JP2014206055A - Control device for internal combustion engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a control device for an internal combustion engine that can suppress injection amount of a high-octane fuel without undermining an inhibiting effect on knocking etc. in a low-rotation high-load region.SOLUTION: A port injector injects first fuel that is separated from mixed fuel of alcohol and gasoline and that has a higher octane number than that of the mixed fuel into an intake port. A direct injection injector injects second fuel that is separated from the mixed fuel and that has a lower octane number than that of the mixed fuel into a cylinder of an engine. When a driving state specified by a cylinder interior intake amount of the engine and engine speed is within a first region of low rotation and high load, an ECU allows injection of the first fuel by the port injector and then allows divided injection of the second fuel by the direct injection injector into an intake process and a compression process.

Description

  The present invention relates to a control device for an internal combustion engine. More specifically, the present invention relates to a control device for an internal combustion engine using a high octane fuel and a low octane fuel.

  As fuel for internal combustion engines, alcohol fuel that can be produced from many crops such as sugar cane, corn, and potatoes has attracted attention. Particularly in recent years, mixed fuels in which alcohol fuel is added to gasoline are in circulation and are expected to become more widespread in the future. There are various types of alcohol fuel, such as ethanol and methanol. In the following, ethanol, which is the most popular alcohol fuel, will be described as an example.

  Along with the widespread use of such mixed fuels, research on a separation device that separates the mixed fuel supplied from the outside into a fuel with a high gasoline concentration and a fuel with a high ethanol concentration on the vehicle is also underway. Since gasoline and ethanol have various differences in fuel properties such as octane number and calorific value, for example, they are separated again on the vehicle rather than using the mixed fuel supplied from the outside as it is, It may be preferable to use ethanol separately.

  For example, Patent Document 1 describes a technique in which a fuel mainly composed of ethanol and a fuel mainly composed of gasoline separated from mixed fuel by two injectors are injected into an intake port. Ethanol has a higher octane number than gasoline and is excellent in knocking resistance. The technique of Patent Document 1 pays attention to this point, and controls the injection ratio of the fuel from the two injectors so that the octane number of the fuel in the cylinder becomes high at the time of low rotation and high load operation, so that knocking, pre-ignition, etc. The occurrence of abnormal combustion is suppressed.

JP2012-163002A

  As described above, it has been conventionally known to inject ethanol having a relatively high octane number in order to suppress knocking and the like. However, when ethanol is injected into the intake port, there are the following problems. Therefore, it is preferable to suppress the injection amount as much as possible.

  First, when a large amount of ethanol is injected from an injector provided in the intake port (hereinafter referred to as “port injector”), the flow of fresh air introduced into the cylinder is obstructed (so-called air block). As a result, the required output cannot be obtained sufficiently. Ethanol has a lower calorific value than gasoline. Therefore, when ethanol is injected from the port injector and when gasoline is injected, it is necessary to inject a larger amount of ethanol in order to obtain the same output in both cases. For this reason, when ethanol is injected from the port injector, the problem of insufficient torque due to the occurrence of such an air block becomes more prominent.

  Secondly, when a large amount of ethanol is injected to obtain a suppression effect such as knocking, the fuel adheres to the wall surface, and the amount of unnecessary exhaust from the internal combustion engine increases, thereby purifying the exhaust gas. There is a risk that the burden on the plant will increase. This second problem becomes more prominent when ethanol having a calorific value smaller than that of gasoline is used.

  Third, since the ethanol content of the most popular mixed fuel is about 10%, the amount of ethanol stored in the running vehicle is less than that of gasoline. Therefore, the amount of ethanol stored during running tends to be insufficient.

  The present invention injects high octane fuel obtained by separating from mixed fuel into the intake port, and suppresses the injection amount of high octane fuel without impairing the suppression effect such as knocking in the low rotation and high load region. An object of the present invention is to provide a control device for an internal combustion engine.

  (1) A control device for an internal combustion engine (for example, engine 1 described later) according to the present invention supplies a first fuel having an octane number higher than that of the mixed fuel separated from the mixed fuel of alcohol and gasoline to an intake port (for example, A first fuel injection device (for example, a port injector 62 described later) that injects the fuel into a later-described intake port 15), and a second fuel having a lower octane number than the mixed fuel separated from the mixed fuel (for example, a cylinder of the internal combustion engine) , A second fuel injection device (for example, a direct injection injector 63 described later) that is injected into a cylinder 13 described later, and an injection control that controls the fuel injection timing and the injection amount from the first and second fuel injection devices. An apparatus (for example, an ECU 7 described later). The injection control device injects the first fuel when the operating state specified by the engine load parameter and the rotational speed is within a low rotation and high load region (for example, a first region described later). Later, the second fuel is divided and injected into an intake process and a compression process.

