JP4075862B2 - Engine ignition timing control device - Google Patents

Description

  The present invention relates to an ignition timing control device for an engine (internal combustion engine), and more particularly to an ignition timing control device that controls a minimum ignition advance value (so-called MBT) necessary for generating a maximum shaft torque.

The combustion speed in the combustion period in which the air-fuel mixture burns in the combustion chamber is divided into a plurality of times, the turbulent flame speed in each divided combustion period is calculated based on the operating conditions, and the MBT is calculated based on the plurality of turbulent flame speeds. Is calculated as the basic ignition timing (see Patent Document 1).
JP 2003-148236 A

  By the way, according to the technique described in Patent Document 1, the laminar flame speeds SL1 and SL2, which are flame propagation speeds in the absence of gas flow, are multiplied by the average turbulence strengths ST1 and ST2 during the combustion period. Thus, turbulent flame speeds FLAME1 and FLAME2, which are flame propagation speeds in a state where there is a gas flow, are calculated.

  In this case, when calculating the average turbulence strengths ST1 and ST2 during the combustion period, the configuration is merely a value corresponding to the rotational speed NRPM of the engine, and the influence of the squish is not dealt with in particular.

  Here, the squish is a gas flow in the radial direction of the cylinder caused by the air-fuel mixture existing in a narrow gap between the piston head and the cylinder head lower surface being pushed into the main space of the combustion chamber near the end of the compression stroke. Therefore, this squish affects the average turbulence intensity during the combustion period. In addition, this influence differs for each engine model.

  For this reason, according to the technique described in Patent Document 1, it is necessary to adapt the average turbulence intensity during the combustion period every time the engine model is different, and the number of man-hours to be increased is increased.

  For this reason, the present inventors conducted various experiments on the relationship between the squish and the average turbulence intensity during the combustion period, and gained new knowledge, whether the number of man-hours can be reduced even if the engine model is different.

  Therefore, the present invention is based on this new knowledge, and an object of the present invention is to provide an apparatus that minimizes the number of man-hours required for calculating the average turbulence intensity during the combustion period.

The present invention divides the combustion period in which the air-fuel mixture burns in the combustion chamber into a plurality of parts, calculates the combustion speed (FLAME1, FLAME2) in each of the divided combustion periods (BURN1, BURN2) based on the operating conditions, An ignition timing control device for an engine that calculates the ignition timing at which MBT is obtained based on the combustion speed of the engine as a basic ignition timing and performs spark ignition at this basic ignition timing. The combustion speed in the combustion period including the squish generation period is the turbulent flame speed in the combustion period including the squish generation period, and the flame speed correction amount (Kq) by the squish is set to at least the engine specifications. based calculated, turbulent flame speed in the combustion period including the period of occurrence of the squish And configured to calculate, based on the at least one of the operating conditions of the laminar flame speed and the engine in the combustion period including the period of occurrence of the squish, the flame speed correction amount according to squish and (Kq).

Further, according to the present invention, the combustion period in which the air-fuel mixture burns in the combustion chamber is divided into an initial combustion period (BURN1) and a main combustion period (BURN2), and the combustion speed in each of the divided combustion periods (BURN1, BURN2). (FLAME1, FLAME2) is calculated based on the operating conditions, and the ignition timing at which MBT is obtained is calculated as the basic ignition timing MBTCAL based on the plurality of combustion speeds (FLAME1, FLAME2), and spark ignition is performed at this basic ignition timing MBTCAL. In the engine ignition timing control apparatus, the main combustion period includes a period in which squish occurs, the combustion speed in the main combustion period is a turbulent flame speed in the main combustion period, and flame speed correction by squish The amount (Kq) is calculated based on at least the engine specifications, and the main fuel The turbulent flame speed (FLAME2) in the period, and at least one of the operating conditions of the laminar flame speed and the engine at the main combustion period, as calculated based on the flame speed correction amount and (Kq) by the squish Configure.

A new experimental result has been obtained that the extent to which the squish affects the combustion speed is determined by the shape of the squish. That is, since the shape of the squish is determined by the engine specifications, the degree to which the squish influences the turbulent flame speed during the combustion period including the squish generation period and the turbulent flame speed during the main combustion period, depending on the engine specifications. It means that it can be estimated almost. That is, according to the present invention, any one combustion period includes a period during which squish occurs, a flame speed correction amount (Kq) by squish is calculated based on at least engine specifications, and combustion including the period during which squish occurs The turbulent flame speed (FLAME2) in the period is based on at least one of a laminar flame speed and an engine operating condition in a combustion period including a period in which the squish is generated, and a flame speed correction amount (Kq) by the squish. In addition, according to the present invention, the main combustion period includes a period during which squish occurs, and the flame speed correction amount (Kq) by squish is calculated based on at least the engine specifications, and the turbulent flow during the main combustion period flame speed (FLAME2), luck laminar flame speed and the engine at the main combustion period And at least one of the conditions, so configured to calculate, based on the flame speed correction amount (Kq) by the squish, there is no need to engine models again adapted for different, it can be reduced man-hour for adaptation.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings.

  FIG. 1 is a schematic diagram for explaining the system of the present invention.

  The air is stored in the intake collector 2 and then introduced into the combustion chamber 5 of each cylinder via the intake manifold 3. Fuel is injected and supplied from a fuel injector 21 disposed in the intake port 4 of each cylinder. The fuel injected into the air is vaporized and mixed with the air to form a gas (air mixture) and flows into the combustion chamber 5. This air-fuel mixture is confined in the combustion chamber 5 when the intake valve 15 is closed, and is compressed by the rise of the piston 6.

  In order to ignite this compressed air-fuel mixture with a high-pressure spark, an ignition device 11 of an electronic power distribution system is provided in which an ignition coil with a built-in power transistor is arranged in each cylinder. That is, the ignition device 11 includes an ignition coil 13 that stores electrical energy from the battery, a power transistor that energizes and shuts off the primary side of the ignition coil 13, and a primary current of the ignition coil 13 that is provided on the ceiling of the combustion chamber 5. It includes a spark plug 14 that receives a high voltage generated on the secondary side of the ignition coil 13 due to interruption of the spark coil 13 and performs spark discharge.

  When a spark is blown off by the spark plug 14 slightly before the compression top dead center and the compressed mixture is ignited, the flame spreads and then explosively burns, and the gas pressure by this combustion works to push down the piston 6. This work is taken out as the rotational force of the crankshaft 7. The combusted gas (exhaust gas) is discharged into the exhaust passage 8 when the exhaust valve 16 is opened.

A three-way catalyst 9 is provided in the exhaust passage 8. When the air-fuel ratio of the exhaust gas is in a narrow range (window) centered on the stoichiometric air-fuel ratio, the three-way catalyst 9 can efficiently remove harmful three components such as HC, CO, and NOx contained in the exhaust gas simultaneously. Since the air-fuel ratio is the ratio of the intake air amount and the fuel amount, the intake air amount introduced into the combustion chamber 5 per one cycle of the engine (crank angle 720 ° section in a four-cycle engine) and the fuel injector 21 The engine controller 31 uses the intake air flow rate signal from the air flow meter 32 and the fuel from the fuel injector 21 based on the signals from the crank angle sensors (33, 34) so that the ratio to the fuel injection amount becomes the stoichiometric air-fuel ratio. The injection amount is determined, and the air-fuel ratio is feedback controlled based on a signal from an O ₂ sensor 35 provided upstream of the three-way catalyst 9.

  A so-called electronically controlled throttle 22 in which a throttle valve 23 is driven by a throttle motor 24 is provided upstream of the intake collector 2. Since the torque required by the driver appears in the amount of depression of the accelerator pedal 41 (accelerator opening), the engine controller 31 determines a target torque based on a signal from the accelerator sensor 42, and a target air for realizing this target torque. The amount is determined, and the opening degree of the throttle valve 23 is controlled via the throttle motor 24 so as to obtain this target air amount.

  Cam sprockets and crank sprockets are attached to the front portions of the intake valve camshaft 25, the exhaust valve camshaft 26 and the crankshaft 7, respectively, and a timing chain (not shown) is hung around these sprockets so that the camshaft 25 and 26 are driven by the crankshaft 7 of the engine, and are interposed between the cam sprocket and the intake valve camshaft 25 to continuously adjust the phase of the intake valve cam with a constant operating angle. An exhaust valve cam control phase (hereinafter referred to as an “intake VTC mechanism”) 27 that can be controlled, and a cam sprocket and an exhaust valve camshaft 26 are interposed between the camshaft 26 and the exhaust valve cam. Exhaust valve timing control mechanism (hereinafter referred to as “exhaust VTC machine”) "I referred to.) And a 28. When the opening / closing timing of the intake valve 15 and the opening / closing timing of the exhaust valve 16 are changed, the amount of the inert gas remaining in the combustion chamber 5 changes. As the amount of the inert gas in the combustion chamber 5 increases, the pumping loss decreases and the fuel consumption improves. Therefore, the target intake valve closing timing and the target exhaust gas indicate how much inert gas should remain in the combustion chamber 5 depending on the operating conditions. The valve closing timing is determined in advance, and the engine controller 31 determines the target intake valve closing timing and the target exhaust valve closing timing from the operating conditions (engine load and rotation speed) at that time, so that these target values can be obtained. The intake valve closing timing and the exhaust valve closing timing are controlled via the actuators of the intake VTC mechanism 27 and the exhaust VTC mechanism 28.

  Furthermore, a variable valve mechanism (hereinafter referred to as “VEL mechanism”) 51 configured by a multi-node link-like mechanism that continuously and variably controls the valve lift amount and the operating angle of the intake valve 15 is provided. Since specific configurations of the VEL mechanism 51 and the intake VTC mechanism 27 are known from Japanese Patent Laid-Open No. 2003-3872, detailed description thereof is omitted.

  If the valve lift is reduced, the flow rate of the intake air that passes through the periphery of the intake valve 15 and flows into the combustion chamber 5 is increased. Therefore, the increase in the flow rate of the intake air improves the atomization of the injected fuel. In 31, since the HC discharged from the engine tends to increase, the atomization of the injected fuel is promoted for a while from the start and the HC is reduced. Therefore, the intake valve 15 is operated by operating the VEL mechanism 51 via the actuator of the VEL mechanism 51. Reduce the amount of valve lift. Thereafter, the VEL mechanism 51 is deactivated, the valve lift amount of the intake valve 15 is restored, and a large amount of intake air is caused to flow into the combustion chamber 5.

  An intake air temperature signal from the intake air temperature sensor 43, an intake air pressure signal from the intake air pressure sensor 44, an exhaust gas temperature signal from the exhaust air temperature sensor 45, and an exhaust gas pressure signal from the exhaust air pressure sensor 46 are output from the water temperature sensor 37. The engine controller 31 that is input together with the coolant temperature signal controls the ignition timing that is the primary current cutoff timing of the spark plug 14 via the power transistor 13.

  FIG. 2 is a block diagram of the ignition timing control performed in the engine controller 31. The ignition timing control section 51 and the ignition timing control section 61 are mainly composed. The ignition timing calculator 51 further includes an initial combustion period calculator 52, a main combustion period calculator 53, a combustion period calculator 54, a basic ignition timing calculator 55, and a previous combustion start timing calculator 56.

  The initial combustion period calculation unit 52 calculates a period from when the air-fuel mixture is ignited until flame nuclei are formed as the initial combustion period BURN1. The main combustion period calculation unit 53 calculates a period from when the flame kernel is formed until the combustion pressure reaches the maximum value Pmax as the main combustion period BURN2. The combustion period calculation unit 54 calculates the sum of the initial combustion period BURN1 and the main combustion period BURN2 as the combustion period BURN from the start of combustion to the maximum combustion pressure Pmax. The basic ignition timing calculation unit 55 calculates an ignition timing (this ignition timing is referred to as “basic ignition timing”) MBTCAL from which MBT is obtained based on the combustion period BURN.

  The ignition timing control unit 61 uses the basic ignition timing calculated in this way as an ignition timing command value, and the ignition plug 14 is ignited to the ignition coil 13 so that the ignition plug 14 ignites the air-fuel mixture in the combustion chamber 5 with this command value. The energization angle and the non-energization angle are controlled.

  As described above, the combustion period BURN is calculated by dividing the combustion period BURN into the initial combustion period BURN1 and the main combustion period BURN2, and the basic ignition timing MBTCAL is obtained according to the combustion period BURN, based on the result obtained from the combustion analysis. Is. Hereinafter, the ignition timing control based on the combustion analysis will be further described.

  As shown in FIG. 3, the crank angle at which the combustion pressure of the air-fuel mixture reaches the maximum value Pmax when the air-fuel mixture is ignited with MBT (minimum advance angle value at which maximum torque is obtained) is defined as a reference crank angle θPMAX [degATDC]. The reference crank angle θPMAX is substantially constant regardless of the combustion method, and is generally in the range of 12 to 15 degrees after compression top dead center, and at most 10 to 20 degrees after compression top dead center.

  FIG. 4 shows changes in the combustion mass ratio BR (combustion gas mass ratio) obtained by the combustion analysis in the combustion chamber in the spark ignition engine. The combustion mass ratio BR representing the ratio of the combustion mass to the fuel supplied to the combustion chamber is 0% at the time of ignition, and reaches 100% by complete combustion. Experiments have confirmed that the combustion mass ratio at the reference crank angle θPMAX is constant and about 60%.

  The combustion period corresponding to the change allowance until the combustion mass ratio BR reaches about 60% corresponding to the reference crank angle θPMAX from 0% is a period in which there is almost no change in both the combustion mass ratio and the combustion pressure immediately after the start of combustion. It is divided into an initial combustion period and a main combustion period in which the combustion mass ratio and the combustion pressure increase rapidly. The initial combustion period is a stage from the start of combustion until flame nuclei are formed, and the flame nuclei are formed when the combustion mass ratio changes from 0% to 2% to 10%. During this initial combustion period, the rate of increase in combustion pressure and combustion temperature is small, and the initial combustion period is long with respect to changes in the combustion mass ratio. The length of the initial combustion period is susceptible to changes in temperature and pressure in the combustion chamber.

  On the other hand, in the main combustion period, the flame propagates from the flame kernel to the outside, and the flame speed (that is, the combustion speed) increases rapidly. Therefore, the change in the combustion mass ratio during the main combustion period is larger than the change in the combustion mass ratio during the initial combustion period.

  In the engine controller 31, the period until the combustion mass ratio reaches 2% (changes) is set as the initial combustion period BURN1 [deg], and the period from the end of the initial combustion period BURN1 to the reference crank angle θPMAX (combustion chamber volume ratio) In other words, the main combustion period BURN2 [deg] is distinguished from 2% to about 60%. Then, a combustion period BURN [deg] that is the sum of the initial combustion period BURN1 and the main combustion period BURN2 is calculated, a reference crank angle θPMAX [degATDC] is subtracted from the combustion period BURN, and an ignition dead time equivalent crank described later is further calculated. The crank angle position to which the angle IGNDEAD [deg] is added is set as the basic ignition timing MBTCAL [degBTDC], which is the ignition timing at which MBT is obtained.

  The pressure and temperature in the combustion chamber 5 during the initial combustion period in which flame nuclei are formed are almost equivalent to the pressure and temperature at the time of ignition, but the ignition timing is calculated from this, but accurate ignition is performed from the beginning. The time cannot be set. Therefore, as shown in FIG. 2, the previous combustion start timing calculation unit 56 calculates the previous value of the basic ignition timing as the previous combustion start timing MBTCYCL [degBTDC], and gives this value to the initial combustion period calculation unit 52. In addition, the initial combustion period calculation unit 52 cyclically repeats the calculation of the initial combustion period so as to obtain a highly accurate result without a time delay.

