JP2007092616A - Ignition timing control method of engine and ignition timing control device of engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To accurately calculate the ignition timing for providing MBT even in an engine having a variable valve gear capable of particularly changing a lift characteristic of an intake valve. <P>SOLUTION: An engine controller 31 executes a combustion pressure maximum crank angle calculation processing procedure of calculating a crank angle of becoming maximum in combustion pressure on the basis of a loss in the engine and a basic ignition timing calculation processing procedure of calculating the basic ignition timing for providing the MBT on the basis of the crank angle of becoming maximum in this combustion pressure. <P>COPYRIGHT: (C)2007,JPO&INPIT

Description

  The present invention relates to an ignition timing control method for an engine (internal combustion engine) and an ignition timing control device for the engine.

The reference crank angle is set to the crank angle θpmax at which the combustion pressure of the air-fuel mixture becomes maximum, and the crank angle at which the reference crank angle and the combustion mass ratio representing the ratio of the fuel mass to the fuel supplied to the combustion chamber reach a certain ratio Assuming that the position (specifically, the crank angle position at which the fuel mass ratio reaches 60%, θmb 60%) is equal, the crank angle position measured from the crank angle position to the advance side for the combustion period BURN is MBT. Is calculated as the ignition timing obtained (see Patent Document 1).
JP 2004-301045 A

  By the way, in recent years, development of a variable valve gear that can change the lift characteristics of the intake valve has been actively promoted, and the ignition timing control method of Patent Document 1 is applied to an engine equipped with such a variable valve gear. Experiments have shown that it is rare that the combustion mass ratio is always constant at the crank angle θpmax at which the combustion pressure is maximum. Therefore, simply applying the ignition timing control method of Patent Document 1 to an engine equipped with a variable valve device may cause the calculated ignition timing to deviate from the ignition timing at which MBT is obtained, resulting in deterioration of fuel consumption. .

  On the other hand, according to the results examined on the desk, the crank angle θpmax at which the combustion pressure becomes maximum fluctuates in relation to the engine loss, and the crank angle θpmax at which the combustion pressure becomes maximum is delayed as the engine loss increases. it is conceivable that. In this case, engine loss includes engine cooling loss Qcool and engine time loss Qtime.

  Here, the engine cooling loss Qcool is the amount of heat taken away by the engine coolant from the combustion heat generated in the combustion chamber. Typical values for estimating the engine cooling loss Qcool include the engine load and the rotational speed Ne. If the engine cooling loss Qcool is calculated using the Wassini equation, the engine rotational speed Ne is the same. If the engine load increases, the engine cooling loss Qcool increases. If the engine load is the same, the engine cooling loss Qcool decreases as the engine speed Ne increases. Therefore, when the engine speed Ne is the same, the crank angle θpmax at which the combustion pressure becomes maximum is delayed as the engine load increases, and when the engine load Ne is the same, the combustion pressure becomes maximum as the engine speed Ne increases. The crank angle θpmax is advanced. The crank angle θpmax at which the combustion pressure is maximized as the engine is heavily loaded is delayed because the amount of heat deprived by the cooling water out of the amount of heat that should contribute to the combustion pressure (that is, the cooling loss) is larger as the engine is heavily loaded. It is to become.

  Next, the engine time loss is a loss that occurs because combustion does not end instantaneously and a combustion period of about 40 to 60 deg is required at the crank angle. In other words, the longer the combustion period, the lower the actual compression ratio and the lower the thermal efficiency. This decrease in thermal efficiency is expressed by engine time loss. Typical values for estimating engine time loss include combustion speed and combustion period. If the combustion speed decreases, the thermal efficiency decreases, so the engine time loss increases and the combustion period increases (engine The engine speed loss is increased because the thermal efficiency is also reduced. For this reason, if the engine speed is the same, the crank angle θpmax at which the combustion pressure becomes maximum is delayed as the combustion speed is reduced, and if the combustion speed is the same, the combustion period is longer (the engine speed is lower). The crank angle θpmax at which becomes the maximum is delayed. If the engine speed is the same, the crank angle θpmax at which the combustion pressure becomes maximum is delayed as the combustion speed is decreased because the combustion angle is less likely to contribute to the smaller combustion speed.

  However, in the technique of Patent Document 1, the crank angle θpmax at which the combustion pressure becomes maximum is set as a constant value or is calculated according to the rotation speed. Therefore, now, when considering the case where the engine load is increased without changing the rotational speed from the operating conditions of the engine load and the rotational speed when the crank angle θpmax at which the combustion pressure is maximized is set, the combustion pressure is considered. The crank angle θpmax at which the maximum value is maximized is calculated as an ignition timing assuming that the crank angle θpmax does not deviate from that before the load increase, so that the calculated ignition timing is advanced from the ignition timing at which MBT is obtained. It becomes the value on the corner side. On the other hand, when considering the case where the rotational speed does not change and only the load becomes smaller than the operating condition of the engine load and the rotational speed when the crank angle θpmax at which the combustion pressure is maximized is set, the combustion pressure is reduced. The maximum crank angle θpmax is calculated so that the ignition timing is not deviated in spite of the fact that the crank angle θpmax is deviated from before the decrease in load. Therefore, the calculated ignition timing is retarded from the ignition timing at which MBT is obtained. It becomes the value of the side.