  (2) In this case, in the injection control device, the operating state of the engine is slightly higher than the low rotation and high load region or slightly lower load region (for example, a second region described later). In this case, it is preferable that the second fuel is injected only in the intake process, and the injection timing is delayed as the operating state of the engine approaches the low rotation / high load region.

  (3) A control device for an internal combustion engine of the present invention includes a first fuel injection device that injects a first fuel having a higher octane number than a mixed fuel separated from a mixed fuel of alcohol and gasoline into an intake port of the internal combustion engine; A second fuel injection device for injecting into the cylinder of the internal combustion engine a second fuel having a lower octane number than the mixed fuel separated from the mixed fuel; and fuel injection timing and injection from the first and second fuel injection devices An injection control device for controlling the amount. The injection control device injects the second fuel after injecting the first fuel, and the operation state specified by the load parameter and the rotational speed of the engine changes to the low rotation side or the high load side. 2 Delay the fuel injection timing.

  (4) In this case, the injection control device compresses the second fuel from the intake step when the operating state of the engine is in a low rotation and high load region (for example, a first region described later). It is preferable to spray over the process.

  (1) In the present invention, after the first fuel injection device injects the high-octane first fuel into the intake port, the second fuel injection device directly injects the low-octane second fuel into the cylinder. Particularly in the present invention, when the operating state of the internal combustion engine is in a low-rotation and high-load region where knocking or the like is likely to occur, the second fuel that is directly injected into the cylinder is divided into an intake process and a compression process. Thereby, the in-cylinder cooling effect by the latent heat of vaporization of the second fuel directly injected into the cylinder can be enhanced. In addition, by ensuring a certain amount of in-cylinder cooling effect by split injection of the second fuel, the amount of injection of the first fuel necessary for suppressing abnormal combustion such as knocking in the low-rotation and high-load region is correspondingly increased. Can be reduced. In addition, by suppressing the knocking in the low rotation and high load region using the first fuel having a high octane number in this way, the expansion ratio (geometric compression ratio) of the internal combustion engine can be set higher accordingly, Fuel consumption can be improved in all areas other than the low rotation and high load areas.

  By the way, in a general internal combustion engine equipped with a fuel injection device that directly injects fuel into a cylinder, the exhaust valve closing timing (EVC) and the intake valve There is a case where the valve overlap amount at the opening timing (IVO) (that is, the length of the period during which both the exhaust valve and the intake valve are open) is increased. When the valve overlap amount is increased, scavenging is promoted, the internal EGR amount is reduced, and a lot of fresh air is taken into the cylinder, so that the in-cylinder cooling effect is improved and the charging efficiency is also improved. However, in the present invention involving the injection of the first fuel into the intake port, the first fuel injected before the second fuel may blow through to the exhaust side together with fresh air, which may increase the burden on the exhaust purification catalyst. The valve overlap amount is limited. On the other hand, in the present invention, the combination of the injection of the first fuel into the intake port and the divided injection of the second fuel can improve the knocking resistance without increasing the valve overlap amount.

  (2) In the present invention, when the operating state of the internal combustion engine is in a region outside the low-rotation and high-load region where the second fuel is divided and injected, the closer to the low-rotation and high-load region, the closer to the first in the intake process. 2 Delay the fuel injection timing. Thus, by increasing the in-cylinder cooling effect by retarding the injection timing of the second fuel until the divided injection region is reached, the injection amount of the first fuel outside the low rotation high load region is reduced. Can be reduced.

  (3) In the present invention, after the first fuel injection device injects the high-octane first fuel into the intake port, the second fuel injection device directly injects the low-octane second fuel into the cylinder. In particular, according to the present invention, the second fuel injected directly into the cylinder is delayed by delaying the injection timing of the second fuel as the operating state of the internal combustion engine changes to the low rotation side or the high load side and knocking or the like is more likely to occur. The in-cylinder cooling effect by the latent heat of vaporization can be increased. In this way, by ensuring a certain amount of in-cylinder cooling effect by retarding the injection timing of the second fuel, the amount of injection of the first fuel necessary for suppressing abnormal combustion such as knocking is reduced accordingly. be able to. In addition, by suppressing knocking in the low rotation high load region using the first fuel having a high octane number in this way, the expansion ratio of the internal combustion engine can be set higher accordingly, so that the other than the low rotation high load region Fuel consumption can be improved in all areas.

  (4) In the present invention, the second fuel is injected from the intake process to the compression process within a predetermined low rotation high load region. That is, in the present invention, the sufficient in-cylinder cooling effect can be enhanced by delaying the injection timing of the second fuel to near the bottom dead center of the intake process in the low rotation and high load region. Therefore, the injection amount of the first fuel can be reduced without impairing the suppression effect such as knocking.