  Next, the calculation of the basic ignition timing MBTCAL executed by the engine controller 31 will be described in detail with reference to the following flowchart.

  FIG. 5 is for calculating various physical quantities necessary for calculating the ignition timing, and is executed at regular time intervals (for example, every 10 msec).

  First, in step 11, the intake valve closing timing IVC [degBTDC], the collector internal temperature TCOL [K] detected by the temperature sensor 43, the collector internal pressure PCOL [Pa] detected by the pressure sensor 44, and the temperature sensor 45 are detected. Exhaust temperature TEXH [K], internal inert gas rate MRESFR [%], cooling water temperature TWK [K] detected by temperature sensor 37, target equivalent ratio TFBYA, engine rotational speed NRPM [rpm] detected by crank angle sensor ], Dead ignition time DEADTIME [μsec] is read.

  Here, the crank angle sensor includes a position sensor 33 for detecting the position of the crankshaft 7 and a phase sensor 34 for detecting the position of the intake camshaft 25. The engine is based on signals from the two sensors 33 and 34. The rotational speed NRPM [rpm] is calculated.

  The intake valve closing timing IVC is known from the command value given to the intake VTC mechanism 27. Alternatively, the actual intake valve closing timing may be detected by the phase sensor 34.

  The internal inert gas ratio MRESFR is a value obtained by dividing the amount of inert gas remaining in the combustion chamber by the total amount of gas in the combustion chamber, and the calculation thereof will be described later. The ignition dead time DEADTIME is a constant value.

  The target equivalent ratio TFBYA is calculated in a fuel injection amount calculation flow (not shown). The target equivalent ratio TFBYA is an unnamed number, and is a value represented by the following expression when the theoretical air-fuel ratio is 14.7.

TFBYA = 14.7 / target air-fuel ratio (1)
For example, from equation (1), when the target air-fuel ratio is the stoichiometric air-fuel ratio, TFBYA = 1.0, and when the target air-fuel ratio is a lean value such as 22.0, TFBYA is a positive value less than 1.0. is there.

In step 12, the volume of the combustion chamber 5 at the intake valve closing timing IVC (that is, the volume at the compression start timing) VIVC [m ³ ] is calculated. The volume VIVC of the combustion chamber 5 when the intake valve is closed is determined by the stroke position of the piston 6. The stroke position of the piston 6 is determined by the crank angle position of the engine.

  Referring to FIG. 6, consider a case where the rotation center 72 of the crankshaft 71 of the engine is offset from the center axis 73 of the cylinder. Assume that the connecting rod 74, the joint point 75 between the connecting rod 74 and the crankshaft 71, and the piston pin 76 that connects the connecting rod 74 and the piston are in the relationship shown in the figure. At this time, the volume VIVC of the combustion chamber 5 at the closing timing of the intake valve can be expressed by the following equations (2) to (6).

VIVC = f1 (θivc)
= Vc + (π / 4) D ² · Hivc (2)
Vc = (π / 4) D ² · Hx / (ε−1) (3)
^{Hivc = {(CND + ST 2} /2) - (CRoff-PISoff) 2} 1/2
− {(ST / 2) · cos (θivc + θoff)} + (CND ² −X ² ) ^1/2
(4)
X = (ST / 2) · sin (θivc + θoff) −CRoff + PISoff
... (5)
θoff = arcsin {(CRoff−PISoff) / (CND · (ST / 2))}
(6)
Where Vc: gap volume [m ³ ],
ε: compression ratio,
D: cylinder bore diameter [m],
ST: Full piston stroke [m],
Hivc: From the TDC of the piston pin 76 when the intake valve is closed
Distance [m],
Hx: Maximum value and minimum value of the distance from the TDC of the piston pin 76
Difference [m],
CND: length of connecting rod 74 [m],
CRoff: offset distance of the nodal point 75 from the cylinder center axis 73 [m],
PISoff: offset distance [m] of the crankshaft rotation center 72 from the cylinder center axis 73,
θivc: Intake valve closing timing crank angle [degATDC],
θoff: piston pin 76 and crankshaft rotation center 72
The angle [deg] between the line connecting the two and the vertical line in TDC,
X: horizontal distance [m] between the nodal point 75 and the piston pin 76,
As described above, the crank angle θivc at the time of closing the intake valve is known because it is determined by the command signal from the engine controller 31 to the intake VTC mechanism 27. By substituting the crank angle θivc (= IVC) at this time into the equations (2) to (6), the volume VIVC of the combustion chamber 5 at the closing timing of the intake valve can be calculated. Therefore, in practice, the volume VIVC of the combustion chamber 5 at the intake valve closing timing is set by a table using the intake valve closing timing IVC as a parameter. When the intake VTC mechanism 27 is not provided, a constant value can be given.

  In step 13, the temperature (that is, the compression start timing temperature) TINI [K] of the combustion chamber 5 at the intake valve closing timing IVC is calculated. The temperature of the gas flowing into the combustion chamber 5 is a temperature of a gas in which the fresh air flowing into the combustion chamber 5 and the inert gas remaining in the combustion chamber 5 are mixed. The temperature of the fresh air flowing into the combustion chamber 5 is Since the temperature of the inert gas equal to the fresh air temperature TCOL in the intake collector 2 and remaining in the combustion chamber 5 can be approximated by the exhaust temperature TEXH in the vicinity of the exhaust port, the temperature of the combustion chamber 5 at the intake valve closing timing IVC. TINI is the following from the fresh air temperature TCOL in the intake collector 2, the exhaust gas temperature TEXH, and the internal inert gas ratio MRESFR that is the ratio of the inert gas remaining in the combustion chamber 5 at the timing when the intake valve closing timing IVC is reached. It can be obtained by an expression.

TINI = TEXH × MRESFR + TCOL × (1−MRESFR)
... (7)
In step 14, the pressure (that is, compression start timing pressure) PINI [Pa] at the intake valve closing timing IVC of the combustion chamber 5 is calculated. That is, the collector internal pressure PCOL at the timing when the intake valve closing timing IVC is reached is taken in as the pressure PINI at the intake valve closing timing IVC.

  In step 15, a reaction probability RPROBA [%] representing the flammability of the air-fuel mixture in the combustion chamber 5 is calculated. The reaction probability RPROBA is a dimensionless value and depends on the three parameters of the residual inert gas ratio MRESFR, the cooling water temperature TWK [K], and the target equivalent ratio TFBYA, and can be expressed by the following equation.

RPROBA = f3 (MRESFR, TWK, TFBYA) (8)
More specifically, the maximum value of the reaction probability obtained by the combination of the three parameters MRESFR, TWK, and TFBYA is set to 100%, the relationship between these parameters and the reaction probability RPROBA is experimentally obtained, and the obtained reaction probability RPROBA is obtained. Are stored in advance in the memory of the engine controller 31 as a table corresponding to the parameters. In step 15, the reaction probability RPROBA is obtained by searching this table according to the parameters.

  Specifically, a table of water temperature correction coefficients having characteristics as shown in FIG. 7 according to the cooling water temperature TWK, a table of internal inert gas rate correction coefficients (not shown) set similarly, and a target equivalent ratio A table of equivalence ratio correction coefficients having characteristics as shown in FIG. 8 according to TFBYA is stored in the memory in advance. The maximum value of each correction coefficient is 1.0, and the reaction probability RPROBA is calculated by multiplying the product of the three types of correction coefficients by the maximum value of 100% of the reaction probability.

  Explaining each table, the water temperature correction coefficient shown in FIG. 7 becomes larger as the cooling water temperature TWK is higher, and becomes 1.0 when the cooling water temperature TWK is 80 ° C. or higher. The equivalence ratio correction coefficient shown in FIG. 8 is the maximum value of 1.0 when the target equivalence ratio TFBYA is 1.0, that is, the stoichiometric air-fuel ratio. The ratio correction factor decreases. Although the internal inert gas rate correction coefficient is not shown, it is 1.0 when the internal inert gas rate MRESFR is zero.

  In step 16, a reference crank angle θPMAX [degATDC] is calculated. As described above, the reference crank angle θPMAX does not fluctuate very much, but it still tends to advance as the engine speed NRPM increases. Therefore, the reference crank angle θPMAX can be expressed by the following equation as a function of the engine speed NRPM. it can.

θPMAX = f4 (NRPM) (9)
Specifically, the reference crank angle θPMAX is obtained by searching a table of characteristics shown in FIG. 9 stored in advance in the memory of the engine controller 31 from the engine speed NRPM. In order to facilitate calculation, the reference crank angle θPMAX can be regarded as constant.

  Finally, in step 17, the ignition dead time equivalent crank angle IGNDEAD [deg] is calculated. The ignition dead time equivalent crank angle IGNDEAD is a crank angle section from the timing at which the engine controller 31 outputs a signal for cutting off the primary current of the ignition coil 13 until the ignition plug 14 actually ignites, and can be expressed by the following equation. .

IGNDEAD = f5 (DEADTIME, NRPM) (10)
Here, the ignition dead time DEADTIME is set to 200 μsec. Equation (10) is for calculating the ignition dead time equivalent crank angle IGNDEAD that is the crank angle corresponding to the ignition dead time DEADTIME from the engine speed NRPM.

  FIG. 10 is for calculating the initial combustion period BURN1 [deg], and FIG. 12 is for calculating the main combustion period BURN2 [deg], which is executed at regular intervals (for example, every 10 msec). 10 and 12 are executed following FIG. Either of FIGS. 10 and 12 may be executed first.

First, referring to FIG. 10, in step 161, the previous combustion start timing MBTCYCL [degBTDC], the volume VIVC [m ³ ] at the intake valve closing timing of the combustion chamber 5 calculated in step 12 of FIG. 5, and the step of FIG. The temperature TINI [K] at the intake valve closing timing of the combustion chamber 5 calculated in 13, the pressure PINI [Pa] at the intake valve closing timing of the combustion chamber 5 calculated in Step 14 of FIG. 5, and the engine speed NRPM [Rpm], the reaction probability RPROBA [%] calculated in step 15 of FIG. 5 is read.

  Here, the previous combustion start timing MBTCYCL is a value one cycle before [degBTDC] of the basic ignition timing MBTCAL, and the calculation thereof will be described later with reference to FIG.

In step 162, the volume V0 [m ³ ] at the combustion start timing of the combustion chamber 5 is calculated. As described above, the ignition timing (combustion start timing) here is not the basic ignition timing MBTCAL calculated in the current cycle but a value one cycle before the basic ignition timing. That is, the volume V0 at the combustion start timing of the combustion chamber 5 is calculated from MBTCYCL, which is a value one cycle before the basic ignition timing, by the following equation.

V0 = f6 (MBTCYCL) (11)
Specifically, the volume V0 of MBTCYCL in the combustion chamber 5 is calculated from the stroke position of the piston 6 at the previous combustion start timing MBTCYCL and the bore diameter of the combustion chamber 5. In step 12 of FIG. 5, the volume VIVC of the combustion chamber 5 at the intake valve closing timing IVC is obtained by searching a table of intake valve closing timing volumes using the intake valve closing timing as a parameter. Here, MBTCYCL is set as a parameter. The volume V0 of the combustion chamber 5 at the previous combustion start time MBTCYCL may be obtained by searching the table of the previous combustion start time volume.

  In step 163, an effective compression ratio Ec at the combustion start timing is calculated. The effective compression ratio Ec is a dimensionless value, and is a value obtained by dividing the volume V0 of the combustion chamber 5 at the combustion start timing by the volume VIVC of the combustion chamber 5 at the intake valve closing timing, as shown in the following equation.

Ec = f7 (V0, VIVC) = V0 / VIVC (12)
In step 164, the temperature increase rate TCOMP in the combustion chamber 5 from the intake valve closing timing IVC to the combustion start timing is calculated based on the effective compression ratio Ec as shown in the following equation.

TCOMP = f8 (Ec) = Ec ^ (κ−1) (13)
Where κ: specific heat ratio,
Equation (13) is an equation for the rate of temperature rise of the adiabatic compressed gas. Note that “^” on the right side of the equation (13) represents power calculation. This symbol is also used in the formula described later.

  κ is a value obtained by dividing the constant pressure specific heat of the gas adiabatically compressed by the constant volume specific heat. If the gas adiabatically compressed is air, κ = 1.4, and this value may be used simply. However, the calculation accuracy can be further improved by experimentally determining the value of κ for the air-fuel mixture.

  FIG. 11 illustrates equation (13). Therefore, it is possible to obtain the temperature increase rate TCOMP by storing a table of such characteristics in advance in the memory of the engine controller 31 and searching the table based on the effective compression ratio Ec.

In step 165, the temperature T0 [K] at the combustion start timing of the combustion chamber 5 is multiplied by the temperature TINI at the intake valve closing timing of the combustion chamber 5 by the temperature increase rate TCOMP, that is, T0 = TINI × TCOMP (14)
It is calculated by the following formula.

  Steps 166 and 167 are the same as steps 164 and 165. That is, at step 166, the pressure increase rate PCOMP in the combustion chamber 5 from the intake valve closing timing IVC to the combustion start timing is calculated based on the effective compression ratio Ec as shown in the following equation.

PCOMP = f9 (Ec) = Ec ^ κ (41)
Where κ: specific heat ratio,
The equation (41) is also an equation for the rate of increase in pressure of gas that is adiabatically compressed in the same manner as the equation (13). “^” On the right side of the equation (41) represents power calculation as in the equation (13).

  κ is the same as the value used in the above equation (13). If the gas to be adiabatically compressed is air, κ = 1.4, and this value may be used simply. However, the calculation accuracy can be further improved by obtaining the value of κ from the composition and temperature of the air-fuel mixture.

  It is also possible to obtain a pressure increase rate PCOMP by storing a table having the same characteristics as in FIG. 11 in advance in the memory of the engine controller 31 and searching the table based on the effective compression ratio Ec.

In step 167, the pressure P0 [Pa] at the combustion start timing of the combustion chamber 5 is multiplied by the pressure PINI at the intake valve closing timing of the combustion chamber 5 by the pressure increase rate PCOMP, that is, P0 = PINI × PCOMP (42)
It is calculated by the following formula.

  In step 168, a laminar flame speed SL1 [m / sec] in the initial combustion period is calculated by the following equation (known).

SL1 = f10 (T0, P0)
= SLstd × {(T0 × Tstd) ^ 2.18}
× {(P0 / Pstd) ^ (− 0.16)} ... (15)
Where Tstd: reference temperature [K],
Pstd: Reference pressure [Pa]
SLstd: reference laminar flame velocity at reference temperature Tstd and reference pressure Pstd
[M / sec],
T0: temperature [K] at the combustion start timing of the combustion chamber 5;
P0: pressure [Pa] at the combustion start timing of the combustion chamber 5;
The laminar flame velocity is a flame propagation velocity in the absence of gas flow, and the temperature and pressure of the combustion chamber 5 are independent of the compression velocity in the combustion chamber 5 and the intake air flow velocity in the combustion chamber 5. Since it is known that the laminar flame speed in the initial combustion period is a function of the combustion start temperature T0 and the combustion start pressure P0, and the laminar flame speed in the main combustion period as described later, It is a function of the compression top dead center temperature TTDC and the compression top dead center pressure PTDC. This is because the laminar flame speed generally varies depending on the engine load, the inert gas ratio in the combustion chamber 5, the intake valve closing timing, the specific heat ratio, and the intake air temperature. Since it is a factor that affects T and pressure P, the laminar flame speed can be finally defined by the temperature T and pressure P in the combustion chamber 5.

  In the above equation (15), the reference temperature Tstd, the reference pressure Pstd, and the reference laminar flame speed SLstd are values determined in advance by experiments.

