JP6268863B2 - Control device for compression ignition engine - Google Patents

Description

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

  For example, in Patent Document 1, when the operating state of the engine is in an operating region of a predetermined switching load or less, the air-fuel mixture in the cylinder is subjected to compression ignition combustion, while in the operating region where the load is higher than the switching load. In addition, an engine is described that is configured to perform combustion by forcibly igniting an air-fuel mixture in a cylinder with an ignition plug. In addition, when performing compression ignition combustion, this engine opens the exhaust valve again during the intake stroke, so that a part of the exhaust gas discharged to the exhaust side is introduced into the cylinder, so-called double exhaust opening. Do. The introduction of the internal EGR gas by the double opening of the exhaust raises the compression end temperature and improves the ignitability and combustion stability of the compression ignition. Furthermore, in this engine, when performing spark ignition combustion, cooled external EGR gas is introduced into the cylinder, thereby reducing cooling loss, raw NOx, and the like.

  Patent Document 2 also discloses an engine that performs compression ignition combustion when the operating state of the engine is in a predetermined region of low rotation and low load, and performs spark ignition combustion when it is in a region other than the predetermined region. Is described. This engine also improves the ignitability of compression ignition by adding ozone to the intake air in a region where compression ignition combustion is performed.

JP 2012-172665 A JP 2002-276404 A

  By the way, in the engine described in Patent Document 1, in the region where spark ignition combustion is performed, in order to introduce as much exhaust gas as possible into the cylinder, the throttle valve is left fully open while the EGR valve is opened. By adjusting the degree, the ratio between the amount of fresh air introduced into the cylinder and the amount of exhaust gas is adjusted. In order to make the air-fuel ratio of the air-fuel mixture constant at the stoichiometric air-fuel ratio with respect to the engine load level, the fuel injection amount decreases as the engine load decreases and the required new air amount also decreases. The lower the value, the higher the EGR rate (that is, the ratio of exhaust gas to the total gas amount in the cylinder). Thus, continuously changing the EGR rate with respect to the engine load is preferable in terms of engine control because the state quantity control is made continuous with respect to the engine load.

  However, a combustion mode that propagates a flame generated by forcibly igniting an air-fuel mixture, such as spark ignition combustion, reduces combustion stability when a large amount of exhaust gas is introduced into the cylinder. The maximum rate is limited to a relatively low value (so-called EGR limit). Therefore, as described in Patent Document 1, as a result of limiting the EGR rate in the low load region adjacent to the switching load in the region where forced ignition combustion is performed, for example, throttle valve opening control or In some cases, it is necessary to reduce the amount of fresh air introduced into the cylinder through control of the closing timing of the intake valve.

  In addition, since the region adjacent to the low load side with respect to the switching load corresponds to the region on the high load side in the region where compression ignition combustion is performed, the compression ignition combustion causes premature ignition where the pressure rise (dp / dθ) is severe. Can be burned. Since compression ignition combustion allows a high EGR rate, it is desirable to increase the EGR rate and introduce as much exhaust gas as possible into the cylinder in this region to slow down the combustion. However, as described above, since the spark ignition combustion is performed in the region adjacent to the high load side with respect to the switching load, the EGR rate is limited to a low level. The difference in the EGR rate becomes large between the two sides.

  The technology disclosed herein has been made in view of the above points, and the purpose thereof is to perform forced ignition combustion in a compression ignition engine that switches between compression ignition combustion and forced ignition combustion according to engine load. It is to allow a high EGR rate within an area to be performed, thereby allowing the EGR rate to continuously change with respect to the engine load.

  The technique disclosed here pays attention to the fact that introducing ozone into a cylinder increases the combustibility of the air-fuel mixture, and is adjacent to a switching load for switching to compression ignition combustion in a region where forced ignition combustion is performed. In a specific area, we decided to introduce ozone into the cylinder. By introducing ozone, combustion stability is ensured even if the EGR rate is set high. As a result, the EGR rate can be continuously changed with respect to the engine load.

  Specifically, a control device for a compression ignition engine disclosed herein includes an engine body having a cylinder, an exhaust gas recirculation system configured to introduce exhaust gas into the cylinder, and an air-fuel mixture in the cylinder. When the ignition device configured to perform forced ignition, the ozone introducer configured to introduce ozone into the cylinder, and the operating state of the engine body is in an operating region below a predetermined switching load, The air-fuel mixture in the cylinder is subjected to compression ignition combustion, and when the operating state of the engine body is in an operation region higher than the switching load, the ignition device is operated to forcibly ignite the air-fuel mixture in the cylinder. And a controller configured to operate the engine body by burning.

The exhaust gas recirculation system includes an external EGR system having an EGR passage configured to communicate an exhaust passage and an intake passage of the engine body and to flow exhaust gas from the exhaust side toward the intake side.

  When the operating state of the engine main body is in an operating region where the engine main body is higher in load than the switching load, the controller reduces the amount of the exhaust gas relative to the total gas amount in the cylinder as the load on the engine main body is lower. The EGR rate is continuously changed according to the level of the load so that the EGR rate, which is the ratio of the load, and in the specific region adjacent to the switching load, ozone is introduced into the cylinder by the ozone introducer. Is introduced.

