JP2009293530A - Control device for internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a control device for an internal combustion engine surely reducing exhaust emission while avoiding the degradation of fuel economy and a fall of output. <P>SOLUTION: The internal combustion engine 1 constituted to change and control a valve overlap quantity in common for all cylinders according to an operating state includes a cylinder internal pressure sensor 15 detecting a cylinder internal pressure, and an ECU 20. The ECU 20 detects a combustion form for each cylinder and computes a deviation between the detected combustion form of each cylinder and a target combustion form. Based on the deviation for each cylinder, a fuel injection form and a fuel injection rate to the cylinder are changed to eliminate the deviation. <P>COPYRIGHT: (C)2010,JPO&INPIT

Description

  The present invention relates to an internal combustion engine, and more particularly to a control device for an internal combustion engine in which the valve overlap amount is changed and controlled in common for all cylinders in accordance with operating conditions.

  Conventionally, an internal combustion engine in which the valve overlap amount is changed and controlled in common for all cylinders in accordance with the operating state is well known.

  In addition, the combustion start timing (spark ignition timing or compression ignition timing) of the air-fuel mixture in the combustion chamber of the internal combustion engine is set to an optimal timing (MBT: Minimum advance for Best Torque) that can obtain a large torque and does not cause knocking. In addition, there is known one that advances or retards the combustion start timing based on the combustion ratio in the combustion chamber (see, for example, Patent Document 1).

  Furthermore, there is also known a method of controlling the ignition timing and the valve overlap amount so that the combustion ratio in the combustion chamber becomes the target combustion ratio in correspondence with the required torque required for the internal combustion engine (for example, Patent Documents). 2).

Japanese Patent Laid-Open No. 9-189281 JP 2007-32531 A

  By the way, when controlling the valve overlap amount according to the above-described operation state, in the case of an internal combustion engine having a variable valve timing mechanism for each cylinder, control of the valve overlap amount to satisfy the target combustion mode for each cylinder. Can be done.

  However, in the case of an internal combustion engine having a variable valve timing mechanism that changes the valve timing uniformly for all cylinders, if the valve overlap amount is controlled in accordance with a certain cylinder, the valve overlap amount is common to all cylinders. As a result, when there is a machine difference variation between the cylinders, the target combustion mode can be satisfied for a certain cylinder, but the remaining cylinders may not be sufficiently satisfied. In other words, the combustion mode over all the cylinders cannot be accurately controlled, and there is a possibility that the fuel consumption deteriorates, the output decreases, or the exhaust emission increases.

  Accordingly, an object of the present invention is to provide a control device for an internal combustion engine that can reliably reduce exhaust emission while avoiding deterioration of fuel consumption and output.

  In order to achieve the above object, one form of a control apparatus for an internal combustion engine according to the present invention is a combustion engine for each cylinder in an internal combustion engine in which the valve overlap amount is changed and controlled in common for all cylinders according to the operating state. It is calculated by the combustion form detecting means for detecting the form, the combustion form deviation calculating means for calculating the deviation between the combustion form for each cylinder detected by the combustion form detecting means, and the combustion form deviation calculating means. And a fuel injection mode changing means for changing the fuel injection mode to the cylinder in order to eliminate the deviation based on the deviation for each cylinder.

  According to one aspect of the control apparatus for an internal combustion engine according to the present invention, when the valve overlap amount is changed in common for all the cylinders according to the operating state, the combustion form is detected for each cylinder by the combustion form detecting means. The detected combustion mode for each cylinder is compared with the target combustion mode set corresponding to the operating state by the combustion mode deviation calculating means, and the deviation from the target combustion mode is calculated. Further, the fuel injection form changing means changes the fuel injection form to the cylinder so as to eliminate the deviation based on the deviation for each cylinder calculated by the combustion form deviation calculating means. As a result of the change to the fuel injection mode that eliminates this deviation, the cylinder in which the deviation exists becomes the target combustion mode, so that the variation between the cylinders is eliminated, and the exhaust gas is avoided while avoiding the deterioration of fuel consumption and output. Emissions can be reduced reliably.