  Under a pressure of 2 bar or more, which is a normal pressure in the combustion chamber 5, the pressure term (P0 / Pstd) ^ (−0.16) in the equation (15) becomes a small value. Accordingly, it is possible to define the reference laminar flame speed SLstd only by the reference temperature Tstd with the pressure term (P0 / Pstd) ^ (− 0.16) as a constant value.

  Accordingly, when the reference temperature Tstd is 550 [K], the reference laminar flame speed SLstd is 1.0 [m / sec], and the pressure term is 0.7, the temperature T0 and the laminar flame speed at the combustion start timing. The relationship with SL1 can be approximately defined by the following equation.

SL1 = f11 (T0)
= 1.0 × 0.7 × (T0 / 550) ^ 2.18 ... (16)
In step 169, the average turbulence intensity ST1 of the gas flow in the initial combustion period is calculated. The average turbulence strength ST1 of the gas flow is a dimensionless value, which will be described later in detail.

  In step 170, the gas turbulent flame speed FLAME1 [m / sec] (combustion speed) in the initial combustion period is calculated from the laminar flame speed S1 in the initial combustion period and the average turbulence intensity ST1 of the gas flow in the initial combustion period. Calculate by the formula.

FLAME1 = SL1 × ST1 (18)
If there is gas turbulence in the combustion chamber 5, the flame speed of the gas changes. Equation (18) calculates the turbulent flame speed, which is the flame speed in a state where there is a gas flow, in consideration of the contribution (influence) to the flame speed associated with this gas turbulence.

  In step 171, the initial combustion period BURN1 [deg] is calculated by the following equation based on the turbulent flame speed FLAME1 thus calculated.

BURN1 = {(NRPM × 6) × (BR1 × V0)}
/ (RPROBA × AF1 × FLAME1) (19)
However, AF1: Reaction area (fixed value) of flame kernel [m ² ],
The equation (19) and the equation (22) described later are derived from the following basic equation that the combustion period is obtained by dividing the mass of the combustion gas by the turbulent flame velocity (combustion velocity). ), And the numerator and denominator on the right side of equation (22), which will be described later, do not immediately represent the mass of the combustion gas and the combustion rate.

Combustion period [sec] = total mass in cylinder [g]
/ (Unburned gas density [g / m ³ ]
× Flame surface area [m ² ] × Turbulent flame velocity [m / sec])
... (Supplement 1)
Since the unburned gas density of the right side denominator of (Supplement 1) is a value obtained by dividing the unburned gas mass [g] by the unburned gas volume [m ³ ], the conventional apparatus (Japanese Patent Laid-Open No. 10-30535) Thus, it can not be said that the unburned gas density can be accurately calculated by the function of only the charging efficiency ITAC corresponding to the mass. Therefore, experimental formulas shown in the above formula (19) and formula (22) to be described later are obtained for the first time by introducing a predetermined approximation to the formula (complement 1) while checking the experimental results.

  Here, BR1 on the right side of the equation (19) is a change amount of the combustion mass ratio from the combustion start timing to the end timing of the initial combustion period BURN1, and here BR1 = 2% is set. (NRPM × 6) on the right side of the equation (19) is a process for converting the unit from rpm to crank angle (deg). The reaction area AF1 of the flame kernel is set experimentally.

  Further, it can be assumed that the combustion chamber volume does not change during the initial combustion period. Therefore, when calculating the initial combustion period BURN1, the combustion chamber volume V0 at the start of combustion, which is the first combustion chamber volume, is employed.

Next, the flow of FIG. 12 is followed. In step 181, similarly to step 161 of FIG. 10, the volume VIVC [m ³ ] at the intake valve closing timing of the combustion chamber 5 calculated in step 12 of FIG. The temperature TINI [K] at the intake valve closing timing of the combustion chamber 5 calculated in step 13 of FIG. 5, the pressure PINI [Pa] at the intake valve closing timing of the combustion chamber 5 calculated in step 14 of FIG. The speed NRPM [rpm], the reaction probability RPROBA [%] calculated in step 15 of FIG. 5 is read, and the cylinder fresh air amount MACYL [g], the target equivalent ratio TFBYA, the internal inert gas amount MRES [g], The external inert gas amount MEGR [g] is read.

  Here, although an external EGR device is not shown in FIG. 1, as far as FIG. 12 is concerned, an explanation will be given on the premise of an engine equipped with an external EGR device. In this case, the external inert gas amount MEGR may be calculated using, for example, a known method (see Japanese Patent Laid-Open No. 10-141150). It should be noted that when an engine that does not include an external EGR device as in the present embodiment shown in FIG. 1 is used, it is sufficient to handle the external inert gas amount MEGR = 0. The calculation of the cylinder fresh air amount MACYL and the internal inert gas amount MRES will be described later with reference to FIG.

Steps 182 and 183 are the same as steps 163 and 164 in FIG. That is, in step 182, the effective compression ratio Ec at the compression top dead center time. 2 is calculated. Effective compression ratio Ec 2 is also a dimensionless value like the effective compression ratio Ec of the above equation (12), and the volume VTDC at the time of compression top dead center of the combustion chamber 5 is calculated at the closing timing of the intake valve of the combustion chamber 5 as shown in the following equation. The value divided by the volume VIVC.

Ec 2 = f13 (VTDC, VIVC) = VTDC / VIVC
... (43)
In equation (43), the volume VTDC at the time of compression top dead center of the combustion chamber 5 is constant regardless of the operating conditions, and may be stored in the memory of the engine controller 31 in advance.

In step 183, the temperature increase rate TCOMP due to adiabatic compression in the combustion chamber 5 during the period from the intake valve closing timing IVC to the compression top dead center. 2 is an effective compression ratio Ec as shown in the following equation: 2 is calculated.

TCOMP 2 = f14 (Ec 2)
= Ec 2 ^ (κ-1) (44)
Where κ: specific heat ratio,
A table having the same characteristics as in FIG. 11 is stored in advance in the memory of the engine controller 31 and the effective compression ratio Ec. By searching the table from 2, the temperature rise rate TCOMP 2 can also be obtained.

  In step 184, the total gas mass MGAS [g] in the combustion chamber 5 is calculated from the cylinder fresh air amount MACYL, the target equivalent ratio TFBYA, the internal inert gas amount MRES, and the external inert gas amount MEGR by the following equation.

MGAS = MACYL × (1 + TFBYA / 14.7) + MRES + MEGR
... (45)
“1” in parentheses on the right side of the equation (45) is a fresh air, and “TFBYA / 14.7” is a fuel.

  In step 185, the total gas mass MGAS of the combustion chamber 5, the cylinder fresh air amount MACYL, and the target equivalent ratio TFBYA are used to calculate the temperature increase amount (combustion increase temperature) TBURN [K] due to the combustion of the air-fuel mixture using the following equation. .

TBURN = {MACYL × TFBYA / 14.7 × BRk × Q}
/ (Cv × MGAS)
... (46)
Where Q is the constant calorific value of the fuel,
BRk: Combustion mass ratio of fuel in cylinder,
Cv: constant volume specific heat,
The numerator on the right side of the equation (46) means the total heat generated by the fuel in the cylinder [J], and the denominator means the temperature increase rate per unit generated heat [J / K]. That is, the equation (46) is an approximate equation applied to the thermodynamic formula.

  Here, the combustion mass ratio BRk of the in-cylinder fuel is adapted in advance through experiments or the like. For example, 60% / 2 = 30% is set. In this embodiment, since the combustion mass ratio BR reaches about 60% as the combustion period, the intermediate 30% is set as BRk.

  Since the constant calorific value Q of fuel varies depending on the type of fuel, it is obtained in advance by experiments or the like according to the type of fuel. The constant volume specific heat Cv is a value of 2 to 3, and the representative value is adapted beforehand by an experiment or the like. However, the calculation accuracy can be further improved by obtaining the value of the constant volume specific heat Cv from the composition and temperature of the air-fuel mixture.

In step 186, the temperature TTDC [K] at the compression top dead center of the combustion chamber 5 is changed from the temperature TINI at the intake valve closing timing of the combustion chamber 5 to the temperature increase rate TCOMP to the compression top dead center. Multiply by 2 and add the combustion rise temperature TBURN to the multiplication value, that is, the following equation is used.

TTDC = TINI × TCOMP 2 + TBURN
... (47)
In step 187, the temperature TTDC and the volume VTDC at the compression top dead center of the combustion chamber 5 and the pressure PINI, the volume VIVC and the temperature TINI at the closing timing of the intake valve of the combustion chamber 5 are The pressure PTDC [K] is calculated.

PTDC = PINI × VIVC × TTDC / (VTDC × TINI)
... (48)
Equation (48) is obtained using the equation of state. That is, the following equation of state is established using the pressure, volume and temperature (PINI, VIVC, TINI) at the intake valve closing timing.

PINI x VIVC = n · R · TINI (Supplement 2)
Where n is the number of moles
R: gas constant,
Since the volume is almost equal in the vicinity of the compression top dead center, the following equation of state is established using the pressure, volume and temperature (PTDC, VTDC, TTDC) at the compression top dead center.

PTDC × VTDC = n · R · TTDC (Supplement 3)
When n · R is eliminated from both (complement 3) and (complement 2) and PTPT is solved, the above equation (48) is obtained.

  In step 188, similarly to step 168 in FIG. 10, a laminar flame speed SL2 [m / sec] in the main combustion period is calculated by the following equation (known).

SL2 = f15 (TTDC, PTDC)
= SLstd × {(TTDC × Tstd) ^ 2.18}
× {(PTDC / Pstd) ^ (− 0.16)}
... (49)
Where Tstd: reference temperature [K],
Pstd: Reference pressure [Pa]
SLstd: reference laminar flame velocity at reference temperature Tstd and reference pressure Pstd
[M / sec],
TTDC: temperature [K] at the compression top dead center of the combustion chamber 5;
PTDC: pressure [Pa] at the compression top dead center of the combustion chamber 5
The explanation of the equation (49) is the same as the equation (16). That is, the reference temperature Tstd, the reference pressure Pstd, and the reference laminar flame speed SLstd in the equation (49) are values determined in advance by experiments. Under a pressure of 2 bar or more, which is a normal pressure in the combustion chamber 5, the pressure term (PTDC / Pstd) ^ (−0.16) in the equation (49) becomes a small value. Therefore, it is possible to define the reference laminar flame speed SLstd only by the reference temperature Tstd with the pressure term (PTDC / Pstd) ^ (− 0.16) as a constant value. Therefore, when the reference temperature Tstd is 550 [K], the reference laminar flame speed SLstd is 1.0 [m / sec], and the pressure term is 0.7, the temperature TTDC and the laminar flame at the compression top dead center The relationship with the speed SL2 can be approximately defined by the following equation.

SL2 = f16 (TTDC)
= 1.0 × 0.7 × (TTDC / 550) ^ 2.18
... (50)
In step 189, the average turbulence intensity ST2 of the gas flow during the main combustion period is calculated. The average turbulence intensity ST2 of the gas flow during the main combustion period will be described in detail later, as is the average turbulence intensity ST1 of the gas flow during the initial combustion period.

  In step 190, from the laminar flame speed SL2 [m / sec] in the main combustion period and the average turbulence intensity ST2 of the gas flow in the main combustion period, the turbulent flame speed FLAME2 [m / sec] (combustion in the main combustion period) Speed) is calculated by the following formula.

FLAME2 = SL2 × ST2 (21)
However, SL2: Laminar flame speed [m / sec],
In the same way as equation (18), equation (21) considers the contribution to the turbulent flame velocity associated with gas turbulence and calculates the turbulent flame velocity, which is the flame velocity in the presence of gas flow. Is.

  In step 191, based on the turbulent flame speed FLAME2 in the main combustion period calculated in this way, the main combustion period BURN2 [deg] is calculated by the following equation similar to the above equation (19).

BURN2 = {(NRPM × 6) × (BR2 × VTDC)}
/ (RPROBA × AF2 × FLAME2) (22)
However, AF2: Reaction area of flame kernel [m ² ]
Here, BR2 on the right side of the equation (22) is a change amount of the combustion mass ratio from the start timing to the end timing of the main combustion period. Since the combustion mass ratio BR becomes 2% at the end of the initial combustion period, and then the main combustion period starts and the combustion mass ratio BR reaches 60% and the main combustion period ends, BR2 = 60 % -2% = 58% is set. AF2 is an average reaction area in the growth process of the flame kernel, and is set to a fixed value experimentally determined in advance, like AF1 in the equation (19).

  During the main combustion period, the combustion chamber volume changes with the compression top dead center interposed therebetween. That is, it can be considered that the compression top dead center position exists at the approximate center between the start timing of the main combustion period and the end timing of the main combustion period. In addition, the combustion chamber volume does not change much in the vicinity of the compression top dead center even if the crank angle changes. Therefore, the combustion chamber volume in the main combustion period is represented by the combustion chamber volume VTDC at the compression top dead center.

  FIG. 13 is for calculating the basic ignition timing MBTCAL [degBTDC], and is executed at regular intervals (for example, every 10 msec). This is executed following the flow that is executed later in FIGS.

  In step 41, the initial combustion period BURN1 calculated in step 171 in FIG. 10, the main combustion period BURN2 calculated in step 191 in FIG. 12, and the ignition timing dead time equivalent crank calculated in step 17 in FIG. The angle IGNDEAD, the reference crank angle θPMAX calculated in step 16 of FIG. 5 is read.

  In step 42, the sum of the initial combustion period BURN1 and the main combustion period BURN2 is calculated as the combustion period BURN [deg].

  In step 43, the basic ignition timing MBTCAL [degBTDC] is calculated by the following equation.

MBTCAL = BURN−θPMAX + IGNDEAD (23)
In step 44, the value obtained by subtracting the ignition dead time equivalent crank angle IGNDEAD from the basic ignition timing MBTCAL is calculated as the previous combustion start timing MBTCYCL [degBTDC].

  The basic ignition timing MBTCAL calculated in this way is transferred to the ignition register as an ignition timing command value, and an ignition signal for cutting off the primary current from the engine controller 31 at a timing when the actual crank angle coincides with the ignition timing command value. It is output to the ignition coil 13.

  If the basic ignition timing MBTCAL calculated in step 43 is used as the ignition timing command value for the current cycle, the previous combustion start timing MBTCYCL calculated in step 44 is displayed until the ignition timing for the next cycle is reached. Ten steps 162 are used.

  Next, FIG. 14 is for calculating the internal inert gas ratio MRESFR in the combustion chamber 5 and is executed at regular intervals (for example, every 10 msec). This flow is executed prior to the flow of FIG.

  In step 51, the output of the air flow meter 32 and the target equivalent ratio TFBYA are read. In step 52, based on the output of the air flow meter 32, a new air amount (cylinder fresh air amount) MACYL flowing into the combustion chamber 5 is calculated. As a method for calculating the cylinder fresh air amount MACYL, a known method may be used (see JP 2001-50091 A).

  In step 53, an internal inert gas amount MRES in the combustion chamber 5 is calculated. The calculation of the internal inert gas amount MRES will be described with reference to the flow of FIG.

  In FIG. 15 (subroutine of step 53 in FIG. 14), in step 61, an inert gas amount MRESCYL at the exhaust valve closing timing EVC in the combustion chamber 5 is calculated. The calculation of the inert gas amount MRESCYL will be further described with reference to the flowchart of FIG.

  In FIG. 16 (subroutine of step 61 in FIG. 15), in step 71, the exhaust valve closing timing EVC [degBTDC], the exhaust temperature TEXH [K] detected by the temperature sensor 45, the exhaust pressure PEXH [kPa] detected by the pressure sensor 46 ].

  Here, just as the intake valve closing timing IVC is known from the command value given to the intake VTC mechanism 27, the exhaust valve closing timing EVC is also known from the command value given to the exhaust VTC mechanism 28.

