amount of the preceding valve opening of the intake valve 21 is increased. As described above, this is performed by adjusting the opening / closing of the solenoid valve 7251 of the variable drive mechanism 72. In other words, when the EGR rate is changed from high to low as the load of the engine 1 increases, the timing of switching from opening to closing of the solenoid valve 7251 is delayed with respect to the start of operation of the cam lobe 7221 to advance the valve. Reduce the lift amount. Conversely, when the EGR rate is changed from low to high as the load of the engine 1 decreases, the lift amount of the preceding valve is opened by bringing the switching timing from opening to closing of the solenoid valve 7251 closer to the start of operation of the cam lobe 7221. Increase

  In the low / medium load range in the CAI mode, the excess air ratio λ of the air-fuel mixture is made larger than 1. The excess air ratio λ of the air-fuel mixture is preferably made to be lean at least 2.4. Making the air-fuel mixture lean is advantageous for improving the fuel efficiency by improving the thermal efficiency, and by making the excess air ratio λ 2.4 or more, the generation of RawNOx is suppressed. This makes it possible to ensure the exhaust emission performance in the engine 1 that does not include the NOx purification catalyst.

  In the high load range in the CAI mode, the air-fuel ratio of the air-fuel mixture is set to the stoichiometric air-fuel ratio (λ≈1). As a result, a three-way catalyst can be used, and exhaust emission performance can be ensured.

  In this high load region, the temperature state in the cylinder 18 is further increased. Therefore, if hot EGR gas is introduced into the cylinder 18 while maintaining the air-fuel ratio of the air-fuel mixture at the stoichiometric air-fuel ratio (λ≈1), the temperature state in the cylinder 18 becomes too high and pre-ignition etc. May occur, or the pressure increase (dP / dθ) in the cylinder 18 may become steep during CAI combustion, resulting in a problem of combustion noise. Therefore, in the high load region of the CAI mode, the cooled EGR gas is introduced into the cylinder 18 together with the hot EGR gas. The cooled EGR gas is basically an external EGR gas cooled by passing through the EGR cooler 52. The external EGR gas that bypasses the EGR cooler 52 may be included in the cooled EGR gas.

  Thus, in the region on the high load side of the CAI mode, the temperature state in the cylinder 18 is a combination of a small amount of hot EGR gas and a decrease in temperature and the introduction of the cooled EGR gas into the cylinder 18. Is prevented from becoming too high, which is advantageous for avoiding abnormal combustion and combustion noise. In the CAI mode, unlike the SI mode described later, there is no restriction on the EGR rate related to combustion stability. Therefore, it is possible to make the EGR rate as high as possible while setting the excess air ratio λ of the air-fuel mixture to substantially 1. Increasing the EGR rate is also advantageous for avoiding abnormal combustion and combustion noise in the high load region of the CAI mode.

  Although details will be described later, in a low load region in the CAI mode, the port injector 68 is used to inject fuel into the intake port 16. This forms a homogeneous mixture layer in the cylinder 18. Further, in the high load region in the CAI mode, the port injector 68 and the direct injection injector 67 are used to supply fuel into the cylinder 18.

  In the SI mode, the hot EGR gas is substantially zero and only the cooled EGR gas is introduced into the cylinder 18. Note that the fact that the hot EGR gas is substantially zero means that the internal EGR control is not performed. Since a part of the combustion gas remains in the cylinder 18 due to the structure, the hot EGR gas can remain in the cylinder 18 even in the SI mode.

  In the SI mode, basically, the opening of the throttle valve 36 is kept fully open, and the opening of the EGR valve 511 is changed according to the engine load. Thus, in the SI mode, the EGR rate is set to the maximum under the condition that the air-fuel ratio of the air-fuel mixture is set to the theoretical air-fuel ratio (λ≈1). This is advantageous for reducing pump loss. In addition, setting the air-fuel ratio of the air-fuel mixture to the stoichiometric air-fuel ratio enables the use of a three-way catalyst. Specifically, the EGR valve 511 gradually closes as the engine load increases, and closes at the fully open load. This is advantageous in improving controllability because the gas composition in the cylinder 18 is continuously changed when the engine load changes continuously.