  Here, the combustion mode detecting means detects the in-cylinder pressure for each cylinder, and the combustion for calculating the combustion ratio at a predetermined timing based on the in-cylinder pressure detected by the in-cylinder pressure detecting means. It is preferable to include a ratio calculation unit.

  The combustion mode deviation calculating means selects a maximum combustion period among the combustion periods for each cylinder based on the combustion ratio calculated by the combustion ratio calculating means, and the maximum combustion period and the target combustion period coincide with each other. A minimum valve overlap amount control means for controlling to the minimum valve overlap amount, and a deviation calculation means for calculating a deviation from the target combustion period for each cylinder in an operation state based on the minimum valve overlap amount. It is preferable.

  Further, the fuel injection mode changing means includes an in-cylinder fuel injection means that is provided for each cylinder and directly injects fuel into the cylinder, and an injection ratio between the initial stage and the latter stage of the fuel supplied by the in-cylinder fuel injection means It is preferable to include a fuel injection ratio control means for controlling the cylinder for each cylinder.

  The combustion ratio calculating means is based on a product value of an in-cylinder pressure detected by the in-cylinder pressure detecting means and a value obtained by raising the cylinder volume at the time of detection of the cylinder pressure by a predetermined exponent. It is preferable to calculate the ratio.

  Further, the predetermined timing is between a first timing set after the intake valve is closed and before the start of combustion, and a second timing set after the start of the combustion and before the exhaust valve is opened. The combustion ratio calculating means is configured to determine the difference between the product values between the first timing and the second timing, and the product between the first timing and the predetermined timing. It is preferable to calculate the combustion ratio based on the difference between the values.

  According to the present invention, in an internal combustion engine in which the valve overlap amount is changed and controlled in common to all cylinders according to the operating state, exhaust emission is reliably reduced while avoiding deterioration of fuel consumption and output. It is possible to realize a control device for an internal combustion engine.

  Hereinafter, the best mode for carrying out the present invention will be specifically described with reference to the drawings.

  FIG. 1 is a schematic configuration diagram showing an internal combustion engine to which a control device according to the present invention is applied. The internal combustion engine 1 shown in FIG. 1 generates power by burning a mixture of fuel and air in a combustion chamber 3 formed in a cylinder block 2 and reciprocating a piston 4 in the combustion chamber 3. Is. The internal combustion engine 1 is configured as a multi-cylinder engine, and the internal combustion engine 1 of the present embodiment is configured as a four-cylinder engine having, for example, # 1 to # 4 cylinders.

  The intake port of each combustion chamber 3 is connected to an intake pipe (intake manifold) 5, and the exhaust port of each combustion chamber 3 is connected to an exhaust pipe 6 (exhaust manifold). In addition, an intake valve Vi and an exhaust valve Ve are provided for each combustion chamber 3 in the cylinder head of the internal combustion engine 1. Each intake valve Vi opens and closes a corresponding intake port, and each exhaust valve Ve opens and closes a corresponding exhaust port. Each intake valve Vi and each exhaust valve Ve are opened and closed by a variable valve mechanism VM including a variable valve timing mechanism. Note that the variable valve timing mechanism in the present embodiment changes the valve overlap amount common to all cylinders by, for example, changing the phase of a camshaft (not shown) that opens and closes each intake valve Vi and each exhaust valve Ve. Can be changed and controlled. Further, the internal combustion engine 1 has a number of spark plugs 7 corresponding to the number of cylinders, and the spark plugs 7 are disposed in the cylinder heads so as to face the corresponding combustion chambers 3.

