JP2016031067A - Compression ignition engine control device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To realize both the improvement in ignitability of compression ignition and the improvement in exhaust emission performance of a compression ignition engine.SOLUTION: A compression ignition engine control device injects a predetermined quantity of fuel from an injector 68 provided in a specific intake passage that is one of a plurality of intake passages (intake ports 161, 162) connected to one cylinder 18 into the specific intake passage and adjusts a lift amount of an intake valve 21 so that an equivalent ratio of an air-fuel mixture introduced into the cylinder from the specific intake passage is within a range from 1.0 to 1.4 when an operating state of an engine body (engine 1) is in a compression ignition combustion region. On the other hand, intake air that does not contain the fuel is introduced into the cylinder from the intake passages other than the specific intake passage.SELECTED DRAWING: Figure 6

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. This engine is equipped with a so-called direct injection injector that directly injects fuel into the cylinder. In a low-load region in the compression ignition combustion region, fuel is injected into the cylinder during the intake stroke. A homogeneous air-fuel mixture is formed.

  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, and the rotation in the low load operation region is stabilized by stabilizing the small amount of fuel injection.

JP 2014-51928 A JP 2012-163028 A

  An air-fuel mixture provided with a direct injection injector as described in Patent Document 1 and formed in a cylinder is combusted by compression ignition (CI) or controlled auto ignition (CAI). In the engine, so-called single-stage injection, in which fuel is injected into a cylinder by one injection, is such that a large amount of air-fuel mixture is formed in the cylinder with a low local equivalent ratio (equivalent ratio φ is less than 1). Thus, there is a problem that the ignitability of compression ignition is lowered and 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.

  The technology disclosed herein has been made in view of the above points, and the object of the technology is to achieve both improvement in ignitability of compression ignition and improvement in exhaust emission performance in a compression ignition type engine. It is in.

  The technique disclosed here uses a so-called port injector instead of directly injecting fuel into a cylinder by a direct injection injector, and uses one or a plurality of intake passages connected to the cylinder (however, the cylinder A predetermined amount of fuel is injected into a specific intake passage (not all of the intake passages connected to the engine), and the amount of fuel is controlled by controlling the valve opening operation of the intake valve for the specific intake passage injected with fuel. It was decided to introduce a corresponding amount of fresh air into the cylinder. In this way, a homogeneous mixture having an equivalence ratio φ of 1 or more can be introduced into the cylinder through the specific intake passage, and a homogeneous mixture layer can be formed in the cylinder 18. As a result, the exhaust emission performance can be improved by optimizing the combustion temperature.

  Specifically, 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 preset compression ignition combustion region. An engine body configured to perform compression ignition combustion of the air-fuel mixture in the cylinder, an intake valve provided in each of the plurality of intake paths and configured to open and close the intake path, and at least a lift amount An intake valve drive mechanism configured to drive the intake valve, and a specific intake passage that is a part of a plurality of intake passages connected to the one cylinder, and the specific An injector configured to inject fuel into the intake passage; and a control unit configured to operate the engine body through valve opening control of the intake valve and injection control of the injector. That.

  The control unit injects a predetermined amount of fuel from the injector into the specific intake passage when the operating state of the engine body is in the compression ignition combustion region, and from the specific intake passage to the inside of the cylinder. The lift amount of the intake valve is adjusted so that the equivalence ratio of the air-fuel mixture introduced into the exhaust gas is in the range of 1.0 to 1.4, while the fuel is introduced into the cylinder from the intake passage other than the specific intake passage. Introducing air intake that does not contain

  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 injector may be disposed only in a specific intake path among the plurality of intake paths, or may be disposed in all of the plurality of intake paths.

  According to this configuration, when the operating state of the engine body is in the compression ignition combustion region, the control unit injects a predetermined amount of fuel from the injector into the specific intake passage. The fuel injection amount is set according to the operating state of the engine body. By injecting fuel into the intake passage, the air-fuel mixture introduced into the cylinder is homogenized.