  In step 72, the volume VEVC of the combustion chamber 5 at the exhaust valve closing timing EVC is calculated. This may be obtained by searching a table using the exhaust valve closing timing as a parameter, similarly to the volume VIVC at the intake valve closing timing IVC. That is, when the exhaust valve VTC mechanism 28 is provided, the volume VEVC at the exhaust valve closing timing EVC of the combustion chamber 5 may be obtained by searching the table shown in FIG. 23 from the exhaust valve closing timing EVC. When the exhaust VTC mechanism 28 is not provided, a constant value can be given.

  Although not shown, when a mechanism for changing the compression ratio is provided, the combustion chamber volume VEVC at the exhaust valve closing timing corresponding to the amount of change in the compression ratio is obtained from the table. When a mechanism for changing the compression ratio in addition to the exhaust VTC mechanism 28 is provided, the combustion chamber volume at the exhaust valve closing timing is obtained by searching a map corresponding to the exhaust valve closing timing and the compression ratio change amount.

  In step 73, the gas constant REX of the inert gas in the combustion chamber 5 is obtained by searching the table shown in FIG. 24 from the target equivalent ratio TFBYA. As shown in FIG. 24, the gas constant REX of the inert gas is the smallest when the target equivalent ratio TFBYA is 1.0, that is, the stoichiometric air-fuel ratio, and becomes larger whether it is larger or smaller.

  In step 74, the temperature TEVC of the combustion chamber 5 at the exhaust valve closing timing EVC is estimated based on the exhaust temperature TEXH. For simplicity, the exhaust temperature TEXH may be set as TEVC as it is. Note that the temperature TEVC of the combustion chamber 5 at the closing timing of the exhaust valve changes depending on the amount of heat corresponding to the fuel injection amount of the injector 21. Therefore, if such characteristics are taken into consideration, the calculation accuracy of TEVC is improved.

  In step 75, the pressure PEVC at the exhaust valve closing timing EVC of the combustion chamber 5 is calculated based on the exhaust pressure PEXH. Simply, the exhaust pressure PEXH may be set to PEVC.

  In step 76, from the volume VEVC at the exhaust valve closing timing EVC of the combustion chamber 5, the temperature TEVC at the exhaust valve closing timing EVC, the pressure PEVC at the exhaust valve closing timing EVC, and the inert gas gas constant REX, the exhaust valve of the combustion chamber 5 is obtained. The inert gas amount MRESCYL at the closing timing EVC is calculated by the following equation.

MRESCYL = (PEVC × VEVC) / (REX × TEVC) (24)
When the calculation of the inert gas amount MRESCYL at the exhaust valve closing timing EVC of the combustion chamber 5 is completed in this way, the processing returns to FIG. 15, and in step 62, the overlap of the intake and exhaust valves 15 and 16 (“O / L” in the figure). The amount of inactive gas MRESOL during the overlap is calculated, which is the amount of inert gas that is blown back from the exhaust side to the intake side.

  The calculation of the inert gas amount MRESOL will be described with reference to the flowchart of FIG.

  In FIG. 17 (subroutine of step 62 in FIG. 15), in step 81, the intake valve opening timing IVO [degBTDC], the exhaust valve closing timing EVC [degBTDC], and the exhaust valve of the combustion chamber 5 calculated in step 74 of FIG. The temperature TEVC at the closing timing EVC is read.

  Here, since the intake valve opening timing IVO is a timing earlier than the intake valve closing timing IVC by the opening angle of the intake valve 15, the opening angle of the intake valve 15 (which is known in advance) from the intake valve closing timing IVC. Can be sought.

  In step 82, the overlap amount VTCOL [deg] of the intake and exhaust valves is calculated from the following equation from the intake valve opening timing IVO and the exhaust valve closing timing EVC.

VTCOL = IVO + EVC (25)
For example, the intake valve opening timing IVO is at the intake top dead center position when the intake VTC mechanism 27 actuator is not energized, and the intake valve opening timing is advanced from the intake top dead center when the intake VTC mechanism 27 actuator is energized. The exhaust valve closing timing EVC is at the exhaust top dead center when the exhaust VTC mechanism 28 actuator is not energized, and the exhaust valve closing timing EVC is at the exhaust top dead center when the exhaust valve VTC mechanism 28 actuator is energized. In the case of more advanced characteristics, the sum of IVO and EVC becomes the overlap amount VTCOL of the intake and exhaust valves.

  In step 83, the accumulated effective area ASUMOL during the overlap is calculated by searching the table shown in FIG. 25 from the overlap amount VTCOL of the intake and exhaust valves. As shown in FIG. 25, the integrated effective area ASUMOL during the overlap is a value that increases as the overlap amount VTCOL of the intake and exhaust valves increases.

  Here, FIG. 26 is an explanatory diagram of the integrated effective area ASUMOL during the overlap of the intake and exhaust valves, where the horizontal axis indicates the crank angle, and the vertical axis indicates the respective opening areas of the intake valve 12 and the exhaust valve 15. Yes. The effective opening area at any time during the overlap is the smaller of the exhaust valve opening area and the intake valve opening area. The integrated effective area ASUMOL in the entire period during the overlap is an integral value (hatched portion in the figure) during the period in which the intake valve 15 and the exhaust valve 16 are open.

  By calculating the accumulated effective area ASUMOL during the overlap in this way, the overlap amount between the intake valve 15 and the exhaust valve 16 can be approximated as one orifice (outflow hole), and the state of the exhaust system and The gas flow rate passing through the virtual orifice can be simply calculated from the state of the intake system.

  In step 84, the specific heat ratio SHEATR of the inert gas remaining in the combustion chamber 5 is obtained by searching the map shown in FIG. 27 from the target equivalent ratio TFBYA and the temperature TEVC at the exhaust valve closing timing EVC of the combustion chamber 5. calculate. As shown in FIG. 27, the specific heat ratio SHEATR of the inert gas remaining in the combustion chamber is the smallest when the target equivalent ratio TFBYA is in the vicinity of 1.0, and it is larger or smaller than that. Further, under the condition where the target equivalent ratio TFBYA is constant, the temperature TEVC at the exhaust valve closing timing EVC of the combustion chamber 5 becomes smaller as the temperature becomes higher.

  In step 85, a supercharging determination flag TBCRG and a choke determination flag CHOKE are set. The setting of the supercharging determination flag TBCRG and the choke determination flag CHOKE will be described with reference to the flowchart of FIG.

  In FIG. 18 (subroutine of step 85 in FIG. 17), in step 101, the intake pressure PIN detected by the intake pressure sensor 44 and the pressure PEVC at the exhaust valve closing timing EVC of the combustion chamber 5 calculated in step 75 of FIG. Is read.

  In step 102, an intake exhaust pressure ratio PINBYEX is calculated from the intake pressure PIN and the pressure PEVC at the exhaust valve closing timing EVC of the combustion chamber 5 by the following equation.

PINBYEX = PIN / PEVC (26)
This intake / exhaust pressure ratio PINBYEX is an unknown number, and 1 is compared with this in step 103. When the intake / exhaust pressure ratio PINBYEX is 1 or less, it is determined that there is no supercharging, and the routine proceeds to step 104 where the supercharging determination flag TBCRG (initially set to zero) = 0.

  If the intake / exhaust pressure ratio PINBYEX is greater than 1, it is determined that there is supercharging, and the routine proceeds to step 105 where the supercharging determination flag TBCRG = 1.

  In Step 106, the specific heat ratio MIXAIRSHR of the mixture is obtained by searching the table shown in FIG. 28 from the target equivalent ratio TFBYA read in Step 51 of FIG. Replace with SHEATR. As shown in FIG. 28, the specific heat ratio MIXAIRSHR of the air-fuel mixture is a value that increases as the target equivalent ratio TFBYA decreases.

  In steps 106 and 107, the specific heat ratio SHEATR of the inert gas is replaced with the specific heat ratio MIXAIRSHR of the air-fuel mixture in consideration of supercharging such as turbocharging or inertial supercharging. That is, during supercharging, the gas flow during the overlap of the intake / exhaust valve is directed (blows through) from the intake system to the exhaust system. In this case, the specific heat ratio of the gas passing through the virtual orifice is the specific heat of the inert gas. By changing the ratio to the specific heat ratio of the air-fuel mixture, the amount of gas blown through is accurately estimated, and the amount of internal inert gas is accurately calculated.

  In Step 108, based on the specific heat ratio SHEATR of the inert gas calculated in Step 84 of FIG. 17 or Steps 106 and 107 of FIG. 18, the minimum and maximum choke determination thresholds SLCHOKE and SLCHOKEH are calculated by the following equations. To do.

SLCHOKER = {2 / (SHEATR + 1)}
^ {SHEATR / (SHEATR-1)}
... (27a)
SLCHOKEH = {− 2 / (SHEATR + 1)}
^ {-SHEATR / (SHEATR-1)}
... (27b)
These choke determination threshold values SLCHOKE and SLCHOKEH calculate the limit value for choking.

  If it is difficult to calculate each power of (27a) right side and (27b) right side in step 108, the calculation results of equations (27a) and (27b) are used as the minimum choke determination threshold value SLCHOKEL table and maximum choke determination. The threshold value SLCHOKEH table may be stored in advance in the memory of the engine controller 31 and obtained by searching the table from the specific heat ratio SHEATR of the inert gas.

  In step 109, it is determined whether or not the intake / exhaust pressure ratio PINBYEX is within the range of not less than the minimum choke determination threshold value SLCHOKEEL and not more than the maximum choke determination threshold value SLCHOKEH, that is, not in the choke state. If the intake / exhaust pressure ratio PINBYEX is within the range, it is determined that there is no choke, and the routine proceeds to step 110 where the choke determination flag CHOKE (initially set to zero) = 0.

  If the intake / exhaust pressure ratio P1NBYEX is not within the range, it is determined that choke is present, and the routine proceeds to step 111 where the choke determination flag CHOKE = 1.

  When the setting of the supercharging determination flag and the choke determination flag is thus completed, the process returns to FIG. 17 and the following four cases are performed in steps 86 to 88.

<1> When supercharging determination flag TBCRG = 0 and choke determination flag CHOKE = 0 <2> When supercharging determination flag TBCRG = 0 and choke determination flag CHOKE = 0 <3> Supercharging determination flag TBCRG = 0 and choke When determination flag CHOKE = 1 <4> When supercharging determination flag TBCRG = 1 and choke determination flag CHOKE = 0 When the above <1>, the routine proceeds to step 89 and overlaps when there is no supercharging and no choke If the average blown back inert gas flow rate MRESOLtmp1 is <2>, the process proceeds to step 90, and the blown back inert gas flow rate MRESOLtmp2 during the overlap when there is no supercharging and choke is present is step 91 when <3>. Go on to overlap when there is supercharging and no choke The average blowback inert gas flow rate MRESOLtmp3 in the case of <4> is advanced to step 92 to calculate the blowback inert gas flow rate MRESOLtmp4 with supercharging and with choke, respectively. Move to active gas flow rate MRESOLtmp.

Here, calculation of the inert gas flow rate MRESOLtmp1 during the overlap when there is no supercharging and no choke will be described with reference to the flow of FIG. 19. In step 121 of FIG. 19 (subroutine of step 89 in FIG. 17), step 73 in FIG. , 75, the gas constant REX of the inert gas, and the pressure PEVC when the exhaust valve of the combustion chamber 5 is closed are read.

  In step 122, based on the gas constant REX of the inert gas and the temperature TEVC at the exhaust valve closing timing of the combustion chamber 5 read in step 81 of FIG. 17, a density term MRSOLD used in a gas flow rate calculation formula described later. Is calculated by the following equation.

MRSOLD = SQRT {1 / (REX × TEVC)} (28)
Here, “SQRT” on the right side of equation (28) is a function that calculates the square root of the value in the parenthesis on the right.

  If the square root of the density term MRSOLD is difficult to calculate, the calculation result of equation (28) is stored in advance in the memory of the engine controller 31 as a map, and the gas constant REX and the temperature of the combustion chamber 5 at the exhaust valve closing timing are stored. You may obtain | require by searching the map from TEVC.

  In step 123, based on the specific heat ratio SHEATR of the inert gas calculated in step 84 of FIG. 17 and the intake / exhaust pressure ratio PINBYEX calculated in step 102 of FIG. The pressure difference term MRSOLP used is calculated by the following equation.

MRSOLP = SQRT [SHEATR / (SHEATR-1)
× {PTNBYEX ^ (2 / SHEATR)
-PTNBYEX ^ ((SHEATR + 1) / SHEATR)}] (29)
In step 124, from the density term MRSOLD, the pressure difference term MRSOLP, and the pressure PEVC at the exhaust valve closing timing of the combustion chamber 5, the blow-back inert gas flow rate MRESOLtmp1 during overlap without supercharging and without choke is expressed by the following equation. In step 125, the calculated value is transferred to the blown back inert gas flow rate MRESOLtmp.

MRESOLtmp1 = 1.4 × PEVC × MRSOLD × MRSOLP
... (30)
Next, the calculation of the blown back inert gas flow rate when there is no supercharging and when there is a choke will be described with reference to the flow of FIG. 20. In FIG. 20 (subroutine of step 90 of FIG. 17), in steps 131 and 132, steps 121 and 122 of FIG. In the same manner as described above, the gas constant REX of the inert gas and the pressure PEVC at the time of closing the exhaust valve of the combustion chamber 5 are read, and the density term MRSOLD is calculated from the above equation (28).

  In step 133, the choke pressure difference term MRSOLPC is calculated by the following equation based on the specific heat ratio SHEATR of the inert gas calculated in step 84 of FIG.

MRSOLPC = SQRT [SHEATR × {2 / (SHEATR + 1)} ^ {(SHEATR + 1) / [SHEATR-1)}]
... (31)
If it is difficult to calculate the power and square root of equation (31), the calculation result of equation (31) is stored in advance in the memory of the engine controller 31 as a table of choke pressure difference term MRSOLPC. Alternatively, it may be obtained by searching the table from the specific heat ratio SHEATR of the inert gas.

  In step 134, from the density term MRSOLD, the choke pressure difference term MRSOLPC, and the pressure PEVC when the combustion chamber 5 is closed, the inert gas flow rate MRESOLtmp2 blown back during overlap without supercharging and with choke is calculated. In step 135, the calculated value is transferred to the blown back inert gas flow rate MRESOLtmp.

MRESOLtmp2 = PEVC × MRSOLD × MRSOLPC
... (32)
Next, calculation of the blowback gas flow rate with supercharging and without choke will be described with reference to the flow of FIG. 21. In FIG. 21 (subroutine of step 91 in FIG. 17), in step 141, the intake pressure PIN detected by the intake pressure sensor 44 is explained. Is read.

  In step 142, the supercharging pressure difference term MRSOLPT is calculated from the specific heat ratio SHEATR of the inert gas calculated in steps 106 and 107 in FIG. 18 and the intake / exhaust pressure ratio PINBYEX calculated in step 102 in FIG. 18. Is calculated by the following equation.

MRSOLPT = SQRT [SHEATR / (SHEATR-1)
× {PINBYEX ^ (-2 / SHEATR)
-PINBYEX ^ (-(SHEATR + 1) / SHEATR)}] (33)
If the power calculation and the square root calculation of Expression (33) are difficult, the calculation result of Expression (33) is stored in advance in the memory of the engine controller 31 as a map of the supercharging pressure difference term MRSOLPT, You may obtain | require by searching the map from the specific heat ratio SHEATR of an inert gas, and the intake-exhaust pressure ratio PINBYEX.

  In step 143, based on the pressure difference term at the time of supercharging MRSOLPT and the intake pressure PIN, the inactive gas flow rate MRESOLtmp3 during the overlap with supercharging and without choke is calculated by the following equation, and the calculated value is calculated in step 143. At 144, the flow returns to the blown back inert gas flow rate MRESOLtmp.