  The intake pipe 5 is connected to a surge tank 8 as shown in FIG. An air supply line L1 is connected to the surge tank 8, and the air supply line L1 is connected to an air intake port (not shown) via an air cleaner 9. A throttle valve 10 (in this embodiment, an electronically controlled throttle valve) 10 and an air flow meter 13 are incorporated in the air supply line L1 (between the surge tank 8 and the air cleaner 9). On the other hand, as shown in FIG. 1, for example, a front-stage catalyst device 11 a including a three-way catalyst and a rear-stage catalyst device 11 b including a NOx storage reduction catalyst are connected to the exhaust pipe 6.

  Furthermore, the internal combustion engine 1 has a plurality of injectors 12, and each injector 12 is arranged in a cylinder head so that fuel can be directly injected into the corresponding combustion chamber 3 as shown in FIG. 1. Yes. Each piston 4 of the internal combustion engine 1 is configured as a so-called deep dish top surface type, and has a concave portion 4a on its upper surface. In the internal combustion engine 1, fuel such as gasoline is directly injected from each injector 12 toward the recess 4 a of the piston 4 in each combustion chamber 3 in a state where air is sucked into each combustion chamber 3. As a result, in the internal combustion engine 1, the fuel / air mixture layer is formed (stratified) in the vicinity of the spark plug 7 so as to be separated from the surrounding air layer. And stable stratified combustion can be performed.

  Each of the spark plugs 7, the throttle valve 10, the injectors 12, the variable valve mechanism VM and the like described above are electrically connected to an ECU 20 that functions as a control device for the internal combustion engine 1. The ECU 20 includes a CPU, a ROM, a RAM, an input / output port, a storage device, and the like, all not shown. As shown in FIG. 1, various sensors such as an air flow meter 13 and a crank angle sensor 14 are electrically connected to the ECU 20. The ECU 20 uses various maps and the like stored in the storage device according to the driving state and also obtains a spark plug 7, a throttle valve 10, and an injector 12 so as to obtain a desired output based on detection values of various sensors. The variable valve mechanism VM is controlled.

  The internal combustion engine 1 has in-cylinder pressure sensors (in-cylinder pressure detecting means) 15 including semiconductor elements, piezoelectric elements, magnetostrictive elements, optical fiber detecting elements, and the like corresponding to the number of cylinders. Each in-cylinder pressure sensor 15 is disposed on the cylinder head so that the pressure receiving surface faces the corresponding combustion chamber 3, and is electrically connected to the ECU 20 via an A / D converter (not shown). Each in-cylinder pressure sensor 15 outputs the pressure (in-cylinder pressure) applied to the pressure receiving surface in the combustion chamber 3 as a relative value to the atmospheric pressure, and corresponds to the pressure (in-cylinder pressure) applied to the pressure receiving surface. A voltage signal (a signal indicating a detected value) is supplied to the ECU 20.

  Further, the internal combustion engine 1 has an intake pressure sensor that detects the pressure (intake pressure) of intake air in the surge tank 8 as an absolute pressure. The intake pressure sensor is also electrically connected to the ECU 20 via an A / D converter (not shown) or the like, and gives a signal indicating the detected absolute pressure of the intake air in the surge tank 8 to the ECU 20. The detection values of the air flow meter 13, the crank angle sensor 14, and the intake pressure sensor are sequentially given to the ECU 20 every minute time, and are stored and held in a predetermined storage area (buffer) of the ECU 20 by a predetermined amount. The detection value (in-cylinder pressure) of each in-cylinder pressure sensor 15 is subjected to absolute pressure correction based on the detection value of the intake pressure sensor, and then stored and held in a predetermined storage area (buffer) of the ECU 20 by a predetermined amount. The

  Next, an example of the control procedure of the combustion mode control executed by the ECU 20 in the present embodiment will be described.