  The control unit also adjusts the lift amount of the intake valve so that the equivalence ratio of the air-fuel mixture introduced into the cylinder from the specific intake passage into which the fuel is injected is in the range of 1.0 to 1.4. A known mechanical or hydraulic variable drive mechanism can be adopted as the intake valve drive mechanism. Thereby, a rich and homogeneous air-fuel mixture can be formed in the cylinder.

  On the other hand, intake air that does not contain fuel is introduced into the cylinder from intake passages other than the specific intake passage. As a result, the above-mentioned rich and homogeneous air-fuel mixture is stratified in the cylinder. In order to increase the stratification of the air-fuel mixture, it is preferable that the gas flow in the cylinder, particularly the flow in the rotation direction around the cylinder axis (that is, swirl) is weak.

  Since this homogeneous mixture has a rich equivalence ratio, the ignitability of compression ignition increases. Further, the homogeneous mixture does not include a rich mixture having an excessively high local equivalent ratio or a lean mixture having a low stoichiometric ratio, and the combustion temperature can be kept within an appropriate range. As a result, exhaust emission performance is improved.

  Further, even if the required fuel injection amount changes according to the operating state of the engine body, the fuel corresponding to the required injection amount is injected in the specific intake passage, and from the specific intake passage into the cylinder, By introducing a quantity of fresh air corresponding to the amount of fuel injected, a rich homogeneous mixture layer having an equivalence ratio of 1 or more can be formed in the cylinder. In other words, when the engine load is low and the fuel injection amount is relatively small, by reducing the amount of fresh air introduced into the cylinder from the specific intake passage, a rich homogeneous mixture layer with an equivalence ratio of 1 or more is created in the cylinder. Conversely, when the engine load is high and the fuel injection amount is relatively large, increasing the amount of fresh air from the specific intake passage into the cylinder increases the equivalence ratio to 1 or more. It is possible to form a uniform homogeneous mixture layer in the cylinder.

  The intake valve drive mechanism introduces a part of the combustion gas in the cylinder into the intake passage by performing a prior valve opening that opens the intake valve during a period from an expansion stroke to an exhaust stroke, and at least Combustion gas and fresh air are introduced into the cylinder by opening the intake valve in the intake stroke, and the control unit opens the intake valve in the specific intake passage in advance. Accordingly, a part of the combustion gas in the cylinder may be introduced into the specific intake path, while the intake valves other than the specific intake path may not perform the preceding valve opening.

  By doing so, a part of the combustion gas (burned gas) in the cylinder is introduced into the specific intake passage for injecting the fuel. For this reason, the air-fuel mixture introduced into the cylinder from this specific intake passage becomes a rich homogeneous and high-temperature air-fuel mixture having an equivalence ratio of 1 or more as described above. In this way, the temperature in the vicinity of the homogeneous air-fuel mixture layer stratified in the cylinder is increased, so that the ignitability of compression ignition is further enhanced.

  The gas introduced into the cylinder from the specific intake passage includes fuel, fresh air, and combustion gas, while the gas introduced into the cylinder from the intake passage other than the specific intake passage includes fresh air. It is. Here, in the specific intake passage, the amount of fresh air introduced into the cylinder is limited so as to correspond to the fuel amount. Therefore, when the combustion gas is not included, the amount of gas introduced into the cylinder from the specific intake passage is There is a difference between the amount of gas introduced into the cylinder from the intake passage other than the specific intake passage, and swirl may occur in the cylinder due to this gas amount difference. . The swirl stirs a layer of a rich homogeneous mixture having an equivalence ratio of 1 or more. This can make the air-fuel mixture in the cylinder as a whole lean, leading to a decrease in ignitability and exhaust emission performance.

  On the other hand, when the gas introduced into the cylinder from the specific intake path includes fuel, fresh air, and combustion gas, the amount of gas introduced into the cylinder from the specific intake path increases. The difference from the amount of gas introduced into the cylinder from other intake passages becomes smaller or disappears. As a result, it is possible to suppress swirl and to form a rich homogeneous and high-temperature air-fuel mixture layer with an equivalence ratio of 1 or more as described above, improving the ignitability of compression ignition, and exhaust emission performance It is advantageous for the improvement of both.