  In SI combustion, if the amount of exhaust gas introduced into the cylinder 18 is too large, the combustion stability is lowered. Therefore, there is a maximum EGR rate (that is, an EGR limit) that can be set in SI combustion. As described above, since the air-fuel ratio of the air-fuel mixture is set to the stoichiometric air-fuel ratio (λ≈1), the EGR rate changes continuously according to the engine load level, and when the engine load is low in the SI mode, the fuel amount Although the EGR rate can be increased by reducing the amount of fresh air and the amount of fresh air, the EGR rate is limited by the EGR limit in the low load region of the SI mode.

  In the SI mode, fuel is directly injected into the cylinder 18 mainly using the direct injection injector 67. As described above, the engine 1 has a relatively high geometric compression ratio and may cause abnormal combustion such as pre-ignition during SI combustion. However, at a later time from the compression stroke to the expansion stroke, By injecting the fuel into the fuel 18 at a high fuel pressure, it is possible to avoid such abnormal combustion.

(Mode of mixture formation in CAI mode)
The engine 1 uses the combination of the variable drive mechanism 72 of the intake valve 21 and the port injector 68 described above in the low load region of the CAI mode to stratify the concentration of the air-fuel mixture formed in the cylinder 18. Therefore, the improvement of the ignitability of the compression ignition combustion and the improvement of the exhaust emission performance are achieved at the same time. On the other hand, in the region on the high load side in the CAI mode, the mixture of the air-fuel mixture formed in the cylinder 18 is utilized by utilizing a combination of the variable drive mechanism 72 and the external EGR of the intake valve 21, the port injector 68 and the direct injection injector 67. Concentration and temperature stratification are achieved to achieve slow combustion while maintaining ignitability. Hereinafter, this configuration will be described in detail with reference to the drawings.

  First, FIG. 6 shows an air-fuel mixture formation procedure when the operating state of the engine 1 is on the low load side in the CAI mode. Here, the low load side of the CAI mode is a low load side where no external EGR gas is introduced, and the air-fuel ratio of the air-fuel mixture is leaner than the stoichiometric air-fuel ratio as a whole in the combustion chamber. In step P61 of FIG. 6, the intake valve 21 is opened in advance by the variable drive mechanism 72. The intake valve 21 to be opened in advance is only the intake valve 21 of the first intake port 161 that performs fuel injection as described later. Thereby, a part of the combustion gas in the cylinder 18 is discharged to the first intake port 161. When the set internal EGR rate is high and a relatively large amount of combustion gas must be discharged to the intake port, the intake valves 21 of both the first and second intake ports 161 and 162 are opened in advance. You may let them. In that case, the lift amount of the preceding valve opening may be the same in the first and second intake ports 161 and 162, or the first intake port 161 may be relatively larger. The first intake port 161 may be relatively small.

  In the subsequent process P62, a predetermined amount of fuel is injected from the port injector 68 disposed in the first intake port 161. Since the load of the engine 1 is relatively low and the amount of fuel is small accordingly, it is possible to inject fuel from one port injector 68 to one intake port 161. Thus, the fuel is injected into the first intake port 161 whose temperature has been increased by the introduction of the combustion gas.

  In step P62, each of the intake valve 21 of the first intake port 161 and the intake valve 21 of the second intake port 162 is opened (that is, the main valve is opened). At this time, the intake valve 21 of the first intake port 161 is configured so that the equivalent ratio φ of the air-fuel mixture introduced from the intake port 161 into the cylinder 18 falls within the range of 1.0 to 1.4. Adjust the lift amount (decrease the lift amount). Note that when the load of the engine 1 is low, the combustion temperature is prevented from increasing even if the equivalence ratio φ of the air-fuel mixture formed in the cylinder 18 is set high. On the other hand, the intake valve 21 of the second intake port 162 does not adjust the lift amount. That is, the intake valve 21 is opened with the lift amount set in the operating state of the engine 1. As a result, a necessary amount of fresh air is introduced into the cylinder 18.