  First, in this embodiment, according to the main routine shown in the flowchart of FIG. 2, first, in step S210, the operating state of the internal combustion engine 1 is obtained by reading parameters such as the intake air amount and the engine speed. In the next step S220, basic control values such as the ignition timing SA, the total fuel injection amount T, and the valve overlap amount OL corresponding to the operating state are read from the basic setting map stored in the ECU 20. Is set. In step S230, the control amount of each actuator is determined based on the basic control value and the correction value as necessary, and each actuator is operated corresponding to the control amount. In the following description, it is assumed that this correction is not performed for ease of understanding and simplification of explanation. In addition, the actuator includes an igniter (not shown) of the spark plug 7 for the ignition timing SA, the injector 12 for the total fuel injection amount T, and the camshaft phase position of the variable valve timing mechanism for the valve overlap amount OL. Corresponds to a hydraulic actuator.

  Regarding the valve overlap amount OL, a basic control value is set to be as large as possible within a range not accompanied by an increase in emission due to deterioration of combustion in order to reduce pumping loss and increase efficiency. When the valve overlap amount OL is controlled so as to increase, the valve overlap amount OL is advanced from the reference position by the variable valve timing mechanism so that the phase of the camshaft for driving the intake valves Vi to open and close is increased. And / or by retarding the phase of the camshaft for opening / closing each exhaust valve Ve from the reference position. In addition, what is necessary is just to reverse to make it small.

  In the next step S240, the combustion mode is detected for each of the cylinders # 1 to # 4. Details of detection of this combustion mode will be described later. Further, the process proceeds to step S250, and the detected combustion mode for each cylinder is compared with the target combustion mode set corresponding to the above-described operation state, and the deviation from the target combustion mode is calculated. Then, based on the calculated deviation for each cylinder in the next step S260, the fuel injection mode to the cylinder is changed so as to eliminate this deviation. More specifically, the injection ratio between the initial stage and the late stage of the fuel is changed for each cylinder in accordance with the magnitude of the deviation.

  Here, the method of detecting the combustion mode in step S240 described above will be described. Detection of the combustion mode in the present embodiment is performed by calculating the combustion rate MFB at a predetermined timing for a certain combustion chamber.

Therefore, when the crank angle is θ, the in-cylinder pressure detected by the in-cylinder pressure sensor 15 as the in-cylinder pressure detecting means is P (θ), and the in-cylinder volume when the crank angle is θ is V (θ). When the specific heat ratio is κ, the product value P (θ) · V of the in-cylinder pressure P (θ) and the value V ^κ (θ) obtained by raising the in-cylinder volume V (θ) to the specific heat ratio κ The change pattern of ^κ (θ) (hereinafter referred to as “PV ^κ ” as appropriate) and the change pattern of the heat generation amount Q in the combustion chamber of the internal combustion engine with respect to the crank angle have a correlation as shown in FIG. In FIG. 3, the solid line shows the in-cylinder pressure detected at a predetermined minute crank angle in a predetermined model cylinder, and the value obtained by raising the in-cylinder volume at the time of detection of the in-cylinder pressure by a predetermined specific heat ratio κ. It plots the product value PV ^kappa of the dashed line on the basis of the heat generation amount Q in the model cylinder following equation (1) is obtained by calculating and plotted as Q = ∫dQ / dθ · Δθ. In any case, for the sake of simplicity, the specific heat ratio κ = 1.32. In FIG. 3, −360 °, 0 °, and 360 ° correspond to the top dead center, and −180 ° and 180 ° correspond to the bottom dead center.

As can be seen from the results shown in FIG. 3, the change pattern of the heat generation amount Q with respect to the crank angle and the change pattern of the product value PV ^κ with respect to the crank angle are substantially the same (similar), and in particular, Before and after the start of combustion of the air-fuel mixture (at the time of spark ignition for a gasoline engine and at the time of compression ignition for a diesel engine) (for example, a range from about −180 ° to about 135 ° in FIG. 3) It can be seen that the change pattern of the product value PV ^κ agrees very well.