It is a figure which shows the structure of the engine which concerns on 1st Embodiment of this invention, and its control apparatus. It is a functional block diagram concerning execution of intake air amount control and fuel injection / ignition control. It is a flowchart which shows the specific procedure which determines the basic value of the injection time and injection timing of the 1st and 2nd fuel. It is an example of the map which determines an operation area | region based on an engine speed and in-cylinder intake air amount. It is an example of the map which determines the basic value of the injection time and injection timing of a 1st fuel, and the basic value of the injection time and injection timing of a 2nd fuel. It is a figure which shows typically the injection aspect of a 2nd fuel. It is a figure which shows the relationship between an engine speed and the torque of an engine. It is an example of the map which determines the basic value of the injection time and injection timing of the 1st fuel which concerns on 2nd Embodiment of this invention, and the injection time and injection timing of the 2nd fuel.

<First Embodiment>
Hereinafter, a first embodiment of the present invention will be described with reference to the drawings.
FIG. 1 is a diagram showing a configuration of an internal combustion engine (hereinafter simply referred to as “engine”) 1 and a control device thereof according to the present embodiment.

  The engine 1 is supplied with fuel to an intake pipe 3 through which intake air flows, an exhaust pipe 4 through which exhaust gas flows, a supercharger 5 that pumps intake air in the intake pipe 3 to the engine 1, and an engine 1 and an intake port 15. A fuel supply system 6 and these electronic control units (hereinafter referred to as “ECU (Electronic Control Unit)”) 7 are provided.

  The engine 1 is a multi-cylinder engine including a plurality of cylinders 13. FIG. 1 representatively shows one of them. The engine 1 is configured by combining a cylinder block 11 in which a cylinder 13 is formed and a cylinder head 12. A piston 14 is slidably provided in the cylinder 13. A combustion chamber 1 a of the engine 1 is formed by the top surface of the piston 14 and the surface of the cylinder head 12 on the cylinder 13 side. The piston 14 is connected to a crankshaft (not shown) via a connecting rod. That is, a crankshaft (not shown) rotates according to the reciprocating motion of the piston 14 in the cylinder 13.

  The cylinder head 12 is formed with an intake port 15 that connects the combustion chamber 1 a and the intake pipe 3, and an exhaust port 16 that connects the combustion chamber 1 a and the exhaust pipe 4. An intake opening facing the combustion chamber 1 a in the intake port 15 is opened and closed by an intake valve 17. An exhaust opening facing the combustion chamber 1 a in the exhaust port 16 is opened and closed by an exhaust valve 18.

  The cylinder head 12 is provided with an ignition plug 23 facing the combustion chamber 1a, an intake camshaft 19 that drives the intake valve 17 to open and close, and an exhaust camshaft 20 that drives the exhaust valve 18 to open and close. The spark plug 23 is connected to the ECU 7 via an igniter (not shown) and its driver. The ignition timing of the spark plug 23 is controlled by fuel injection / ignition control described later.

  The intake camshaft 19 is provided with a plurality of intake cams 19a for opening and closing the intake valve 17. Similarly, the exhaust camshaft 20 is provided with a plurality of exhaust cams 20a for driving the exhaust valve 18 to open and close. FIG. 1 shows only one of the plurality of cams 19a and 20a. The camshafts 19 and 20 are connected to a crankshaft via a connecting mechanism (not shown) such as a sprocket or a timing chain, and are rotated by the crankshaft. More specifically, when the crankshaft rotates twice, the camshafts 19 and 20 rotate once. When the camshafts 19 and 20 rotate, the valves 17 and 18 advance and retract in a manner corresponding to the profiles of the cams 19a and 20a.

  A cam phase variable mechanism (VTC) 21 for changing the cam phase of the intake cam 19a with respect to the crankshaft is provided at one end of the intake camshaft 19. The intake side VTC 21 advances or retards the intake valve 17 opening / closing timing (that is, the opening timing (IVO) and the closing timing (IVC)) by advancing or retarding the cam phase of the intake camshaft 19 steplessly. it can. In the present embodiment, the operation in the Atkinson cycle (Miller cycle) in which the effective compression ratio of the combustion chamber 1a of the engine 1 can be variably controlled by the intake side VTC 21 is realized. That is, the effective compression ratio can be lowered by correcting the closing timing of the intake valve 17 to the advance side or the retard side with respect to the bottom dead center and reducing the intake amount.

  In addition, the cam phase of the exhaust cam 20a with respect to the crankshaft is also changed at one end of the exhaust camshaft 20, and the opening / closing timing of the exhaust valve 18 (that is, the opening timing (EVO) and the closing timing (EVC)) is advanced or delayed. A VTC 22 is provided. In this embodiment, the valve overlap amount between the closing timing of the exhaust valve 18 and the opening timing of the intake valve is variably controlled by advancing or delaying the closing timing of the exhaust valve 18 by the exhaust side VTC 22.