  In this way, in the process P63, it is possible to form a homogeneous and high-temperature air-fuel mixture layer 181 having an equivalent ratio φ in the vicinity of 1 (more precisely, richer than 1) in the cylinder 18. In order to realize such stratification, it is desirable to avoid swirl in the cylinder 18 as much as possible. For example, the intake port 16 is not a so-called swirl port that imparts angular momentum to the gas flowing into the cylinder 18, but has a shape that does not impart angular momentum to the gas flowing into the cylinder 18 (for example, a straight port). It is preferable to do. Also, internal EGR gas, fresh air, and fuel are introduced from the first intake port 161 into the cylinder 18, while only fresh air (or fresh air and internal air are introduced into the cylinder 18 from the second intake port 162. EGR gas) is introduced. Although the amount of fresh air introduced into the cylinder 18 from the first intake port 161 is limited, the amount of gas introduced into the cylinder 18 from the first intake port 161 by introducing the internal EGR gas, The difference from the amount of gas introduced into the cylinder 18 from the two intake ports 162 is reduced or eliminated. The occurrence of swirl in the cylinder 18 due to the difference in the amount of gas introduced from the two intake ports 161 and 162 is avoided. As a result, the air-fuel mixture can be stratified in the cylinder 18. It should be noted that tumble is allowed to occur in the cylinder 18 as long as the air-fuel mixture layer 181 is maintained. Therefore, the intake port 16 may be a tumble port as long as no angular momentum in the swirl direction is given to the gas flowing into the cylinder 18.

  Thus, in step P64, when the piston 14 reaches the vicinity of the compression top dead center, the homogeneous air-fuel mixture having an equivalent ratio φ of about 1 is compressed and ignited and burned. Since the air-fuel mixture layer 181 is rich in the equivalence ratio φ, the ignitability of compression ignition is enhanced.

  Here, FIG. 7 is a diagram for comparing the distribution frequency of the local equivalence ratio of the air-fuel mixture formed in the cylinder 18 when the fuel injection modes are different. The upper left diagram in FIG. 7 shows the distribution frequency of the local equivalence ratio when the direct injection injector 67 is used to form a mixture by performing one fuel injection into the cylinder 18. When the number of injections is one, the injected fuel is dispersed, and as a result, a locally lean air-fuel mixture having a local equivalent ratio φ smaller than 1 is formed. This lowers the ignitability of compression ignition, and as a result of the combustion temperature becoming relatively low, HC and CO can be generated and exhaust emission performance can be reduced.

  In order to improve the ignitability of compression ignition, as shown in the upper right of FIG. In the multi-stage injection, the fuel injection amount per one time is decreased, and the fuel spray injected later collides with the fuel spray injected earlier, so that the equivalence ratio is considered to be locally increased. As described above, locally forming an air-fuel mixture having a high equivalence ratio φ is advantageous in improving the ignitability of compression ignition. However, as a result of the formation of a large amount of air-fuel mixture with a high equivalence ratio φ, a locally excessively rich air-fuel mixture tends to generate soot, and the lean region near the stoichiometric region has NOx generated due to an increase in combustion temperature in an oxygen-excess atmosphere. Likely to happen.

  As described above, in the configuration using the direct injection injector 67, the ignitability of compression ignition is low and the exhaust emission performance is deteriorated. On the other hand, even if the ignitability of compression ignition is increased, the exhaust emission performance is deteriorated. I can not escape.

  On the other hand, by using the intake port injection described above and stratifying the concentration of the air-fuel mixture in the cylinder 18, as shown in the lower right of FIG. It is possible to form a homogeneous air-fuel mixture layer 181 having a rich equivalence ratio φ by the air-fuel mixture introduced into the gas, and locally a lean air-fuel mixture in the vicinity of stoichiometry or a locally excessively rich air-fuel mixture. Can be avoided. As a result, while improving the ignitability of compression ignition, the combustion temperature is optimized, HC and CO that can be generated when the combustion temperature is low, NOx that can be generated when the combustion temperature is high due to excessive oxygen, and local It is possible to suppress wrinkles that can occur when the equivalent ratio is too high. In the figure, the fresh air introduced into the cylinder 18 through the second intake port 162 is shown with an equivalence ratio of zero. Here, the equivalence ratio φ of the homogeneous air-fuel mixture is desirably 1.0 to 1.4. By doing this, it is possible to suppress the increase in the combustion temperature while improving the ignitability of the compression ignition. This is advantageous in suppressing the generation of NOx.