In a preferred embodiment of the present invention, the in-cylinder pressure detected by the in-cylinder pressure sensor 15 and the in-cylinder pressure are detected using the correlation between the heat generation amount Q in the combustion chamber and the product value PV ^κ . Based on the product value PV ^κ with the in-cylinder volume, a combustion ratio MFB that is a ratio of a heat generation amount up to a predetermined timing between the two points with respect to a total heat generation amount between the two points is obtained.

This is because if the combustion rate MFB in the combustion chamber is calculated based on the product value PV ^κ , the combustion rate in the combustion chamber can be obtained with high accuracy without requiring high-load calculation processing. That is, as shown in FIG. 4, the combustion rate obtained based on the product value PV ^κ (see the solid line in the figure) is substantially the same as the combustion rate obtained based on the heat generation rate (see the broken line in the figure). To do.

  In FIG. 4, the solid line represents the combustion ratio MFB at the timing when the crank angle = θ in the model cylinder described above, calculated according to the following equation (2), and plotted based on the detected in-cylinder pressure P (θ). It is a thing. However, for simplicity, the specific heat ratio κ = 1.32. In FIG. 4, the broken line indicates the combustion ratio MFB at the timing when the crank angle = θ in the model cylinder described above in accordance with the above equation (1) and the following equation (3), and the detected in-cylinder pressure P (θ ) And plotted. Also in this case, for simplicity, κ = 1.32.

  Further, the predetermined timing is set between a first timing set after the intake valve is closed and before the start of combustion and a second timing set after the start of combustion and before the exhaust valve is opened. Preferably, the combustion ratio calculation means is based on the difference between the product values between the first timing and the second timing and the difference between the product values between the first timing and the predetermined timing. It is preferable to calculate the combustion ratio MFB.

In this case, if the crank angle at the predetermined timing is θ ₀ , the combustion ratio MFB at the predetermined timing at which the crank angle = θ ₀ is the product value PV ^κ between the first timing and the predetermined timing. Difference {P (θ ₀ ) · V ^κ (θ ₀ ) −P (θ ₁ ) · V ^κ (θ ₁ )} as a difference between product values PV ^κ between the first timing and the second timing. It can be obtained by dividing by {P (θ ₂ ) · V ^κ (θ ₂ ) −P (θ ₁ ) · V ^κ (θ ₁ )} and multiplying by 100. Thereby, it becomes possible to obtain | require a combustion ratio accurately based on the in-cylinder pressure detected in three points, and a calculation load can be reduced significantly.

Therefore, when calculating the combustion ratio MFB as the combustion mode in step S240, the ECU 20 first starts from the predetermined storage area for the combustion chamber 3 of the target cylinder from the first storage after the intake valve Vi is closed and before ignition. In-cylinder pressure P (θ ₁ ) at the timing (timing at which the crank angle becomes θ ₁ ) and in-cylinder pressure P at the second timing (timing at which the crank angle becomes θ ₂ ) after ignition and before opening the exhaust valve In-cylinder at a predetermined timing that is predetermined between (θ ₂ ) and the first timing and the second timing, and the crank angle = θ ₀ (where θ ₁ <θ ₀ <θ ₂ ). Read the pressure P (θ ₀ ).

The crank angle θ ₁ is preferably set to a timing sufficiently before the time point at which combustion starts in the combustion chamber 3 (at the time of ignition), and is set to −60 °, for example. The crank angle θ ₂ is preferably set to a timing at which combustion of the air-fuel mixture in the combustion chamber 3 is almost completed, for example, 90 °. In addition, the predetermined timing between the first timing and the second timing has a crank angle θ ₀ = 8 ° (experimentally and empirically known that the combustion ratio MFB is approximately 50% ( The timing is set to 8 ° after top dead center.

  Note that the crank angle at which the combustion ratio MFB is approximately 50% varies depending on the cooling loss of the internal combustion engine, and slightly varies from 8 ° after top dead center depending on the model.