  As described above, if the crank angle θpmax at which the combustion pressure is maximized is set to a constant value, or only the sensitivity of the rotational speed is given, there may occur a situation where the ignition timing at which MBT is obtained cannot be calculated with high accuracy.

  Accordingly, the present invention provides an ignition timing control method and an ignition timing control device capable of accurately calculating an ignition timing at which MBT can be obtained even in an engine having a variable valve device that can change the lift characteristics of an intake valve. The purpose is to provide.

  The present invention is configured to calculate a crank angle θpmax at which the combustion pressure is maximum based on engine loss, and to calculate a basic ignition timing at which MBT is obtained based on the crank angle θpmax at which this combustion pressure is maximum. To do.

  Although the crank angle θpmax at which the combustion pressure is maximum varies depending on the operating conditions, the value is determined by the engine loss, and according to the present invention, the crank angle θpmax at which the combustion pressure is maximum is used as the engine loss. Since the basic ignition timing at which MBT is obtained is calculated based on the crank angle θpmax at which the combustion pressure is maximized, even if the crank angle θpmax at which the combustion pressure is maximized varies depending on operating conditions. Since the fluctuation θpmax can be estimated with high accuracy, the prediction accuracy of the basic ignition timing from which the MBT can be obtained can be improved.

  In particular, an engine having a variable valve device that can change the lift characteristics of the intake valve is particularly effective for such an engine because θpmax varies more greatly than an engine that does not have a variable valve device. However, it is needless to say that the present invention is applicable to an engine that does not include a variable valve operating device.

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

  FIG. 1 shows a schematic configuration of an ignition timing control apparatus for an engine that is directly used for carrying out an ignition timing control method for an engine.

  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 electric energy from the battery, a power transistor (not shown) that supplies and shuts off the primary side of the ignition coil 13, and an ignition coil that is provided on the ceiling of the combustion chamber 5. 13 includes a spark plug 14 that receives a high voltage generated on the secondary side of the ignition coil 13 by cutting off the primary current 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.

The exhaust passage 8 includes three-way catalysts 9 and 10. When the air-fuel ratio of the exhaust is within a narrow range (window) centered on the stoichiometric air-fuel ratio, the three-way catalysts 9, 10 can efficiently remove harmful three components such as HC, CO, and NOx contained in the exhaust 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.

  The intake valve 15 and the exhaust valve 16 are driven to open and close by operation of cams provided on the intake side camshaft 25 and the exhaust side camshaft 26, respectively, with the crankshaft 7 as a power source. On the intake side, a variable valve mechanism (hereinafter referred to as “VEL mechanism”) 28 configured with 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. The VEL mechanism 28 is provided with an operating angle sensor (not shown) for detecting the valve lift amount and operating angle of the intake valve 15.

  Similarly, on the intake side, a variable valve timing mechanism (hereinafter referred to as “VTC mechanism”) that continuously and variably controls the rotational phase difference between the crankshaft 7 and the intake camshaft 25 to advance and retard the valve timing of the intake valve 15. 27). A cam angle sensor 34 for detecting the rotational position of the intake side camshaft 25 is also provided at the other end of the intake side camshaft 25.

  Since the specific configurations of the VEL mechanism 28 and the VTC mechanism 27 (variable valve operating device) are known from Japanese Patent Application Laid-Open No. 2003-3872, detailed description thereof is omitted.

  When the actuators of the VEL mechanism 28 and the VTC mechanism 27 are commanded to change the lift characteristics of the intake valve 15 (valve timing (opening / closing timing) and valve lift amount of the intake valve 15), the inert gas remaining in the combustion chamber 5 is reduced. The amount 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 valve indicate how much inert gas should remain in the combustion chamber 5 depending on the operating conditions. The engine controller 31 determines a target intake valve closing timing and a target valve lift amount based on the operating conditions (engine load and rotational speed) at that time, and VTC so that these target values can be obtained. The closing timing and valve lift amount of the intake valve 15 are controlled via the actuators of the mechanism 27 and the VEL mechanism 28.

  In the engine controller 31 to which an intake air temperature signal from the intake air temperature sensor 43, an intake air pressure signal from the intake air pressure sensor 44, and an exhaust gas temperature signal from the exhaust air temperature sensor 45 are also input, the spark plug 14 is connected via the power transistor 13. The ignition timing, which is the primary current cutoff timing, is controlled.

  Now, when the air-fuel mixture is ignited with MBT (minimum advance value for obtaining the maximum torque) and the crank angle θpmax at which the combustion pressure of the air-fuel mixture reaches the maximum value Pmax is the reference crank angle, the reference crank angle is the combustion method. Regardless, it is almost constant. Further, according to the combustion analysis in the combustion chamber, the combustion mass ratio representing the ratio of the fuel mass to the fuel supplied to the combustion chamber 5 is 0% at the time of ignition and reaches 100% by complete combustion. Then, assuming that the combustion mass ratio at the reference crank angle is constant and about 60%, the combustion speed is obtained from the operating conditions (engine load and rotation speed) at that time, and the combustion period BURN is calculated based on this combustion speed. Has proposed an ignition timing control method in which the crank angle position on the advance side is calculated as the basic ignition timing MBTCAL for the total crank angle section of the combustion period BURN and the ignition delay time equivalent angle IGNDEAD from the reference crank angle (special feature). No. 2004-332647).

