JP6019935B2 - Spark ignition direct injection engine - Google Patents

Description

  The present invention relates to a spark ignition type direct injection engine that performs compression self-ignition combustion in a low load region and performs spark ignition combustion in a high load region.

  In compression self-ignition combustion (hereinafter also referred to as CI combustion) in which the air-fuel mixture in the cylinder is compressed self-ignited (Homogenious Charge Compression Ignition), the advantage that the fuel efficiency can be improved by increasing thermal efficiency, Although it is short but does not cause intense combustion, there is an advantage that the production of nitrogen oxides is remarkably reduced. For this reason, conventionally, even in a spark ignition type gasoline engine, spark ignition combustion (hereinafter also referred to as SI combustion) is performed in an operation region on a high load side that is equal to or higher than a predetermined load, and on the lower load side than a predetermined load. In the operation region, it is known to improve fuel efficiency and exhaust emission by performing CI combustion (for example, Patent Document 1).

  Further, in this type of engine, in order to improve the ignitability in the low load side region where the temperature in the cylinder hardly rises due to the small amount of fuel injection, etc., and to reduce the pumping loss, It is known to introduce internal EGR gas into a cylinder.

  Here, as a technique for introducing the internal EGR gas into the cylinder, the burned gas remains in the combustion chamber by changing the closing timing of the exhaust valve and the opening timing of the intake valve in a predetermined operation region. What is known (so-called negative overlap) is known. However, when the burned gas is left in the cylinder due to the negative overlap, the burnt gas is lost due to the heat of the burned gas being taken away by the cylinder or the burned gas is re-expanded. There is a problem that the temperature is lowered and a problem that a pumping loss occurs when the burned gas is recompressed.

As a technique for solving such a problem, for example, in Patent Document 2, the exhaust valve is lifted (opened) in the exhaust stroke , and burned gas is once expelled to the exhaust port, and then the exhaust valve is lifted again in the intake stroke. It is disclosed that the ignitability in the CI combustion region is improved by introducing the burned gas into the cylinder at a high temperature (so-called exhaust twice).

JP 2007-154859 A JP 2011-214479 A

  By the way, the fuel injection amount increases as the engine load increases, and the intake air amount increases as the fuel injection amount increases. For example, a variable mechanism capable of changing the lift amount and opening / closing timing of the exhaust valve is provided. If the engine is not open, it will be difficult to inhale into the cylinder the amount of fresh air corresponding to a load higher than the predetermined level if the exhaust double opening that re-lifts the exhaust valve in the intake stroke is continued. It is necessary to stop opening the exhaust twice before it becomes.

  Here, the entire region composed of the CI combustion region and the SI combustion region (excluding the throttle valve fully open region) is divided into a low load region with a low load, a medium load region with a higher load, and a high load with a higher load. If the CI combustion region is set to a low load region, the cylinder internal temperature has not reached so high in the low load region. (Including opening) and introducing high-temperature internal EGR gas into the cylinder, the ignitability is improved and the combustion of the engine can be stabilized.

  On the other hand, since CI combustion is a combustion mode capable of improving fuel consumption, it is preferable to expand the CI combustion region to the high load side, but the CI combustion region exceeds the low load region, and When expanded to the load region or the high load region, the temperature in the cylinder is high in these regions, so that CI combustion becomes combustion accompanied by a sharp pressure rise (dP / dt), which may increase combustion noise. There is. Thus, when the exhaust is opened twice, the high temperature internal EGR gas is introduced into the cylinder, so that the CI combustion becomes a combustion accompanied by a more rapid pressure increase, which may further increase the combustion noise. There is.

  Therefore, in order to improve fuel efficiency and suppress a steep pressure increase, the CI combustion region is expanded to the above-mentioned middle load region, but the region where the exhaust is twice opened is extended to the intermediate load in the CI combustion region (low load in the entire region). It is conceivable to introduce an external EGR gas whose temperature is equal to or lower than the internal EGR into the cylinder in a region exceeding the intermediate load.

  However, in the region of the same combustion form, in other words, even if it is the same CI combustion region, if the method of introducing EGR gas into the cylinder is changed in the middle, before and after the change (before and after opening the exhaust twice) Therefore, there is a possibility that a difference in the ratio of the EGR gas amount to the total gas amount or a torque difference may occur, and it is accompanied by a special difficulty to perform control to eliminate these steps. There is a problem of becoming.

  On the other hand, the CI combustion region is extended to the low load region, the exhaust is opened twice over the entire CI combustion region, and external EGR gas is introduced into the cylinder when switching from CI combustion to SI combustion. In this way, although these problems do not occur, there is a problem that the effect of improving the fuel efficiency due to the improvement of the thermal efficiency is reduced.

  The present invention has been made in view of such points, and the object of the present invention is to provide a CI combustion region in a spark ignition engine that performs CI combustion in a low load region and performs SI combustion in a high load region. The present invention provides a technique that eliminates a torque step or the like in the CI combustion region even when expanding to a high load side, or omits complicated control for eliminating such a torque step or the like.

  In order to achieve the above object, in the spark ignition direct injection engine according to the present invention, even when the CI combustion region is expanded to the high load side, by suppressing combustion accompanied by a steep pressure rise, The exhaust double opening (or the intake double opening) can be executed over the entire region.

Specifically, the first invention has an engine main body having a cylinder forming a combustion chamber at the top and configured to be supplied with fuel mainly composed of gasoline, and opens to the combustion chamber side. An intake valve and an exhaust valve for opening and closing the intake port opening and the exhaust port opening, respectively, a fuel injection valve configured to inject the fuel into the combustion chamber, and a pressure of fuel injected by the fuel injection valve A variable fuel pressure mechanism configured to change, a spark plug disposed to face the combustion chamber, and configured to ignite an air-fuel mixture in the combustion chamber, and introduces exhaust gas into the cylinder. for the introduction of internal EGR means for executing at least one of the opening of the exhaust valve in the opening the intake stroke of the intake valve in the exhaust stroke, at least the fuel injection valve, the fuel pressure varying mechanism, the ignition plug及A controller configured to operate the engine body by controlling the internal EGR introduction means, and in the compression self-ignition combustion region on a load side lower than a predetermined load by the control of the controller. The air-fuel mixture in the combustion chamber is combusted by compression self-ignition, while the air-fuel mixture in the combustion chamber is combusted by spark ignition using the spark plug in the spark ignition combustion region on the high load side above the predetermined load. A spark ignition direct injection engine that operates the internal EGR introduction means in a compression self-ignition combustion region is intended.

In the compression self-ignition combustion region, the controller determines that the fuel injection end time closest to the ignition timing is from the initial intake stroke to the middle of the compression stroke in the predetermined first region of the compression self-ignition combustion region. On the other hand, in the second region of the compressed self-ignition combustion region, which is on the higher load side than the first region , the fuel injection valve is set so as to be within the period from the late stage of the compression stroke to the beginning of the expansion stroke. In the second region , the fuel injection pressure of the fuel injection valve is set to a predetermined pressure of 30 MPa or more using the variable fuel pressure mechanism in the second region , while the fuel injection pressure is lower than the predetermined pressure in the first region. set, the compression self-ignition combustion region in addition, the proportion of the internal EGR gas amount to the total amount of gas introduced into the cylinder, smaller towards the second region than the first region As described above, the internal EGR introduction means is driven so as to increase the intake flow rate as the load increases, and the combustion mode is switched from the compression self-ignition combustion region to the spark ignition combustion region. The operation of the internal EGR introduction means is stopped.

  Here, the “predetermined load” refers to the entire region composed of the CI combustion region and the SI combustion region (excluding the throttle valve fully open region), a low load region with a low load, and a medium load region with a higher load than this, Furthermore, when it is divided into three equally into a high load region with a high load, it may be set to a medium load region.

  Further, “fuel injection closest to the ignition timing” means batch injection in the case of batch injection, and subsequent injection in the case of split injection.

Further, the “ first region of the compression self-ignition combustion region” divides the CI combustion region into a low load region with a low load, a medium load region with a higher load, and a high load region with a higher load. In this case, the low load region and the medium load region may be used, and the “ second region of the compression self-ignition combustion region ” may be a high load region when the CI combustion region is divided into three.

  In addition, “internal EGR gas” refers to exhaust gas that is once introduced from the combustion chamber into the cylinder again without passing through the exhaust passage or the like.

In the first aspect of the invention, in the CI combustion region, since the internal EGR introduction means for executing the opening of the intake valve in the exhaust stroke and / or the opening of the exhaust valve in the intake stroke is operated, the combustion chamber is operated in the exhaust stroke. The high-temperature exhaust gas once pushed out to the intake port and / or the exhaust port is again introduced into the cylinder at a high temperature. Thus, in the first region of the CI combustion region, the in-cylinder temperature does not reach a high temperature, so that the ignitability is improved by introducing the high-temperature internal EGR gas into the cylinder. Further, in the first region in the CI combustion region, the fuel injection end timing closest to the ignition timing is set within the period from the initial stage of the intake stroke to the middle stage of the compression stroke. In other words, the mixture formation period is increased. Therefore, it is possible to promote vaporization of fuel mainly composed of gasoline and to generate a homogeneous air-fuel mixture as a whole to stabilize combustion.

On the other hand, in the second region of the CI combustion region, the temperature in the cylinder is high, so that the CI combustion is likely to become a combustion accompanied by a sharp pressure increase (dP / dt), and the operation of the internal EGR introduction means continues. Thus, when a large amount of high-temperature internal EGR gas is introduced into the cylinder, there is a risk that CI combustion becomes combustion accompanied by a more rapid pressure increase and increases combustion noise. In addition, in the second region in the CI combustion region, the inside of the cylinder is extremely hot as described above. When injection is performed in the cylinder from the initial stage of the intake stroke to the middle stage of the compression stroke, the fuel is discharged. When exposed to high temperature air, abnormal combustion such as pre-ignition may occur.