When the ECU ₂₀ reads the in-cylinder pressure P (θ ₁ ), the in-cylinder pressure P (θ ₀ ), and the in-cylinder pressure P (θ ₂ ), the product value when the crank angle becomes θ ₁ , θ _0, and θ _2. P (θ ₁ ) · V ^κ (θ ₁ ), P (θ ₀ ) · V ^κ (θ ₀ ) and P (θ ₂ ) · V ^κ (θ ₂ ) are calculated. That is, the ECU 20 determines the in-cylinder pressure P (θ ₁ ) and the in-cylinder pressure P (θ ₁ ), that is, the in-cylinder volume V (θ ₁ ) when the crank angle is θ ₁ , the specific heat ratio κ. In this embodiment, a product value P (θ ₁ ) · V ^κ (θ ₁ ), which is a product of the value raised to the power of κ = 1.32, is calculated. Similarly, the ECU 20 obtains a product value P (product which is a product of the in-cylinder pressure P (θ ₀ ) and the value obtained by raising the in-cylinder volume V (θ ₀ ) when the crank angle becomes θ ₀ by the specific heat ratio κ. θ ₀ ) · V ^κ (θ ₀ ), the in-cylinder pressure P (θ ₂ ), and the value obtained by raising the in-cylinder volume V (θ ₂ ) when the crank angle is θ ₂ by the specific heat ratio κ A product value P (θ ₂ ) · V ^κ (θ ₂ ), which is a product, is calculated.
The values of V ^κ (θ ₁ ), V ^κ (θ ₀ ), and V ^κ (θ ₂ ) are calculated in advance and stored in the storage device.

Then, the ECU ₂₀ determines the product values P (θ ₁ ) · V ^κ (θ ₁ ), P (θ ₀ ) · V ^κ (θ ₀ ) and P (when the crank angles are θ ₁ , θ ₀ and θ _2. Using θ ₂ ) · V ^κ (θ ₂ ), the combustion ratio MFB at the timing when the crank angle becomes θ ₀ is calculated from the following equation (4). As a result, the combustion ratio MFB is accurately obtained based on the in-cylinder pressures P (θ ₁ ), P (θ ₀ ), and P (θ ₂ ) detected at the three points.

  Next, the relationship between the above-described combustion ratio MFB as the combustion mode, the combustion period Θ, and the valve overlap amount OL will be described with reference to the graph of FIG.

  FIG. 5 is a graph showing a state in which the combustion ratio MFB changes according to the change in the valve overlap amount OL, with the combustion ratio MFB on the vertical axis and the crank angle CA on the horizontal axis. Is small (indicated by the curve a), the rising gradient of the combustion ratio MFB is large, and when the valve overlap amount OL is large (indicated by the curve b), the rising gradient of the combustion ratio MFB is small. SA is the ignition timing, that is, the combustion start timing, and TDC is the top dead center.

Here, the combustion period Θ can also be referred to as a crank angle period from the combustion start timing when the combustion ratio MFB is 0% to the combustion end timing when the combustion ratio MFB is 100%. The crank angle period from the combustion start timing when the combustion ratio MFB is used to the combustion end timing when the combustion ratio MFB exceeds 90% is used. Therefore, in the example shown in FIG. 5, the combustion period when the valve overlap amount OL indicated by the curve a is small is Θ _S , and the combustion period when the valve overlap amount OL indicated by the curve b is large is Θ _L. It is. From this, it can be seen that the combustion period Θ becomes longer as the valve overlap amount OL increases.

Thus, the detection of the combustion mode for each cylinder executed in step S240 described above is performed based on the in-cylinder pressure detected by the in-cylinder pressure sensor 15 at three points including a predetermined timing in the present embodiment. The ratio MFB is obtained, and is obtained as Θi (Θ _{# 1} , Θ _{# 2} , Θ _{# 3} , Θ _{# 4} ) for each cylinder having the combustion period Θ of # 1 to # 4.