In the specific region where ozone is introduced, the EGR rate is adjusted by the external EGR system, and the EGR rate set in the specific region where ozone is introduced is set higher than the EGR limit during forced ignition combustion. Has been.

  According to this configuration, the air-fuel mixture in the cylinder is subjected to compression ignition combustion when the operating state of the engine main body is in an operating region of a predetermined switching load or less. As a result, both improvement in exhaust emission performance and improvement in thermal efficiency are achieved. Further, when the operating state of the engine body is in an operating region where the load is higher than the switching load, the ignition device is operated to forcibly ignite and burn the air-fuel mixture in the cylinder. Combustion noise can be avoided by not performing compression ignition combustion when the load on the engine body is relatively high.

  When the operating state of the engine body is in an operating region where forced ignition combustion is performed, the EGR rate is continuously changed according to the level of the load so that the EGR rate increases as the load on the engine body decreases. Thereby, the state quantity control of the engine body is continued. For example, the exhaust gas as much as possible may be introduced into the cylinder after setting the air-fuel ratio of the air-fuel mixture to the stoichiometric air-fuel ratio. By doing so, pump loss and cooling loss are reduced, and generation of Raw NOx is also suppressed.

  Thus, ozone is introduced into the cylinder by the ozone introducer in a specific region adjacent to the switching load that is an operation region higher than the switching load. The specific region adjacent to the switching load corresponds to a low load region in the operation region where the forced ignition combustion is performed, and the EGR rate to be changed according to the load of the engine body is set to be relatively high. For this reason, although the combustion stability of forced ignition combustion is reduced, by introducing ozone into the cylinder in a specific region where the EGR rate is set high, the combustibility of the air-fuel mixture is improved and high combustion stability is achieved. It becomes possible to secure. In this way, by introducing ozone into the cylinder in the specific region, it is possible to continuously change the EGR rate over the entire operation region where the forced ignition combustion is performed. Note that ozone may be introduced into the cylinder when ozone is introduced into the cylinder, or ozone may be added to the gas within the cylinder after the intake air has been introduced into the cylinder. Good.

Further, as described above, in a specific region, the combustibility of forced ignition combustion is increased by introducing ozone into the cylinder. For this reason, even if a high EGR rate exceeding the EGR limit during forced ignition combustion is set, it is possible to ensure combustion stability. This makes it possible to continuously change the EGR rate over the entire operation region where the forced ignition combustion is performed. In particular, in a specific region where ozone is introduced into the cylinder, the control of the throttle valve and the intake valve The control for reducing the amount of fresh air introduced into the cylinder through the control of the valve closing timing is eliminated.

  In addition, by setting the EGR rate in the specific region higher than the EGR limit during forced ignition combustion, the difference in EGR rate between the high load side and the low load side across the switching load for switching the combustion mode is reduced. Or can be substantially eliminated.

Further, exhaust gas cooled by the external EGR system can be introduced into the cylinder, and forced ignition combustion is performed by combining the introduction of the cooled exhaust gas into the cylinder and the introduction of ozone in a specific region. It is possible to continuously change the EGR rate with respect to the load level of the engine body over the entire region in the load direction.

  The controller has a lower load on the engine body over an operating region that is higher in load than the switching load and a predetermined region that is lower than the switching load and that is adjacent to the operating region of high load. The EGR rate may be continuously changed according to the load level so that the EGR rate, which is the ratio of the exhaust gas amount to the total gas amount in the cylinder, becomes high.

  By doing so, the EGR rate is continuous over the high load side region where forced ignition combustion is performed and the low load side region where compression ignition combustion is performed (that is, the predetermined region) which are adjacent to each other with the switching load interposed therebetween. Therefore, the state quantity control can be continued while switching the combustion mode. Thus, at the time of switching the combustion mode, it is only necessary to substantially switch the ignition device between operation and non-operation, smoothing the switching of the combustion mode and suppressing the occurrence of torque shock or the like.

  Further, in the region where the compression ignition combustion below the switching load is performed, the predetermined region adjacent to the switching load has a relatively high EGR rate although the load on the engine body is relatively high, so that the compression ignition combustion is slowed down. This is advantageous for avoiding combustion noise.

The external EGR system, the exhaust gas further comprises a configured cooler to cool, the external EGR system also includes at least in the operating region of a high load than the switching load, adjusting the EGR rate It is good also as.

  In this way, in the operation region where the load is higher than the switching load, the exhaust gas cooled by the cooler is introduced into the cylinder, so that abnormal combustion such as pre-ignition and knocking can be avoided relatively. A large amount of exhaust gas can be introduced into the cylinder. In this way, the introduction of the cooled exhaust gas into the cylinder and the introduction of ozone in a specific region are combined, and the load on the engine body is reduced over the entire load direction in the region where forced ignition combustion is performed. It is possible to continuously change the EGR rate.

  Further, even in the predetermined region adjacent to the switching load and on the lower load side than the switching load, if the exhaust gas cooled by the external EGR system is introduced into the cylinder, the compression end temperature becomes too high. Is avoided, and compression ignition combustion is slowed down. This is advantageous for avoiding combustion noise.