Here, in the first invention, in the CI combustion region, the load ratio is such that the ratio of the EGR gas amount to the total gas amount introduced into the cylinder is smaller in the second region than in the first region . Since the intake flow rate is increased with the increase, in other words, the amount of internal EGR gas introduced into the cylinder is decreased with the increase of the load, so that the CI combustion region is expanded to the high load side. However, it is possible to suppress the CI combustion from being accompanied by a sharp pressure increase.

In addition, in the second region in the CI combustion region, in order to avoid abnormal combustion such as premature ignition, injection is performed within a period from the late stage of the compression stroke to the early stage of the expansion stroke with a predetermined pressure of 30 MPa or more. Yes. That is, abnormal combustion such as pre-ignition is suppressed by delaying (retarding) the end timing of fuel injection closest to the ignition timing from the latter half of the compression stroke to the beginning of the expansion stroke. Also, during the period from the late stage of the compression stroke to the early stage of the expansion stroke, the piston is located near the compression top dead center, so the combustion chamber is extremely narrow. By injecting the fuel, the intensity of turbulence in the combustion chamber can be increased, the vaporization of the fuel mainly composed of gasoline can be promoted, and the combustion can be stabilized.

As described above, according to the first invention, as the load increases, the amount of internal EGR gas introduced into the cylinder is reduced, and with a predetermined pressure of 30 MPa or more, from the latter stage of the compression stroke to the initial stage of the expansion stroke. Even when the CI combustion region is expanded to the high load side in combination with the injection within the period up to, the intake valve is not opened during the exhaust stroke without causing combustion with a sharp pressure rise. The operation of the internal EGR introduction means that opens the valve and / or the exhaust valve during the intake stroke can be continuously performed over the entire CI combustion region.

  Thus, since the operation of the internal EGR introduction means can be continuously performed over the entire CI combustion region, in other words, the timing for stopping the operation of the internal EGR introduction means is determined by the CI combustion. Since it is possible to coincide with the time when the combustion mode is changed from SI to SI combustion, it is possible to consistently introduce the EGR gas into the cylinder in the highly common CI combustion region. Not only can the control to eliminate the step be omitted, but also the occurrence of a torque step or the like can be suppressed.

  According to a second aspect of the present invention, in the first aspect, the predetermined load is a low load region with a low load, the entire region including the compression self-ignition combustion region and a spark ignition combustion region excluding the throttle valve full open region. And when it is equally divided into a medium load region having a higher load and a high load region having a higher load, the medium load region is set.

  Here, the “throttle valve fully open region” means a region where the ratio of the fresh air amount to the total gas amount introduced into the cylinder becomes 100%. As a result of introducing EGR gas into the cylinder, the throttle valve is It does not mean an area in which intake can be performed with the throttle valve opened, not the throttled state.

  According to the second invention, since the CI combustion region having excellent thermal efficiency is expanded beyond the low load region to the middle load region, it is possible to improve the thermal efficiency of the entire engine, thereby improving fuel efficiency. be able to.

According to a third aspect of the present invention, in the first or second aspect of the present invention, the apparatus further comprises external EGR gas recirculation means configured to be able to introduce low-temperature cooled EGR gas in which the exhaust gas is cooled by heat exchange into the cylinder. The controller drives the external EGR recirculation means so that the cooled EGR gas is introduced into the cylinder in the second region in the compression self-ignition combustion region.

In the second region in the CI combustion region, the inside of the cylinder becomes extremely hot with an increase in load or the like, and there is a high possibility that the CI combustion becomes a combustion with a sharp pressure increase (dP / dt). According to the invention, since the cooled EGR gas that has been actively cooled using, for example, an EGR cooler or the like is introduced into the cylinder, the temperature in the cylinder is lowered to reduce dP / dt during CI combustion. Is possible. Thereby, even when the CI combustion region is expanded to the high load side, it is possible to further suppress the occurrence of combustion accompanied by a sharp pressure increase.

  A fourth invention further includes a variable valve timing mechanism capable of changing a valve opening timing of the intake valve in any one of the first to third inventions, wherein the internal EGR introduction means is provided during an intake stroke. When the exhaust valve is opened, a variable valve timing mechanism is used to adjust the amount of EGR gas introduced into the cylinder.

In the fourth invention, since the internal EGR introduction means opens the exhaust valve during the intake stroke, the high-temperature exhaust gas once pushed out from the combustion chamber to the exhaust port during the exhaust stroke is again in the cylinder at a high temperature. To be introduced. At this time, by adjusting the opening timing of the intake valve by the variable valve timing mechanism, the intake amount introduced into the cylinder is determined. In other words, the EGR gas amount relative to the total gas amount sucked into the cylinder is determined. Since the ratio is inevitably determined, the amount of EGR gas introduced into the cylinder can be adjusted without providing a variable valve timing mechanism on the exhaust side of the valve operating system . It is possible to secure the intake air flow rate in the two regions .

  According to the spark ignition direct injection engine according to the present invention, the amount of internal EGR gas is reduced in accordance with an increase in load in the CI combustion region, and therefore, even when the CI combustion region is expanded to the high load side, the CI It is possible to suppress the combustion from being accompanied by a sharp pressure increase.

Further, in the second region in the CI combustion region, injection is performed within a period from the latter stage of the compression stroke to the early stage of the expansion stroke at a predetermined pressure of 30 MPa or more, thereby suppressing abnormal combustion such as pre-ignition and combustion. It is possible to increase the intensity of turbulence in the room and promote vaporization and atomization of fuel mainly composed of gasoline to stabilize combustion.

  As described above, the reduction of the internal EGR gas amount as the load increases and the latter half of the compression stroke injection with a predetermined pressure of 30 MPa or more are performed, so that the CI combustion region is moved to the high load side. Even in the case of expansion, it is necessary to continue the operation of the internal EGR introduction means over the entire CI combustion region, that is, when the operation to stop the operation of the internal EGR introduction means is switched from the CI combustion to the SI combustion. Thus, not only can the control for eliminating the torque step be omitted, but also the occurrence of the torque step itself can be suppressed.

It is the schematic which shows the structure of the spark ignition type direct injection gasoline engine which concerns on this embodiment. It is a block diagram concerning control of a spark ignition direct injection gasoline engine. It is sectional drawing which expands and shows a combustion chamber. It is a figure which illustrates the operation area | region of a spark ignition type direct injection gasoline engine. FIG. 5A shows an example of the fuel injection timing in the case where the intake stroke injection is performed in the CI combustion region, and an example of the heat generation rate of the CI combustion associated therewith, and FIG. 5B shows the high pressure retarded injection in the CI combustion region. An example of the fuel injection timing in the case of performing the fuel injection, an example of the heat generation rate of the CI combustion associated therewith, FIG. An example of the heat generation rate of SI combustion associated therewith, FIG. 4D shows an example of the fuel injection timing and ignition timing when split injection of intake stroke injection and high pressure retarded injection is performed in the SI combustion region, and the accompanying SI It is an illustration of the heat release rate of combustion. It is a figure which compares the state of SI combustion by a high pressure retarded injection, and the state of conventional SI combustion. These are (a) the gas composition in the cylinder, (b) the compression start temperature, (c) the oxygen concentration, and (d) the external EGR ratio during intake to the difference in engine load. These are (a) the gas composition in the cylinder, (d) the ratio of the external EGR during intake, (e) the exhaust valve timing, (f) the intake valve timing, and (g) the intake valve lift amount with respect to the difference in engine load. Changes in (a) cylinder gas composition, (d) external EGR ratio during intake, (h) throttle opening, (i) EGR valve opening, (j) EGR cooler bypass valve opening with respect to engine load differences It is. Changes in (a) gas composition in the cylinder, (k) fuel injection start timing, (l) fuel pressure, and (m) ignition timing with respect to differences in engine load. It is a figure which shows the relationship between the opening / closing timing of an intake / exhaust valve, and an internal EGR rate. It is a figure which shows the relationship between an EGR rate and engine load in a predetermined rotation speed.

  Embodiments of a spark ignition direct injection gasoline engine according to the present invention will be described below in detail with reference to the drawings.

(overall structure)
FIG. 1 is a schematic diagram showing a configuration of a spark ignition direct injection gasoline engine according to the present embodiment, and FIG. 2 is a block diagram relating to control of the spark ignition direct injection gasoline engine. This engine (engine main body) 1 is a spark ignition direct injection gasoline engine mounted on a vehicle and supplied with fuel mainly composed of gasoline. The engine 1 includes a cylinder block 11 provided with a plurality of cylinders 18, each of which forms a combustion chamber 19 at the top, a cylinder head 12 disposed on the cylinder block 11, and a lower side of the cylinder block 11. And an oil pan 13 in which lubricating oil is stored. In FIG. 1, only one cylinder 18 is illustrated, but the cylinder block 11 is provided with, for example, four cylinders in series.

  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 top surface (crown surface) of the piston 14, as shown in an enlarged view in FIG. 3, a cavity 141 like a reentrant type in a diesel engine is formed. When located in the vicinity, it is opposed to a direct injection injector 67 described later. Thus, a combustion chamber 19 is defined by the cylinder head 12, the cylinder 18, and the piston 14 having the cavity 141. The shape of the combustion chamber 19 is not limited to the shape shown in the figure. 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 are changed as appropriate. It is possible.

  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.

  An intake port 16 and an exhaust port 17 are formed in the cylinder head 12 for each cylinder 18, and the intake port 16 and the exhaust port 17 have peripheral portions of the intake port opening and the exhaust port opening on the combustion chamber 19 side. An intake valve 21 and an exhaust valve 22 that open and close each port opening by abutting a valve seat (not shown) fixed to each other are provided.

  Among the valve systems that drive the intake valve 21 and the exhaust valve 22 respectively, the exhaust side valve system switches the operation mode of the exhaust valve 22 between a normal mode and a special mode, for example, a hydraulically operated variable mechanism. (Variable Valve Lift, hereinafter also referred to as VVL) 71 is provided (see FIG. 2). 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 The lost motion mechanism is configured to selectively transmit the operating state of one of the second cams to the exhaust valve 22.