Then, the calculation of the deviation of the combustion mode for each cylinder from the target combustion mode executed in step S250 described above is executed according to the subroutine shown in the flowchart of FIG. 6 in the present embodiment. First, in step S2501, the maximum combustion period Θmax among the combustion periods Θi (Θ _{# 1} , Θ _{# 2} , Θ _{# 3} , Θ _{# 4} ) for each cylinder of # 1 to # 4 is selected. Then, in the next step S2502, the valve overlap amount is set so that the maximum combustion period Θmax and the target combustion period Θt as the target combustion mode set corresponding to the operation state read in step S210 coincide. The OL is feedback controlled. As a result of this feedback control, the valve overlap amount becomes the minimum valve overlap amount OLmin.

  Therefore, in the next step S2503, the combustion period Θi for each cylinder of # 1 to # 4 is obtained again in the operation state of the next cycle with the minimum valve overlap amount OLmin. In step S2504, the cylinder deviation αi between the target combustion period Θt and the combustion period Θi is calculated for each cylinder by αi = Θt−Θi.

Here, the relationship between the target combustion period Θt and the cylinder deviation αi will be further described with reference to the graph of FIG. In FIG. 7, when the vertical axis represents the combustion period Θ and the horizontal axis represents the valve overlap amount OL, and there is a variation between the cylinders, that is, there is a difference, the valve overlap amount OL is set to a predetermined value due to the difference. It is the graph which showed a mode that the combustion period (THETA) will differ when it is a value. However, in FIG. 7, as a result of feedback control of the valve overlap amount OL to the minimum valve overlap amount OLmin so that the above-described maximum combustion period Θmax and the target combustion period Θt coincide with each other, as an example, # 3 cylinder The case where α _{# 3} = 0, in which the combustion period Θ _{# 3} of the same coincides with the target combustion period Θt is shown. Note that the graph of FIG. 7 is for explanation only, and the cylinder deviation αi with respect to the target combustion period Θt caused by the difference between the cylinders is exaggerated. In the graph of FIG. 7, the cylinder deviation in the # 2 cylinder is specifically exemplified as α _{# 2} = Θt−Θ _{# 2} , but the same applies to the other cylinders # 1 and # 4.

In the fuel injection mode change process executed in step S260 described above, the fuel injection mode is changed so as to eliminate the deviation based on the calculated cylinder deviation αi for each cylinder. Needless to say, there is no need to change the # 3 cylinder in which the deviation α _{# 3} = 0.

In the present embodiment, the change of the fuel injection mode is changed for each cylinder in accordance with the magnitude of the deviation αi, that is, the injection ratio between the initial stage and the late stage of the fuel injected from the injector 12. The Therefore, if β is determined as the injection ratio conversion coefficient, the total fuel injection amount T set corresponding to the operation state read in step S210 is injected in the initial and late stages as shown in the following equation (5). It is divided and injected from the injector 12.
Total fuel injection amount T = (1−αi × β) × T + αi × β × T (5)

  Here, the previous term on the right side of the above equation (5) is the initial fuel injection amount, and the latter term is the late fuel injection amount. The timing of the initial fuel injection is a compression stroke before the intake stroke or the ignition timing SA, and the timing of the late fuel injection may be an appropriate timing after the start of combustion, and is preferably after top dead center. The injection ratio conversion coefficient β is set so as to appropriately extend the combustion period Θ in accordance with the magnitude of the cylinder deviation αi. However, as shown in FIG. 8, if the ratio of the late fuel injection is increased too much (for example, more than 20%) in order to lengthen the combustion period Θ, the initial fuel injection amount becomes too small and the spark plug 7 Since there is a possibility of so-called misfire that does not start even by ignition (the misfire region is indicated by hatching in FIG. 8), an appropriate value is obtained in advance by experiments and stored in the ROM of the ECU 20 as a map.