  The control unit introduces the air-fuel mixture so that the equivalence ratio is in the range of 1.0 to 1.4 from the specific intake passage to the cylinder in the low load region where the load in the compression ignition combustion region is low. The intake air that does not contain the fuel is introduced into the cylinder from the intake passages other than the specific intake passage, and the control unit also has a high load region that is higher in load than the low load region in the compression ignition combustion region. Injecting a predetermined amount of fuel in the specific intake passage and adjusting the lift amount of the intake valve to introduce an air-fuel mixture into the cylinder so as to achieve a predetermined equivalence ratio, while the specific intake passage By injecting the remaining fuel in the intake passage other than the above, an air-fuel mixture leaner than the predetermined equivalent ratio may be introduced from the intake passage into the cylinder.

  Since the fuel injection amount is relatively small in the low load region where the load on the engine body is low, even if the fuel is injected only into the specific intake passage, the equivalent ratio of the air-fuel mixture introduced into the cylinder from the specific intake passage It is possible to fall within the range of 1.0 to 1.4. Therefore, as described above, by not injecting fuel into the intake passages other than the specific intake passage, it becomes possible to form a rich and homogeneous mixture layer in the cylinder.

  On the other hand, since the fuel injection amount is relatively large in a high load region where the load of the engine body is high, if the fuel is injected only into the specific intake passage, the equivalent ratio falls within the range of 1.0 to 1.4. It becomes difficult. Therefore, in a high load region where the load on the engine body is high, fuel is also injected into the intake passage other than the specific intake passage.

  Specifically, a predetermined amount of fuel is injected in the specific intake passage, and the lift amount of the intake valve is adjusted. As a result, the air-fuel mixture is introduced into the cylinder so as to obtain a predetermined equivalent ratio. The predetermined amount of fuel may be set as the minimum fuel capable of compression ignition. Further, the fuel injection amount may be set in a range in which the combustion noise does not increase in consideration of the combustion noise accompanying the compression ignition combustion. As described above, the predetermined equivalence ratio may be in the range of 1.0 to 1.4, and when the load on the engine body is high, the temperature state in the cylinder is high and the ignitability of compression ignition is increased. Then, the predetermined equivalent ratio may be set lower than the range.

  On the other hand, the remaining fuel is injected into intake passages other than the specific intake passage. Further, in the intake passages other than the specific intake passage, it is not necessary to adjust the fresh air amount. The intake valve may be opened with the lift amount set in the operation state.

  Thus, the air-fuel mixture introduced into the cylinder from the specific intake passage becomes a relatively rich homogeneous air-fuel mixture, and the air-fuel mixture introduced into the cylinder from the intake passage other than the specific air intake passage is relatively lean and homogeneous. The mixture becomes a mixture, and concentration stratification of the mixture is achieved in the cylinder. Since a relatively rich homogeneous mixture has high ignitability, it is ignited at an early stage, and the relatively rich homogeneous mixture is also sequentially ignited by the heat generated by it and the rise in cylinder pressure. . Thus, when the load on the engine body is high, it becomes possible to sequentially ignite the air-fuel mixture in the cylinder while ensuring ignitability, and compression ignition combustion is slowed down. This is advantageous in preventing an increase in combustion noise by suppressing a sharp increase in pressure in the cylinder.

  As described above, according to the control apparatus for a compression ignition engine, when the operating state of the engine body is in the compression ignition combustion region, a predetermined amount of fuel is injected into the specific intake passage and the specific intake air is injected. While adjusting the lift amount of the intake valve so that the equivalence ratio of the air-fuel mixture introduced from the road into the cylinder is in the range of 1.0 to 1.4, from the intake path other than the specific intake path to the cylinder By introducing the intake air that does not contain fuel, it becomes possible to form a rich and homogeneous air-fuel mixture layer in the cylinder, so that the ignitability of compression ignition is enhanced and the exhaust emission performance is improved.