The exhaust valve 22 operates in a normal mode in which the valve is opened only once during the exhaust stroke when the operating state of the first cam is transmitted, while the exhaust valve 22 operates when the operating state of the second cam is transmitted. It operates in a special mode that opens not only in the stroke but also in the intake stroke, so-called “double exhaust opening”, and these normal mode and special mode can be switched according to the operating state of the engine . Specifically, the special mode is used in the control related to the internal EGR. Here, “internal EGR” means that the burned gas expelled from the combustion chamber 19 to the exhaust port 17 is introduced into the cylinder 18 at a high temperature. In the present embodiment, the internal EGR is performed by opening the exhaust valve 22 in the intake stroke by opening the exhaust valve 22 in the intake stroke by opening the exhaust twice, that is, in the exhaust stroke. However, the present invention is not limited to this. For example, the internal EGR control may be performed by opening the intake valve 21 twice or by opening the intake valve twice. As a result, the intake valve 21, the exhaust valve 22, the VVL 71, the VVT 72, the CVVL 73, and the like, as referred to in the present invention, open the intake valve 21 and the intake stroke in the exhaust stroke in order to introduce the exhaust gas into the cylinder 18. The internal EGR introduction means is configured to execute at least one of the opening of the exhaust valve 22 in FIG.

  In the internal EGR control in which the burned gas remains in the cylinder 18 by 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, the combustion heat of the burned gas is reduced to the cylinder or the like. This is not adopted in the present embodiment because the temperature of the burned gas decreases due to the loss of the burned gas or the burnt gas re-expands, and a pumping loss occurs when the burned gas is recompressed. .

  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 with an electromagnetic actuator may be employed.

  On the other hand, on the intake side of the valve operating system, as shown in FIG. 2, a variable valve timing (hereinafter also referred to as VVT) capable of changing the rotation phase of the intake camshaft with respect to the crankshaft 15. ) 72 and a continuously variable valve lift (hereinafter also referred to as CVVL) 73 capable of continuously changing the lift amount of the intake valve 21. The VVT 72 may adopt a known hydraulic, electromagnetic or mechanical structure as appropriate, and the CVVL 73 may adopt various known structures as appropriate. The detailed structure is not shown. To do. With these VVT 72 and CVVL 73, the intake valve 21 can change its valve opening timing, valve closing timing, and lift amount.

  In addition, a direct injection injector (fuel injection valve) 67 that directly injects fuel into the combustion chamber 19 is attached to the cylinder head 12 for each cylinder 18. As shown in an enlarged view in FIG. 3, the direct injection injector 67 is disposed so that its injection hole faces the inside of the combustion chamber 19 from the central portion of the ceiling surface of the combustion chamber 19. The direct injection injector 67 supplies fuel into the combustion chamber 19 at an injection timing set according to the operating state of the engine 1 and by an amount set according to the operating state of the engine 1 in accordance with a command from the PCM 10 described later. Inject directly. In this example, the direct injection injector 67 is a multi-injection type direct injection injector having a plurality of injection holes, although the detailed illustration is omitted, whereby the fuel spray diffuses radially from the center position of the combustion chamber 19. Inject fuel as is. As indicated by arrows in FIG. 3, the fuel spray that is 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 formed on the top surface of the piston 14. It flows along the wall surface of the cavity 141. In other words, the cavity 141 is formed so as to contain the fuel spray injected at the timing when the piston 14 is positioned near the compression top dead center. This combination of the multi-injection type direct injection injector 67 and the cavity 141 is advantageous in shortening the mixture formation period and the combustion period after fuel injection. The direct injection injector 67 is not limited to a multi-injection type direct injection injector, and may employ an outer valve-open type direct injection injector.

  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 (fuel pressure variable mechanism) 62 including a fuel pump 63 and a common rail 64 that can supply fuel to the direct injection injector 67 at a relatively high fuel pressure is interposed on the fuel supply path. Has been. The fuel pump 63 pumps fuel from the fuel tank to the common rail 64, and the common rail 64 is configured to store the pumped fuel at a relatively high fuel pressure. Then, 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 makes it possible to supply fuel with a high fuel pressure of 30 MPa or more to the direct injection injector 67, and the fuel pressure is set to a maximum of about 120 MPa. Can do. The pressure of the fuel supplied to the direct injection injector 67 is changed according to the operating state of the engine 1 as will be described later. The high pressure fuel supply system 62 is not limited to this configuration.

  Further, as shown in FIG. 3, a spark plug 25 for igniting the air-fuel mixture in the combustion chamber 19 is attached to the cylinder head 12. In this example, the spark plug 25 penetrates through the cylinder head 12 so as to extend obliquely downward from the exhaust side of the engine 1, and its tip faces the cavity 141 of the piston 14 located at the compression top dead center. Are arranged as follows.

  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, while the combustion of each cylinder 18 is connected to the other side of the engine 1. An exhaust passage 40 for discharging burned gas (exhaust gas) from the chamber 19 is connected.

  An air cleaner 31 for filtering the intake air is disposed at the upstream end of the intake passage 30, while a surge tank 33 is disposed near the downstream end of the intake passage 30. Further, the downstream intake passage 30 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. Further, an intercooler bypass passage 35 that bypasses the intercooler / warmer 34 is connected to the intake passage 30, and the intercooler bypass passage 35 is connected to an intercooler for adjusting the flow rate of air passing through the passage 35. A cooler bypass valve 351 is provided. The temperature of fresh air introduced into the cylinder 18 is adjusted by adjusting the ratio between the flow rate of the intercooler bypass passage 35 and the flow rate of the intercooler / warmer 34 through the opening adjustment of the intercooler bypass valve 351. It is possible.

  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. As a result, the EGR passage 50 including the main passage 51 and the EGR cooler bypass passage 53, the EGR cooler 52, the EGR valve 511, the EGR cooler bypass valve 531 and the like are used in the present invention to exchange the exhaust gas by heat exchange. An external EGR gas recirculation means configured to be able to introduce the cooled low-temperature cooled EGR gas into the cylinder 18 is configured.

  The engine 1 configured as described above 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 controller as referred to in the present invention.

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, on the downstream side of the air cleaner 31, an air flow sensor SW1 that detects the flow rate of fresh air, an intake air temperature sensor SW2 that detects the temperature of fresh air, and a downstream side of the intercooler / warmer 34, the intercooler / warmer 34 is disposed. A second intake air temperature sensor SW3 for detecting the temperature of fresh air after passing, an EGR gas temperature sensor SW4 for detecting the temperature of the external EGR gas, arranged near the connection portion of the EGR passage 50 with the intake passage 30, an intake port 16, an intake port temperature sensor SW5 that detects the temperature of the intake air just before flowing into the cylinder 18, an in-cylinder pressure sensor SW6 that detects the pressure in the cylinder 18, and an EGR in the exhaust passage 40. An exhaust temperature sensor that is disposed near the connection portion of the passage 50 and detects the exhaust temperature and the exhaust pressure, respectively. Sa SW7 and exhaust pressure sensor SW8, disposed upstream of the direct catalyst 41, the linear O ₂ sensor SW9 for detecting the oxygen concentration in the exhaust gas, is arranged between the direct catalyst 41 and underfoot catalyst 42, Operation of a lambda O ₂ sensor SW10 that detects the oxygen concentration in the exhaust, a water temperature sensor SW11 that detects the temperature of engine cooling water, a crank angle sensor SW12 that detects the rotation angle of the crankshaft 15, and an accelerator pedal (not shown) of the vehicle An accelerator opening sensor SW13 for detecting an accelerator opening corresponding to the amount, cam angle sensors SW14 and SW15 on the intake side and exhaust side, and a common rail 64 of the high-pressure fuel supply system 62 are supplied to a direct injection injector 67. The fuel pressure sensor SW16 detects the fuel pressure.

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

(Outline of engine control)
FIG. 4 shows an example of the operating region of the engine. The engine 1 has a relatively low engine load, specifically, a lower load than a predetermined load (predetermined load T5 in FIG. 7A) for the purpose of improving fuel consumption and reducing emissions. In the region on the load side, compression ignition combustion (hereinafter also referred to as CI combustion) in which combustion is performed by compression self-ignition (Homogenious Charge Compression Ignition) without ignition by the spark plug 25 is performed. However, as the load on the engine 1 is increased, the combustion becomes too steep in the CI combustion, which may cause problems such as combustion noise. Therefore, in this engine 1, in a high load region where the engine load is relatively high, specifically, in a region on the high load side that is equal to or higher than a predetermined load, CI combustion is stopped and spark ignition (Spark) using the spark plug 25 is performed. Ignition) spark ignition combustion (hereinafter also referred to as SI combustion). In other words, the engine 1 is configured to switch between a CI mode in which compression ignition combustion is performed and an SI mode in which spark ignition combustion is performed 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.

  The CI mode is further divided into three areas according to the engine load. Specifically, in the region (1) where the load is lowest 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 CI combustion. As will be described in detail later, the introduction of the hot EGR gas is performed by turning on the VVL 71 and opening the exhaust valve 22 twice during the intake stroke. Such introduction of hot EGR gas is advantageous in increasing the compression end temperature in the cylinder 18 and improving the ignitability and stability of the compression ignition combustion in the light load region (1). Further, in the region (1), as shown in FIG. 5A, the direct injection injector 67 injects fuel into the cylinder 18 at least during the period from the initial stage of the intake stroke to the middle stage of the compression stroke. A lean mixture is formed. The excess air ratio λ of the air-fuel mixture may be set to, for example, 2.4 or more (for example, 2.5), and as a result, the combustion temperature is lowered, the generation of Raw NOx is suppressed, and the exhaust emission performance is reduced. Can be increased. Then, as shown in FIG. 5A, the lean air-fuel mixture undergoes compression self-ignition near the compression top dead center.