  The supercharger 5 includes a turbine (not shown) provided in the exhaust pipe 4 and a compressor (not shown) provided in the intake pipe 3. The turbine is driven by the kinetic energy of the exhaust flowing through the exhaust pipe 4. The compressor is rotationally driven by the turbine and pumps the intake air in the intake pipe 3 to the engine 1.

  The intake pipe 3 is provided with a compressor of the supercharger 5, an intercooler 31, and a throttle valve 32 in order from the upstream side to the downstream side. The intercooler 31 cools the intake air compressed by the supercharger 5 and improves the intake air charging efficiency of the engine 1. The throttle valve 32 controls the flow rate of air supplied to the combustion chamber 1 a of the engine 1 (hereinafter referred to as “intake flow rate”). The throttle valve 32 is connected to the ECU 7 via a driver (not shown). That is, the throttle valve 32 is called a so-called DBW (Drive By Wire) throttle, which is mechanically disconnected from an accelerator pedal (not shown) operated by the driver's vehicle. The throttle valve 32 is controlled to an appropriate opening degree by the intake air amount control described later.

  Further, the intake pipe 3 is provided with a bypass passage 33 that communicates with the front and rear of the compressor of the supercharger 5 and a blow-off valve 34 that opens and closes the bypass passage 33. When the blow-off valve 34 is opened, the front and rear of the compressor of the supercharger 5 communicate with each other, and the pressure between the compressor and the throttle valve 32 (hereinafter referred to as “supercharging pressure”) is released. The blow-off valve 34 is connected to the ECU 7 via a driver (not shown). The blow-off valve 34 is opened and closed at an appropriate time by intake air amount control described later.

  The exhaust pipe 4 is provided with a turbine of the supercharger 5 and an exhaust purification catalyst 41 for purifying exhaust gas in order from the upstream side to the downstream side. The exhaust purification catalyst 41 is, for example, a three-way catalyst, and purifies HC, CO, NOx, etc. in the exhaust.

  Further, the exhaust pipe 4 is provided with a bypass passage 42 that communicates between the front and rear of the turbine of the supercharger 5 and a waste gate valve 43 that opens and closes the bypass passage 42. When this waste gate valve 43 is opened, the front and rear of the turbine of the supercharger 5 communicate with each other, limiting the rotational speed of the turbine and consequently limiting the supercharging pressure. The waste gate valve 43 is connected to the ECU 7 via a driver (not shown). The waste gate valve 43 is opened and closed at an appropriate time by intake air amount control described later.

  The fuel supply system 6 includes a fuel storage device 61 that stores fuel, a port injector 62 that injects the fuel stored in the fuel storage device 61 into the intake air of the engine 1, and the fuel stored in the fuel storage device 61 in the combustion chamber 1a. A direct injection injector 63 for direct injection.

  The fuel storage device 61 was separated by a main tank 611 into which a mixed fuel of ethanol and gasoline supplied from the fuel filler F was introduced, a separation device 612 for separating the mixed fuel in the main tank 611, and a separation device 612. A sub-tank 613 for storing fuel. The separation device 612 separates ethanol and aromatics from the mixed fuel stored in the main tank 611 and supplies it to the sub tank 613.

  The separation device 612 uses, for example, a mixed fuel stored in the main tank 611 by a pervaporation method (pervaporation method), a first fuel having a higher ethanol concentration than the mixed fuel, and a lower ethanol than the mixed fuel. The fuel is separated into a second fuel having a concentration. The separation device 612 supplies the first fuel separated from the mixed fuel to the sub tank 613 and returns the remaining second fuel to the main tank 611.

  More specifically, the separation device 612 includes a separation membrane 612a that selectively permeates specific components in the mixed fuel, and a high-pressure chamber 612b and a low-pressure chamber 612c partitioned by the separation membrane 612a. In this separation device 612, when the fuel stored in the main tank 611 is circulated in the high pressure chamber 612b and the low pressure chamber 612c is decompressed by a pump (not shown), a part of the mixed fuel circulating in the high pressure chamber 612b evaporates, The light passes through the low pressure chamber 612c and is supplied to the sub tank 613. As a result, the sub tank 613 stores the first fuel mainly composed of ethanol and aromatic and having a higher octane number than the first fuel supplied from the outside. The main tank 611 stores a second fuel that is mainly made of gasoline and has a lower octane number than the first mixed fuel supplied from the outside.

  The direct injection injector 63 directly injects the second fuel in the main tank 611 supplied via a fuel pump (not shown) into the combustion chamber 1 a of the engine 1. The direct injection injector 63 is connected to the ECU 7 via a driver (not shown). The valve opening timing and valve opening time of the direct injection injector 63, that is, the injection timing and injection time of the second fuel into the combustion chamber 1a are controlled by the ECU 7.