  As described above, the compression ignition type engine control apparatus introduces ozone into the cylinder by the ozone introducer in a specific region adjacent to the switching load that is an operation region higher in load than the switching load. Thus, combustibility can be enhanced in a low load region in the region where forced ignition combustion is performed. As a result, it is possible to continuously change the EGR rate according to the level of the load so that the EGR rate becomes higher as the load on the engine body is lower over the entire operation region where the load is higher than the switching load. To do.

It is the schematic which shows the structure of a compression ignition type engine. It is a block diagram concerning control of a compression ignition type engine. It is sectional drawing which expands and shows a combustion chamber. It is a conceptual diagram which illustrates the structure of an ozone generator. An example of an intake valve lift curve configured to be able to continuously switch the lift amount, and an exhaust valve lift curve configured to be able to switch between a normal valve opening operation and a special operation that restarts during the intake stroke, and This is an example. It is a figure which illustrates the operation control map of an engine. (A) is an example of the fuel injection timing when the intake stroke injection is performed in the CI mode, and an example of the heat generation rate of the CI combustion associated therewith, and (b) is the fuel injection when performing the high pressure retarded injection in the CI mode. It is an example of the heat release rate of an example of time and the accompanying CI combustion. It is a figure which illustrates the relationship of the EGR rate with respect to high and low of engine load, and the relationship of ozone concentration with respect to high and low of engine load.

  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. 1 and 2 show a schematic configuration of an engine (engine body) 1. The engine 1 is a spark ignition gasoline engine that is mounted on a vehicle and supplied with 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. A cavity 141 like a reentrant type in a diesel engine is formed on the top surface of the piston 14 as shown in an enlarged view in FIG. The cavity 141 is opposed to an 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 19. The shape of the combustion chamber 19 is not limited to the illustrated shape. For example, the shape of the cavity 141, the top surface shape of the piston 14, the shape of the ceiling portion of the combustion chamber 19, and the like can be changed as appropriate.

  The 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 the compression ignition combustion described later. In addition, what is necessary is just to set a geometric compression ratio suitably in the range of about 15-20.

  The cylinder head 12 is provided with an intake port 16 and an exhaust port 17 for each cylinder 18. The intake port 16 and the exhaust port 17 have an intake valve 21 and an exhaust for opening and closing the opening on the combustion chamber 19 side. Each valve 22 is disposed.

  Among the valve systems that drive the intake valve 21 and the exhaust valve 22, respectively, on the exhaust side, the operation mode of the exhaust valve 22 is switched between a normal mode and a special mode, for example, a hydraulically operated variable mechanism (see FIG. 2). Hereinafter, a VVL (Variable Valve Lift) 71 is provided. Although the detailed illustration of the configuration of the VVL 71 is omitted, two types of cams having different cam profiles, a first cam having one cam peak and a second cam having two cam peaks, and the first And a lost motion mechanism that selectively transmits the operating state of one of the second cams to the exhaust valve 22. When the operating state of the first cam is transmitted to the exhaust valve 22, as illustrated by a solid line in FIG. 5, the exhaust valve 22 operates in the normal mode in which the valve is opened only once during the exhaust stroke. When the operating state of the second cam is transmitted to the exhaust valve 22, as illustrated by a broken line in FIG. 5, the exhaust valve 22 opens during the exhaust stroke and also opens during the intake stroke. It operates in a special mode that opens the exhaust twice. The normal mode and the special mode of the VVL 71 are switched according to the operating state of the engine. Specifically, the special mode is used in the control related to the internal EGR. In the following explanation, operating the VVL 71 in the normal mode and not opening the exhaust twice is referred to as “turning off the VVL 71”, and operating the VVL 71 in the special mode and opening the exhaust twice. , “Turn on VVL 71”. In order to enable switching between the normal mode and the special mode, an electromagnetically driven valve system that drives the exhaust valve 22 by an electromagnetic actuator may be employed.

  Note that the execution of the internal EGR is not realized only by opening the exhaust gas twice. For example, it is possible to perform internal EGR control by opening the intake valve 21 twice, or by opening the intake valve twice, and providing a negative overlap period in which both the intake valve 21 and the exhaust valve 22 are closed in the exhaust stroke or the intake stroke. It is also possible to perform internal EGR control that causes the burned gas to remain in the cylinder 18. However, as will be described later, in order to increase the compression end temperature, it is most preferable to open the exhaust twice.

  As shown in FIG. 2, the exhaust side valve system having the VVL 71 is arranged on the intake side as shown in FIG. 2. 72) and a lift variable mechanism (hereinafter referred to as CVVL (hereinafter referred to as CVVL)) that can continuously change the lift amount of the intake valve 21, as exemplified by a solid line, a broken line, and an alternate long and short dash line in FIG. (Continuously Variable Valve Lift) 73). The VVT 72 may employ a hydraulic, electromagnetic, or mechanical structure as appropriate, and illustration of the detailed structure is omitted. The CVVL 73 can also adopt various known structures as appropriate, and the detailed structure is not shown. By the VVT 72 and the CVVL 73, the intake valve 21 can change its valve opening timing, valve closing timing, and lift amount.