  Although details will be described later, in the region (1) where the load is high (region where the predetermined load T1 is greater than or equal to the predetermined load T2 in FIG. 7A), at least in the period from the intake stroke to the middle of the compression stroke, the cylinder 18 Although the fuel is injected into the air-fuel ratio, the air-fuel ratio of the air-fuel mixture is set to the stoichiometric air-fuel ratio (λ≈1). In this way, by setting the stoichiometric air-fuel ratio, the three-way catalyst can be used, and as will be described later, the control at the time of switching between the SI mode and the CI mode is simplified, and the CI mode is further increased. It also contributes to the ability to expand to the load side.

  In the CI mode, in the region (2) where the load is higher than the region (1), the fuel is injected into the cylinder 18 during the period from the initial stage of the intake stroke to the middle stage of the compression stroke, as in the high load side of the region (1). The fuel is injected (see FIG. 5A) to form a homogeneous air / fuel ratio (λ≈1).

  In the region (2), the temperature in the cylinder 18 naturally increases as the engine load increases, so the hot EGR gas amount is reduced to avoid pre-ignition. Although this will be described in detail later, this is due to the adjustment of the amount of internal EGR gas introduced into the cylinder 18.

  Further, in the region (2), the ratio of the cooled EGR gas amount to the total gas amount of the external EGR gas (cooled EGR gas) cooled mainly through the EGR cooler 52 and introduced into the cylinder 18 is relatively low. Is introduced into the cylinder 18 so as to gradually increase as the load increases. Thus, by introducing the hot hot EGR gas and the cold cooled EGR gas into the cylinder 18 at an appropriate ratio, the compression end temperature in the cylinder 18 is made appropriate, and the rapid ignition while ensuring the ignitability of the compression ignition. Avoid combustion and stabilize compression ignition combustion. The EGR rate as a ratio of the EGR gas to the total amount of gas introduced into the cylinder 18 including the hot EGR gas and the cooled EGR gas is possible under the condition that the air-fuel ratio of the mixture is set to λ≈1. The EGR rate is set as high as possible. Therefore, in the region (2), the fuel injection amount increases as the engine load increases, and the intake air amount increases accordingly, so that the EGR rate gradually decreases.

The region (1) and the region (2) correspond to the “ first region of the compression self-ignition combustion region ” as referred to in the present invention, and the region (3) described later refers to the present invention. , “ The second region of the compression self-ignition combustion region ”.

  Here, if the CI combustion region is set in the low load region (region (1) and region (2)), even if the exhaust is opened twice over the entire CI combustion region, the high temperature The internal EGR gas improves the ignitability and stabilizes the combustion of the engine, so there is no problem, but if the CI combustion region is extended beyond the low load region to the high load side (region (3)), the following There is a problem like this. That is, in the region (3) where the load is highest in the CI mode including the boundary line between the CI mode and the SI mode, the compression end temperature in the cylinder 18 is further increased, so that the CI combustion has a sharp pressure rise (dP / Dt), the combustion noise may increase. Thus, when the exhaust is opened twice, the high-temperature internal EGR gas is introduced into the cylinder 18, so that the CI combustion becomes combustion accompanied by a more rapid pressure rise, and the combustion noise is further increased. There is a fear. In addition, if fuel is injected into the cylinder 18 in the period from the initial stage of the intake stroke to the middle stage of the compression stroke as in the region (1) and the region (2), abnormal combustion such as pre-ignition occurs. On the other hand, if a large amount of cool EGR gas having a low temperature is introduced to lower the compression end temperature in the cylinder 18, the ignitability of the compression ignition is deteriorated.

  Here, in order to suppress a steep pressure rise while improving fuel efficiency, the CI combustion region is expanded to the region (3), but the region where the exhaust is opened twice is limited to the region (2), and the region (3) In this case, it is conceivable to introduce an external EGR gas whose temperature is equal to or lower than the internal EGR into the cylinder. However, if the method of introducing the hot EGR gas into the cylinder 18 is changed in the middle of the same CI combustion region, Before and after the change (before and after opening the exhaust twice), there may be a step in the ratio of the EGR gas amount to the total gas amount or a torque step, and these steps may be eliminated. There is a problem in that control involves special difficulties. On the other hand, the CI combustion region is limited to region (2), the exhaust is opened twice over the entire CI combustion region, and external EGR gas is introduced into the cylinder 18 when switching from CI combustion to SI combustion. By doing so, although these problems do not occur, there is a problem that the effect of improving the fuel efficiency due to the improvement of thermal efficiency is reduced.

Therefore, even when the CI combustion region is expanded to the region (3), the exhaust and double combustion can be executed over the entire region of the CI combustion region while suppressing the combustion accompanied by the steep pressure rise and the abnormal combustion.
In this region (3), the ratio of the high-temperature internal EGR gas amount to the total gas amount introduced into the cylinder 18 is smaller in the region (3) than in the region (1) and the region (2). In addition to increasing the intake air flow rate as the load increases, the fuel injection pressure of the direct injection injector 67 is set to a predetermined pressure of 30 MPa or more using the high-pressure fuel supply system 62, and FIG. As shown in FIG. 6, the direct injection injector 67 is driven so that the end timing of fuel injection closest to the ignition timing is within the period from the latter half of the compression stroke to the early stage of the expansion stroke. This characteristic fuel injection mode is hereinafter referred to as “high pressure retarded injection” or simply “retarded injection”. Details of the high-pressure retarded injection will be described later.

  Thus, in the engine 1, in the CI mode, the ratio of the hot EGR gas amount to the total gas amount introduced into the cylinder 18 is reduced as the load increases, so the CI combustion region is shifted to the high load side. Even in the case of expansion, it is possible to suppress the CI combustion from being accompanied by a sharp pressure increase. In addition, by delaying (retarding) the end timing of the fuel injection closest to the ignition timing from the late compression stroke period to the early expansion stroke period, abnormal combustion such as premature ignition is suppressed. Thus, during the period from the late stage of the compression stroke to the early stage of the expansion stroke, the piston 14 is located near the compression top dead center, and thus the combustion chamber is extremely narrow. By injecting fuel at a high pressure, the intensity of turbulence in the combustion chamber can be increased, fuel atomization of fuel mainly composed of gasoline can be promoted, and combustion can be stabilized.

  As described above, in combination with reducing the amount of internal EGR gas introduced into the cylinder 18 as the load increases, and performing the latter half of the compression stroke with a predetermined pressure of 30 MPa or more, Even when the CI combustion region is expanded to the high load side, it is possible to continuously open the exhaust gas twice over the entire CI combustion region without causing combustion with a sharp pressure increase. It becomes.

  In this way, since it is possible to continuously perform the exhaust double opening over the entire CI combustion region, in other words, the timing for stopping the exhaust double opening is switched from the CI mode to the SI mode. Since it is possible to match the timing, it is possible to consistently introduce the EGR gas into the cylinder in the highly common CI combustion region, thereby eliminating the control to eliminate the torque step. Not only can this be done, but it is also possible to suppress the occurrence of torque steps and the like.

  In the region (3), the air-fuel ratio of the air-fuel mixture is set to the stoichiometric air-fuel ratio (λ≈1) as in the region (2). This makes it possible to use a three-way catalyst, which is advantageous for improving the emission performance. Thus, in the region (3), the hot EGR gas amount (internal EGR gas amount) is reduced as the load increases to avoid premature ignition as described above, while the cooled EGR gas amount (external EGR gas amount). ) And the hot hot EGR gas and the cold cooled EGR gas are introduced into the cylinder 18 at an appropriate ratio. As a result, the compression end temperature in the cylinder 18 is appropriately set to stabilize the compression ignition combustion. The EGR rate as a ratio of the EGR gas to the total amount of gas introduced into the cylinder 18 including the hot EGR gas and the cooled EGR gas is possible under the condition that the air-fuel ratio of the mixture is set to λ≈1. The EGR rate is set as high as possible. Therefore, also in the region (3), the fuel injection amount increases as the engine load increases, and the intake air amount increases accordingly, so the EGR rate gradually decreases.

  The CI mode is divided into three regions according to the engine load level, while the SI mode is divided into two regions (4) and (5) according to the engine speed. ing. The region (4) corresponds to a low speed region when the operation region of the engine 1 is divided into a low speed and a high speed in the illustrated example, and the region (5) corresponds to a high speed region. Further, the boundary between the region (4) and the region (5) is inclined in the rotational speed direction with respect to the load level in the operation region shown in FIG. 4, but the region (4) and the region (5) The boundary is not limited to the illustrated example.

  In each of the region (4) and the region (5), the air-fuel mixture is set to the stoichiometric air-fuel ratio (λ≈1) as in the regions (2) and (3). Therefore, the air-fuel ratio of the air-fuel mixture is made constant at the theoretical air-fuel ratio (λ≈1) across the boundary between the CI mode and the SI mode. This allows the use of a three-way catalyst. In the region (4) and the region (5), the details will be described later, but basically, the throttle valve 36 is fully opened, and the fresh air introduced into the cylinder 18 by adjusting the opening of the EGR valve 511. Adjust the amount and external EGR gas amount. In this way, by adjusting the ratio of the gas introduced into the cylinder 18 so that the air-fuel ratio of the air-fuel mixture is maintained constant, the pump loss is reduced and a large amount of EGR gas is introduced into the cylinder 18. Therefore, the combustion temperature of the spark ignition combustion is kept low, and the cooling loss is also reduced. In the region (4) and the region (5), the external EGR gas cooled mainly through the EGR cooler 52 is introduced into the cylinder 18. This is advantageous for avoiding abnormal combustion and also has an advantage of suppressing generation of Raw NOx. 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 region (4) and the region (5) of the SI mode, abnormal combustion is avoided by performing the above-described high-pressure retarded injection. More specifically, in the region (4), fuel is injected into the cylinder 18 with a high fuel pressure of 30 MPa or more and within the retard period from the late compression stroke to the early expansion stroke, as shown in FIG. 5 (c). Only high-pressure retarded injection is performed. On the other hand, in the region (5), as shown in FIG. 5 (d), a part of the fuel to be injected is injected into the cylinder 18 within the intake stroke period in which the intake valve 21 is open. The remaining fuel is injected into the cylinder 18 within the retard period. That is, in the region (5), fuel split injection is performed. Here, the intake stroke period during which the intake valve 21 is open is not a period defined based on the piston position, but a period defined based on the opening and closing of the intake valve 21, and the intake stroke referred to here is: Depending on the closing timing of the intake valve 21 that is changed by the CVVL 72 or the VVT 73, the piston 14 may deviate from the time when the piston 14 has reached the intake bottom dead center.