MRESOLtmp3 = −0.152 × PIN × MRSOLPT (34)
Here, by setting the blown back inert gas flow rate MRESOLtmp3 in the equation (34) to a negative value, the gas flow rate of the air-fuel mixture blown from the intake system to the exhaust system during the overlap can be expressed.

Next, the calculation of the flow rate of the inert gas blown back during overlap with supercharging and with choke will be described with reference to the flow of FIG. 22. In FIG. 22 (subroutine of step 92 in FIG. 17), in steps 151 and 152, the step of FIG. Similarly to 141, the intake pressure PIN detected by the intake pressure sensor 44 is read, and the choke pressure difference term MRSOLPC is calculated by the aforementioned equation (31) as in step 132 of FIG.

  In step 153, based on the choke pressure difference term MRSOLPC and the intake pressure PIN, the overlap blow-back gas flow rate MRESOLtmp4 with supercharging and with choke is calculated by the following equation, and the calculated value is exceeded in step 154. Transfer to blown inert gas flow rate MRESOLtmp during lap.

MRESOLtmp4 = −0.108 × PIN × MRSOLPC (35)
Here, the blown back inert gas flow rate MRESOLtmp4 of the equation (35) can also represent a gas flow rate of the air-fuel mixture blown from the intake side to the exhaust side during the overlap, similarly to MRESOLtmp3.

  In this way, when the calculation of the blown back inert gas flow rate MRESOLtmp during the overlap divided according to the combination of the presence or absence of supercharging and the presence or absence of choke is completed, the process returns to FIG. From the inert gas flow rate MRESOLtmp and the integrated effective area ASUMOL during the overlap period, the blown back inert gas amount MRESOL during the overlap is calculated by the following equation.

MRESOL = (MRESOLtmP × ASUMOL × 60)
/ (NRPM × 360) (36)
When the calculation of the blown back inert gas amount MRESSOL is completed in this way, the process returns to FIG. 15, and in step 63, the inert gas amount MRESCYL at the exhaust valve closing timing EVC in the combustion chamber 5 and the blowback during overlap are returned. The internal inert gas amount MRES is calculated by adding the gas amount MRESOL, that is, the following equation.

MRES = MRESCYL + MRESSOL (37)
As described above, the flow rate of the inert gas blown back during overlap (MRESOLtmp3, MRESOLtmp4) becomes negative when there is supercharging, and therefore, the amount of blown inert gas MRESOL during overlap of the above equation (36) also becomes negative. According to the equation (37), the internal inert gas amount is reduced by the amount of the blown back inert gas amount MRESOL during the overlap.

  When the calculation of the internal inert gas amount MRES is completed in this way, the flow returns to FIG. 14. In step 54, using this internal inert gas amount MRES and the target equivalent ratio TFBYA, the internal inert gas rate MRESFR is calculated by the following equation. (Ratio of the amount of internal inert gas to the total amount of gas in the combustion chamber 5) is calculated.

MRESFR = MRES
/{MRES+MACYL×(1+TFBYA/14.7)}
... (38)
This completes the calculation of the internal inert gas ratio MRESFR.

  As described above, according to the present embodiment, the internal inert gas amount MRES is configured by the inert gas amount MRESCYL when the exhaust valve of the combustion chamber 5 is closed and the blowback gas amount MRESOL during the overlap of the intake and exhaust valves. In this case, the temperature TEV and pressure PEVC at the closing timing of the exhaust valve in the combustion chamber 5 are calculated (steps 74 and 75 in FIG. 16), and these temperature TEVC, pressure PEVC and inert gas are calculated. Since the inert gas amount MRESCYL at the exhaust valve closing timing of the combustion chamber 5 is calculated from the state constant (Equation (24) above) based on the gas constant REX (see step 76 of FIG. 16), Even during transient operation in which the amount of state (PEVC, VEVC, TEVC) inside the combustion chamber 5 changes every moment, the combustion chamber 5 can be accurately obtained regardless of the operating conditions. The inert gas amount MRESCYL in the exhaust valve closing timing can be calculated (estimated).

  In addition, based on the temperature TEVC and pressure PEVC of the combustion chamber 5 when the exhaust valve is closed, the gas constant REX and specific heat ratio SHEATR of the inert gas, and the intake pressure PIN, the blow-back inert gas flow rate (MRESOLtmp1, MRESOLtmp2) during the overlap is calculated. This was calculated (see FIGS. 19 and 20), and this gas flow rate was multiplied by the integrated effective area ASUMOL during the overlap to calculate the amount of blown back gas MRESOL during the overlap (see step 93 in FIG. 17). Therefore, the overlapped blow-back gas amount MRESOL can be calculated (estimated) with high accuracy.

  As described above, if both the inert gas amount MRESCYL and the overlapped blow-back gas amount MRESSOL in the combustion chamber 5 at the closing timing of the exhaust valve can be calculated (estimated) with high accuracy, the internal inert gas amount MRES, which is the sum of them, can also be accurately calculated. The internal inert gas ratio MRESFR calculated based on the internal inert gas amount MRES that can be estimated with high accuracy can be calculated (estimated). By making use of the temperature TINI at the intake valve closing timing IVC (see step 13 in FIG. 5), the temperature TINI at the intake valve closing timing IVC in the combustion chamber 5 can be accurately calculated. In addition, the engine can be appropriately controlled by making use of the internal inert gas amount MRES that can be accurately estimated for the fuel injection amount, valve opening / closing timing (overlap amount), and the like. .

  Further, since the gas constant REX of the inert gas and the specific heat ratio SHEATR of the inert gas are values corresponding to the target equivalent ratio TFBYA (see FIGS. 24 and 27), the operation is performed at an air-fuel ratio that deviates from the theoretical air-fuel ratio. (E.g. during lean operation where the air / fuel ratio is leaner than the stoichiometric air / fuel ratio, in order to stabilize the engine's originally unstable state, such as during cold start, the engine is on the rich side of the stoichiometric air / fuel ratio. Immediately after starting the engine operating at the air-fuel ratio, since the same large output is required, the exhaust valve closing timing of the combustion chamber 5 is also applied during full load operation where the air-fuel ratio is richer than the stoichiometric air-fuel ratio. Inert gas amount MRESCYL, overlapped blow-back gas amount MRESSOL, total internal inert gas amount MRES, and internal inert gas ratio MRESFR based on this are accurately calculated. It can be.

  Further, since the accumulated effective area ASUMOL during the overlap period is defined as the area of the virtual orifice, and it is assumed that the exhaust gas blows through the virtual orifice from the combustion chamber 5 to the intake system, the amount of the blown back inert gas amount MRESOL during the overlap is assumed. Calculation is simplified.

  Next, FIG. 29 is for calculating the average turbulence intensity ST1 in the initial combustion period.

  In FIG. 29 (subroutine of step 169 in FIG. 10), in step 201, the engine speed NRPM [rpm] is read.

  In steps 202 and 203, the intake valve passage average flow velocity Vo in the state of the intake valve opening timing IVO and the intake valve passage average flow velocity Vc in the state of the intake valve closing timing IVC are calculated. The calculation of the intake valve passage average flow velocity Vo in the state of the intake valve opening timing IVO is explained by the flow of FIG. 30, and the calculation of the intake valve passage average flow velocity Vc in the state of the intake valve closing timing IVC is explained by the flow of FIG. To do.

  First, from one intake valve passage average flow velocity Vo, in FIG. 30 (subroutine of step 202 in FIG. 29), in step 211, the intake port pressure Po [Pa] at the intake valve opening timing IVO, the state of the intake valve opening timing IVO The intake valve lift amount Lo [mm] and the engine rotational speed NRPM [rpm] in the state of the valve opening crank angle section θo [deg] at the intake valve opening timing IVO are read.

  Here, as the intake port pressure Po at the intake valve opening timing, the collector internal pressure PCOL detected by the pressure sensor 44 at the intake valve opening timing IVO may be sampled.

  In an engine provided with the VEL mechanism 51, the intake valve lift amount Lo and the valve opening crank angle section θo in the state of the intake valve opening timing IVO, and the intake valve lift amount Lc and the valve opening crank angle in the state of the intake valve closing timing IVC. Since the section θc may be different, these θo, Lo, θc, and Lc are calculated as follows. That is, FIG. 31 shows that the VEL mechanism 51 is operating at the intake valve opening timing IVO, and therefore the valve lift of the intake valve 15 is small (see the solid line), but the VEL mechanism 51 immediately after the intake valve opening timing. Is inactive, and the intake valve 15 is shifted to a state in which the valve lift of the intake valve 15 is large (see the chain line) before the intake valve closing timing IVC.

  Here, the intake valve lift amount is an average lift amount during the valve opening period, and the intake valve lift amount Lo is smaller in the state of the intake valve opening timing IVO (when the VEL mechanism 51 is operated). When the intake valve closing timing IVC is in the state (when the VEL mechanism 51 is not in operation), the intake valve lift amount Lc is larger. This is the average lift amount (that is, Lb in the figure) during the valve opening period of the characteristic.

  Further, in the state of the intake valve opening timing IVO (when the VEL mechanism 51 is operated), the valve opening crank angle section θo is a valve opening crank angle section (that is, θa in the drawing) having a small lift characteristic, and the intake valve closing timing. In the IVC state (when the VEL mechanism 51 is not in operation), the valve opening crank angle section θc is a valve opening crank angle section (that is, θb in the drawing) having a larger lift characteristic.

  Since the two intake valve lift characteristics shown in the figure are determined by the specifications of the VEL mechanism 51 and the intake valve cam, the values La, Lb, θa, and θb shown in the figure can be calculated in advance. Accordingly, the intake valve lift amount Lo in the state of the intake valve opening timing IVO, the intake valve lift amount Lc in the state of the intake valve closing timing IVC, the valve opening crank angle section θo in the state of the intake valve opening timing IVO, the intake valve As for the valve opening crank angle section θc in the state of the closing timing IVC, when the VEL mechanism 51 is activated and deactivated, La and θa are observed when the VEL mechanism 51 is activated, and Lb when the VEL mechanism 51 is not activated. And θb may be used (La, Lb, θa, and θb are all constant values).

  Or, by calculation, the intake valve lift amount in the intake valve open timing state, the intake valve lift amount in the intake valve close timing state, the valve opening crank angle section in the intake valve open timing state, the intake valve close timing state It is also possible to obtain the valve opening crank angle section at. For example, the instantaneous intake valve lift amount is integrated over the crank angle section in which the intake valve 15 is open, the integrated value is divided by the valve opening crank angle section to obtain an average value, and the average value is the state of the intake valve opening timing. And the intake valve lift amount Lc at the intake valve opening timing. In addition, the crank angle at the intake valve opening timing and the crank angle at the intake valve closing timing are detected, and the crank angle interval between them is the valve opening crank angle section θo in the intake valve opening timing state or the intake valve closing timing state. The valve opening crank angle section θc may be used.

  On the other hand, when the intake valve lift characteristic is one of the characteristics shown in FIG. 31 and the operating state of the VEL mechanism 51 does not change during the crank angle section from the intake valve opening timing IVO to the intake valve closing timing IVC, for example, If the VEL mechanism 51 is operating throughout the crank angle section from the valve opening timing IVO to the intake valve closing timing IVC, the intake valve lift amount Lc in the state of the intake valve closing timing IVC and the intake valve opening timing IVO In this state, the intake valve lift amount Lo is both equal to La shown in FIG. 31, and the valve opening crank angle section θc in the state of the intake valve closing timing IVC and the valve opening crank in the state of the intake valve opening timing IVO. Both of the angular sections θo are equal to θa shown in FIG. On the contrary, when the VEL mechanism 51 is inactive throughout the crank angle section from the intake valve opening timing IVO to the intake valve closing timing IVC, the intake valve lift amount Lc in the state of the intake valve closing timing IVC, The intake valve lift amount Lo in the state of the intake valve opening timing IVO is both equal to Lb shown in FIG. 31, and the valve opening crank angle interval θc in the state of the intake valve closing timing IVC and the state of the intake valve opening timing IVO The valve opening crank angle section θo at is equal to θb shown in FIG.

In step 212, when it is assumed that the intake port pressure during the intake stroke maintains the intake port pressure Po at the intake valve opening timing, the amount of intake air flowing into the combustion chamber 5 (the amount of intake air flowing into the combustion chamber 5 is the following) It is referred to as “cylinder intake air amount.”) Qo [mm ³ ] is calculated in proportion to the intake port pressure Po at the intake valve opening timing, that is, by the following equation.

Qo = Po × f21 (θo, NRPM) (51)
Here, the correction coefficient f21 (θo, NRPM) in the equation (51) represents a function of the intake valve lift amount θo and the rotational speed NRPM in the state of the intake valve opening timing. Even if the intake port pressure Po at the intake valve opening timing is the same, the correction coefficient f21 (θo, NRPM) causes the cylinder intake air amount Qo to vary depending on the valve opening crank angle section θo and the rotational speed NRPM at the intake valve opening timing. This is to be taken into account as it changes. The correction coefficient f21 (θo, NRPM) may be obtained not by a function but by a map search. FIG. 32 shows the map characteristics in this case. As shown in the figure, the correction coefficient f21 (θo, NRPM) increases as the valve opening crank angle section θo in the state of the intake valve opening timing increases, and the rotational speed NRPM increases. It is a value that becomes larger.

In step 213, the average intake valve opening area So [mm ² ] from the intake valve opening timing IVO to the intake valve closing timing IVC in the intake valve opening timing IVO is set to the intake valve lift amount Lo in the intake valve opening timing state. As a function, that is, calculated by the following formula.

So = f22 (Lo) (52)
Here, the average intake valve opening area So in the state of the intake valve opening timing may be obtained by a table search instead of being given by a function. FIG. 33 shows the table characteristics in this case, and the average intake valve opening area So in the state of the intake valve opening timing is a value that increases as the intake valve lift amount Lo in the state of the intake valve opening timing increases as shown in the figure. is there.

  In step 214, the intake valve passage average flow velocity Vo [mm / s] in the state of the intake valve opening timing IVO is calculated by the following equation.

Vo = Qo / (So · NRPM) (53)
That is, the intake valve passing average flow velocity Vo in the intake valve opening timing state increases in proportion to the cylinder intake air amount Qo in the intake valve opening timing state, and the average intake valve opening in the intake valve opening timing state It is a value that decreases in inverse proportion to the area So and the rotational speed NRPM.

  When the calculation of the intake valve passage average flow velocity Vo in the state of the intake valve opening timing is thus completed, the process returns to step 203 of FIG. 29 to calculate the intake valve passage average flow velocity Vc in the state of the intake valve closing timing IVC. The calculation of the intake valve passage average flow velocity Vc will be described with reference to the flow of FIG. The calculation method of the intake valve passage average flow velocity Vc in the state of the intake valve closing timing is the same as the calculation method of the intake valve passage average flow velocity Vo in the state of the intake valve opening timing described above with reference to FIG.

  34 (subroutine of step 203 in FIG. 29), in step 221, the intake port pressure Pc [Pa] at the intake valve closing timing IVC, the valve opening crank angle section θc [deg] in the state of the intake valve closing timing IVC, the intake valve The intake valve lift amount Lc [mm] and the engine speed NRPM [rpm] in the state of the closing timing IVC are read.

  Here, as the intake port pressure Pc at the intake valve closing timing, the collector internal pressure PCOL detected by the pressure sensor 44 at the intake valve closing timing IVC may be sampled.