  FIG. 8 shows an air-fuel mixture formation procedure when the operating state of the engine 1 is on the high load side in the CAI mode, as described above. Here, the high load side of the CAI mode is the high load side for introducing the external EGR gas, and the air-fuel ratio of the air-fuel mixture is the stoichiometric air-fuel ratio as a whole of the combustion chamber. As the load on the engine 1 increases, the fuel injection amount increases, so that fuel is injected in both the port injector 68 and the direct injection injector 67.

  First, in step P81, the intake valve 21 is opened in advance by the variable drive mechanism 72 in the same manner as described above. The intake valve 21 to be opened in advance is only the intake valve 21 of the first intake port 161. Since the load on the engine 1 is high, the internal EGR rate is set to be relatively low. If necessary, the intake valve 21 of the second intake port 162 may also be opened in advance.

  In the subsequent process P82, a predetermined amount of fuel is injected from the port injector 68 disposed in the first intake port 161, and the remaining fuel is injected from the direct injection injector 67. The fuel injection amount of the first intake port 161 is the minimum injection amount necessary for compression ignition. The injection amount is set in a range where combustion noise does not increase. As a result, the amount of fuel injected into the first intake port 161 is smaller than the amount of fuel directly injected into the cylinder 18. The fuel injection amount of the first intake port 161 decreases as the load on the engine 1 increases. By doing so, an increase in combustion noise can be avoided in advance. On the other hand, since the temperature state in the cylinder increases as the load on the engine 1 increases, the ignitability of compression ignition is ensured even if the fuel injection amount is reduced. While the fuel injection amount of the first intake port 161 is reduced, the total fuel injection amount increases as the load of the engine 1 increases, so the fuel amount injected into the cylinder by the direct injection injector 67 increases. Note that the fuel injection amount (absolute amount) of the first intake port 161 may be maintained substantially constant as the load on the engine 1 increases. In this case, since the total fuel injection amount increases as the load on the engine 1 increases, the ratio of the fuel injection amount of the first intake port 161 to the total injection amount or the ratio of the direct injection injector 67 to the injection amount is As the load on the engine 1 increases, it decreases.

  The injection timing at the first intake port 161 is set before the intake valve 21 is closed. In contrast, the injection timing in the cylinder 18 is set to the second half of the compression stroke (for example, the second half when the compression stroke is divided into two equal parts, the first half and the second half). Therefore, the fuel injection by the direct injector 67 is actually performed at the timing from the process P82 to the process P83 described later. Thereby, it is avoided that the air-fuel mixture in the air-fuel mixture layer 182 described later ignites early. However, it is preferable that the homogeneous air-fuel mixture layer 182 is formed by the fuel injected by the direct injection injector 67 in the latter half of the compression stroke. In order to further homogenize the air-fuel mixture layer 182, the fuel injection timing by the direct injection injector 67 may be set earlier than the latter half of the compression stroke.

  In step P82, each of the intake valve 21 of the first intake port 161 and the intake valve 21 of the second intake port 162 is opened (that is, the main valve is opened). At this time, the intake valve 21 of the first intake port 161 is configured so that the equivalent ratio φ of the air-fuel mixture introduced from the intake port 161 into the cylinder 18 falls within the range of 1.0 to 1.4. Adjust the lift amount (decrease the lift amount). Considering that the load of the engine 1 is relatively high and ignitability can be secured, the equivalence ratio φ may be set lower than the above range. On the other hand, the intake valve 21 of the second intake port 162 does not adjust the lift amount. That is, the intake valve 21 is opened with the lift amount set in the operating state of the engine 1. As a result, a necessary amount of fresh air is introduced into the cylinder 18.

  In the high load region of the CAI mode, as described above, the cooled EGR gas (that is, the external EGR gas cooled by the EGR cooler 52) is introduced into the cylinder 18. The cooled EGR gas is introduced into the cylinder 18 through the second intake port 162 to which the EGR passage 50 is connected, as indicated by the white arrow in FIG. Therefore, internal EGR gas, fresh air and fuel are introduced into the cylinder 18 from the first intake port 161, and external EGR gas and fresh air are introduced into the cylinder 18 from the second intake port 162. . Thereafter, fuel injection is performed by the direct injection injector 67 as described above. By injecting fuel in the latter half of the compression stroke in a state where the pressure in the cylinder 18 is increased, the period from fuel injection to ignition can be made relatively short. When fuel is injected in the latter half of the compression stroke, it is preferable to set the fuel pressure high. The fuel injection may be single-stage injection or multistage injection.