  Next, the high pressure retarded injection in the SI mode will be described with reference to FIG. FIG. 6 shows the heat generation rate (upper diagram) and the progress of the unburned mixture reaction in the SI combustion (solid line) by the high pressure retarded injection described above and the conventional SI combustion (broken line) in which fuel injection is performed during the intake stroke. It is a figure which compares the difference in a degree (lower figure). The horizontal axis in FIG. 6 is the crank angle. As a precondition for this comparison, the operating state of the engine 1 is both a high-load low-speed region (that is, region (4)), and the amount of fuel to be injected is the case of SI combustion by high-pressure retarded injection and conventional SI combustion. They are the same as each other.

  First, in the conventional SI combustion, a predetermined amount of fuel is injected into the cylinder 18 during the intake stroke (broken line in the upper diagram). In the cylinder 18, a relatively homogeneous air-fuel mixture is formed after the fuel injection until the piston 14 reaches the compression top dead center. In this example, ignition is executed at a predetermined timing indicated by a white circle after the compression top dead center, thereby starting combustion. After the start of combustion, as shown by the broken line in the upper diagram of FIG. 6, the combustion ends through a peak of the heat generation rate. The period from the start of fuel injection to the end of combustion corresponds to the reaction possible time of the unburned mixture (hereinafter also simply referred to as the reaction possible time). As shown by the broken line in the lower diagram of FIG. Qi reaction gradually progresses. The dotted line in the figure shows the ignition threshold, which is the reactivity with which the unburned mixture reaches ignition, and the conventional SI combustion has a very low reaction time in combination with the low speed range. In the meantime, the reaction of the unburned mixture continues to progress during that time, so the reactivity of the unburned mixture exceeds the ignition threshold before and after ignition, and abnormal combustion such as premature ignition or knocking occurs. cause.

  On the other hand, the high pressure retarded injection aims to shorten the reaction possible time, thereby avoiding abnormal combustion. That is, as shown in FIG. 6, the reaction possible time is a period during which the direct injection injector 67 injects fuel ((1) injection period), and after completion of the injection, a combustible air-fuel mixture is present around the spark plug 25. The sum of the period until formation ((2) mixture formation period) and the period until combustion ended by ignition ((3) combustion period), that is, (1) + (2) + (3). The high-pressure retarded injection shortens the injection period, the mixture formation period, and the combustion period, thereby shortening the reaction time. This will be described in order.

  First, the high fuel pressure relatively increases the amount of fuel injected from the direct injection injector 67 per unit time. For this reason, when the fuel injection amount is constant, the relationship between the fuel pressure and the fuel injection period is generally such that the lower the fuel pressure, the longer the injection period, and the higher the fuel pressure, the shorter the injection period. Therefore, the high pressure retarded injection in which the fuel pressure is set to be significantly higher than the conventional one shortens the injection period.

  Further, the high fuel pressure is advantageous for atomization of the fuel spray injected into the cylinder 18 and makes the flight distance of the fuel spray longer. For this reason, the relationship between the fuel pressure and the fuel evaporation time is generally longer as the fuel pressure is lower, and the fuel evaporation time is longer as the fuel pressure is higher. Further, the time until the fuel spray reaches the fuel pressure and the spark plug 25 is generally longer as the fuel pressure is lower, and the time until the fuel spray is higher as the fuel pressure is higher. The air-fuel mixture formation period is a time obtained by adding the fuel evaporation time and the fuel spray arrival time around the spark plug 25. Therefore, the higher the fuel pressure, the shorter the air-fuel mixture formation period. Therefore, the high pressure retarded injection in which the fuel pressure is set to be significantly higher than the conventional one shortens the mixture formation period as a result of the fuel evaporation time and the fuel spray arrival time around the spark plug 25 being shortened. On the other hand, as shown by white circles in the figure, the conventional intake stroke injection at a low fuel pressure significantly increases the mixture formation period. In the SI mode, the combination of the multi-injection type direct injection injector 67 and the cavity 141 shortens the time until the fuel spray reaches around the spark plug 25 after the fuel is injected. It is effective for shortening the period.

  Thus, shortening 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. Therefore, in the high pressure retarded injection, as shown in the upper diagram of FIG. 6, fuel injection is performed within the retard period from the latter stage of the compression stroke to the early stage of the expansion stroke. As the fuel is injected into the cylinder 18 at a high fuel pressure, the turbulence in the cylinder 18 becomes stronger and the turbulence energy in the cylinder 18 increases. This high turbulence energy is relatively late in fuel injection timing. Combined with the timing setting, it is advantageous for shortening the combustion period.

  That is, when the fuel injection is performed within the retard period, the relationship between the fuel pressure and the turbulent energy in the combustion period is generally lower as the fuel pressure is lower and the turbulent energy is lower as the fuel pressure is higher. Becomes higher. Here, even if the fuel is injected into the combustion chamber 19 at a high fuel pressure, if the injection timing is in the intake stroke, the time until the ignition timing is long, or the cylinder in the compression stroke after the intake stroke The disturbance in the cylinder 18 is attenuated due to the compression in the cylinder 18. As a result, when fuel is injected during the intake stroke, the turbulent energy during the combustion period becomes relatively low regardless of the fuel pressure level.

  In general, the relationship between the turbulent energy and the combustion period in the combustion period is such that the lower the turbulent energy, the longer the combustion period, and the higher the turbulent energy, the shorter the combustion period. Therefore, the relationship between the fuel pressure and the combustion period is such that the lower the fuel pressure, the longer the combustion period, and the higher the fuel pressure, the shorter the combustion period. That is, the high pressure retarded injection shortens the combustion period. In contrast, the conventional intake stroke injection at a low fuel pressure has a longer combustion period. The multi-injection type direct injection injector 67 is advantageous in improving the turbulence energy in the cylinder 18 and is effective in shortening the combustion period. In addition, the multi-injection type direct injection injector 67 and the cavity 141 It is also effective for shortening the combustion period to contain the fuel spray in the cavity 141 by combination.

  As described above, the high pressure retarded injection shortens the injection period, the mixture formation period, and the combustion period, and as a result, as shown in FIG. 6, the fuel injection start timing SOI to the combustion end timing θend are not changed. It is possible to significantly shorten the reaction time of the fuel mixture as compared with the case of fuel injection during the conventional intake stroke. As a result of shortening this reaction possible time, as shown in the upper diagram of FIG. 6, in the conventional intake stroke injection at a low fuel pressure, as shown by a white circle, the reaction progress of the unburned mixture at the end of combustion is shown. However, the ignition threshold value is exceeded and abnormal combustion occurs, whereas high-pressure retarded injection suppresses the progress of the reaction of the unburned mixture at the end of combustion, as shown by the black circle. Combustion can be avoided. It should be noted that the ignition timing is set to the same timing in the white circle and the black circle in the upper diagram of FIG.

  By setting the fuel pressure to, for example, 30 MPa or more, the combustion period can be effectively shortened. 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 is mainly composed of 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. The retarding of the ignition timing suppresses the progress of the reaction by suppressing the increase in the temperature and pressure of the unburned mixture. However, retarding the ignition timing leads to a decrease in thermal efficiency and torque, whereas when performing high-pressure retarded injection, the ignition timing can be advanced by an amount that avoids abnormal combustion by devising the form of fuel injection. Since it is possible, 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 for improving fuel efficiency.

  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 (3) 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 a relatively high load region, for example, while preventing premature ignition during the compression stroke period, as described above. Since a substantially homogeneous air-fuel mixture is quickly formed, it is possible to reliably perform compression ignition 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 dt) becomes steep. Thus, as a result of eliminating the NVH restriction, the CI mode region is expanded to the high load side.

  Returning to the description of the SI mode, as described above, the high pressure retarded injection in the SI mode shortens the reaction time of the unburned mixture by performing the fuel injection within the retard period. In the low speed range where the engine 1 has a relatively low rotational speed, the actual time for the crank angle change is long, which is effective. On the other hand, in the high speed range where the engine 1 has a relatively high rotational speed, the actual Not very effective due to short time. On the contrary, in the retard injection, since the fuel injection timing is set near the compression top dead center, in-cylinder gas not containing fuel, in other words, air having a high specific heat ratio is compressed in the compression stroke. As a result, in the high speed region, the compression end temperature in the cylinder 18 becomes high, and this high compression end temperature causes knocking. Therefore, when only the retard injection is performed in the region (5), it may occur that the ignition timing must be retarded to avoid knocking.

  Therefore, as shown in FIG. 4, in the region (5) where the rotational speed is relatively high in the SI mode, as shown in FIG. 5 (d), a part of the fuel to be injected is cylinder 18 within the intake stroke period. And the remaining fuel is injected into the cylinder 18 within the retard period. In the intake stroke injection, it is possible to lower the specific heat ratio of the in-cylinder gas (that is, the air-fuel mixture containing fuel) during the compression stroke, thereby keeping the compression end temperature low. Thus, since the compression end temperature is lowered, knocking can be suppressed, so that the ignition timing can be advanced.

  Further, by performing the high pressure retarded injection, as described above, the turbulence becomes strong in the cylinder 18 (combustion chamber 19) near the compression top dead center, and the combustion period is shortened. This is also advantageous in suppressing knocking, and the ignition timing can be further advanced. Thus, in the region (5), by performing split injection of intake stroke injection and high pressure retarded injection, it is possible to improve thermal efficiency while avoiding abnormal combustion.