  Such an ignition timing control method based on combustion analysis (this ignition timing control method is hereinafter referred to as “preceding ignition timing control method”) is an engine in which the valve lift amount of the intake valve 15 and the opening / closing timing of the intake valve 15 do not change (conventional engine). As long as it is applied to conventional engines, there is no problem with the preceding ignition timing control method as long as it is applied to conventional engines. However, for engines equipped with VEL mechanism 28 and VTC mechanism 27 However, it has been found that when the preceding ignition timing control method is applied, the basic ignition timing MBTCAL does not match the ignition timing at which MBT is obtained. The reason is that the VEL mechanism 28 and the VTC mechanism 27 are in a non-operating state as long as various values used for ignition timing control are adapted when the VEL mechanism 28 and the VTC mechanism 27 are in a non-operating state. Sometimes the basic ignition timing MBTCAL coincides with the ignition timing at which MBT is obtained.

  In this state, that is, when the VEL mechanism 28 and the VTC mechanism 27 are operated to change the lift characteristics (the valve lift amount of the intake valve 15 and the intake valve closing timing) under the same operating conditions, the tumble that is the gas flow in the combustion chamber 5 Each strength of the swirl changes, and accordingly, the turbulent combustion speed, which is the combustion speed in the turbulent state of the combustion chamber gas, changes. Then, the ignition timing at which MBT is obtained also changes. However, since the operating conditions are the same even when the VEL mechanism 28 and the VTC mechanism 27 are operated, the basic ignition timing MBTCAL does not change. Therefore, as soon as the VEL mechanism 28 and the VTC mechanism 27 are operated, the basic ignition timing MBTCAL deviates from the ignition timing at which MBT is obtained, resulting in poor fuel consumption.

  For example, if the valve lift amount of the intake valve 15 is switched from a large state to a small state by the operation of the VEL mechanism 28, tumble and swirl are less likely to be formed before the switching, and the turbulent combustion speed becomes slower and the combustion period is prolonged. . Accordingly, under the same operating conditions, when the valve lift amount is small, the ignition timing at which MBT is obtained moves to the advance side when the valve lift amount is large, but the valve lift amount is small. Even in the state, if the same combustion period BURN as when the valve lift amount is large is calculated, a combustion period shorter than the actual one is calculated, and the ignition timing at which the basic ignition timing MBTCAL is obtained as MBT. Will be too late.

  The same applies when the intake valve closing timing is advanced from the late state to the early state by the operation of the VTC mechanism 27. That is, when the intake valve closing timing is advanced from a late state to an early state, tumble and swirl are less likely to be formed than before the intake valve is closed, so that the turbulent combustion speed is reduced and the combustion period is lengthened. Therefore, under the same operating conditions, when the intake valve closing timing is early, the ignition timing at which MBT is obtained moves to the advance side when the intake valve closing timing is late. If the combustion period BURN that is the same as when the intake valve closing timing is late even when the timing is early is calculated, a combustion period that is shorter than the actual one is calculated, and the basic ignition timing MBTCAL is an ignition at which MBT is obtained. It will be too late.

  Therefore, in the present embodiment, even if an engine including the VEL mechanism 28 and the VTC mechanism 27 is targeted, the control structure according to the physical model is used and applied to a conventional engine in order to make the control unnecessary for adaptation as much as possible. <1> The method for calculating the combustion speed, <2> The method for setting the reference crank angle, and <3> the method for calculating the ignition delay time with respect to the preceding ignition timing control method.

  Here, not only the above <1> but also <2> and <3> are added as a result of reviewing the control method. First, <2> is also added for the following reason. That is, the relationship between the reference crank angle and the crank angle position when the combustion mass ratio is 60%, which has been assumed to be correct with respect to the preceding ignition timing control method applied to the conventional engine, is expressed as follows. When an experiment is performed on an engine equipped with the VTC mechanism 27, there is a large deviation (crank angle difference) between the reference crank angle and the crank angle position when the combustion mass ratio is 60%, and the deviation is calculated as it is. This is because it becomes an estimation error of the value (MBTCAL), and it is necessary to change the setting method of the reference crank angle position.

  The reason why <3> is added is as follows. The ignition delay time of <3> is a time during which the combustion mass ratio is 0% starting from the ignition timing, but the estimation error of this ignition delay time is large. The reason is that the measurement error when measuring the crank angle position where the combustion mass ratio is 0% by the combustion analysis device is large, and if the measured value is made positive and the ignition delay time DEADTIME is adapted, This is because the estimation error of the period BURN includes this measurement error and the error becomes large. Therefore, the calculation method of the ignition delay time is modified so that the measurement error by the combustion analyzer does not enter.

Hereinafter, the items will be described separately.
<1> Calculation Method of Combustion Rate In the preceding ignition timing control method applied to the conventional engine, the combustion rate Sb is set to a laminar combustion rate SL that is a combustion rate in a laminar flow state of the combustion chamber gas, and The combustion speed Sb is obtained from the sum of the turbulent combustion speed ST, which is the combustion speed in the turbulent state, that is, the following (Supplement 1).

Sb = SL + ST (Supplement 1)
The laminar combustion speed SL and the turbulent combustion speed ST in (Supplement 1) are given by the following (Supplement 2) and (Supplement 3), respectively.