  The intake valve lift amount Lc in the state of the intake valve closing timing IVC and the valve opening crank angle section θc in the state of the intake valve closing timing IVC are set to VEL before the intake valve closing timing IVC as described above with reference to FIG. If the mechanism 51 is deactivated and the intake valve closing timing IVC shifts to the side where the intake valve lift characteristic is large (see the alternate long and short dash line), Lb and θb shown in the figure are used. Further, when the VEL mechanism 51 is operating throughout the crank angle section from the intake valve opening timing IVO to the intake valve closing timing IVC, the intake valve lift amount Lc in the state of the intake valve closing timing IVC is the La shown in the figure. Further, the valve opening crank angle section θc in the state of the intake valve closing timing IVC becomes θa shown in the figure. On the contrary, when the VEL mechanism 51 is inactive throughout the crank angle section from the intake valve opening timing IVO to the intake valve closing timing IVC, the intake valve lift amount Lc in the state of the intake valve closing timing IVc is shown in the figure. The valve opening crank angle section θc in the state of the intake valve closing timing IVC becomes θb shown in FIG.

  In step 222, the cylinder intake air amount Qc when the intake port pressure during the intake stroke is maintained at the intake port pressure Pc at the intake valve closing timing is proportional to the intake port pressure Pc at the intake valve closing timing. Calculated by

Qc = Pc × f21 (θc, NRPM) (54)
Here, the correction coefficient f21 (θc, NRPM) in the equation (54) is the same function as the equation (51). The correction coefficient f21 (θc, NRPM) may be obtained by map search instead of being given by a function. FIG. 32 shows the characteristics of the map in this case. As shown in the figure, the correction coefficient f21 (θc, NRPM) increases as the valve opening crank angle section θc in the intake valve closing timing increases and the rotational speed NRPM increases. It is a value that becomes larger.

In step 223, the average intake valve opening area Sc [mm ² ] from the intake valve opening timing to the intake valve closing timing in the intake valve closing timing IVC state as a function of the intake valve lift amount Lc in the intake valve closing timing state. That is, it is calculated by the following equation.

Sc = f22 (Lc) (55)
Here, the average intake valve opening area Sc in the state of the intake valve closing timing may be obtained by a table search instead of being given by a function. FIG. 33 shows the table characteristics in this case, and is a value that increases as the intake valve lift amount Lc in the state of the intake valve closing timing increases as shown.

  In step 224, the intake valve passage average flow velocity Vc [mm / s] in the state of the intake valve closing timing IVC is calculated by the following equation.

Vc = Qc / (Sc · NRPM) (56)
That is, the intake valve passage flow velocity Vc in the state of the intake valve closing timing increases in proportion to the cylinder intake air amount Qc in the state of the intake valve close timing, and the average intake valve opening area in the state of the intake valve close timing It is a value that decreases in inverse proportion to Sc and the rotational speed NRPM.

  When the calculation of the intake valve passage average flow velocity Vc in the state of the intake valve closing timing is thus completed, returning to FIG. 29, in steps 204 and 205, the average turbulence intensity Uo1 in the initial combustion period in the state of the intake valve opening timing IVO is returned. Then, the average turbulence intensity Uc1 in the initial combustion period in the state of the intake valve closing timing IVC is calculated by the following equations, respectively.

Uo1 = f23 (Vo, NRPM) × Kc × Kt (57)
Uc1 = f23 (Vc, NRPM) × Kc × Kt (58)
Where Kc: swirl control valve correction coefficient,
Kt: Tumble control valve correction coefficient,
The basic values f23 (Vo, NRPM) of average turbulence intensity during the initial combustion period of the equations (57) and (58) are functions of the intake valve passage average flow velocity (Vo, Vc) and the rotational speed NRPM. It represents something. The average turbulence intensity basic values f23 (Vo, NRPM) and f23 (Vc, NRPM) in the initial combustion period change in the average turbulence intensity in the initial combustion period depending on the difference in the intake valve passage average flow velocities Vo and Vc and the rotational speed NRPM. Therefore, this is taken into consideration. The average turbulence intensity basic values f23 (Vo, NRPM) and f23 (Vc, NRPM) in the initial combustion period may be obtained by map search instead of being given by a function. FIG. 35 shows the map characteristics in this case. As shown in the figure, the average turbulence intensity basic values f23 (Vo, NRPM) and f23 (Vc, NRPM) in the initial combustion period are large in the intake valve passage average flow velocities Vo, Vc. The value increases as the rotational speed NRPM increases.

  Here, although not shown in FIG. 1, in order to improve the combustion state in the low load state, the gas flow (swirl) is made in the circumferential direction (direction along the cylinder wall) around the cylinder axis of the combustion chamber 5. There is a case where a swirl control valve for giving a gas or a tumble control valve for giving a gas flow (tumble) swiveling about an axis orthogonal to the cylinder axis is provided in the intake port 4, and the above (57), (58) The swirl control valve correction coefficient Kc and the tumble control valve correction coefficient Kt in the equation consider these additional gas flows. That is, when the swirl control valve and the tumble control valve are operated in the low load region to enhance the gas flow in the combustion chamber 5, the average turbulence intensity during the initial combustion period increases, so that the correction coefficients Kc and Kt are 1. A value exceeding 0 is given. Actually, the correction coefficients Kc and Kt are appropriate values.

  For control purposes, both swirl control valves and tumble control valves are used to provide versatility, but an actual engine has only one of these swirl control valves and tumble control valves. Which one is provided is known from the engine specifications, so the unnecessary correction coefficient is 1.0.

  In step 206, a weighted average coefficient k1 is calculated. The calculation of the weighted average coefficient k1 will be described with reference to the flowchart of FIG.

  36 (subroutine of step 206 in FIG. 29), in step 231, the intake port pressure Po [Pa] at the intake valve opening timing IVO, the valve opening crank angle section θo [deg] in the state of the intake valve opening timing IVO, the intake valve The intake valve lift amount Lo [mm] in the state of the open timing IVO, the intake port pressure Pc [Pa] at the intake valve close timing IVC, the valve opening crank angle section θc [deg] in the state of the intake valve close timing IVC, the intake air The intake valve lift amount Lc [mm], the average intake port pressure Pave [Pa], the average valve opening crank angle interval θave [degATDC], and the average intake valve lift amount Lave [mm] in the state of the valve closing timing IVC are read.

  Here, the average intake port pressure ave, the average intake valve lift amount Lave, and the average valve opening crank angle section θave are calculated by the following equations.

Pave = ΣP (t) / (tc-to) (59)
Lave = ΣL (t) / (tc-to) (60)
θave = Σθ (t) / (tc-to) (61)
Where P (t): instantaneous intake port pressure,
L (t): instantaneous intake valve lift amount,
θ (t): instantaneous valve opening crank angle section,
tc: time of intake valve closing timing,
to: time of intake valve opening timing,
As the instantaneous intake port pressure P (t) in the equation (59), the collector internal pressure PCOL detected by the pressure sensor 44 is used.

  In the VEL mechanism 51, the intake valve lift amount and the valve opening crank angle section are uniquely determined with respect to the control amount given to the VEL mechanism actuator, and the control amount given to the VEL mechanism actuator changes with time. Each table of the intake valve lift amount and the valve opening crank angle section with respect to the control amount is prepared in advance, and the values obtained by searching each table from the control amount given to the VEL mechanism actuator at that time are expressed by the equations (60), respectively. The instantaneous intake valve lift amount L (t) is used as the instantaneous valve opening crank angle interval θ (t) in the equation (61). As the instantaneous intake valve lift amount L (t) in the equation (60), a lift sensor for detecting the instantaneous intake valve lift amount L (t) may be provided, and a signal from the lift sensor may be used.

  In step 232, the intake port pressure Pc at the intake valve closing timing is compared with the intake port pressure Po at the intake valve open timing. When the intake port pressure Pc at the intake valve closing timing is not equal to the intake port pressure Po at the intake valve open timing, the routine proceeds to step 233, and a weighted average coefficient (first coefficient) Kp related to the intake port pressure is calculated by the following equation.

Kp = (Pave−Po) / (Pc−Po) (62)
On the other hand, when the intake port pressure Pc at the intake valve closing timing is equal to the intake port pressure Po at the intake valve open timing, the routine proceeds from step 232 to step 234, where Kp = 1/2.

  In step 235, the valve opening crank angle section θc in the intake valve closing timing state is compared with the valve opening crank angle section θo in the intake valve opening timing state. When the valve opening crank angle section θc in the state of the intake valve closing timing and the valve opening crank angle section θo in the state of the intake valve opening timing are not equal, the routine proceeds to step 236 and the weighted average for the valve opening crank angle section is calculated by the following equation. A coefficient (second coefficient) Kθ is calculated by the following equation.

Kθ = (θave−θo) / (θc−θo) (63)
On the other hand, when the valve opening crank angle section θc in the intake valve closing timing state is equal to the valve opening crank angle section θo in the intake valve opening timing state, the process proceeds from step 235 to step 237 to set Kθ = 1/2. .

  In step 238, the intake valve lift amount Lc in the intake valve closing timing state is compared with the intake valve lift amount Lo in the intake valve open timing state. When the intake valve lift amount Lc in the intake valve closing timing state and the intake valve lift amount Lo in the intake valve open timing state are not equal, the routine proceeds to step 239, where the weighted average coefficient (the third The coefficient KL is calculated by the following equation.

KL = (Lave-Lo) / (Lc-Lo) (64)
On the other hand, when the intake valve lift amount Lc at the intake valve closing timing is equal to the intake valve lift amount Lo at the intake valve opening timing, the routine proceeds from step 238 to step 240, where KL = 1/2.

  In step 241, the three coefficients Kp, Kθ, and KL calculated in this way are multiplied, that is, the weighted average coefficient k1 is calculated by the following equation.

k1 = 4 × Kp × Kθ × KL (65)
When the calculation of the weighted average coefficient k1 is completed in this manner, the process returns to step 207 in FIG. 29, and the average turbulence intensity Uo1 in the initial combustion period in the state of the intake valve opening timing and the intake valve closing are returned using the weighted average coefficient k1. A value obtained by multiplying the average turbulence intensity Uc1 in the initial combustion period in the state of time by the weighted average and the rotational speed NRPM is defined as an average turbulence intensity ST1 in the initial combustion period. That is, the average turbulence intensity ST1 in the initial combustion period is calculated by the following equation.

ST1 = {k1 × Uo1 + (1−k1) × Uc1} × NRPM (66)
It is based on experimental results that the rotational speed NRPM is exposed in the equation (66), and the average turbulence intensity ST1 in the initial combustion period is obtained in proportion to the exposed rotational speed NRPM. Of course, the following equation may be used instead of the equation (66).

ST1 = {k1 × Uo1 + (1−k1) × Uc1} (67)
Next, FIG. 37 is for calculating the average turbulence intensity ST2 in the main combustion period. The calculation method itself of the average turbulence intensity ST2 in the main combustion period is the same as the calculation method of the average turbulence intensity ST2 in the initial combustion period.

  In FIG. 37 (subroutine of step 189 in FIG. 12), in step 251, the engine speed NRPM [rpm] is read.

  In steps 252 and 253, the intake valve passing average flow velocity Vo in the state of the intake valve opening timing IVO and the intake valve passing average flow velocity Vc in the state of the intake valve closing timing IVC are calculated, respectively, as in steps 202 and 203 of FIG. . The calculation of the intake valve passing average flow velocity Vo in the state of the intake valve opening timing IVO is described with reference to the flow of FIG. 30, and the calculation of the intake valve passing average flow velocity Vc in the state of the intake valve closing timing IVC is described with reference to the flow of FIG. did.

  In step 254, a squish correction coefficient Kq (flame speed correction amount by squish) that is not calculated in FIG. 29 is calculated.

  Here, the squish is a cylinder radial direction generated by the air-fuel mixture existing in a narrow gap between the piston 6 head and the cylinder head lower surface being pushed into the main space of the combustion chamber 5 near the end of the compression stroke. Due to the gas flow, the size and position of the combustion chamber 5 are determined by the shape of the combustion chamber 5. This affects the turbulence intensity in the combustion chamber 5.

  Therefore, when the present inventor conducted various experiments and found that swirl, tumble, and squish have different effects on the turbulence intensity in the combustion chamber 5, the present invention is the result of this experiment. Are reflected in the calculation of the average turbulence intensity ST1 during the initial combustion period and the average turbulence intensity ST2 during the main combustion period.

  In other words, it has been found that swirl and tumble affect the average turbulence intensity ST1 in the initial combustion period, but hardly affect the average turbulence intensity ST2 in the main combustion period. Therefore, the swirl control valve correction coefficient Kc and the tumble control are determined. The valve correction coefficient Kt is introduced only in the calculation of the average turbulence intensity ST1 in the initial combustion period as described above in the equations (57) and (58), and the average turbulence intensity ST2 in the main combustion period is calculated as described later. Not introduced.

  On the other hand, squish differs from swirl and tumble in that it does not affect the average turbulence intensity ST1 in the initial combustion period, but affects the average turbulence intensity ST2 in the main combustion period. This is introduced only in the calculation of the average turbulence intensity ST2 in the period, and is not introduced in the calculation of the average turbulence intensity ST1 in the initial combustion period as described above in the equations (57) and (58). This is because squish occurs in the latter half of the compression stroke, that is, the main combustion period, and does not occur in the initial combustion period.

  Further, in the present invention, the squish correction coefficient Kq is calculated based on at least engine specifications such as the squish area volume and the gas blowing direction from the squish area.

  On the other hand, although the squish correction coefficient Kq itself has already been disclosed in the prior application device (see Japanese Patent Application No. 2003-95751), a detailed experiment has not yet been performed at the time of filing of the prior application device. The coefficient Kq is treated in the same row as the swirl control valve correction coefficient Kc and the tumble control valve correction coefficient Kt, and together with Kc and Kt, this squish coefficient Kq is the average turbulence intensity ST1 in the initial combustion period, and the average turbulence intensity in the main combustion period. Introduced in any of ST2 calculations. Further, the squish correction coefficient Kq at this time is given as a conforming value (a constant value).

  Here, a specific example of engine specifications will be described. FIG. 39 is a schematic plan view in which only one cylinder is enlarged for a four-valve engine having two intake valves 15 and two exhaust valves 16 per cylinder. With respect to the combustion chamber shape having such a configuration, in this embodiment, the squish area is divided into four parts except for the part that interferes with the intake valve 15 and the exhaust valve 16. In order to distinguish these four squish areas, numbers are assigned in the clockwise direction from the top in the figure, and they are designated as a first squish area A1, a second squish area A2, a third squish area A3, and a fourth squish area A4. In this case, since the volumes of the squish areas A1 to A4 are determined by the drawings, the first squish area volume, the second squish area volume, the third squish area volume, and the fourth squish area volume determined by the drawings are respectively sqv1 and sqv2. , Sqv3, sqv4 (both are constant values).

  FIG. 38 is a flowchart illustrating an engine having a combustion chamber shape having a plurality of squish areas A1 to A4.

  In FIG. 38 (subroutine of step 254 in FIG. 37), in step 261, the engine speed NRPM [rpm] is read.

  In step 262, 1 is set in the variable i. This variable i indicates the skick area number. That is, the first squish area A1 when i = 1, the second squish area A2 when i = 2, the third squish area A3 when i = 3, and the fourth squish area A4 when i = 4. Instruct.

  In step 263, the variable i is compared with 4, which is the total number of squish areas. At this time, since 1 is included in the variable i, the process proceeds to step 264 to calculate a site-specific squish area volume sqv (i) when i = 1. For example, as shown in FIG. 40, a table that can contrast the variable i with the first to fourth site-specific squish area volumes aqv1 to sqv4 is created in advance, and the squish area volume indicated by the variable i is designated from this table. Just pick up. Since i = 1 at this time, the region-specific squish area volume sqv (1) = first squish area volume sqv1.