  In order to shorten the combustion period in the region (5), a multi-point ignition configuration may be adopted instead of performing high pressure retarded injection. That is, a plurality of ignition plugs are arranged facing the combustion chamber 19, and in the region (5), the intake stroke injection is performed and each of the plurality of ignition plugs is driven to perform multipoint ignition. . By doing so, since the flame spreads from each of the plurality of fire types in the combustion chamber 19, the flame spreads quickly and the combustion period is shortened. As a result, the combustion period is shortened similarly to the case where high pressure retarded injection is employed, which is advantageous for improving the thermal efficiency.

(Specific control procedure)
FIGS. 7 to 10 show examples of control of each parameter of the engine 1 with respect to the engine load in the low speed range. The change of the load in the direction from the low load to the high load is shown in the engine operation map shown in FIG. Is exemplified by a dashed-dotted arrow.

  7A to 7D relate to the state in the cylinder 18, FIG. 7A shows the gas composition (gas ratio) in the cylinder 18, and FIG. 7B shows the inside of the cylinder 18 at the start of compression. (C) in FIG. 2 shows the oxygen concentration. FIG. 7D shows the ratio of external EGR during intake, which can be said to be the amount obtained by removing the internal EGR gas from the EGR gas introduced into the cylinder 18.

  FIGS. 8A and 8D are the same as FIGS. 7A and 7D and show the gas composition in the cylinder 18 and the external EGR ratio during intake, respectively. 8 (e) to 8 (g) relate to the control of the valve operating system, FIG. 8 (e) shows the opening / closing timing of the exhaust valve 22, FIG. 8 (f) shows the opening / closing timing of the intake valve 21, and FIG. g) is the lift amount of the intake valve 21.

  9 (a) and 9 (d) are the same as FIGS. 7 (a) and 7 (d). 9 (h) to 9 (j) relate to the control of the intake / exhaust system, FIG. 9 (h) shows the opening of the throttle valve 36, FIG. 9 (i) shows the opening of the EGR valve 511, and FIG. j) shows the opening of the EGR cooler bypass valve 531.

  Further, FIG. 10A is also the same as FIG. 7A and shows the gas composition in the cylinder 18. 10 (k) to 10 (m) relate to the control of the fuel injection and ignition system, FIG. 10 (k) shows the injection start timing, FIG. 10 (l) shows the fuel pressure, and FIG. 10 (m) shows the ignition timing. Respectively.

  FIG. 7A shows the state in the cylinder 18 as described above. The region on the left side of the figure below the predetermined load T5 is in the CI mode, and the load on the right side of the figure is higher than the predetermined load T5. The area is in SI mode. Although not shown, the fuel amount (total fuel amount) injected into the cylinder 18 is increased as the load increases regardless of the CI mode and the SI mode.

(Up to a predetermined load T1)
In the CI mode, in a region where the load is lower than the predetermined load T1 (this corresponds to the region (1) in the operation map in FIG. 4), fresh air and internal EGR gas are introduced so as to become a lean mixture. The Specifically, the opening of the throttle valve 36 is set to fully open as shown in FIG. 9 (h), while the exhaust VVL 71 is turned on and the exhaust valve 22 is turned on as shown in FIG. 8 (e). The exhaust is opened twice during the intake stroke. Further, as shown in FIG. 8G, the internal EGR rate (the ratio of the internal EGR gas amount introduced into the cylinder 18) becomes the highest by setting the lift amount of the intake valve 21 to the minimum. (See also S1 in FIG. 11). As described above, in the region (1), for example, a lean air-fuel mixture with an excess air ratio of λ ≧ 2.4 may be used, and this is in parallel with introducing a large amount of internal EGR gas into the cylinder 18. As a result, the combustion temperature decreases and the production of Raw NOx is suppressed. Introducing a large amount of EGR gas into the cylinder 18 is also advantageous in reducing pump loss. As shown in FIGS. 10 (k) and 10 (l), in the region (1), fuel injection is executed within the intake stroke period at a relatively low fuel pressure. However, the fuel pressure gradually increases as the engine load increases.

  Up to a predetermined load T1 (more precisely, to a predetermined load T2), a large amount of internal EGR gas is introduced into the cylinder 18, and as shown in FIG. The compression end temperature is increased, which is advantageous for improving the ignitionability of compression ignition and improving the stability of compression ignition combustion. Here, the temperature inside the cylinder 18 is increased by introducing high-temperature internal EGR gas into the cylinder 18, and the combustion temperature is lowered by introducing a large amount of internal EGR gas into the cylinder 18 as described above. It seems that contradicting the generation of Raw NOx is contradictory. However, RAW NOx, when the air temperature exceeds approximately 1800K, the amount of generation suddenly increases. The internal EGR gas burns once, so it hardly contributes to new combustion, and fresh air burns. Since the combustion heat at that time is used to raise a large amount of internal EGR gas (occupying about 80% of the total gas amount), the combustion temperature is lower than when only fresh air is present in the cylinder 18. (For example, about 1500K), the generation of Raw NOx is suppressed, so both are consistent.

  Further, as shown in FIG. 7B, the oxygen concentration gradually decreases as the load increases. Although illustration is omitted, in the low load to medium load region up to a predetermined load T5 where hot EGR gas is introduced into the cylinder 18, the intercooler / warmer 34 is heated by closing the intercooler bypass valve 351. Fresh air may be introduced into the cylinder 18.

(From predetermined load T1 to T2)
At an engine load equal to or higher than the predetermined load T1, the air-fuel ratio of the air-fuel mixture is set to the stoichiometric air-fuel ratio (λ≈1). Therefore, as the amount of injected fuel increases, the amount of fresh air introduced into the cylinder 18 also increases, and the EGR rate decreases accordingly (see FIG. 7A). Even at the predetermined loads T1 to T2, fuel injection is performed within the intake stroke period at a relatively low fuel pressure (see FIGS. 10 (k) and 10 (l)).

  Also, at predetermined loads T1 to T2, as shown in FIG. 9 (h), the throttle opening is basically fully open. On the other hand, as shown in FIG. 8E, with the exhaust VVL 71 turned on, as shown in FIGS. 8F and 8G, the valve opening timing and lift amount of the intake valve 21 are adjusted. As a result, the amount of fresh air introduced into the cylinder 18 and the amount of internal EGR gas are adjusted.

  Specifically, as shown in FIG. 11, if the lift amount of the intake valve 21 is minimized while the exhaust VVL 71 is turned on and the exhaust is opened twice (see S <b> 1 in FIG. 11), the internal EGR The rate is maximized and the amount of fresh air introduced into the cylinder 18 is minimized. This corresponds to the control of the intake valve 21 and the exhaust valve 22 up to a predetermined load T1, as shown in FIGS. 8 (e), 8 (f), and 8 (g).

  As shown in S2 of FIG. 11, if the lift amount of the intake valve 21 is increased while the exhaust is opened twice, the valve opening period of the intake valve 21 and the opening of the exhaust valve 22 when the exhaust valve 22 is opened twice are opened. Since the overlap with the period changes, the internal EGR rate decreases. The closing timing of the intake valve 21 is set to be substantially constant even when the lift amount of the intake valve 21 changes. If the lift amount of the intake valve 21 is continuously changed by the control of the CVVL 73 and the VVT 72, the internal EGR rate can be continuously reduced. Between the predetermined loads T1 and T2, the intake valve is arranged so that the EGR rate is maximized while maintaining the theoretical air-fuel ratio λ≈1, in other words, as much internal EGR gas as possible is introduced into the cylinder 18. The lift amount of 21 is controlled. Specifically, as shown in FIGS. 8 (e), 8 (f), and 8 (g), the lift amount of the intake valve 21 is gradually increased, and accordingly, the valve opening timing (IVO) of the intake valve 21 is increased. ) Is also gradually advanced.

(Predetermined loads T2 to T3)
The engine load equal to or higher than the predetermined load T2 corresponds to the region (2) in the operation map in FIG. Therefore, when the engine load is equal to or higher than the predetermined load T2, the internal EGR gas amount is reduced, and cooled external EGR gas (that is, cooled EGR gas) is introduced into the cylinder 18 instead. That is, as shown in FIG. 9 (i), the opening degree of the EGR valve 511 is gradually increased from the closed state, and thereby the amount of external EGR gas cooled by passing through the EGR cooler 52 becomes the load of the engine 1. The amount is gradually increased along with the increase in. As shown in FIG. 9 (j), the EGR cooler bypass valve 531 remains closed. Thus, the amount of the cooled EGR gas is gradually increased as the engine load increases (see also FIG. 7 (d)).

  On the other hand, as shown in FIG. 7A, the EGR rate including the internal EGR gas and the external EGR gas is equal to the stoichiometric air-fuel ratio (λ≈1) even on the high load side of the predetermined load T2 or higher. Therefore, the load decreases at a predetermined rate with respect to the increase in load. For this reason, the internal EGR gas is reduced with an increase in load at a higher load side that is equal to or higher than the predetermined load T2 (that is, the slope in FIG. 7A increases). Specifically, as shown in FIGS. 8 (e), 8 (f) and 8 (g), the lift amount of the intake valve 21 is increased at a higher rate than the low load side up to the predetermined load T2, As the load increases, the intake valve 21 is gradually increased, and the valve opening timing (IVO) of the intake valve 21 is gradually advanced accordingly.

  Thus, as shown in FIG. 7B, the temperature in the cylinder 18 gradually decreases as the load increases on the high load side of the predetermined load T2 or higher.