  The port injector 62 injects the first fuel in the sub tank 613 supplied via a fuel pump (not shown) into the intake port 15. The port injector 62 is connected to the ECU 7 via a driver (not shown). The ECU 7 controls the opening timing and opening time of the port injector 62, that is, the injection timing and injection time of the first fuel into the intake port 15.

  The ECU 7 is an electronic control unit that controls various devices provided in the engine 1, and includes an electronic circuit such as a CPU, a ROM, a RAM, and various interfaces. A plurality of sensors 81 to 93 are connected to the ECU 7 in order to grasp the state of the engine 1 and the state of the vehicle on which the engine 1 is mounted.

  The supercharging pressure sensor 81 is a detection signal substantially proportional to the supercharging pressure between the intercooler 31 and the throttle valve 32 in the intake pipe 3, more specifically, downstream of the supercharger 5. Is output to the ECU 7. The water temperature sensor 82 outputs a detection signal to the ECU 7 that is approximately proportional to the temperature of cooling water (hereinafter simply referred to as “water temperature”) flowing in the cooling water passage (not shown) of the cylinder block 11 of the engine 1. The ECU 7 uses the water temperature detected by the water temperature sensor 82 as the representative temperature of the engine 1, but is not limited thereto.

  The intake sensor 83 is a sensor that detects the state of intake on the downstream side of the throttle valve 32 in the intake pipe 3. More specifically, the intake sensor 83 includes an intake temperature sensor that outputs a detection signal substantially proportional to the intake air temperature at the target location (hereinafter referred to as “intake air temperature”) to the ECU 7, and the intake pressure ( Hereinafter, it is constituted by an intake pressure sensor or the like that outputs a detection signal substantially proportional to “intake pressure”) to the ECU 7.

The intake side cam sensor 84 outputs a pulse signal to the ECU 7 for each predetermined cam angle as the intake camshaft 19 rotates. The exhaust side cam sensor 85 outputs a pulse signal to the ECU 7 at every predetermined cam angle as the exhaust camshaft 20 rotates. The ECU 7 grasps the actual cam phase of the camshafts 19 and 20 based on the pulse signals transmitted from the cam sensors 84 and 85.
The throttle opening sensor 86 outputs a detection signal to the ECU 7 that is substantially proportional to the position of the throttle valve 32 (hereinafter referred to as “throttle opening”).

Knocking sensor 87 outputs a detection signal corresponding to the vibration of engine 1 caused by knocking to ECU 7. The ECU 7 determines the occurrence of knocking based on the output from the knocking sensor 87.
The air flow meter 88 outputs a detection signal to the ECU 7 that is substantially proportional to the flow rate of the intake air upstream of the compressor of the supercharger 5 in the intake pipe 3.

The accelerator pedal sensor 89 detects the amount of depression of the accelerator pedal operated by the driver, and outputs a detection signal corresponding to the depression amount to the ECU 7.
The vehicle speed sensor 90 is provided on a drive shaft (not shown) of the vehicle, and outputs a pulse signal to the ECU 7 every time the drive shaft rotates a predetermined angle. The vehicle speed is calculated in the ECU 7 by a process (not shown) based on the pulse signal from the vehicle speed sensor 90.

The crank angle sensor 91 outputs a pulse signal to the ECU 7 for each predetermined crank angle in accordance with the rotation of the pulsar 92 fixed to the crankshaft. The ECU 7 grasps the actual engine speed based on the output of the crank angle sensor 91.
The LAF sensor 93 is provided upstream of the exhaust purification catalyst 41 in the exhaust pipe 4 and outputs a detection signal that is substantially proportional to the oxygen concentration (air-fuel ratio) of the exhaust to the ECU 7. The LAF sensor 93 has a linear output characteristic that is substantially proportional to the air-fuel ratio at the detection location.

  FIG. 2 is a functional block diagram of intake air amount control and fuel injection / ignition control executed mainly by the ECU 7.

  The intake air amount control is performed according to the following procedure. First, the required output calculation unit 71 calculates a required output to the engine by searching a predetermined map based on the engine speed, the amount of depression of the accelerator pedal, the vehicle speed, and the like.

  The target intake air amount calculation unit 72 searches for a predetermined map based on the engine speed, the required output, the intake air flow rate, the intake air pressure, the intake air temperature, the fuel injection amount determined in the fuel injection / ignition control, and the like. Thus, the target intake air amount that is the target of the intake air amount to the engine is calculated.