  In addition, for each cylinder 18, an injector 67 that directly injects fuel into the cylinder 18 (direct injection) is attached to the cylinder head 12. As shown in an enlarged view in FIG. 3, the injector 67 is disposed so that its nozzle hole faces the inside of the combustion chamber 19 from the central portion of the ceiling surface of the combustion chamber 19. The injector 67 directly injects an amount of fuel into the combustion chamber 19 at an injection timing set according to the operating state of the engine 1 and according to the operating state of the engine 1. In this example, the injector 67 is a multi-hole injector having a plurality of nozzle holes, although detailed illustration is omitted. Thereby, the injector 67 injects the fuel so that the fuel spray spreads radially from the center position of the combustion chamber 19. As indicated by the arrows in FIG. 3, the fuel spray injected radially from the central portion of the combustion chamber 19 at the timing when the piston 14 is located near the compression top dead center is a cavity formed on the top surface of the piston. It flows along the wall surface of 141. 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-hole 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 injector 67 is not limited to a multi-hole injector, and may be an outside-opening type injector.

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

  As shown in FIG. 3, a spark plug 25 for forcibly igniting the air-fuel mixture in the combustion chamber 19 is 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. As shown in FIG. 3, 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 19 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.

  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.

Further, an ozone generator (O3 generator) 76 for adding ozone to fresh air introduced into the cylinder 18 is interposed between the throttle valve 36 and the surge tank 33 in the intake passage 30.
For example, as shown in FIG. 4, the ozone generator 76 includes a plurality of electrodes arranged in parallel at predetermined intervals in the vertical and horizontal directions on the cross section of the intake pipe 301. The ozone generator 76 generates ozone by silent discharge using oxygen contained in the intake air as a source gas. That is, when a high frequency alternating current high voltage is applied to the electrode from a power source (not shown), silent discharge is generated in the discharge gap, and the air (that is, intake air) passing therethrough is ozonized. The intake air thus added with ozone is introduced into each cylinder 18 from the surge tank 33 via the intake manifold. The ozone concentration in the intake air after passing through the ozone generator 76 is adjusted by changing the voltage application mode to the electrodes of the ozone generator 76 and / or changing the number of electrodes to which the voltage is applied. It is possible. As will be described later, the PCM 10 adjusts the ozone concentration in the intake air introduced into the cylinder 18 through the control of the ozone generator 76.

  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. This PCM 10 constitutes a controller.

  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. An exhaust temperature sensor SW7 and an exhaust pressure sensor SW8 for detecting force, and a linear O2 sensor SW9 for detecting the oxygen concentration in the exhaust gas, upstream of the direct catalyst 41, a direct catalyst 41 and an underfoot catalyst 42, The lambda O2 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 accelerator pedal of the vehicle An accelerator opening sensor SW13 for detecting an accelerator opening degree corresponding to an operation amount (not shown), cam angle sensors SW14 and SW15 on the intake side and exhaust side, and a common rail 64 of the fuel supply system 62, and an injector 67 is a fuel pressure sensor SW16 for detecting the fuel pressure supplied to 67. .

  The PCM 10 determines the state of the engine 1 and the vehicle by performing various calculations based on these detection signals, and accordingly, the injector 67, the spark plug 25, the VVT 72 and the CVVL 73 of the intake valve 21, and the exhaust valve 22 Control signals are output to the VVL 71, the fuel supply system 62, actuators of various valves (throttle valve 36, intercooler bypass valve 351, EGR valve 511, EGR cooler bypass valve 531), and ozone generator 76. Thus, the PCM 10 operates the engine 1.

  FIG. 6 shows an example of the operation control map of the engine 1. This engine 1 is a compression ignition combustion in which combustion is performed by compression self-ignition without ignition by the spark plug 25 in a low load region where the engine load is relatively low for the purpose of improving fuel consumption and exhaust emission performance. I do. However, as the load on the engine 1 increases, in the compression ignition combustion, the combustion becomes too steep and causes problems such as combustion noise. Therefore, in the engine 1, in a high load region where the engine load is relatively high, the compression ignition combustion is stopped and switched to forced ignition combustion (here, spark ignition combustion) using the spark plug 25. As described above, the engine 1 has a CI (Compression Ignition) mode in which compression ignition combustion is performed and an SI (Spark Ignition) mode in which spark ignition combustion is performed in accordance with the operation state of the engine 1, in particular, the load of the engine 1. It is configured to switch. However, the boundary line for mode switching is not limited to the illustrated example.

  The CI mode is further divided into two areas depending on the engine load. Specifically, in the region (1) corresponding to the low and medium load in the CI 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 compression ignition combustion. . As will be described in detail later, this is because the VVL 71 on the exhaust side is turned on and the exhaust valve 22 is opened twice during the intake stroke. The introduction of the hot EGR gas is advantageous in increasing the compression end temperature in the cylinder 18 and improving the ignitability and combustion stability of the compression ignition in the region (1). In the region (1), as shown in FIG. 7A, the injector 67 injects fuel into the cylinder 18 at least during the period from the intake stroke to the middle of the compression stroke, thereby forming a homogeneous mixture. To do. As shown in FIG. 7A, the homogeneous air-fuel mixture undergoes compression self-ignition near the compression top dead center.