(Predetermined loads T3 to T4)
In order to unify the method of introducing the hot EGR into the cylinder 18 over the entire CI mode, the exhaust double opening is continuously performed until the switching boundary (predetermined load T4) between the CI mode and the SI mode. Therefore, in order to lower the temperature in the cylinder 18, as shown in FIG. 9 (i), the opening degree of the EGR valve 511 is gradually increased at the same increase rate from the predetermined load T2 to T3, thereby the cooled EGR gas amount. However, the amount is gradually increased as the load of the engine 1 increases, as shown in FIG. As shown in FIG. 9 (j), the EGR cooler bypass valve 531 remains closed.

  On the other hand, as shown in FIG. 7A, the EGR rate including the internal EGR gas and the external EGR gas is equal to the stoichiometric air-fuel ratio (λ≈1), even on the high load side of the predetermined load T3 or higher. Therefore, as the load increases, the load decreases at the same predetermined rate as the predetermined loads T2 to T3. Further, as shown in FIG. 7B, the temperature in the cylinder 18 gradually decreases as the load increases on the high load side of the predetermined load T3 or higher, while the oxygen concentration is shown in FIG. 7C. Thus, it gradually increases as the load increases.

  As described above, the cooled EGR gas is introduced into the cylinder 18 to suppress premature ignition even on the high load side of the predetermined load T3 or higher. However, in the engine load of the CI mode higher than the predetermined load T3, the cooled EGR is used. By adjusting the introduction ratio of the gas and the hot EGR gas, it becomes difficult to achieve both the ignitability of compression ignition and the avoidance of abnormal combustion such as premature ignition. Perform high-pressure retarded injection. This corresponds to the region (3) in the operation map of FIG.

  As shown in FIG. 10 (k), the fuel injection start time is largely changed from the time during the intake stroke in the regions (1) and (2) to the time near the compression top dead center. Further, as shown in FIG. 10 (l), the fuel pressure is also greatly changed from the low fuel pressure in the region (1) and the region (2) to the high fuel pressure of 30 MPa or more. Thus, although the fuel injection mode is largely changed between the region (2) and the region (3), the gas composition in the cylinder 18 continuously changes. 22, the opening degree of the throttle valve 36, the opening degree of the EGR valve 511, and the opening degree of the EGR cooler bypass valve 531 do not change suddenly (FIGS. 8E and 8F). ), FIG. 8 (g), FIG. 9 (h), FIG. 9 (i), and FIG. 9 (j)). This is advantageous in suppressing the occurrence of a torque shock or the like during the transition between the region (2) and the region (3), and simplification of the control is achieved.

  On the high load side that is equal to or higher than the predetermined load T3, the start timing of the fuel injection as the high pressure retarded injection is gradually retarded as the engine load increases, as shown in FIG. 10 (k). The fuel pressure is also set higher as the engine load increases, as shown in FIG. As the engine load increases, pre-ignition and the like are more likely to occur, and the pressure increase can be more severe. Therefore, these are effectively avoided by delaying the injection start timing of combustion and setting the fuel pressure higher.

(Predetermined loads T4 to T5)
The predetermined load T4 is related to switching between the CI mode and the SI mode, and the SI mode is set on the high load side exceeding the predetermined load T4.

  Here, as described above, the amount of internal EGR gas introduced is adjusted by adjusting the degree of overlap of the valve opening period of the intake valve 21 with respect to the valve opening period of the exhaust valve 22 opened within the intake stroke period. Basically, it is based on the control of the intake CVVL 73 and VVT 73. Thereby, as shown by the solid line arrow in FIG. 11, the introduction amount of the internal EGR gas can be continuously reduced up to a predetermined amount (see S1 and S2 in FIG. 11). However, if the high-temperature internal EGR is continuously introduced into the cylinder 18, there is a concern about abnormal combustion such as pre-ignition, or when the exhaust VVL 71 whose lift amount cannot be adjusted is opened, the lift amount of the CVVL 73 is maximized. However, the exhaust VVL 71 must be turned off somewhere to stop opening the exhaust twice because, for example, fresh air of a predetermined amount or more cannot be sucked into the cylinder 18. For this reason, as shown in S3 and S4 in the figure, as the exhaust VVL 71 is switched on / off, the amount of internal EGR gas introduced decreases discontinuously (see the dashed line arrow in FIG. 11). .

  On the other hand, if the method of introducing hot EGR gas into the cylinder 18 is changed in the middle of the same combustion mode region (same CI combustion region), all changes are made before and after the change (before and after the exhaust double opening is stopped). There may be a step in the ratio of the EGR gas amount to the gas amount, or a torque step may occur, and it will be accompanied by exceptional difficulty to perform control to eliminate these steps.

  Therefore, at a predetermined load T4 related to switching between the CI mode and the SI mode, the introduction of the internal EGR gas into the cylinder 18 is stopped, that is, as shown in FIG. 8 (e), the exhaust VVL 71 is turned off. The exhaust is opened twice.

  That is, in the engine 1, combustion accompanied by a steep pressure increase is combined with a decrease in the amount of internal EGR gas introduced into the cylinder 18 as the load increases and a high-pressure retarded injection. By making it possible to continuously open the exhaust gas twice in the region (3) without causing it, the timing for stopping the exhaust gas double opening is made coincident with the timing for switching from the CI mode to the SI mode. It is possible.

  Thus, by making the timing for stopping the exhaust double opening coincide with the timing for switching from the CI mode to the SI mode, for example, as shown in FIGS. 8 (f) and 8 (g), the operation of the exhaust VVL 71 is performed. When the engine is stopped at a predetermined load T3 (see the two-dot chain line in the figure), there is no need to suddenly change the valve opening timing and the lift amount of the intake valve 21, and as shown in FIG. As in the case where the operation of the exhaust VVL 71 is stopped at the predetermined load T3 (see the two-dot chain line in the figure), it is not necessary to suddenly change the opening degree of the EGR valve 511. This enhances controllability against an increase in engine load. The point that high pressure retarded injection is performed with an engine load equal to or higher than the predetermined load T4 is the same as described above (see FIGS. 10 (k), 10 (l), and 10 (m)).

  In addition, since the air-fuel ratio of the air-fuel mixture is set to the stoichiometric air-fuel ratio (λ≈1) on each of the low load side and the high load side across the boundary related to switching between the CI mode and the SI mode, The EGR rate is set so as to continuously decrease from the CI mode to the SI mode. This means that when switching from the CI mode to the SI mode where the combustion mode is switched, there is no significant change other than starting spark ignition, switching from the CI mode to the SI mode, or vice versa. Each can be made smooth to suppress the occurrence of torque shock or the like. In particular, since control responsiveness related to the recirculation of exhaust gas through the EGR passage 50 is relatively low, control that does not cause a sudden change in the EGR rate is advantageous in improving controllability.

  Further, as described above, in the CI mode, as the EGR rate is set as high as possible, the EGR rate becomes high in the low load region in the SI mode near the boundary with the CI mode. A high EGR rate is advantageous for reducing pump loss, but may be disadvantageous for combustion stability in the SI mode.

  Therefore, high temperature external EGR gas is introduced into the cylinder 18 in the low load region in the SI mode, specifically, in the low load side of the predetermined load T5. That is, the external EGR gas that has passed through the EGR cooler bypass passage 53 and is not cooled is introduced into the cylinder 18. As a result, as shown in FIG. 7B, the temperature in the cylinder 18 is set higher, the ignition delay time is shortened, and the stability of the spark ignition combustion in the environment of the high EGR rate is improved. ing.

  Specifically, as shown in FIG. 9 (i), the opening degree of the EGR valve 511 is gradually decreased as the load increases so as to continue from the CI mode, while as shown in FIG. 9 (j). The EGR cooler bypass valve 531 is opened at a predetermined load T4 and gradually decreased as the load increases (in addition, when the opening of the EGR valve 511 and the opening of the EGR cooler bypass valve 531 are compared, the EGR cooler bypass valve 531 The rate of decrease is higher for the opening degree). As a result, the cooled EGR gas increases, the hot EGR gas decreases, and the EGR rate including the cooled EGR gas and the hot EGR gas gradually decreases as the engine load increases. Therefore, the amount of fresh air increases. At this time, the EGR valve 511 is open. Further, the opening degree of the throttle valve is maintained fully open between the predetermined loads T4 and T5.

  On the other hand, as shown in FIG. 10 (k), the start timing of fuel injection is gradually retarded as the engine load increases, and the fuel pressure also increases as the engine load increases as shown in FIG. 10 (l). Increase gradually. Further, as shown in FIG. 5 (m), the ignition timing gradually retards with the start of fuel injection as the engine load increases. In the SI mode, spark ignition is performed by operating the spark plug 25 at a predetermined ignition timing in the low load side region of the predetermined loads T4 to T5. However, the combustion mode is that flame nuclei are generated by spark ignition. In addition, the flame is not limited to the form in which the flame is propagated, and there may be a form in which the low-temperature oxidation reaction is promoted by spark ignition and self-ignition occurs.

(Predetermined load T5 or more)
In the SI mode, on the high load side that is equal to or higher than the predetermined load T5, the temperature in the cylinder 18 is increased, so that the combustion stability is increased. Therefore, the EGR cooler bypass valve 531 is closed. As shown, the hot EGR gas amount becomes zero and only the cooled EGR gas is introduced into the cylinder 18. Although illustration is omitted, on the high load side of the predetermined load T5 or more, the intercooler bypass valve 351 is opened (for example, the opening degree is gradually increased as the engine load increases), thereby causing the intercooler / warmer 34 to be opened. The amount of fresh air that bypasses the engine may be increased, and the temperature of fresh air introduced into the cylinder 18 may be lowered. This is advantageous in avoiding abnormal combustion such as pre-ignition and knocking by lowering the temperature in the cylinder 18 in the high load side region.

  Further, as shown in FIG. 9 (h), the opening degree of the throttle valve 36 is kept fully open, and as shown in FIG. 9 (i), the EGR valve 511 is gradually closed as the engine load increases and fully opened. Close the valve with a load. On the other hand, as shown in FIGS. 8 (f) and 8 (g), the lift amount of the intake valve 21 is gradually increased as the engine load increases, and the maximum lift amount is obtained with the fully open load. Thus, the amount of fresh air introduced into the cylinder 18 is increased as the engine load increases, so that the torque on the high load side in the operating region of the engine 1 is improved.