SL = SL0 * (T / 298) ^ n * (p / 101.325) ^ d
× (1-b × MRESR ^ k) (Supplement 2)
ST = St · Ne (Supplement 3)
However, SL0: Laminar burning velocity in standard state [m / sec],
T: unburned gas temperature [K],
p: pressure in the combustion chamber [kPa],
MRESR: Internal inert gas ratio [%]
St: Turbulent combustion rate coefficient,
n, d, b, k: coefficients,
The turbulent combustion speed ST in (Supplement 3) is determined by the gas turbulence intensity in the combustion chamber 5. Further, as shown in (Supplement 3), the turbulent combustion speed St is used as a function of the engine speed Ne because the turbulence intensity of the gas in the combustion chamber 5 depends on the valve lift amount of the intake valve 15 and the opening / closing timing. In the same case, since it is proportional to the piston speed (engine speed), it is considered that it is sufficient to give the sensitivity of the engine speed Ne to the estimation of the turbulent combustion speed St.

  However, in an engine provided with the VEL mechanism 28 and the VTC mechanism 27, unlike the conventional engine, the valve lift amount and the opening / closing timing of the intake valve 15 vary greatly depending on the operating conditions, and the combustion chamber is caused by changes in the valve lift amount and the opening / closing timing. In this embodiment, the turbulent combustion speed is calculated as follows instead of the above (Supplement 3). That is, in order to cope with the change in the valve lift amount of the intake valve 15 due to the operation of the VEL mechanism 28 and the change in the intake valve closing timing accompanying the operation of the VTC mechanism 27, the influence of tumble is particularly used in the estimation formula of the combustion speed. Add.

  First, the combustion speed Sb [m / sec] is expressed by the following equation (1). The second term on the right side of the equation (1) is an equation that gives the turbulent combustion speed ST, and it is well known that the turbulent combustion speed ST is given by an exponential function of this variable with u / SL as a variable.

Sb = SL + b · (u / SL) ^ a (1)
Where SL is the laminar burning velocity,
u: Disturbance strength,
a, b: fitness coefficient,
The turbulence intensity u [m / sec], which is one characteristic value of the turbulence associated with the gas flow in the combustion chamber of the equation (1), is expressed by the following equation (2). This is because the turbulence strength u of the combustion chamber gas is proportional to the rotational speed Ne, and the proportionality constant is known to correlate with swirl strength and tumble strength. It is expressed in the formula.

u = c · It · Ne (2)
However, It: Tumble strength [anonymous number]
Ne: Engine rotation speed [rpm],
c: conformity factor,
Here, in this embodiment, the swirl strength is not included in the equation (2). This is the flow swirling in the circumferential direction with respect to the axis perpendicular to the central axis of the piston 6 in the engine which is the object this time. This is because it is judged that the effect on the turbulent combustion speed is less than that of tumble, and is not considered this time.

  However, swirl strength is not excluded. When the influence of swirl on the turbulent combustion speed is larger than that of tumble in the target engine, the swirl intensity Is may be estimated by an estimation method similar to the estimation method of tumble intensity It described later.

  Next, a method for estimating the tumble intensity It (gas flow in the combustion chamber) of the equation (2) will be described. First, FIGS. 2 (a) to 2 (e) illustrate tumble generation to extinction in the combustion chamber 5 in accordance with the movement of the piston 6. To explain in order, FIG. 2A shows the intake valve opening timing IVO, that is, the moment when the intake valve 15 is opened, and the gas flows into the combustion chamber 5 in various directions. FIG. 2B shows the vicinity of 45 deg (45 degATDC) after the intake top dead center where the piston 6 is descending, and the tumble starts to be formed from the vicinity of this 45 degATDC. FIG. 2 (c) shows the vicinity of the intake valve closing timing IVC, and the tumble strength increases to the vicinity of the intake valve closing timing IVC. Therefore, the later the intake valve closing timing IVC is, the stronger the tumble becomes.

  FIG. 2 (d) shows the vicinity of 315deg after the intake top dead center, that is, during the compression stroke in which the piston 6 ascends, and the tumble is gradually converted into small vortices (disturbances) during the compression process. FIG. 2 (e) shows the vicinity of 315 deg after the intake top dead center, and the tumble disappears near this 315 deg ATDC and all are converted into small vortices.

  On the other hand, FIG. 2 (f) and FIG. 2 (g) show how the tumble formation differs depending on the valve lift amount of the intake valve 15 that accompanies the operation and non-operation of the VEL mechanism 28. Of these, FIG. 2 (f) shown on the left side shows a case where the valve lift amount of the intake valve 15 is small. When the valve lift amount is small, the intake air flows separately on both the cylinder wall side and the cylinder center side, and flows to each other. Tumble is difficult to form. On the other hand, FIG. 2G shown on the right side shows a case where the valve lift amount of the intake valve 15 is large. When the valve lift amount is large, most of the intake air flows toward the center of the cylinder, so that it is easy to form a tumble.

  Now, it is described in SAE paper that the tumble strength It [unnamed number] is calculated as the sum of the angular momentum of the small vortex by the following equation (SAE 981048).

It = Σm _i r _i × v _i (3)
Where _mi : the mass of the gas in the i-th small vortex,
r _i : distance from the center of the cylinder to the i-th small vortex,
v _i : gas flow velocity at the i-th small vortex,
However, since it is currently impossible for the engine controller 31 to calculate these values m _i , r _i , and v _i for all small vortices in the combustion chamber online, the following equation (3) is used instead. The tumble strength It [anonymous number] is approximated by the equation (4).