  In the high load region (2) in the CI mode including the switching boundary line between the CI mode and the SI mode (that is, the switching load), the temperature environment in the cylinder 18 becomes high. For this reason, the amount of hot EGR gas is reduced in order to suppress premature ignition, while cooled EGR gas cooled by passing through the EGR cooler 52 is introduced into the cylinder 18.

  The engine 1 also expands the CI mode region to the high load side as much as possible by setting the switching load as high as possible. In the high load region (2), the intake stroke is changed to the middle of the compression stroke. If fuel is injected into the cylinder 18 within the period up to this time, abnormal combustion such as pre-ignition may occur. On the other hand, if a large amount of cooled EGR gas having a low temperature is introduced to lower the compression end temperature in the cylinder, the ignitability of the compression ignition is deteriorated. That is, in the region (2), the compression ignition combustion cannot be stably performed only by the temperature control in the cylinder 18. Therefore, in this region (2), in addition to the temperature control in the cylinder 18, the compression ignition combustion is stabilized while avoiding abnormal combustion such as pre-ignition by devising the fuel injection mode. Specifically, this fuel injection mode has a fuel pressure that is significantly higher than that of the prior art, and as shown in FIG. 7B, at least a period from the latter stage of the compression stroke to the initial stage of the expansion stroke (hereinafter referred to as this The fuel is injected into the cylinder 18 within a period of time (referred to as a retard period). This characteristic fuel injection mode is hereinafter referred to as “high pressure retarded injection” or simply “retarded injection”. By such high-pressure retarded injection, compression ignition combustion is stabilized while avoiding abnormal combustion in the region (2). Details of the high-pressure retarded injection will be described later.

  In contrast to the CI mode, the SI mode is not clearly shown in FIG. 6, but the introduction of the hot EGR gas is continued while the VVL 71 on the exhaust side is turned off to stop the introduction of the hot EGR gas. . In the SI mode, as will be described in detail later, while the throttle valve 36 is fully opened, the amount of fresh air introduced into the cylinder 18 and the amount of external EGR gas are adjusted by adjusting the opening of the EGR valve 511. Adjusting the gas ratio introduced into the cylinder 18 in this way reduces pump loss, avoids abnormal combustion by introducing a large amount of cooled EGR gas into the cylinder 18, and keeps the combustion temperature of spark ignition combustion low. Can suppress the generation of Raw NOx and the cooling loss. In the fully open load range, the external EGR is set to zero by closing the EGR valve 511.

  As described above, the geometric compression ratio of the engine 1 is set to 15 or more (for example, 18). Since the high compression ratio increases the compression end temperature and the compression end pressure, it is advantageous for stabilizing the compression ignition combustion in the CI mode, particularly in a low load region (for example, the region (1)). On the other hand, the high compression ratio engine 1 has a problem that abnormal combustion such as pre-ignition and knocking is likely to occur in the SI mode which is a high load region.

  Therefore, in the engine 1, in the SI mode, abnormal combustion is avoided by performing the high-pressure retarded injection described above. More specifically, high-pressure retarded injection is performed in which fuel is injected into the cylinder 18 within the retard period from the latter half of the compression stroke to the early stage of the expansion stroke with a high fuel pressure of 30 MPa or more. In the SI mode, in addition to the high pressure retarded injection in the retard period, a part of the injected fuel is injected into the cylinder 18 in the intake stroke period in which the intake valve 21 is opened. (That is, split injection may be performed).

  Here, the high-pressure retarded injection in the SI mode will be briefly described. For example, as described in detail in Patent Document 1 (Japanese Patent Laid-Open No. 2012-172665) filed earlier by the applicant of the present application, High-pressure retarded injection aims to shorten the possible reaction time from the start of fuel injection to the end of combustion, thereby avoiding abnormal combustion. That is, the reaction possible time includes a period during which the injector 67 injects fuel ((1) injection period) and a period after the end of injection until a combustible mixture is formed around the spark plug 25 ((2) mixture) (The formation period) and the period until the combustion started by ignition is completed ((3) combustion period), that is, (1) + (2) + (3). In the high pressure retarded injection, fuel is injected into the cylinder 18 at a high pressure, thereby shortening the injection period and the mixture formation period. The shortening of the injection period and the mixture formation period makes it possible to set the fuel injection timing, more precisely, the injection start timing to a relatively late timing. Fuel injection is performed within the retard period.

  As fuel is injected into the cylinder 18 with high fuel pressure, the turbulence in the cylinder becomes stronger, and the turbulence energy in the cylinder 18 increases. By setting the fuel injection timing to a relatively late timing, it becomes possible to start combustion by performing spark ignition while maintaining high turbulent energy. This shortens the combustion period.

  In this way, the high pressure retarded injection shortens the injection period, the mixture formation period, and the combustion period, and as a result, the reaction time of the unburned mixture is compared with the case of the fuel injection in the conventional intake stroke. Can be significantly shortened. As a result of shortening the possible reaction time, it is possible to suppress the progress of the reaction of the unburned mixture at the end of combustion and to avoid abnormal combustion.