  In the process P83, a rich, homogeneous and high-temperature air-fuel mixture layer (first layer) 181 having an equivalence ratio φ of 1 or more in the cylinder 18 and an equivalence ratio φ leaner than that of the first layer 181. It is possible to form an air-fuel mixture layer (second layer) 182 that extends uniformly in the cylinder 18 at a low temperature.

  Since the first layer 181 has a relatively high equivalence ratio and a high temperature, the first layer 181 starts to ignite by compression first. The low temperature second layer 182 is also ignited in sequence. In particular, since the second layer 182 contains a low-temperature external EGR gas, the air-fuel mixture is sequentially ignited (see process P84). Thus, when the load on the engine 1 is high, combustion can be slowed while ensuring ignitability. This suppresses the increase in the combustion temperature and suppresses the generation of RawNOx. Further, it is possible to prevent the combustion noise from being increased by suppressing the pressure rise in the cylinder 18 from becoming steep during combustion.

  FIG. 9 shows a modified example related to air-fuel mixture formation in a high load region in the CAI mode. In this modified example, a port injector 68 is provided in each of the first and second intake ports 161 and 162. This is different from the configuration described above. In the modification, when the operating state of the engine 1 is on the high load side in the CAI mode, fuel is injected from the port injector 68 disposed in the second intake port 162 instead of the direct injection injector 67. Accordingly, in the modification, fuel is injected into both the first intake port 161 and the second intake port 162.

  First, in step P91, the intake valve 21 is opened in advance by the variable drive mechanism 72 in the same manner as described above. The intake valve 21 to be opened in advance is only the intake valve 21 of the first intake port 161. Since the load on the engine 1 is high, the internal EGR rate is set to be relatively low. If necessary, the intake valve 21 of the second intake port 162 may also be opened in advance.

  In the subsequent process P92, a predetermined amount of fuel is injected from the port injector 68 provided in the first intake port 161, and the remaining fuel is supplied from the port injector 68 provided in the second intake port 162. Spray. The fuel injection amount of the first intake port 161 is the minimum injection amount necessary for compression ignition, and is set in a range where combustion noise does not increase. As the load on the engine 1 increases, the fuel injection amount of the first intake port 161 decreases. The amount of fuel injected into the first intake port 161 is smaller than the amount of fuel injected into the second intake port 162. The injection timing in the first intake port 161 and the second intake port 162 is set before the intake valve 21 is closed.

  In step P92, each of the intake valve 21 of the first intake port 161 and the intake valve 21 of the second intake port 162 is opened (that is, the main valve is opened). At this time, the intake valve 21 of the first intake port 161 is configured so that the equivalent ratio φ of the air-fuel mixture introduced from the intake port 161 into the cylinder 18 falls within the range of 1.0 to 1.4. Adjust the lift amount (decrease the lift amount). In view of the fact that the load of the engine 1 is relatively high and the ignitability can be ensured, the equivalent ratio φ may be set lower than the above range, which is the same as the configuration example of FIG. On the other hand, the intake valve 21 of the second intake port 162 does not adjust the lift amount. That is, the intake valve 21 is opened with the lift amount set in the operating state of the engine 1. As a result, a necessary amount of fresh air is introduced into the cylinder 18.

  The cooled EGR gas is introduced into the cylinder 18 through the second intake port 162 to which the EGR passage 50 is connected, as indicated by a white arrow in FIG. Therefore, internal EGR gas, fresh air and fuel are introduced into the cylinder 18 from the first intake port 161, and external EGR gas, fresh air and fuel are introduced into the cylinder 18 from the second intake port 162. Is done.

  Thus, in step P93, a rich, homogeneous and high-temperature air-fuel mixture layer (first layer) 181 having an equivalent ratio φ of 1 or more in the cylinder 18 and an equivalent ratio φ leaner than that of the first layer 181. It becomes possible to form a homogeneous and low-temperature air-fuel mixture layer (second layer) 183.