  In the above-described embodiment, the intake air amount is read in step S210 to grasp the operation state. However, this intake air amount is not compressed by the air flow meter 13 but is used during the compression stroke using the in-cylinder pressure sensor 15. The amount of air taken into the combustion chamber may be calculated based on the in-cylinder pressure detected at a predetermined timing before the start of combustion.

It is a schematic block diagram which shows the internal combustion engine to which the control apparatus of this invention was applied. It is a flowchart for demonstrating the control procedure of the fuel injection form change by the control apparatus of this invention. It is a graph which shows the correlation with the product value PV ( ^kappa) used in this invention, and the amount of heat generation in a combustion chamber. It is a graph which shows the correlation of the combustion rate calculated | required based on product value PV ( ^kappa) , and the combustion rate calculated ^| required based on a heat release rate. It is a graph for demonstrating the relationship between the combustion ratio MFB as a combustion form, combustion period (THETA), and valve overlap amount OL. It is a flowchart for demonstrating the control procedure which calculates the deviation with the target combustion form by the control apparatus of this invention. It is a graph which shows a mode that combustion period (THETA) differs when valve overlap amount OL is a predetermined value resulting from the variation between cylinders. It is a graph which shows the relationship between the ratio of late fuel injection, and a misfire area | region.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Internal combustion engine 3 Combustion chamber 7 Spark plug 12 Injector 14 Crank angle sensor 15 In-cylinder pressure sensor 20 ECU
Ve Exhaust valve Vi Intake valve VM Variable valve mechanism

Claims (6)

In an internal combustion engine that changes and controls the valve overlap amount in common to all cylinders according to the operating state,
Combustion mode detection means for detecting the combustion mode for each cylinder;
Combustion mode deviation calculating means for calculating a deviation from the target combustion mode of the combustion mode for each cylinder detected by the combustion mode detecting unit;
Fuel injection mode changing means for changing the fuel injection mode to the cylinder in order to eliminate the deviation based on the deviation for each cylinder calculated by the combustion mode deviation calculating unit;
A control device for an internal combustion engine, comprising:   The combustion form detecting means includes an in-cylinder pressure detecting means for detecting an in-cylinder pressure for each cylinder, and a combustion ratio calculating means for calculating a combustion ratio at a predetermined timing based on the in-cylinder pressure detected by the in-cylinder pressure detecting means. The control device for an internal combustion engine according to claim 1, comprising:   The combustion mode deviation calculating means selects the maximum combustion period among the combustion periods for each cylinder based on the combustion ratio calculated by the combustion ratio calculating means so that the maximum combustion period and the target combustion period coincide with each other. Minimum valve overlap amount control means for controlling to the minimum valve overlap amount, and deviation calculation means for calculating a deviation from the target combustion period for each cylinder in the operation state by the minimum valve overlap amount. The control apparatus for an internal combustion engine according to claim 2, wherein the control apparatus is an internal combustion engine.   The fuel injection mode changing means includes an in-cylinder fuel injection means that is provided for each cylinder and directly injects fuel into the cylinder, and an injection ratio between the initial period and the latter period of the fuel supplied by the in-cylinder fuel injection means. And a fuel injection ratio control means for controlling each of them.   The combustion ratio calculation means calculates the combustion ratio based on a product value of an in-cylinder pressure detected by the in-cylinder pressure detection means and a value obtained by raising the cylinder volume at the time of detection of the cylinder pressure by a predetermined index. 5. The control device for an internal combustion engine according to claim 2, wherein the control device calculates the internal combustion engine.   The predetermined timing is set between a first timing set after the intake valve is closed and before the start of combustion, and a second timing set after the start of the combustion and before the exhaust valve is opened. The combustion ratio calculating means calculates the difference between the product values between the first timing and the second timing, and the product value between the first timing and the predetermined timing. 6. The control device for an internal combustion engine according to claim 5, wherein the combustion ratio is calculated based on the difference.

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