The target throttle opening calculation unit 73 calculates a target throttle opening that is a target of the throttle valve opening by searching a predetermined map based on the target intake air amount, the engine speed, and the like.
The target supercharging pressure calculation unit 74 calculates a target supercharging pressure that is a target for the supercharging pressure by searching a predetermined map based on the target intake air amount, the engine speed, and the like.

  The target intake side VTC angle calculation unit 75 calculates a target effective compression ratio that is a target of the effective compression ratio by searching a predetermined map based on the target intake air amount, the engine speed, and the like. A target intake side cam phase that is a target of the cam phase of the intake camshaft for realizing the effective compression ratio is calculated. Note that this target effective compression ratio is set so as to decrease as it changes from the high load side to the low load side. Thereby, the operation using the Otto cycle can be performed on the high load side, and the operation using the Atkinson cycle can be performed on the low load side.

  The target exhaust side VTC angle calculation unit 76 calculates a target overlap amount that is a target of the valve overlap amount by searching a predetermined map based on the target intake air amount, the engine speed, and the like. A target exhaust side cam phase that is a target of the cam phase of the exhaust camshaft for realizing the overlap amount is calculated. In the present embodiment, not only the second fuel is directly injected into the cylinder by the direct injection injector, but also the first fuel is injected into the intake port by the port injector during the exhaust process. For this reason, if the valve overlap amount is excessively increased, a part of the first fuel injected from the port injector may be blown through without being used for combustion. Therefore, the optimum target overlap amount in the present embodiment is limited to be smaller than the optimum target overlap amount in the conventional internal combustion engine having only the direct injection injector. Therefore, this map for determining the target valve overlap amount takes into consideration the injection amount of the first fuel from the port injector so that the amount of unburned HC discharged from the engine does not exceed a predetermined reference value. Set experimentally.

  The feedback control unit 77 is a known feedback algorithm so that there is no deviation between the actual throttle opening, the actual boost pressure, the actual intake side cam phase, the actual exhaust side cam phase, and the target values set for each. Based on the control input, control inputs for the throttle valve, waste gate valve, intake side VTC, and exhaust side VTC are determined and input to the driver.

  The fuel injection / ignition control is performed according to the following procedure. First, the fuel injection / ignition control unit 78 has basic values for the injection time and injection timing of the first fuel from the port injector, the injection time and injection timing of the second fuel from the direct injection injector, and the ignition timing of the spark plug. Is calculated. The basic values of parameters (injection time, injection timing, and ignition timing) that define these injection modes and ignition modes are engine speed, intake flow rate, intake pressure, intake air temperature, water temperature, throttle opening, intake side cam phase, And by searching a predetermined map based on the exhaust-side cam phase and the like. Of the basic values of these parameters, a specific procedure for determining the injection time and injection timing of the first fuel and the injection time and injection timing of the second fuel will be described later with reference to FIGS. explain.

  The fuel injection / ignition control unit 78 adds or multiplies the correction coefficient calculated based on the outputs of the knock sensor and the LAF sensor to the basic value of the parameter that defines the injection mode and the ignition mode calculated as described above. To determine the injection mode and the ignition mode. The fuel injection / ignition control unit 78 determines control inputs for the port injector, the direct injection injector, and the spark plug so as to realize the determined injection mode and ignition mode, and inputs the control inputs to the respective drivers.

  Next, a procedure for determining specific injection modes (basic values of injection time and injection timing, presence / absence of divided injection, etc.) of the first fuel and the second fuel will be described with reference to FIGS. .

  FIG. 3 is a flowchart showing a specific procedure for determining the basic values of the injection time and injection timing of the first and second fuels. This process is executed at predetermined intervals in the ECU.

In S1, the ECU calculates an in-cylinder intake amount as a load parameter corresponding to the amount of intake air introduced into the cylinder, and proceeds to S2. The in-cylinder intake amount is searched for a predetermined map based on, for example, the engine speed, the rate of change of the throttle opening (the amount of change in the throttle opening per unit time), the intake air flow rate, the intake air temperature, and the like. It is calculated by.
In S2, the ECU determines to which operating region the operating state of the engine specified based on the engine speed and the in-cylinder intake air amount belongs.

  FIG. 4 is an example of a map for determining the operation region based on the engine speed and the in-cylinder intake air amount. In FIG. 4, the horizontal axis represents the engine speed, and the vertical axis represents the in-cylinder intake amount. As shown in FIG. 4, the operation region is roughly divided into three states. In the present embodiment, the region of the lowest rotation and the highest load is defined as the first region within the range that the engine speed and the in-cylinder intake amount can take. Within the range that can be taken by the engine speed and the in-cylinder intake amount, a region slightly higher than the first region or slightly lower load side is defined as a second region. In addition, within the range that the engine speed and the in-cylinder intake amount can take, a region of high rotation or low load excluding the first and second regions is defined as a third region.