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 for demonstrating the basic concept of the technique disclosed here. 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. FIG. 7 is a diagram corresponding to FIG. 6 illustrating the process of forming an air-fuel mixture when the engine load is low. FIG. 7 is a diagram corresponding to FIG. 6 illustrating the process of forming an air-fuel mixture when the engine load is high.

  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.

  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. 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. 8 and 9, the port injector 68 is provided in each 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.

  A portion between the surge tank 33 and the throttle valve 36 in the intake passage 30 and a portion upstream of the direct catalyst 41 in the exhaust passage 40 are used for returning a part of the exhaust gas to the intake passage 30. They are connected via a passage 50. 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. ing. 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.

  In the CAI mode, basically, the port injector 68 is used to inject fuel into the intake port 16. As a result, a homogeneous air-fuel mixture is formed in the cylinder 18. The details of the air-fuel mixture formation in the cylinder including the fuel injection control will be described later. Even in the CAI mode, fuel may be injected into the cylinder 18 by using the direct injection injector 67 in a particularly high load region.

  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 to achieve concentration stratification of the air-fuel mixture formed in the cylinder 18, thereby achieving compression ignition combustion. Improving ignitability and exhaust emission performance. Hereinafter, this configuration will be described in detail with reference to the drawings.

  FIG. 6 shows the basic concept of the procedure for forming the air-fuel mixture in the cylinder 18. As shown in the figure, the first and second intake ports 161 and 162 communicate with one cylinder 18. Only the first intake port 161 is provided with the port injector 68, and the second intake port 162 is not provided with the port injector 68. FIG. 6 particularly shows a case where the load of the engine 1 is low and the fuel injection amount is relatively small.

  First, as shown in step P61 of FIG. 6, a predetermined amount of fuel is injected into the first intake port 161 through the port injector 68. The predetermined amount of fuel here corresponds to the load of the engine 1.

  In the subsequent process P62, the first intake port 161 is adjusted by adjusting the lift amount of the intake valve 21 that opens and closes the first intake port 161 into which fuel is injected using the variable drive mechanism 72. The fresh air amount is adjusted so that the equivalence ratio φ of the air-fuel mixture flowing into the cylinder 18 through 1 is approximately 1 (that is, the fresh air amount is reduced). For example, the equivalence ratio φ of the air-fuel mixture may fall within the range of 1.0 to 1.4. On the other hand, the lift amount of the intake valve 21 that opens and closes the second intake port 162 is not adjusted. That is, the intake valve 21 that opens and closes the second intake port 162 opens with a preset lift amount so that the fresh air amount set in the operating state is introduced into the cylinder 18. .

  Thus, a rich and homogeneous mixture layer 181 is formed 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. Note that tumble is allowed to occur in the cylinder 18 as long as the concentration stratification of the air-fuel mixture is maintained.

  Then, 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 (process P63). Since the air-fuel mixture layer is rich in the equivalence ratio φ, the ignitability of compression ignition is enhanced. Further, the air-fuel mixture layer is homogeneous by utilizing port injection, and does not include a gas mixture having a locally high equivalent ratio or a gas mixture having a locally low equivalent ratio. As a result, HC and CO that can be generated when combustion temperature is optimized and combustion temperature is low, NOx that can be generated when combustion temperature is high due to excessive oxygen, and local equivalent ratio are too high. You can suppress the wrinkles you get.

  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 gas mixture layer with a rich equivalence ratio φ by the air-fuel mixture introduced into the local area, and locally the lean air-fuel mixture near the stoichiometric or locally excessively rich air-fuel mixture can be formed. It becomes possible to avoid formation. Thereby, the improvement of the ignitability of compression ignition and the improvement of exhaust emission performance can both be achieved. 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.

  Next, a procedure for forming an air-fuel mixture when applied to the engine 1 will be described with reference to FIGS. FIG. 8 shows a procedure for forming an air-fuel mixture when the operating state of the engine 1 is on the low load side of the CAI mode, and FIG. 9 is a diagram when the operating state of the engine 1 is on the high load side of the CAI mode. 2 shows a procedure for forming an air-fuel mixture. Compared to FIG. 6, the first and second intake ports 161 and 162 are different in that a port injector 68 is provided.