It = a. (M ^ b.Ne ^ c) (4)
Where m: gas mass in the combustion chamber [g],
Ne: Engine rotation speed [rpm],
a: coefficient,
b, c: fitness coefficient,
In the equation (4), m ^ b reflects the influence of the combustion chamber gas mass m on the tumble strength, and Ne ^ c reflects the influence of the rotational speed Ne on the tumble strength. Here, the gas mass m in the combustion chamber is the mass of gas in the combustion chamber 5 at the intake valve closing timing IVC.

  Further, as shown in FIGS. 2 (a) to 2 (e), the tumble becomes stronger as the intake valve closing timing IVC is later than around 45 degATDC, and in FIGS. 2 (f) and 2 (g). As shown, the larger the valve lift amount of the intake valve 15 is, the stronger it is, and the coefficient a in the equation (4) is newly expressed by the following equation (5).

a = kVEL · (IVC−θ0t) (5)
However, kVEL: tumble strength valve lift amount correction coefficient [1 / deg],
IVC: Intake valve closing timing [degATDC],
θ0t: Tumble formation start angle [degATDC],
Accordingly, the tumble strength It is approximated by the following equation (6) in which the equation (5) is substituted into the equation (4).

It = kVEL · (IVC−θ0t) · m ^ b · Ne ^ c (6)
However, kVEL: tumble strength valve lift amount correction coefficient [1 / deg],
IVC: Intake valve closing timing [degATDC],
θ0t: Tumble formation start angle [degATDC],
b, c: fitness coefficient,
Since the unit of the intake valve closing timing IVC in the equation (6) is a value [degATDC] measured from the intake top dead center to the retard side, the value of IVC increases as the intake valve closing timing IVC is delayed from the intake top dead center. Therefore, the tumble strength It becomes larger than the equation (6). Further, the valve lift amount correction coefficient kVEL in the equation (6) is a function of the valve lift amount. When the valve lift amount is large, the valve lift amount is larger than when the valve lift amount is small. Therefore, the tumble strength It is larger than the equation (6). Become.

  In order to confirm whether or not the new idea that approximates the tumble strength It by the equation (6) is valid, the VEL mechanism 28 is in the non-operating state, that is, the valve lift amount of the intake valve 15 is determined for the engine including the VEL mechanism 28 and the VTC mechanism 27. As a result of an experiment in which only the VTC mechanism 27 was operated and the intake valve closing timing IVC was changed with constant, a result showing the relationship between the tumble strength and the intake valve closing timing IVC was obtained as shown in FIG. FIG. 3 confirms that the tumble strength It increases as the intake valve closing timing IVC increases (that is, IVC lags behind the intake top dead center) under the condition that the valve lift amount of the intake valve 15 is constant.

  The tumble formation start angle θ0t in the equation (6) is a constant value (for example, 45 degATDC). The tumble formation start angle θ0t depends on the engine specifications, and does not depend on the presence or absence of the VEL mechanism 28 and the VTC mechanism 27. From FIG. 2, the intake valve closing timing IVC is always a value on the retard side from the tumble formation start angle θ0t. Therefore, it is considered that the value of (IVC−θ0t) in the above equation (6) is always a positive value.

  The valve lift amount correction coefficient kVEL for the tumble strength in the above equation (6) is determined according to the valve lift amount Lift [m] (or the valve operating angle). Here, for simplicity, for example, the valve lift amount of the intake valve 15 can be switched between two stages of large and small by the operation and non-operation of the VEL mechanism 28, and the VEL mechanism 28 is set under normal operating conditions (operating condition 1). When the operation condition 1 is changed to the predetermined operation condition (operation condition 2), the VEL mechanism 28 is activated, and the broken line valve lift is shown in FIG. If the valve lift amount is reduced to the characteristic, the valve lift amount becomes the first valve lift amount Lift1 [m] when the VEL mechanism 28 is not operated, and the valve lift amount becomes the first valve lift amount when the VEL mechanism 28 is operated. The two-valve lift amount Lift2 [m] is obtained (Lift2 <Lift1). Therefore, as shown in the upper part of FIG. 4, when the VEL mechanism 28 is not operated (or when the operating condition is 1), the first valve lift amount correction coefficient kVEL1 corresponding to the first valve lift amount Lift1 is When the VEL mechanism 28 is activated (or when the operating condition is 2), the second valve lift amount correction coefficient kVEL2 corresponding to the second valve lift amount Lift2 is set as the valve lift amount correction coefficient kVEL. Thus, when the valve lift amount of the intake valve 15 is reduced from the first valve lift amount Lift1 to the second valve lift amount Lift2 by the operation of the VEL mechanism 28, the valve lift amount correction coefficient kVEL is set to the first valve lift amount. The reason why the second valve lift amount correction coefficient kVEL2 is made smaller than the amount correction coefficient kVEL1 is that the tumble becomes weaker (therefore, the tumble strength It becomes smaller) when the valve lift amount is small.

Thus, according to the equation (6), the tumble strength It is newly set as a function of the valve lift amount of the intake valve 15, the intake valve closing timing IVC, the combustion chamber gas mass m, and the engine rotational speed Ne.
<2> Reference Crank Angle Setting Method In the preceding ignition timing control method applied to the conventional engine, the crank angle θpmax at which the combustion pressure is maximized and the combustion mass ratio is 60 as shown in the following (Supplement 4) equation. The crank angle position at which the combustion pressure becomes% is set, and the crank angle θpmax at which the combustion pressure is maximum is used as the reference crank angle. The crank angle position when the combustion mass ratio is x% is hereinafter expressed as “θmb x%”. For example, the crank angle position when the combustion mass ratio is 60% is “θmb 60%”. The starting point of θmb x% is the compression top dead center.