  Further, when the operating state of the engine changes near the boundary between the first region and the second region, it is adjacent to the first region in the second region in order to prevent frequent changes in the injection mode accordingly. An area having a predetermined width (area surrounded by a broken line in the second area in FIG. 4) is defined as a marginal area. That is, if the engine operating state changes from the second region to the first region from the previous S2 determination to the current S2 determination, thereafter, unless the engine changes from the marginal region to the second region. The engine operating state is determined to belong to the first region.

  Returning to FIG. 2, in S <b> 3, the ECU determines a basic value of the injection time and the injection timing according to the determined operation region by searching a predetermined map, and ends this processing.

  FIG. 5 is an example of a map for determining the basic values of the injection time and injection timing of the first fuel and the basic values of the injection time and injection timing of the second fuel. FIG. 5 shows an example when the engine speed is 1500 rpm. Since other engine speeds do not change qualitatively, illustration and description thereof are omitted.

  When the operating state of the engine belongs to the third region, the first fuel injection is stopped and the second fuel is injected at a predetermined time during the intake process for a predetermined time.

  When the engine operating state belongs to the second region, the first fuel is injected at a predetermined time during the exhaust process for a predetermined time, and then the second fuel is injected at a predetermined time during the intake process for a predetermined time. To do. Here, in the second region, the closer to the first region, the earlier the start timing of the first fuel injection and the longer the injection time. In addition, the closer to the first region, the later the injection start timing of the second fuel and the longer the injection time.

  When the operating state of the engine belongs to the first region, the first fuel is injected for a predetermined time at a predetermined time in the exhaust process, and then the second fuel is dividedly injected into the intake process and the compression process. Here, in the first region, the injection start time of the first fuel is advanced and the injection time is lengthened as the operating state of the engine changes to the higher load side or the lower rotation side. As for the second fuel, the injection start timing during the intake process is delayed and the injection start timing during the compression process is made substantially constant as the operating state of the engine changes to the higher load side or the lower rotation side.

  Note that the quantitative configuration of the map of FIG. 4, that is, the specific values of the injection timing and injection time of the first and second fuels, the suppression effect such as knocking, the amount of HC discharged from the engine, the combustion stability And the air intake efficiency are determined experimentally.

FIG. 6 is a diagram schematically illustrating an injection mode of the second fuel according to the present embodiment. More specifically, in FIG. 6, the injection timing of the second fuel is shown in shades.
As shown in FIG. 6, in the present embodiment, in the first region and the second region, the second fuel injection timing is delayed as it changes to the high load side or the low rotation side. In the first region, in addition to retarding the injection timing of the second fuel, split injection of the intake process and the compression process is executed.

Next, the effect by optimizing the injection mode of the second fuel in the first region and the second region as described above will be described.
FIG. 7 is a graph showing the relationship between engine speed and engine torque (net average effective pressure (BMEP [kPa])). More specifically, the left side of FIG. 7 shows a case where both the retarding of the injection timing of the second fuel and the divided injection of the second fuel are not performed. The right side of FIG. 7 shows the case where the injection mode of the second fuel is optimized as described above.

  As shown in FIG. 7, by optimizing the injection mode of the second fuel, the in-cylinder cooling effect can be enhanced regardless of the first fuel having a high octane number. As a result, the output can be increased with less first fuel. Further, by reducing the amount of the first fuel injected into the intake port, more fresh air can be taken in accordingly. Accordingly, since the intercept point of the supercharger can be further reduced, the torque can be improved particularly in the low rotation range.

Second Embodiment
Next, a second embodiment of the present invention will be described with reference to the drawings. This embodiment differs from the first embodiment in the injection mode of the second fuel.

  FIG. 8 is an example of a map for determining basic values of the injection time and injection timing of the first fuel and the injection time and injection timing of the second fuel. FIG. 8 shows an example when the engine speed is 1500 rpm. Since other engine speeds do not change qualitatively, illustration and description thereof are omitted.

  When the operating state of the engine belongs to the third region, the first fuel injection is stopped and the second fuel is injected at a predetermined time during the intake process for a predetermined time.

  When the engine operating state belongs to the second region, the first fuel is injected at a predetermined time during the exhaust process for a predetermined time, and then the second fuel is injected at a predetermined time during the intake process for a predetermined time. To do. Here, in the second region, the closer to the first region, the earlier the start timing of the first fuel injection and the longer the injection time. In addition, the closer to the first region, the later the injection start timing of the second fuel and the longer the injection time.

  When the operating state of the engine belongs to the first region, the first fuel is injected for a predetermined time at a predetermined time in the exhaust process, and then the second fuel is injected from the intake process to the compression process. Here, in the first region, the injection start time of the first fuel is advanced and the injection time is lengthened as the operating state of the engine changes to the higher load side or the lower rotation side. For the second fuel, the injection start timing during the intake process is delayed as the operating state of the engine changes to the higher load side or the lower rotation side.