  First, in step P81 of FIG. 8, 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 P82, 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 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). 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.

  Thus, in the process P83, it is possible to form a homogeneous and high-temperature air-fuel mixture layer 181 having an equivalent ratio φ of about 1 in the cylinder 18. Internal EGR gas, fresh air and fuel are introduced into the cylinder 18 from the first intake port 161, while only fresh air (or fresh air and internal EGR gas) is introduced into the cylinder 18 from the second intake port 162. ) 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.

  Thus, in the process P84, compression ignition combustion is reached. As described above, since the air-fuel mixture layer is rich with an equivalence ratio φ of 1 or more, the ignitability of compression ignition is further enhanced. Further, referring to FIG. 6, the exhaust emission performance can be improved without generating a locally excessively rich mixture, locally, in the vicinity of stoichiometric, or excessively, lean mixture. As described above.

  FIG. 9 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. As the load on the engine 1 increases, the fuel injection amount increases, and as a result, it becomes difficult to inject fuel only into the first intake port 161. Therefore, when the CAI mode is on the high load side, 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. 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 injected into the second intake port 162.

  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). 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) is introduced into the cylinder 18. The cooled EGR gas flows into each of the first intake port 161 and the second intake port 162 as shown by the white arrow in FIG. Therefore, internal EGR gas, external 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. And fuel is introduced.

  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 is possible to form a homogeneous and low-temperature air-fuel mixture layer (second layer) 182.

  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 this way, when the load of the engine 1 is high, it is possible to ignite sequentially while ensuring ignitability, so that the combustion is slowed down. This suppresses an abrupt increase in pressure in the cylinder 18 during combustion, thereby avoiding 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.

  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 62 Fuel supply system 67 Injector 72 Variable drive mechanism (intake valve drive mechanism)
L11, L12, L13 Advance valve opening

Claims (3)

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;
An intake valve drive mechanism configured to drive the intake valve so that at least a lift amount can be changed; and
An 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 control unit configured to operate the engine body through valve opening control of the intake valve and injection control of the injector; and
The control unit injects a predetermined amount of fuel into the specific intake passage from the injector and introduces the specific intake passage into the cylinder when the operating state of the engine body is in the compression ignition combustion region. The lift amount of the intake valve is adjusted so that the equivalent ratio of the air-fuel mixture is in the range of 1.0 to 1.4, while the cylinder contains the fuel from the intake passage other than the specific intake passage Control device for a compression ignition engine that introduces no intake air. The control apparatus for a compression ignition engine according to claim 1,
The intake valve drive mechanism introduces a part of the combustion gas in the cylinder into the intake passage by performing a prior valve opening that opens the intake valve during a period from an expansion stroke to an exhaust stroke, and at least It is configured to introduce combustion gas and fresh air into the cylinder by opening the intake valve in the intake stroke.
The control unit introduces a part of the combustion gas in the cylinder into the specific intake path by opening the intake valve of the specific intake path in advance, while intake air in intake paths other than the specific intake path A valve is a control device for a compression ignition engine that does not perform the preceding valve opening. The control apparatus for a compression ignition engine according to claim 1 or 2,
The control unit introduces the air-fuel mixture so that the equivalence ratio is in the range of 1.0 to 1.4 from the specific intake passage to the cylinder in the low load region where the load in the compression ignition combustion region is low. Introducing intake air that does not contain the fuel into the cylinder from intake passages other than the specific intake passage,
The control unit also injects a predetermined amount of fuel in the specific intake passage and adjusts a lift amount of the intake valve in a high load region where the load is higher than the low load region in the compression ignition combustion region. Then, while introducing the air-fuel mixture into the cylinder so as to achieve a predetermined equivalence ratio, the remaining fuel is injected into the intake passage other than the specific intake passage, so that the predetermined fuel is injected into the cylinder from the intake passage. A control device for a compression ignition engine that introduces an air-fuel mixture leaner than the equivalent ratio.

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