θmb60% = θpmax (Supplement 4)
Further, the crank angle θpmax at which the combustion pressure becomes maximum is given as a function of the engine speed Ne as in the following (Supplement 5).

θpmax = h · (Ne) ^ i (Supplement 5)
Where h and i are conformity factors,
On the other hand, when an experiment was conducted on an engine including the VEL mechanism 28 and the VTC mechanism 27, as shown in FIG. 5, the crank angle θpmax at which the combustion pressure becomes maximum under the condition where the rotational speed Ne is constant. And an experimental result representing the relationship between θmb x%. However, in FIG. 5, θmb is 0%, which is the crank angle position when the combustion mass ratio is 0% as θmb x%, θmb is 10%, which is the crank angle position when the combustion mass ratio is 10%, and the combustion mass ratio is Only three cases of θmb 60%, which is the crank angle position at 60%, are shown. In FIG. 5, the left is θpmax and θmb0%, the center is θpmax and θmb10%, and the right is θpmax and θmb60%. It is a sort of relationship. Actually, experimental results at the crank angle position when the combustion mass ratio is a value other than 0%, 10%, and 60% (20%, 30%, 40%, and 50%) are also obtained (see FIG. 7). ).

  The crank angle θpmax at which the combustion pressure is maximum and the crank angle position θmb 60% when the combustion mass ratio is 60% have a maximum error of 2.6 deg as shown on the right side of FIG. It can be seen that this is the same as the left side of FIG. 5 showing the error between θmb and 0%.

  Next, FIG. 6 shows the relationship between the crank angle θpmax at which the combustion pressure becomes maximum and the engine rotational speed Ne under the condition that the intake air amount (engine load) is constant. As shown in the figure, as the rotation speed Ne increases, θpmax shifts to the advance side, but the variation of θpmax with respect to the rotation speed Ne is large, and the crank angle θpmax at which the combustion pressure becomes maximum is the above (Supplement 5). As can be seen from the equation, the engine rotation speed Ne alone cannot be expressed.

  7, the horizontal axis represents the combustion mass ratio, and the vertical axis represents the square value of R when θmb x%, which is the crank angle position when the combustion mass ratio becomes x%, is linearly approximated by θpmax. The three experimental results at θmb 0%, θmb 10%, and θmb 60% shown in FIG. 6 and the remaining experimental results at θmb 20%, θmb 30%, θmb 40%, and θmb 50%, which are not shown in FIG. It is a thing. The square value of R on the vertical axis indicates that there is a correlation between θpmax and θmb x% (x = 0, 10, 20, 30, 40, 50, 60) as this value approaches 1.0. Therefore, according to FIG. 7, it can be seen that the crank angle θpmax at which the combustion pressure is maximum is most correlated in the vicinity of θmb 10%.

  FIG. 7 shows the characteristics of two types of engines having different engine specifications even in an engine including the VEL mechanism 28 and the VTC mechanism 27. The difference between the two engines is mainly in the shape of the combustion chamber, and the second engine has a shape in which the combustion chamber is flatter than the first engine. In the first engine, the crank angle θpmax at which the combustion pressure is maximum has a wide correlation with the crank angle position at which the combustion mass ratio is 10% from the crank angle position at which the combustion mass ratio is 10%. In the second engine, the crank angle θpmax at which the combustion pressure becomes maximum has a correlation only in the vicinity of the crank angle position when the combustion mass ratio becomes 10%. Accordingly, the crank angle θpmax at which the combustion pressure is maximized has a high correlation common to the two engines only when the crank angle position is near the combustion mass ratio of 10%. As described above, even when the engine models are different, the result that the crank angle θpmax at which the combustion pressure becomes maximum is most correlated in the vicinity of the crank angle position when θmb is 10% is obtained for the first time.

  In response to such experimental results, in the present embodiment, based on the crank angle θpmax at which the combustion pressure is maximum, θmb10%, which is the crank angle position when the combustion mass ratio becomes 10%, is calculated, and this θmb10% is used as a reference. The crank angle is set, that is, θmb10% [degATDC] is calculated by the following equation (7).

θmb10% = a · θpmax−b (7)
Where a: fitness coefficient [anonymous number],
b: Applicable crank angle [deg],
Equation (7) is an equation for obtaining θmb10% by shifting the crank angle θpmax at which the combustion pressure is maximum to the advance side by a crank angle of b (positive value). As a result, as shown in FIG. 8, the reference crank angle (= θmb 10%) in the present embodiment is more advanced than the reference crank angle (= θmb 60%) in the preceding ignition timing control method applied to the conventional engine. It will be difficult.

  In the engine including the VEL mechanism 28 and the VTC mechanism 27, the crank angle θpmax at which the combustion pressure becomes maximum is greater than the compression top dead center as the intake valve closing timing IVC is later and the valve lift amount of the intake valve 15 is larger. It is considered to be delayed, and is determined to be influenced by the gas mass (load) in the combustion chamber. Further, it is determined from FTA (Fault Tree Analysis) that the gas flow velocity (tumble, swirl) in the combustion chamber has an influence, and the following ( The crank angle θpmax [degATDC] at which the combustion pressure is maximized is calculated from the equation (8). Therefore, if the equation (8) is compared with the above (complement 5), k1VEL · (IVC−θt0) · m ^ d is newly introduced.