  Here, the combustion period can be effectively shortened by setting the fuel pressure to, for example, 30 MPa or more. Moreover, the fuel pressure of 30 MPa or more can effectively shorten the injection period and the mixture formation period, respectively. The fuel pressure is preferably set as appropriate according to the properties of the fuel used, which contains at least gasoline. The upper limit may be 120 MPa as an example.

  The high pressure retarded injection avoids the occurrence of abnormal combustion in the SI mode by devising the form of fuel injection into the cylinder 18. Unlike this, it is conventionally known that the ignition timing is retarded for the purpose of avoiding abnormal combustion. While retarding the ignition timing causes a decrease in thermal efficiency and torque, when performing high pressure retarded injection, it is possible to advance the ignition timing by avoiding abnormal combustion by devising the form of fuel injection. As a result, thermal efficiency and torque are improved. That is, the high pressure retarded injection not only avoids abnormal combustion, but also makes it possible to advance the ignition timing by the amount that can be avoided, which is advantageous in improving fuel consumption.

  As described above, the high pressure retarded injection in the SI mode can shorten the injection period, the mixture formation period, and the combustion period, but the high pressure retarded injection performed in the CI mode region (2) It is possible to shorten the injection period and the mixture formation period. In other words, the turbulence in the cylinder 18 is increased by injecting the fuel into the cylinder 18 at a high fuel pressure, so that the mixing performance of the atomized fuel is increased and the fuel is injected at a late timing near the compression top dead center. However, a relatively homogeneous air-fuel mixture can be quickly formed.

  In the high pressure retarded injection in the CI mode, fuel is injected at a late timing near the compression top dead center in an area where the load is relatively high, so that excessive fuel is not injected into the cylinder 18 in the first place. As described above, a substantially homogeneous air-fuel mixture is quickly formed while preventing pre-ignition, so that compression ignition can be reliably performed after the compression top dead center. Thus, in the expansion stroke period in which the pressure in the cylinder 18 gradually decreases due to motoring, the compression ignition combustion is performed, so that the combustion becomes slow, and the pressure increase in the cylinder 18 due to the compression ignition combustion (dP / It is avoided that dθ) becomes steep. This eliminates the NVH restriction, so that the CI mode area is expanded to the high load side.

  FIG. 8 shows a change in the EGR rate with respect to the load of the engine 1 (that is, a change in the gas composition in the cylinder 18). Hereinafter, the change in the EGR rate will be described in order from the high load side to the low load side.

(From maximum load Tmax to switching load T3)
A region where the load is higher than the switching load T3 corresponds to the SI mode. In this SI region, as described above, only the cooled EGR gas is introduced into the cylinder 18. That is, the opening degree of the throttle valve 36 is kept fully open, and the EGR valve 511 is closed at the fully open load, but gradually opens as the engine load decreases. 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 makes it possible to use a three-way catalyst. Since the fuel injection amount decreases as the engine load decreases, the EGR rate increases continuously. This is advantageous in improving controllability because the gas composition in the cylinder 18 is continuously changed when the engine load changes continuously.

  In spark ignition combustion, if the amount of exhaust gas introduced into the cylinder 18 is too large, the combustion stability deteriorates. Therefore, there is a maximum EGR rate (that is, an EGR limit) that can be set in the spark ignition combustion. As described above, the EGR rate continuously increases as the engine load decreases, but at the predetermined load T4, the EGR rate reaches the EGR limit. For this reason, during the period from the predetermined load T4 to the switching load T3, the EGR rate cannot be set higher than that, as indicated by a one-dot chain line in FIG. Further, when the EGR rate is limited by the EGR limit as described above, when setting the air-fuel ratio of the air-fuel mixture to the stoichiometric air-fuel ratio (λ≈1), as shown by a one-dot chain line in FIG. The fresh air introduced into the cylinder 18 must be reduced through the opening degree control and the closing timing control of the intake valve 21.

  On the other hand, in the engine 1, the ozone generator 76 is operated and introduced into the cylinder 18 in a specific region between the predetermined load T4 and the switching load T3 where the EGR rate is limited by the EGR limit. Add ozone to the intake. Introducing ozone into the cylinder 18 enhances the ignitability of spark ignition combustion, so that it is possible to ensure the combustion stability of spark ignition combustion even when the EGR rate is set higher than the EGR limit. . Thus, in this engine 1, by setting the EGR rate higher than the EGR limit between the predetermined load T4 and the switching load T3, the EGR rate is continuously changed with respect to the load level of the engine 1. (Refer to the hatched portion in FIG. 8). As a result, the EGR rate is continuously changed with respect to the load level of the engine 1 over the entire region where the spark ignition combustion is performed up to the switching load T3. Further, it is not necessary to reduce the amount of fresh air introduced into the cylinder 18.

  Here, as shown in the lower diagram of FIG. 8, the ozone concentration may be set so as to increase continuously as the load decreases according to the load of the engine 1. By doing this, it is possible to achieve the minimum ozone concentration necessary to ensure the combustion stability of spark ignition combustion, minimizing the power consumption required for generating ozone, and improving fuel efficiency. Become advantageous. The maximum ozone concentration may be about 50 to 30 ppm, for example. Further, the ozone concentration may be set to increase stepwise as the load decreases according to the load of the engine 1. Further, the ozone concentration may be set to a constant concentration between the predetermined load T4 and the switching load T3 regardless of the load of the engine 1.