  Further, as shown in FIGS. 10 (k), 10 (l), and 10 (m), the fuel injection start timing is gradually retarded as the engine load increases, and the fuel pressure increases as the engine load increases. Gradually set higher. Thus, the ignition timing is gradually retarded as the engine load increases. Although abnormal combustion or the like tends to occur as the engine load increases, it is effectively avoided by retarding the eruption start timing and increasing the fuel pressure.

  The change of each parameter with respect to the level of the engine load has been described above with reference to FIGS. 7 to 10. FIG. 12 shows the relationship between the EGR rate and the engine load. As described above, the air-fuel ratio is set to lean in the light load region where the engine load is low, while in the region where the load is higher than the light load region, the engine load is high or the combustion mode is different. Regardless, the air-fuel ratio is set constant at the theoretical air-fuel ratio (λ≈1). The engine 1 is controlled along a control line indicated by a thick solid arrow in FIG. 12, and the EGR rate is set to the maximum under the condition that the air-fuel ratio is set to the theoretical air-fuel ratio (λ≈1). Therefore, the EGR rate changes continuously regardless of the engine load level and regardless of the switching of the combustion mode. This is advantageous in improving the controllability because the gas composition in the cylinder 18 continuously changes when the engine load changes continuously.

  In a combustion mode in which compression ignition combustion is performed by injecting fuel during the intake stroke while introducing a large amount of EGR gas into the cylinder 18 (that is, corresponding to the region (1) and the region (2)), FIG. As indicated by the alternate long and short dash line in FIG. 12, the engine load exceeding a predetermined value cannot be realized due to the restriction of dP / dt. By performing the high pressure retarded injection, it is possible to perform the compression ignition combustion stably while slowing the combustion and eliminating the restriction of dP / dt. This corresponds to the combustion mode of the region (3) in FIG. 4, and the CI mode can be expanded to the high load side. It can also be said that by providing this region (3), a continuous change in the EGR rate with respect to the engine load level is realized.

  In the SI combustion region where the abnormal combustion such as pre-ignition (pre-ignition) may occur due to the high geometric compression ratio of the engine 1 (see the one-dot chain line in FIG. 12), high-pressure retarded injection is performed. By doing so, it is possible to avoid such abnormal combustion and perform stable spark ignition combustion. The high pressure retarded injection also enhances the combustion stability, and is advantageous in ensuring a predetermined combustion stability even when a high EGR rate is set at a load immediately after switching from the CI mode to the SI mode. This is another factor that makes it possible to continuously change the EGR rate with respect to the engine load.

  Thus, ensuring the continuity of the state quantity in the cylinder 18 with respect to the level of the engine load suppresses torque shocks and the like during mode switching in the engine 1 that involves switching between the SI mode and the CI mode. Will be advantageous.

  Further, in the engine 1 in which the geometric compression ratio is set to be high, the volume of the combustion chamber 19 becomes relatively small at a timing at which fuel is injected by high-pressure retarded injection. Although this may be disadvantageous in terms of the air utilization rate in the combustion chamber 19, high pressure retarded injection injects fuel into the cavity 141 at a high fuel pressure, thereby strengthening the flow in the cavity 141 and increasing the flow of air. Increase utilization. In particular, since the direct injection injector 67 is of a multi-hole type, it is effective for effectively increasing the turbulent energy of the gas in the cavity 141 and improving the air utilization rate.

  As a result, in the region (3) in the CI mode, a relatively homogeneous air-fuel mixture is quickly formed, and the ignitability and stability of the compression ignition combustion are improved. Similarly, abnormal combustion is also avoided in the region (4) in the SI mode.

  Here, when the high pressure retarded injection in the CI mode is compared with the high pressure retarded injection in the SI mode, as shown in FIG. 10 (k), the fuel injection start timing is advanced in the high pressure retarded injection in the CI mode. Set to the side. This is because the region (3) in which high pressure retarded injection is performed in the CI mode can introduce a large amount of EGR gas into the cylinder 18 by performing compression ignition combustion and the load of the engine 1 is relatively low. Thus, combustion can be slowed down by a large amount of EGR gas. Therefore, by making the fuel injection start time earlier as long as abnormal combustion can be avoided, the period of formation of the homogeneous mixture is secured to a certain extent to improve ignitability and combustion stability, and compression ignition By delaying the timing after the compression top dead center, it becomes possible to avoid a sudden pressure increase as well as slow combustion by a large amount of EGR gas.

  On the other hand, in the region (4) (or region (5)) in which high pressure retarded injection is performed in the SI mode, a large amount of EGR gas cannot be introduced into the cylinder 18 from the viewpoint of combustion stability. It is desirable to avoid abnormal combustion due to the effect of retard injection by delaying the start timing as much as possible.

(Other embodiments)
The present invention is not limited to the embodiments, and can be implemented in various other forms without departing from the spirit or main features thereof. That is, it is not limited to application to the engine configuration described above. For example, fuel may be injected into the intake port 16 through the port injector provided separately in the intake port 16 instead of the direct injection injector 67 provided in the cylinder 18 during the intake stroke period.

  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 a predetermined operation range, and the three-way catalyst can be used. If a catalyst (LNT: Lean NOx Trap) is used, the air-fuel ratio of the air-fuel mixture may be set to lean.

  Moreover, the operation area | region shown in FIG. 4 is an illustration, and it is possible to provide various operation areas besides this.

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

  As described above, the above-described embodiment is merely an example in all respects and should not be interpreted in a limited manner. Further, all modifications and changes belonging to the equivalent scope of the claims are within the scope of the present invention.

  As described above, the present invention provides internal EGR introduction means for executing at least one of the opening of the intake valve in the exhaust process and the opening of the exhaust valve in the intake stroke in order to introduce the exhaust gas into the cylinder. It is useful for a spark ignition direct injection engine or the like that operates and operates the internal EGR introduction means in the CI combustion region.

1 Engine (Engine body)
10 PCM (controller)
18 cylinder 19 combustion chamber 21 intake valve 22 exhaust valve 25 spark plug 62 high pressure fuel supply system (fuel pressure variable mechanism)
67 Direct injection injector (fuel injection valve)
71 VVL (internal EGR introduction means)
72 VVT (Variable valve timing mechanism) (Internal EGR introduction means)
73 CVVL (internal EGR introduction means)
511 EGR valve (external EGR gas recirculation means)
531 EGR cooler bypass valve (external EGR gas recirculation means)

Claims (4)

An engine body having a cylinder forming a combustion chamber at the top and configured to be supplied with fuel mainly composed of gasoline;
An intake valve and an exhaust valve for opening and closing the intake port opening and the exhaust port opening that open to the combustion chamber side, and
A fuel injection valve configured to inject the fuel into the combustion chamber;
A variable fuel pressure mechanism configured to change the pressure of fuel injected by the fuel injection valve;
An ignition plug disposed to face the combustion chamber and configured to ignite an air-fuel mixture in the combustion chamber;
Internal EGR introduction means for performing at least one of opening of an intake valve in an exhaust stroke and opening of an exhaust valve in an intake stroke in order to introduce exhaust gas into the cylinder;
A controller configured to operate the engine body by controlling at least the fuel injection valve, the fuel pressure variable mechanism, the spark plug, and the internal EGR introduction means,
By the control of the controller, in the compression self-ignition combustion region on the lower load side than the predetermined load, the air-fuel mixture in the combustion chamber is burned by compression self-ignition, while in the spark ignition combustion region on the high load side above the predetermined load. A spark ignition type direct injection engine that burns the air-fuel mixture in the combustion chamber by spark ignition using the spark plug and operates the internal EGR introduction means in the compression self-ignition combustion region,
The controller is
In the compression self-ignition combustion region, the end timing of fuel injection closest to the ignition timing is within the period from the initial stage of the intake stroke to the middle stage of the compression stroke in the predetermined first region of the compression self-ignition combustion region , The fuel injection valve is driven in the second region on the higher load side than the first region in the compression self-ignition combustion region so as to be within a period from the latter half of the compression stroke to the early stage of the expansion stroke , and the second In the region , the fuel injection pressure of the fuel injection valve is set to a predetermined pressure of 30 MPa or more using the fuel pressure variable mechanism, while in the first region, the fuel injection pressure is set to be lower than the predetermined pressure,
In the compression self-ignition combustion region, the load is increased so that the ratio of the internal EGR gas amount to the total gas amount introduced into the cylinder is smaller in the second region than in the first region. Accordingly, the internal EGR introduction means is driven so as to increase the intake flow rate,
A spark ignition direct injection engine characterized in that the operation of the internal EGR introduction means is stopped in accordance with switching of the combustion mode from the compression self-ignition combustion region to the spark ignition combustion region. The spark ignition direct injection engine according to claim 1,
The predetermined load is the entire region consisting of the compression self-ignition combustion region and the spark ignition combustion region excluding the throttle valve full open region, a low load region with a low load, and a medium load region with a higher load than this, Furthermore, the spark ignition direct injection engine is characterized in that it is set to an intermediate load region when it is divided equally into a high load region with a high load. The spark ignition direct injection engine according to claim 1 or 2,
An external EGR gas recirculation means configured to be able to introduce a low-temperature cooled EGR gas in which the exhaust gas is cooled by heat exchange into the cylinder;
The controller drives the external EGR recirculation means so that the cooled EGR gas is introduced into the cylinder in the second region in the compression self-ignition combustion region. Jet engine. In the spark ignition direct injection engine according to any one of claims 1 to 3,
A variable valve timing mechanism capable of changing the opening timing of the intake valve;
The internal EGR introduction means is configured to adjust the amount of EGR gas introduced into the cylinder using a variable valve timing mechanism when the exhaust valve is opened during the intake stroke. A spark ignition direct injection engine.

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