  The first layer 181 having a relatively high equivalence ratio and a high temperature is first subjected to compression ignition, and is subjected to a rise in heat and in-cylinder pressure generated thereby, and a combustion at a relatively low equivalence ratio and a low temperature. The air-fuel mixture of the second layer 183 containing gas sequentially undergoes compression ignition. Thus, when the load on the engine 1 is high, the temperature of the air-fuel mixture in the cylinder 18 is stratified so that the combustion is slowed while ensuring the ignitability. This suppresses generation of RawNOx and avoids an increase in combustion noise.

  The technique disclosed here is not limited to application to the engine configuration described above.

  The drive mechanism of the intake valve 21 is not limited to the configuration including the VVT 71 and the variable drive mechanism 72. Various mechanisms can be employed as long as the lift amounts of the plurality of intake valves 21 provided in one cylinder 18 can be adjusted independently. For example, the drive mechanism of the intake valve 21 may include a mechanical variable valve lift (CVVL) capable of continuously changing the lift amount of the intake valve and a VVT.

  In the above configuration, the EGR passage 50 is connected to the second intake port 162 of each cylinder 18, but external EGR gas is introduced into each of the first intake port 161 and the second intake port 162. Thus, you may connect the EGR channel | path 50 to the upstream position of the surge tank 33, for example. Even with this configuration, it is possible to form an air-fuel mixture layer having a relatively high temperature in the cylinder 18.

  The operation control map shown in FIG. 5 is an example, and various other maps can be provided.

  In addition, the exhaust passage is provided with only a three-way catalyst, but it is provided with a NOx purification catalyst, so that the excess air ratio λ is smaller than 2.4 and larger than 1.0, and A / F can be operated with Lean. May be.

  Furthermore, the technology disclosed herein can be applied to a diesel engine.

1 Engine (Engine body)
10 PCM (control unit)
16 Intake port (intake path)
161 First intake port 162 Second intake port 18 Cylinder 21 Intake valve 22 Exhaust valve 50 EGR passage (external EGR means)
62 Fuel supply system 67 Direct injection injector 68 Port injector 72 Variable drive mechanism (internal EGR means)
L11, L12, L13 Advance valve opening

Claims (4)

An engine main body having a cylinder to which each of a plurality of intake passages is connected, and configured to perform compression ignition combustion of an air-fuel mixture in the cylinder in a preset compression ignition combustion region;
An intake valve provided in each of the plurality of intake passages and configured to open and close the intake passage;
A port injector disposed in a specific intake path that is a part of a plurality of intake paths connected to the one cylinder, and configured to inject fuel into the specific intake path;
A second injector configured to supply fuel into the cylinder;
By performing a prior valve opening that opens the intake valve in the period from the expansion stroke to the exhaust stroke, and a main valve that opens at least in the intake stroke, a part of the combustion gas is passed through the specific intake passage. Internal EGR means configured to be introduced into the cylinder together with fresh air;
External EGR means configured to introduce a part of the combustion gas into the cylinder via an EGR passage connected at least to an intake passage other than the specific intake passage;
A control unit configured to operate the engine body through injection control of the port injector and the second injector, and
The control unit, when the operating state of the engine body is in a high load region greater than a predetermined load in the compression ignition combustion region,
A part of combustion gas is discharged into the specific intake passage by the internal EGR means, and a predetermined amount of fuel is injected into the specific intake passage by the port injector,
Control of a compression ignition engine through which a part of combustion gas is introduced into the cylinder by the external EGR means through the intake passages other than the specific intake passage and the remaining fuel amount is supplied into the cylinder by the second injector apparatus. The control apparatus for a compression ignition engine according to claim 1,
The control unit is a control device for a compression ignition engine that reduces the amount of fuel injected by the port injector as the load on the engine body increases. The control apparatus for a compression ignition engine according to claim 1 or 2,
The plurality of intake passages including the specific intake passage do not give angular momentum to the gas flowing into the cylinder, so that a swirl is not generated in the cylinder. . In the control apparatus of the compression ignition type engine according to any one of claims 1 to 3,
The second injector is a direct injection injector that injects fuel into the cylinder,
The control unit is a control device for a compression ignition engine that starts fuel injection by the direct injection injector in the latter half of the compression stroke.

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