(From switching load T3 to specific load T1)
As described above, the switching load T3 is involved in switching between the CI mode and the SI mode, and is in the CI mode on the low load side below the switching load T3. In each of the low load side and the high load side across the switching load between the CI mode and the SI mode, the air-fuel ratio of the air-fuel mixture is set to the theoretical air-fuel ratio (λ≈1). For this reason, the EGR rate continuously increases from the CI mode to the SI mode. This means that when switching between the CI mode and the SI mode in which the combustion mode is switched, there is no significant change other than switching between execution and non-execution of spark ignition, switching from the CI mode to the SI mode, Alternatively, it is possible to suppress the occurrence of torque shock or the like by switching the reverse switching smoothly.

  Further, in the region adjacent to the low load side with respect to the switching load T3, a relatively large amount of EGR gas (cooled EGR gas) is put into the cylinder 18 so as to continue from the region adjacent to the high load side with respect to the switching load T3. While being introduced, compression ignition combustion is performed by performing high pressure retarded injection in which fuel is injected at a high fuel pressure of 30 MPa or more and in the vicinity of compression top dead center. This means that the compression ignition combustion is performed stably in a region where the load of the engine 1 is relatively high while slowing down the compression ignition combustion and eliminating the restriction of dP / dθ. Enable.

  In the CI mode, the VVL 71 on the exhaust side is turned on, and internal EGR gas (that is, hot EGR gas) is introduced into the cylinder 18. Therefore, the VVL 71 on the exhaust side is switched on / off with the switching load T3 as a boundary. The EGR rate obtained by adding the hot EGR gas and the cooled EGR gas continuously increases as the load on the engine 1 decreases. Further, the ratio of the cooled EGR gas to the hot EGR gas gradually decreases as the load on the engine 1 decreases, and the hot EGR gas ratio gradually increases. The introduction amount of the cooled EGR gas is adjusted by controlling the opening degree of the EGR valve 511. On the other hand, the amount of hot EGR gas introduced is adjusted by adjusting the degree of overlap of the opening period of the intake valve 21 with respect to the opening period of the exhaust valve 22 that opens within the intake stroke period. Specifically, the introduction amount of the hot EGR gas is adjusted by adjusting the valve opening timing and the valve closing timing of the intake valve 21 by the VVT 72 and the CVVL 73 on the intake side.

  Thus, the introduction of the cooled EGR gas is stopped at the predetermined load T2 between the switching load T3 and the specific load T1, and when the load of the engine 1 is lower than the predetermined load T2, only the hot EGR gas is contained in the cylinder 18. To be introduced. Increasing the amount of hot EGR gas introduced as the load on the engine 1 decreases increases the gas temperature in the cylinder before the start of compression, and accordingly increases the compression end temperature. This is advantageous in improving the ignition quality of compression ignition in the region where the load of the engine 1 is low and improving the stability of compression ignition combustion.

  The EGR rate that continuously increases as the load of the engine 1 decreases is set to the maximum EGR rate rmax at the specific load T1.

(From specific load T1 to minimum load)
Until the specific load T1, as described above, the EGR rate is continuously set higher as the load of the engine 1 decreases. However, when the load of the engine 1 is lower than the specific load T1, the load of the engine 1 Regardless of the height, the EGR rate is made constant at the maximum EGR rate rmax. Thereby, the air-fuel ratio of the air-fuel mixture is set to lean.

  Here, setting the EGR rate so as not to exceed the maximum EGR rate rmax means that if a large amount of exhaust gas is introduced into the cylinder 18 by increasing the EGR rate, the specific heat ratio of the gas in the cylinder 18 This is because, even if the gas temperature before the start of compression is high, the compression end temperature becomes low.

  That is, the exhaust gas contains a large amount of triatomic molecules such as CO2 and H2O, and has a higher specific heat ratio than air containing nitrogen (N2) and oxygen (O2). Therefore, when the EGR rate is increased and the exhaust gas introduced into the cylinder 18 increases, the specific heat ratio of the gas in the cylinder 18 decreases.

  Since the temperature of the exhaust gas is higher than that of fresh air, the higher the EGR rate, the higher the temperature of the gas before starting compression. However, the higher the EGR rate, the lower the specific heat ratio of the gas. Therefore, even if compression is performed, the temperature of the gas does not increase so much. As a result, the compression end temperature becomes the highest at a predetermined EGR rate rmax, and the EGR rate Even if it is made higher than that, the compression end temperature is lowered.