  In the present embodiment, the in-cylinder cooling effect can be ensured regardless of the first fuel by retarding the injection timing of the second fuel as described above. The injection amount can be suppressed.

  Here, the degree of retarding the injection timing of the second fuel in the first region and the second region will be compared between the present embodiment and the first embodiment. In the first embodiment, split injection of the intake process and the compression process is executed in the first region. For this reason, the degree of retardation of the injection timing during the intake process of the second fuel is substantially constant in the first region and the second region (see FIG. 5). In contrast, in the present embodiment, in the first region, the second fuel is injected from the intake process to the compression process. For this reason, it is preferable to increase the degree of retardation in the vicinity of the first region in the second region as shown in FIG.

  In the first embodiment, the split injection of the second fuel is performed only in the first region. For this reason, in consideration of the case where the operating state of the engine fluctuates near the boundary between the first region and the second region, a marginal region is set in the second region, and the hunting in which the number of injections of the second fuel is frequently switched Occurrence was prevented. In contrast, in the second embodiment, the injection timing of the second fuel is continuously retarded from the second region to the first region. For this reason, even if the operating state of the engine fluctuates near the boundary between the first region and the second region, hunting as described above does not occur. For this reason, in the second embodiment, there is no need to set a marginal area in the second area.

The embodiment of the present invention has been described above, but the present invention is not limited to this.
For example, in the above embodiment, the effective compression ratio and the valve overlap amount are variably controlled by changing the intake cam phase and the exhaust cam phase by the intake side VTC and the exhaust side VTC. The means for controlling both is not limited to these. For example, the effective compression ratio and valve overlap amount can be variably controlled using a variable valve lift mechanism (VTEC (registered trademark)) that changes the lift amount of the intake valve or exhaust valve and the angular width of the valve opening period. Good.

  In the first embodiment, the occurrence of hunting is prevented by defining a marginal area in the second area, but means for preventing hunting is not limited to this. For example, even if the engine operating state changes to the second region and continues to be determined to belong to the first region until the predetermined time has elapsed after the change from the second region to the first region, the same effect is obtained. Play.

  In the above-described embodiment, the case where the in-cylinder intake amount estimated based on the engine speed, the change rate of the throttle opening, or the like is used as the parameter indicating the load has been described, but the present invention is not limited to this. . Any parameter may be used as long as the parameter is generally proportional to the engine load.

  Moreover, in the said embodiment, although the use of the mixed fuel of ethanol and gasoline was assumed, this invention is not limited to this. The alcohol component to be mixed with gasoline is not limited to ethanol, but may be methanol or butanol.

1. Engine (internal combustion engine)
13 ... Cylinder
15 ... Intake port 62 ... Port injector (first fuel injection device)
63 ... Direct injection injector (second fuel injection device)
7 ... ECU (Injection Control Device)

Claims (4)

A first fuel injection device that injects a first fuel having a higher octane number than the mixed fuel separated from the mixed fuel of alcohol and gasoline into the intake port of the internal combustion engine;
A second fuel injection device that injects into the cylinder of the internal combustion engine a second fuel having a lower octane number than the mixed fuel separated from the mixed fuel;
An internal combustion engine control device comprising: an injection control device that controls an injection timing and an injection amount of fuel from the first and second fuel injection devices;
The injection control device compresses the second fuel after the first fuel is injected and compressed when the operating state specified by the engine load parameter and the rotational speed is in a low rotation and high load region. A control apparatus for an internal combustion engine, wherein the injection is divided into steps.   The injection control device injects the second fuel only in the intake process when the operating state of the engine is slightly higher than the low rotation and high load region or slightly lower load region, and 2. The control apparatus for an internal combustion engine according to claim 1, wherein the injection timing is delayed as the operating state of the engine approaches the low rotation and high load region. A first fuel injection device that injects a first fuel having a higher octane number than the mixed fuel separated from the mixed fuel of alcohol and gasoline into the intake port of the internal combustion engine;
A second fuel injection device that injects into the cylinder of the internal combustion engine a second fuel having a lower octane number than the mixed fuel separated from the mixed fuel;
An internal combustion engine control device comprising: an injection control device that controls an injection timing and an injection amount of fuel from the first and second fuel injection devices;
The injection control device injects the second fuel after injecting the first fuel, and the operation state specified by the load parameter and the rotational speed of the engine changes to the low rotation side or the high load side. 2. A control device for an internal combustion engine, wherein the fuel injection timing is delayed.   The said injection control apparatus injects the said 2nd fuel from an intake process to a compression process, when the driving | running state of the said engine is in the area | region of a low rotation and a high load. Control device for internal combustion engine.