  In step 265, a region-specific gas blowing direction correction coefficient sqt (i) is calculated by searching a map having the contents shown in FIG. 41 from the gas blowing direction angle from the squish area and the engine speed NRPM.

  Here, it is known that there are two types of gas blowing directions from the squish area as shown in FIG. 42 (the upper (a) shows that the gas blowing direction is not horizontal, and the lower (b) shows the gas blowing direction. The blowing direction is the horizontal direction), which is determined in advance depending on the engine specifications. In the case of (a) in the upper part of FIG. 42, the gas blowing direction changes when the engine rotational speed NRPM deviates from the reference rotational speed. Therefore, if the characteristics as shown in FIG. 41 are obtained in advance by experiments or the like using the gas blowing direction angle determined in advance by the engine specifications (including the reference rotation speed) and the engine rotation speed NRPM as parameters, The region-specific gas blowing direction correction coefficient sqt (i) can be obtained by searching a map having the contents shown in FIG. 41 from the determined gas blowing direction angle and the engine rotational speed NRPM at that time.

  As shown in FIG. 41, the region-specific gas blowing direction correction coefficient sqt (i) is a characteristic that increases as the gas blowing direction angle increases under a constant rotational speed, and takes a peak and then decreases. This is because the gas is blown out from the squish area toward the spark plug 6 and the squish is maximized at the blowing direction angle at the peak position.

  FIG. 41 shows three typical characteristics in the case of the first rotation speed (for example, low rotation speed), the second rotation speed (for example, medium rotation speed), and the third rotation speed (for example, high rotation speed). Although shown, the number of characteristics may be further increased. In FIG. 41, when the actual rotational speed is between three representative rotational speeds (first, second, and third rotational speeds), the region-specific gas blowing direction correction coefficient sqt (i) is calculated by interpolation calculation. Find it.

  In step 266, the squish correction coefficient for each part ksq (i), here, the squish correction coefficient for the first squish area A1 is calculated by the following equation.

ksq (i) = sqv (i) × sqt (i) × k2 (68)
Where k2: coefficient,
Equation (68) is used to calculate the squish correction coefficient ksq in proportion to the squish area volume and also taking into account the blowing direction angle and the rotational speed NRPM. The coefficient k2 in the equation (68) is a coefficient for converting the squish area volume into the turbulence intensity, and is adapted beforehand in the application target engine and stored.

  In step 267, the part-specific squish correction coefficient ksq (1) for the first squish area A1 is stored in the memory, and then the variable i is incremented by 1 in step 268, and the process returns to step 263. At this time, since the operation of steps 264 to 266 is executed because i = 2, the part-specific squish correction coefficient ksq (2) for the second squish area A2 is obtained. At 268, the variable i is incremented by 1, and the process returns to step 263. At this time, since the operation of steps 264 to 266 is executed because i = 3, the site-specific squish correction coefficient ksq (3) for the third squish area A3 is obtained. After this is stored in the memory in step 267, In step 268, the variable i is incremented by 1, and the flow returns to step 263. At this time, since the operation of steps 264 to 266 is executed because i = 4, the region-specific squish correction coefficient ksq (4) for the fourth squish area A4 is obtained. At 268, the variable i is incremented by 1, and the process returns to step 263. At this time, since i = 5, that is, calculation of the squish correction coefficients for all squish areas has been completed, the process proceeds from step 263 to step 269, and the sum of the correction coefficients for all squish areas is used as the squish correction coefficient Ksq. That is, the squish correction coefficient Ksq for the entire cylinder is calculated by the following equation.

Ksq = ksq (1) + ksq (2) + ksq (3) + ksq (4) (69)
Note that the loop operation in steps 263 to 268 ends in an instant.

  When the calculation of the squish correction coefficient Ksq is completed in this manner, the process returns to FIG. 37, and in steps 255 and 256, the average turbulence intensity Uo2 during the main combustion period in the state of the intake valve opening timing IVO and the state of the intake valve closing timing IVC. The average turbulence intensity Uc2 during the main combustion period is calculated according to the following equations.

Uo2 = f24 (Vo, NRPM) × Kq (70)
Uc2 = f24 (Vc, NRPM) × Kq (71)
The average turbulence intensity basic values f24 (Vo, NRPM) and f24 (Vc, NRPM) in the main combustion period of the equations (70) and (71) are also functions of the intake valve passage average flow velocity (Vo or Vc) and the rotational speed NRPM. It represents something. The average turbulence intensity basic values f24 (Vo, NRPM) and f24 (Vc, NRPM) change the average turbulence intensity during the main combustion period depending on the difference in the intake valve passage average flow velocity Vo, Vc and the rotational speed NRPM. Is to be considered. The average turbulence intensity basic values f24 (Vo, NRPM) and f24 (Vc, NRPM) in the main combustion period may be obtained by map search instead of being given by a function. FIG. 43 shows the map characteristics in this case. As shown in the figure, the average turbulence intensity basic values f24 (Vo, NRPM) and f24 (Vc, NRPM) in the main combustion period are large in the intake valve passage average flow velocities Vo, Vc. The value increases as the rotational speed NRPM increases.

  In step 257, the weighted average coefficient k1 is calculated in the same manner as in step 206 of FIG. 29, and in step 258, the average turbulence intensity Uo2 in the main combustion period in the state of the intake valve opening timing is calculated using this weighted average coefficient k1. A value obtained by multiplying the average turbulence intensity Uc2 during the main combustion period in the state of the intake valve closing time by the weighted average and the rotational speed NRPM is defined as an average turbulence intensity ST2 during the main combustion period. That is, the average turbulence intensity ST2 in the main combustion period is calculated by the following equation.

ST2 = {k1 × Uo2 + (1-k1) × Uc2} × NRPM (72)
It is based on the experimental results that the rotational speed NRPM is exposed in the equation (72), and the average turbulence intensity ST2 in the main combustion period is obtained in proportion to the exposed rotational speed NRPM. Of course, the following equation may be used instead of the equation (72).

ST2 = {k1 × Uo2 + (1-k1) × Uc2} (73)
Next, FIG. 44 shows how the intake port pressure, the cylinder intake air amount, the intake valve passage average flow velocity, the average turbulence intensity during the combustion period, and the turbulent flame speed during the combustion period are accelerated from the low load state. It is a wave form diagram which shows whether it changes. However, for the sake of simplicity, it is assumed that the VEL mechanism 51 is in an operating state or a non-operating state throughout acceleration and does not change from an operating state to a non-operating state or vice versa during acceleration.

  Due to the sudden acceleration when the accelerator pedal is depressed, In the two-cylinder intake stroke, the intake port pressure rapidly changes from a small value to a large value. For this reason, no. In the two cylinders, the intake port pressure greatly differs between the intake valve opening timing IVO state and the intake valve closing timing IVC state. Therefore, the intake air amount Qo and the intake valve closing timing IVC state at the intake valve opening timing IVO state. The intake air amount at the intake valve closing timing IVC is different from the intake air amount Qc at the intake valve opening timing IVO as shown in the drawing (see the third stage one-dot chain line). The quantity Qc (see the solid line in the third stage) is larger.

  As for the intake valve passage average flow velocity and average turbulence intensity, the intake valve passage average flow velocity Vo in the state of the intake valve opening timing IVO (see the dashed line in the fourth stage) and average turbulence strength Uo1 and Uo2 as shown in the figure. Intake valve passing average flow velocity Vc (refer to the solid line in the fourth stage), average turbulence strengths Uc1, Uc2 (in the fifth stage) (See the solid line) is larger.

  Therefore, the turbulent flame speed FLAME1o (= Uo1 × in the initial combustion period obtained by multiplying the average turbulence intensity Uo1 in the initial combustion period in the state of the intake valve opening timing IVO by the laminar flame speed SL1 in the initial combustion period. ST1) Initial value obtained by multiplying the average turbulence intensity Uc1 in the initial combustion period in the state of the intake valve closing timing IVC by the laminar flame speed SL1 in the initial combustion period rather than (see the sixth stage two-dot chain line) The turbulent flame velocity FLAME1c (= Uc1 × ST1) in the combustion period (see the solid line in the sixth stage) is larger, and the average turbulence intensity Uo2 in the main combustion period in the state of the intake valve opening timing IVO is the main combustion. Turbulent flame velocity FLAME2o (= Uo2 × ST2) in the main combustion period obtained by multiplying the laminar flame velocity SL2 in the period (the two-dot chain line at the bottom) Turbulent flame speed FLAME2c (= Uc2) in the main combustion period obtained by multiplying the laminar flame speed SL2 in the main combustion period by the average turbulence intensity Uc2 in the initial combustion period in the state of the intake valve closing timing IVC. × ST2) (see bottom solid line) is larger.

  In this case, the state at the intake valve opening timing IVO and the state at the intake valve closing timing IVC are different. When the turbulent flame speeds FLAME1o and FLAME2o in the state of the intake valve opening timing IVO are used as the turbulent flame speed for calculating the initial combustion period and the main combustion period for the two cylinders, the actual initial combustion period, actual In contrast, when the turbulent flame speeds FLAME1c and FLAME2c in the state of the intake valve closing timing IVC are used, the actual initial combustion period and the actual main combustion period are longer. Will also be long.

  In contrast, according to the present embodiment, the average turbulence intensity Uo1 in the initial combustion period in the state of the intake valve opening timing IVO, and the average turbulence intensity Uc1 in the initial combustion period in the state of the intake valve close timing IVC Is calculated as the average turbulence intensity ST1 in the initial combustion period (see step 207 in FIG. 29), and the average turbulence intensity ST1 in the initial combustion period as the weighted average value is calculated as a laminar flow in the initial combustion period. The turbulent flame speed FLAME1 (= SL1 × ST1) in the initial combustion period obtained by multiplying the flame speed SL1 is the turbulent flame speed FLAME1o at the intake valve opening timing IVO and the intake valve closing timing IVC. Between the turbulent flame speed FLAME1c (refer to the broken line in the sixth stage) and during the main combustion period in the state of the intake valve opening timing IVO. A value obtained by weighted averaging the average turbulence intensity Uo2 and the average turbulence intensity Uc2 in the main combustion period in the state of the intake valve closing timing IVC is calculated as the average turbulence intensity ST2 in the main combustion period (step in FIG. 37). 258), and the turbulent flame speed FLAME2 (= SL2 × ST2) in the main combustion period obtained by multiplying the laminar flame speed SL2 in the main combustion period by the average turbulence intensity ST2 in the main combustion period as the weighted average value. Takes a value between the turbulent flame speed FLAME2o in the state of the intake valve opening timing IVO and the turbulent flame speed FLAME2c in the state of the intake valve closing timing IVC (see the bottom broken line).

  As a result, no. Even when there is a cylinder in which the intake port pressure changes during the intake stroke and the intake valve opening timing state and the intake valve closing timing state are different, such as the two cylinders, the combustion period (the initial combustion period, the main combustion period, (Combustion period) can be accurately calculated. Note that the same effect can be obtained during deceleration.

  FIG. 45 shows the change in the pressure in the combustion chamber.

  (A) and (B) are respectively in the low load state and the high load state (both in steady state), and the intake port pressure is higher in the high load state than in the low load state, and the maximum value at the reference crank angle θPMAX. Is high. In any state, ignition is performed at the basic ignition timing MBTCAL, so that the pressure in the combustion chamber is maximum at the position of the reference crank θPmax.

  In these two steady states, the state of the intake valve opening timing and the state of the intake valve closing timing do not change, so the basic ignition timing MBTCAL is higher than the basic ignition timing MBTCAL in a high load state in which combustion proceeds rapidly. The basic ignition timing MBTCAL in the low load state where the combustion is moderate comes to the retard side.

  On the other hand, (D) changes transiently from the low load state to the high load state, and the intake port pressure increases during the intake stroke, that is, the intake valve close timing is higher than the intake port pressure in the intake valve open timing state. Of the combustion chamber pressure when ignition is performed at the basic ignition timing calculated based on the intake port pressure in the intake valve opening timing state for the cylinder (No. 2 cylinder described above) in which the intake port pressure in the state is high A change is causing a knock. This is because the actual average turbulence intensity rises higher than the average turbulence intensity at the intake valve opening timing as the intake port pressure increases after the intake valve opening timing. The flame speed (combustion speed) also rises from the turbulent flame speed calculated from the average turbulence intensity at the intake valve opening timing. Therefore, when the ignition is performed at the basic ignition timing MBTCAL calculated using the average turbulence intensity in the intake valve opening timing state, it is too early than the actual required ignition timing, and the combustion proceeds rapidly and knocks. It will wake you up.

  On the other hand, (C) is a weighted average of the average turbulence strengths Uo1 and Uo2 in the intake valve opening timing state and the average turbulence strengths Uc1 and Uc2 in the intake valve closing timing state according to this embodiment. This shows the change in the pressure in the combustion chamber when the basic ignition timing MBTCAL is calculated based on the values ST1 and ST2 obtained in this way. According to (C), the turbulent flame speed and the maximum value of the pressure in the combustion chamber are located between (A) and (B), and in this case, the pressure in the combustion chamber is maximum at the reference crank θPmax, Knock does not occur.

  FIG. 45 shows the change in the combustion chamber pressure for the cases (A) to (D), while FIG. 46 shows the change in the combustion mass ratio for the cases (A) to (D). It is a thing.

  In FIG. 46, (A) and (B) are the low load state and the high load state (both in steady state), respectively, and the combustion state is better in the high load state because the intake port pressure is higher than in the low load state. Therefore, the rise of the combustion mass ratio is abrupt. In any steady state, ignition is performed at the basic ignition timing MBTCAL, so the combustion mass ratio at the reference crank angle θPmax is a constant value Rmax (= 60%).

  In these two steady states, the state of the intake valve opening timing and the state of the intake valve closing timing do not change, so the basic ignition timing MBTCAL is higher than the basic ignition timing MBTCAL in a high load state in which combustion proceeds rapidly. The basic ignition timing MBTCAL in the low load state where the combustion is moderate comes to the retard side.

  On the other hand, (D) changes transiently from the low load state to the high load state, and the intake port pressure increases during the intake stroke, that is, the intake valve close timing is higher than the intake port pressure in the intake valve open timing state. Mass ratio when ignition is performed at the basic ignition timing MBTCAL calculated based on the intake port pressure in the state of the intake valve open timing for the cylinder (No. 2 cylinder described above) in which the intake port pressure in the state is high The change has caused a knock. This is because the actual average turbulence intensity rises higher than the average turbulence intensity at the intake valve opening timing as the intake port pressure increases after the intake valve opening timing. The flame speed (combustion speed) also rises from the turbulent flame speed calculated from the average turbulence intensity at the intake valve opening timing. Accordingly, when ignition is performed at the basic ignition timing MBTCAL calculated using the average turbulence intensity in the intake valve opening timing state, the ignition timing is too early than the actual required ignition timing, and the reference crank angle θPmax at this time is The combustion mass ratio becomes larger than the above-mentioned constant value Rmax, and knocking occurs.

  On the other hand, (C) is a weighted average of the average turbulence strengths Uo1 and Uo2 in the intake valve opening timing state and the average turbulence strengths Uc1 and Uc2 in the intake valve closing timing state according to this embodiment. The change of the combustion mass ratio when the basic ignition timing MBTCAL is calculated based on the obtained values is shown. According to (C), the combustion mass ratio at the reference crank angle θPmax coincides with the above-mentioned constant value Rmax, and knock does not occur.

  Here, the effect of this embodiment is demonstrated.