  Therefore, in the engine 1, the EGR rate at which the compression end temperature is highest is set to the maximum EGR rate rmax. When the load on the engine 1 is lower than the specific load T1, the EGR rate is set to the maximum EGR rate rmax, thereby avoiding a decrease in the compression end temperature. The maximum EGR rate rmax may be set to 50 to 90%. The maximum EGR rate rmax may be set as high as possible as long as a high compression end temperature can be secured, and is preferably 70 to 90%. In this engine 1, the geometric compression ratio is set to a high compression ratio of 15 or more so that a high compression end temperature can be obtained. Further, in order to introduce exhaust gas having as high a temperature as possible into the cylinder 18, a double exhaust opening is adopted. In other words, in the case of the double exhaust opening, the exhaust gas introduced into the cylinder 18 is once discharged to the exhaust port, so that unlike the configuration in which a negative overlap period is provided, the exhaust gas is compressed during the exhaust stroke to increase the cooling loss. In addition, unlike the double intake opening that exhausts the exhaust gas to the intake port having a relatively low temperature, the temperature drop of the exhaust gas can be suppressed, so that the gas temperature before the start of compression is the highest. Is possible. In the engine 1 configured to ensure as high a compression end temperature as possible, the maximum EGR rate rmax may be set to about 80%, for example. Setting the maximum EGR rate rmax as high as possible is advantageous in reducing the unburned loss of the engine 1. That is, since the unburned loss tends to increase when the load on the engine 1 is low, setting the EGR rate as high as possible when the load on the engine 1 is lower than the specific load T1 improves fuel consumption by reducing the unburned loss. Is extremely effective.

  In this way, in this engine 1, even when the load of the engine 1 is lower than the specific load T1, by ensuring a high compression end temperature, the ignitability and combustion stability of compression ignition combustion are ensured.

  The technique disclosed here is not limited to application to the engine configuration described above. For example, fuel may be injected into the intake port 16 through a port injector provided separately in the intake port 16 instead of the injector 67 provided in the cylinder 18 during the intake stroke period.

  Further, regarding the valve train of the engine 1, a VVL that can switch between a large lift cam that relatively increases the lift amount and a small lift cam that relatively decreases the lift amount is provided instead of the CVVL 73 of the intake valve 21. It may be. In that case, VVT may be provided on the exhaust side so that the amount of hot EGR gas introduced can be adjusted.

  The engine 1 is not limited to an in-line 4-cylinder engine, and may be applied to an in-line 3-cylinder, in-line 2-cylinder, in-line 6-cylinder engine, or the like. Further, the present invention can be applied to various engines such as a V type 6 cylinder, a V type 8 cylinder, and a horizontally opposed 4 cylinder.

  Further, in the above description, the air-fuel ratio of the air-fuel mixture is set to the stoichiometric air-fuel ratio (λ≈1) in the predetermined operation region, but the air-fuel ratio of the air-fuel mixture may be set to lean. However, setting the air-fuel ratio to the stoichiometric air-fuel ratio has the advantage that a three-way catalyst can be used.

  The operation control map shown in FIG. 6 is merely an example, and various other maps can be provided.

  Further, the high-pressure retarded injection may be divided injection as necessary, and similarly, the intake stroke injection may also be divided injection as necessary. In these divided injections, fuel may be injected in each of the intake stroke and the compression stroke.

1 Engine (Engine body)
10 PCM (controller)
18 cylinder 21 intake valve 22 exhaust valve 25 spark plug (ignition device)
50 EGR passage (exhaust gas recirculation system)
51 Main passage (external EGR system, exhaust gas recirculation system)
511 EGR valve (external EGR system, exhaust gas recirculation system)
52 EGR cooler (cooler)
76 Ozone generator (ozone introducer)

Claims (3)

An engine body having a cylinder;
An exhaust gas recirculation system configured to introduce exhaust gas into the cylinder;
An ignition device configured to perform forced ignition on the air-fuel mixture in the cylinder;
An ozone introducer configured to introduce ozone into the cylinder;
When the operating state of the engine body is in an operating region below a predetermined switching load, the air-fuel mixture in the cylinder is subjected to compression ignition combustion, and the operating state of the engine body is in an operating region where the load is higher than the switching load. A controller configured to operate the engine body by operating the ignition device and forcibly igniting and burning the air-fuel mixture in the cylinder; and
The exhaust gas recirculation system includes an external EGR system having an EGR passage configured to communicate an exhaust passage and an intake passage of the engine body and to flow exhaust gas from the exhaust side toward the intake side,
The controller, when the operating state of the engine body is in an operating region of a higher load than the switching load,
The EGR rate is continuously changed according to the level of the load so that the EGR rate, which is the ratio of the amount of the exhaust gas to the total gas amount in the cylinder, becomes higher as the load on the engine body is lower. ,
In a specific region adjacent to the switching load, ozone is introduced into the cylinder by the ozone introducer ,
In the specific region where ozone is introduced, the EGR rate is adjusted by the external EGR system,
The control apparatus for a compression ignition engine in which the EGR rate set in the specific region where ozone is introduced is set higher than an EGR limit during forced ignition combustion . The control apparatus for a compression ignition engine according to claim 1,
The controller has a lower load on the engine body over an operating region that is higher in load than the switching load and a predetermined region that is lower than the switching load and that is adjacent to the operating region of high load. A control device for a compression ignition engine that continuously changes the EGR rate according to the load level so that the EGR rate, which is the ratio of the exhaust gas amount to the total gas amount in the cylinder, becomes high. The control apparatus for a compression ignition engine according to claim 1 or 2 ,
The external EGR system further comprises a cooler configured to cool the exhaust gas ,
The external EGR system is also a control device for a compression ignition engine that adjusts the EGR rate at least in an operation region where the load is higher than the switching load.

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