θpmax = k1VEL · (IVC−θt0) · m ^ d · Ne ^ e
(8)
However, k1VEL: flow velocity coefficient [anonymous number],
IVC: Intake valve closing timing [degATDC],
θt0: tumble formation start angle [degATDC],
m: combustion chamber gas mass [g],
Ne: Engine rotation speed [rpm],
d, e: conformance constant [anonymous number],
In order to confirm whether or not such a new concept for the crank angle θpmax at which the combustion pressure at which the combustion pressure is maximized shown in the equation (8) is maximum is valid, the valve lift amount of the intake valve 15 and the intake valve closing timing IVC are constant, and the combustion pressure is maximized When the relationship between the crank angle θpmax, the intake air amount (representative value of the gas mass m in the combustion chamber), and the engine speed Ne (representative value of the gas flow rate) was tested, the results shown in FIG. According to FIG. 9, when the valve lift amount of the intake valve 15 and the intake valve closing timing IVC are constant and the intake air amount is the same, the engine speed Ne is larger than the engine speed Ne is smaller. The crank angle θpmax at which the combustion pressure is maximized is increased (the amount of retardation from the compression top dead center is increased), and the combustion pressure is maximized as the intake air amount is increased if the engine rotational speed Ne is the same. It was confirmed that the crank angle θpmax is increased (the amount of retardation from the compression top dead center is increased).

  Note that the unit [degATDC] on the vertical axis in FIG. 9 is a crank angle starting from the compression top dead center, and does not start from the intake top dead center.

  The tumble formation start angle θ0t in the equation (8) is the same as that described above in the equation (6). That is, the tumble formation start angle θ0t is a constant value (for example, 45 degATDC). The tumble formation start angle θ0t depends on the engine specifications and does not depend on the presence or absence of the VEL mechanism 28 and the VTC mechanism 27. From FIG. 2, the intake valve closing timing IVC is always a value on the retard side from the tumble formation start angle θ0t. Therefore, the value of (IVC−θ0t) in the equation (8) is always a positive value.

  The flow velocity coefficient k1VEL in the above equation (8) is determined according to the valve lift amount Lift (or valve operating angle) of the intake valve 15. Here, for the sake of simplicity, as described above, the valve lift amount of the intake valve 15 can be switched between large and small by the operation and non-operation of the VEL mechanism 28. Under the operating condition 1, the VEL mechanism 28 is deactivated. When the operation condition 1 shifts to the operation condition 2, the VEL mechanism 28 is operated, and it is assumed that the valve lift amount is reduced from the solid line valve lift characteristic to the broken line valve lift characteristic as shown in the lower part of FIG. As shown in the upper part of FIG. 10, when the VEL mechanism 28 is not in operation (or when the operating condition is 1), the first flow rate coefficient k1VEL1 corresponding to the first valve lift amount Lift1 is set to the VEL mechanism 28. Is set to the second flow velocity coefficient k1VEL2 corresponding to the second valve lift amount Lift2 as the flow velocity coefficient k1VEL. To. As described above, when the valve lift amount is reduced, the flow rate coefficient k1VEL is decreased from the first flow rate coefficient k1VEL1 to the second flow rate coefficient k1VEL2, so that the tumble is weaker when the valve lift amount is small (therefore, This is because θpmax moves more to the retard side.

  In the above equation (8), m ^ d reflects the influence of the in-cylinder gas mass m on θpmax, and Ne ^ e reflects the influence of the rotational speed Ne on θpmax.

Thus, according to the equation (8), the crank angle θpmax at which the combustion pressure becomes maximum is added to the engine speed Ne, the gas mass m in the combustion chamber, the valve lift amount of the intake valve 15 and the intake valve closing timing IVC. It is newly constructed as a function of.
<3> Calculation Method of Ignition Delay Time In the preceding ignition timing control method applied to the conventional engine, an ignition is performed in the interval between the crank angle position θmb 0% when the combustion mass ratio becomes 0% starting from the ignition timing. It is calculated as a delay time DEADTIME.

  However, θmb 0%, which is the crank angle position when the combustion mass ratio becomes 0%, is easily affected by the calculation error of the cooling loss and gas leakage, and is difficult to measure accurately. Therefore, if the value of θmb0% is set to be positive and the ignition delay time DEADTIME is adapted, a measurement error of the ignition delay time is included in the estimation error of the combustion speed. For these reasons, it is rare that the combustion simulation is generally performed with θmb0% being positive.

  On the other hand, in the present embodiment, the time from the ignition timing to θmb2% which is the crank angle position when the combustion mass ratio becomes 2% is regarded as the ignition delay time τ. Therefore, in this embodiment, the crank angle section from θmb2% to θmb10% is the combustion period, which corresponds to the main combustion period BURN2 in the preceding ignition timing control method applied to the conventional engine. In the present embodiment, the combustion period is estimated only for the main combustion period.

  The ignition delay time τ is not a pure ignition delay time but a period from the ignition timing until the fuel burns by 2%. Therefore, since the Arrhenius equation (self-ignition equation) proposed in the preceding ignition timing control method applied to the conventional engine cannot be used, the ignition delay time τ [sec] is newly set in this embodiment. It is calculated by the following equation (9). This equation (9) is calculated using an equation for calculating the combustion speed immediately after ignition.