A new experimental result has been obtained that the extent to which the squish influences the turbulent flame velocity during the main combustion period is determined by the shape of the squish. This means that since the shape of the squish is determined by the engine specifications, the degree to which the squish affects the turbulent flame speed during the main combustion period can be estimated approximately by the engine specifications. That is, according to the present embodiment (the invention described in claims 1 and 2 ), the main combustion period BURN2 includes a period during which squish occurs, and the squish correction coefficient Kq (flame speed correction amount by squish) is set as the squish area volume or squish area. (See steps 264 and 265 in FIG. 38), the combustion turbulent flame speed FLAME2 in the main combustion period is changed to the laminar flame speed SL2 in the main combustion period, and this squish. Since the calculation is based on the correction coefficient Kq (see Steps 189 and 190 in FIG. 12 and Steps 255, 256 and 258 in FIG. 37), it is not necessary to re-fit every time the engine model is different, and the number of man-hours for fitting can be reduced.

According to the experimental results, the initial combustion period excluding the time period of occurrence of squish, thus Squish corresponds to not affect the turbulent flame speed FLAME1 between initial combustion, the present embodiment (according to claim 3 According to the invention, the turbulent flame speed FLAME1 in the initial combustion period is not calculated based on the squish correction coefficient Kq (flame speed correction amount by squish), so that unnecessary calculation is not required.

  According to this embodiment (the invention according to claim 4), the larger the squish area volume, the stronger the average turbulence intensity in the main combustion period and the higher the turbulent flame speed FLAME2 in the main combustion period. Since the squish correction coefficient Kq (flame speed correction amount by squish) is calculated based on the squish area volume (sqv1 to sqv4) that is the engine specification (see step 264 in FIG. 38), regardless of the size of the squish area volume. The turbulent flame speed FLAME2 during the main combustion period can be accurately calculated.

Corresponding to the change of the average turbulence intensity in the main combustion period and the change of the turbulent flame speed FLAME2 in the main combustion period in the gas blowing direction from the squish area, this embodiment (invention according to claim 5 ) is applied. According to this, since the squish correction coefficient Kq (flame speed correction amount by squish) is calculated based on the gas blowing direction from the squish area, which is the engine specification (see step 265 in FIG. 38), the gas blowing direction from the squish area. Regardless of, the turbulent flame speed FLAME2 in the main combustion period can be accurately calculated.

Corresponding to the fact that the shape of the squish area varies from part to part in one cylinder, according to the present embodiment (the invention according to claim 6 ), the squish area is divided into a plurality of parts in one cylinder, A part-specific squish correction coefficient ksq (i) is calculated for each divided part-specific squish area, and the sum of these part-specific squish correction coefficients ksq (i) is used as the squish complement coefficient Kq (depending on the squish) for the entire cylinder. (Refer to step 269 in FIG. 38), the turbulent flame speed FLAME2 in the main combustion period can be accurately calculated even when the shape of the squish area is different for each part in one cylinder.

According to the present embodiment (the invention described in claim 7 ), when the VEL mechanism 51 (variable valve mechanism) changes from operation to non-operation between the intake valve opening timing IVO and the intake valve closing timing IVC. Based on the intake valve passing average flow velocity Vo and the squish correction coefficient Kq (flame speed correction amount by squish) based on the intake valve opening timing IVO, the average turbulence intensity during the main combustion period in the intake valve opening timing IVO state The intake valve is closed based on the means for calculating Uo2 (see step 255 in FIG. 37), the intake valve passing average flow velocity Vc in the state of the intake valve closing timing IVC, and the squish correction coefficient Kq (flame speed correction amount by squish). Means for calculating the average turbulence intensity Uc2 in the main combustion period in the state of the timing IVC (see step 256 in FIG. 37), and calculating these weighted average values in the main combustion period The turbulent flame speed FLAME2 in the main combustion period is calculated from the means for calculating the turbulence intensity ST2 (see step 258 in FIG. 37), the average turbulence intensity ST2 in the main combustion period, and the laminar flame speed SL2 in the main combustion period. In the case where the VEL mechanism 51 changes from the operation to the non-operation during the period from the intake valve opening timing IVO to the intake valve closing timing IVC, as shown in FIG. The average turbulence intensity ST2 during the main combustion period can be accurately calculated.

Corresponding to the experimental result that the average turbulence intensity ST2 in the main combustion period is proportional to the engine speed NRPM, according to this embodiment (the invention according to claim 8 ), the intake valve opening timing A value obtained by further multiplying the weighted average value of the average turbulence intensity Uo2 in the main combustion period in the IVO state and the average turbulence intensity Uc2 in the main combustion period in the state of the intake valve closing timing IVC by the engine speed NRPM. Since the average turbulence intensity ST2 in the main combustion period is calculated (see step 258 in FIG. 37), the average turbulence intensity ST2 in the main combustion period can be easily calculated.

According to the present embodiment (the invention described in claim 9 ), the intake valve opening is based on the intake valve passing average flow velocity Vo and the squish correction coefficient Kq (combustion speed correction amount by squish) in the intake valve opening timing state. Means for calculating the average turbulence intensity Uo2 in the main combustion period in the timing state (see step 255 in FIG. 37), the intake valve passage average flow velocity Vc in the intake valve closing timing state, and the squish correction coefficient Kq (depending on the squish) Means for calculating an average turbulence intensity Uc2 in the main combustion period in the state of the intake valve closing timing based on the combustion speed correction amount) (see step 256 in FIG. 37), and calculating these weighted average values in the main combustion period Means for calculating the average turbulence intensity ST2 (see step 258 in FIG. 37), the average turbulence intensity ST2 in the main combustion period, and the laminar flame in the main combustion period Means for calculating the turbulent flame speed FLAME2 in the main combustion period from the degree SL2 (see step 190 in FIG. 12), so that the VEL mechanism 51 maintains either the operating state or the non-operating state. In this case, the average turbulence intensity ST2 in the main combustion period can be accurately calculated even during acceleration or deceleration when the intake port pressure changes greatly between the intake valve opening timing IVO and the intake valve closing timing IVC.

  In the embodiment, the initial combustion period is changed from zero to 2% (ie, BR1 = 2%) as the change amount of the combustion mass ratio, and the main combustion period is changed from 2 to 60% (ie, BR2 = 58%). However, the present invention is not necessarily limited to this value.

  In the embodiment, the combustion gas mass ratio is described, but the combustion gas mass itself may be used.

In the embodiment, the case where the combustion period is divided into two, that is, the initial combustion period in which squish does not occur and the main combustion period in which squish occurs, has been described, but it may be divided into three or more (claim 1). invention). For example, the change amount of the combustion gas mass combusted in the combustion chamber from the start of combustion to a predetermined crank angle is divided into a plurality of 3 or more, and the divided combustion period corresponding to each divided change amount of the combustion gas mass is calculated. A value obtained by summing all the divided combustion periods may be calculated as a combustion period from the start of combustion to a predetermined crank angle (inventions according to claims 1 and 2 ).

In the embodiment, the case where the VEL mechanism 51 (variable valve mechanism) changes from the operation to the non-operation between the intake valve opening timing and the intake valve closing timing has been described. However, from the intake valve opening timing to the intake valve closing timing. The present invention can also be applied to a case where the VEL mechanism 51 changes from non-operation to operation during the period (the invention according to claim 7 ).

  In the embodiment, the case where the VEL mechanism 51 is provided has been described. However, it is needless to say that the present invention is applicable to a case where the VEL mechanism 51 is not provided.

In the first aspect of the invention, the function of the combustion speed calculation means is the step 170 of FIG. 10 and the step 190 of FIG. 12, and the function of the basic ignition timing calculation means is the step 43 of FIG. 38 is performed by steps 264 and 265 in FIG. 38, and the function of the turbulent flame speed calculation means is performed by steps 189 and 190 in FIG. 12, and steps 254 to 258 in FIG.

In the second aspect of the invention, the function of the combustion speed calculating means is the step 170 of FIG. 10 and the step 190 of FIG. 12, and the function of the basic ignition timing calculating means is the step 43 of FIG. 38 is performed by steps 264 and 265 in FIG. 38, and the function of the turbulent flame speed calculation means is performed by steps 189 and 190 in FIG. 12, and steps 254 to 258 in FIG.

The engine control system figure of one Embodiment. The block diagram of the ignition timing control performed with an engine controller. The pressure change figure of a combustion chamber. The characteristic view explaining the change of a combustion mass ratio. The flowchart for demonstrating calculation of a physical quantity. Diagram explaining the positional relationship between the crankshaft of the engine and the connecting rod. The characteristic diagram of a water temperature correction coefficient. The characteristic view of an equivalence ratio correction coefficient. The characteristic figure of a standard crank angle. The flowchart for demonstrating calculation of an initial combustion period. The characteristic figure of a temperature rise rate. The flowchart for demonstrating calculation of the main combustion period. The flowchart for demonstrating calculation of basic ignition timing. The flowchart for demonstrating calculation of an internal inert gas rate. The flowchart for demonstrating calculation of an internal inert gas amount. The flowchart for demonstrating calculation of the amount of inert gas at the time of EVC. The flowchart for demonstrating calculation of the amount of inert gas blown back during overlap. The flowchart for demonstrating the setting of a supercharging determination flag and a choke determination flag. The flowchart for demonstrating calculation of the inactive gas flow rate during the overlap when there is no supercharging and there is no choke. The flowchart for demonstrating the calculation of the inactive gas flow rate during the overlap at the time of supercharging without a choke. The flowchart for demonstrating the calculation of the inactive gas flow rate during the overlap at the time of supercharging and no choke. The flowchart for demonstrating calculation of the inactive gas flow rate during the overlap at the time of supercharging and a choke. The characteristic figure of the combustion chamber volume in the exhaust valve closing timing. The characteristic figure of the gas constant of an inert gas. The characteristic figure of the integrated effective area during overlap. Explanatory drawing of the integrated effective area during overlap. The characteristic figure of the specific heat ratio of an inert gas. The characteristic figure of the specific heat ratio of air-fuel | gaseous mixture. The flowchart for demonstrating calculation of the average turbulence intensity in an initial combustion period. The flowchart for demonstrating calculation of the intake valve passage average flow velocity in the state of intake valve opening timing. When the VEL mechanism changes to non-operation during the intake stroke, the valve opening crank angle section θa in the intake valve opening timing state, the intake valve lift amount La in the intake valve opening timing state, and the intake valve closing timing FIG. 6 is a waveform diagram for explaining an intake valve lift amount Lb in a state of a valve opening crank angle section θb in a state and an intake valve closing timing; The characteristic figure of a correction coefficient. The characteristic view of an average intake valve opening area. The flowchart for demonstrating calculation of the intake valve passage average flow velocity in the state of intake valve closing timing. The characteristic figure of the average turbulence intensity basic value in an initial combustion period. The flowchart for demonstrating calculation of a weighted average coefficient. The flowchart for demonstrating calculation of the average turbulence intensity in the main combustion period. The flowchart for demonstrating calculation of a squish correction coefficient. The schematic plan view of the engine for demonstrating the division | segmentation of the squish area in one cylinder. The table of squish area volume according to part. The characteristic view of the part specific gas blowing direction correction coefficient. The schematic sectional drawing for demonstrating the gas blowing direction from a squish area. The characteristic figure of the average turbulence intensity basic value in the main combustion period. The wave form diagram which shows the change of the typical value at the time of acceleration. The wave form diagram which shows the change of a combustion chamber pressure. The wave form diagram which shows the change of a combustion mass ratio.

Explanation of symbols

1 Engine 5 Combustion chamber 11 Ignition device (spark ignition means)
15 Intake valve 31 Engine controller 51 VEL mechanism (variable valve mechanism)

Claims (9)

A combustion speed calculating means for dividing a combustion period in which the air-fuel mixture burns in the combustion chamber into a plurality of parts, and calculating a combustion speed in each of the divided combustion periods based on operating conditions;
Basic ignition timing calculation means for calculating an ignition timing at which MBT is obtained based on the plurality of combustion rates as a basic ignition timing;
In an engine ignition timing control device comprising spark ignition means for performing spark ignition at this basic ignition timing,
Any one of the combustion periods includes a period during which squish occurs;
The combustion speed in the combustion period including the period in which the squish occurs is the turbulent flame speed in the combustion period including the period in which the squish occurs,
And flame speed correction amount calculating means for calculating, based on at least engine specifications flame speed correction amount by squish,
The turbulent flame speed in the combustion period including the squish occurrence period is determined by at least one of the laminar flame speed and the engine operating condition in the combustion period including the squish occurrence period, and the flame speed correction amount by the squish. Turbulent flame speed calculating means for calculating based on
Ignition timing control apparatus for an engine, characterized in that it comprises a. A combustion rate calculating means for dividing a combustion period in which the air-fuel mixture burns in the combustion chamber into an initial combustion period and a main combustion period, and calculating a combustion speed in each divided combustion period based on operating conditions;
Basic ignition timing calculation means for calculating an ignition timing at which MBT is obtained based on the plurality of combustion rates as a basic ignition timing;
Spark ignition means for performing spark ignition at this basic ignition timing and
In an engine ignition timing control device comprising:
The main combustion period includes a period during which squish occurs;
The combustion speed in the main combustion period is the turbulent flame speed in the main combustion period;
Flame speed correction amount calculating means for calculating a flame speed correction amount by squish based on at least engine specifications;
The turbulent flame speed for calculating the turbulent flame speed in the main combustion period based on at least one of the laminar flame speed and engine operating conditions in the main combustion period and the flame speed correction amount by the squish With calculation means
Ignition timing control system characteristics and to Rue engine that comprises a. 3. The engine ignition timing control device according to claim 2, wherein the turbulent flame speed in the initial combustion period is not calculated based on a flame speed correction amount by the squish . The engine ignition timing control device according to claim 1 or 2, wherein a flame speed correction amount by the squish is calculated based on a squish area volume which is a specification of the engine. 3. The engine ignition timing control device according to claim 1, wherein a turbulent flame speed correction amount by the squish is calculated based on a gas blowing direction from a squish area which is a specification of the engine. The squish area is divided into a plurality of parts in one cylinder, and the squish combustion speed correction amount is calculated for each divided squish area, and the squish flame speed correction amount for each squish area is calculated. 6. The engine ignition timing control device according to claim 4, wherein the total value is used as a flame speed correction amount by squish for the entire cylinder . A variable valve mechanism that continuously and variably controls the valve lift amount and operating angle of the intake valve,
When this variable valve mechanism changes from operation to non-operation or vice versa between intake valve opening timing and intake valve closing timing,
Means for calculating an average turbulence intensity in the main combustion period in the intake valve opening timing state based on the intake valve passing average flow velocity in the intake valve opening timing state and the combustion speed correction amount by the squish;
Means for calculating an average turbulence intensity in the main combustion period in the intake valve closing timing state based on the intake valve passing average flow velocity in the intake valve closing timing state and the combustion speed correction amount by the squish;
Means for calculating these weighted average values as the average turbulence intensity during the main combustion period;
Means for calculating the turbulent flame speed in the main combustion period from the average turbulence intensity in the main combustion period and the laminar flame speed in the main combustion period;
Ignition timing control device for an engine according to claim 2, characterized in that it contains. 8. The engine ignition timing control apparatus according to claim 7 , wherein a value obtained by further multiplying the weighted average value by an engine speed is calculated as an average turbulence intensity in the main combustion period . Means for calculating an average turbulence intensity in the main combustion period in the intake valve opening timing state based on the intake valve passing average flow velocity in the intake valve opening timing state and the combustion speed correction amount by the squish;
Means for calculating an average turbulence intensity in the main combustion period in the intake valve closing timing state based on the intake valve passing average flow velocity in the intake valve closing timing state and the combustion speed correction amount by the squish;
Means for calculating these weighted average values as the average turbulence intensity during the main combustion period;
Means for calculating the turbulent flame speed in the main combustion period from the average turbulence intensity in the main combustion period and the laminar flame speed in the main combustion period;
Ignition timing control device for an engine according to claim 2, characterized in that it contains.

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