τ = (Dkernel−D0)
/ 2 [(Tad / T) · SL + {(2/3) · k} ^ (1/2)]
... (9)
However, Dkernel: Flame diameter [m] when θmb 2%,
D0: Flame diameter [m] immediately after the ignition timing,
Tad: Flame temperature [K],
T: unburned gas temperature [K],
SL: Laminar burning velocity [m / sec],
k: Kinetic energy of combustion chamber gas,
Since the flame diameter Dkernel in equation (9) is the flame diameter when the combustion mass ratio becomes 2%, it is newly calculated by the following equation (10).

Dkernel = {(Vcyl · [1− (1−xb) / (3 · xb + 1) ^ (1 / κ)]
・ 6) / π} ^ (1/3) (10)
Vcyl: combustion chamber volume [m ³ ],
xb: combustion mass ratio (= 0.02),
κ: polytropic index (= 1.34),
Equation (10) is derived as follows. Assuming that the flame shape is a sphere, the equation giving the flame diameter Dkernel is the following equation (11a).

Dkernel = {(Vb · 6) / π} ^ (1/3) (11a)
Where Vb: flame volume,
From the thermodynamic formula, the equation of volumetric combustion ratio is expressed by the following equation (11b) (known).

Volume combustion ratio = 1- (1-xb) / (3 · xb + 1) ^ (1 / κ)
... (11b)
This equation (11b)
Vb = Vcyl · volume combustion ratio (11c)
The following equation (11d) is obtained by using

Vb = Vcyl · [1- (1-xb) / (3 · xb + 1) ^ (1 / κ)]
... (11d)
Substituting this equation (11d) into equation (11a) yields the above equation (10).

  The diameter D0 of the flame immediately after the ignition timing of the above equation (9) is 1 mm. Tad / T, which is the ratio between the flame temperature Tad and the unburned gas temperature T, is a conforming constant. The calculation method of the laminar combustion speed SL is the same as the preceding ignition timing control method applied to the conventional engine.

  The kinetic energy k of the combustion chamber gas in the above equation (9) is calculated by the following equation.

k = (1/2) m (f · Ne) ^ 2 (12)
Where m: gas mass in the combustion chamber [g],
Ne: Engine rotation speed [rpm],
f: conformity constant,
This completes the itemization explanation.

  Next, a new ignition timing control method after changing the above <1> to <3> (this new ignition timing control method is hereinafter referred to as “changed ignition timing control method”) is summarized below. However, the coefficient of conformity has been reviewed again.

  The ignition timing at which MBT is obtained (this ignition timing is referred to as “basic ignition timing”) MBTCAL [degBTDC] is calculated by the following equation.

MBTCAL = − | (τ + BT) · Ne · 6-θmb10% | ... (13)
Where BT: combustion time [sec]
τ: ignition delay time [sec],
θmb 10%: Reference crank angle [degATDC],
This is illustrated in FIG. In the post-change ignition timing control method, the ignition delay time τ [sec], which is the time from the ignition timing to θmb2%, and the combustion time BT [sec], which is the time from θmb2% to θmb10%, are added. The value [sec] is multiplied by Ne [rpm] · 6 to convert to the crank angle section [deg]. This converted crank angle section ((τ + BT) · Ne · 6) is a crank angle section from the ignition timing to θmb 10% (= θpmax). Therefore, the value on the advance side for the converted crank angle section is calculated as the basic ignition timing MBTCAL from θmb10% [degATDC] which is the reference crank angle (= θpmax).

  Note that in equation (13), the absolute value is taken and a minus sign is added, whereas the unit of θmb 10% is the crank angle [degATDC] measured on the retard side from the compression top dead center, whereas MBTCAL This is because the unit of θmb 10% needs to be converted into the unit of MBTCAL because the unit of is the crank angle [degBTDC] measured from the compression top dead center to the advance side.

  Here, the reference crank angle θmb10% [degATDC] is calculated by the following equation (14).

θmb10% = k1VEL · (IVC−θt0) · m ^ c1 · Ne ^ c2 + c3
... (14)
However, k1VEL: flow velocity coefficient [anonymous number],
IVC: Intake valve closing timing [degATDC],
θt0: tumble formation start angle [degATDC],
m: combustion chamber gas mass [g],
Ne: Engine rotation speed [rpm],
c1 to c3: conformity factors,
As described above, the unit of the intake valve closing timing IVC and the tumble formation start angle θt0 starts from the intake top dead center, while θmb 10% starts from the compression top dead center. Therefore, the conversion from the unit starting from the intake top dead center to the unit starting from the compression top dead center is performed by the adaptation coefficient c3. For example, in a four-cylinder engine, there is a shift of 180 deg between the intake top dead center and the compression top dead center.

  The combustion time BT [sec] in equation (13) is calculated by the following equation.

BT = 0.08 · Vcyl / {((3 · xb + 1) ^ (1 / κ)) · Ab · Sb}
= 0.0707 · Vcyl / (Ab · Sb) (15)
Vcyl: combustion chamber volume [m ³ ],
xb: average combustion mass ratio (= 0.06) at the time of burning time calculation,
κ: polytropic index,
Ab: Flame surface area [m ² ],
Sb: burning rate [m / sec],
Equation (15) is basically the same as the equation in the preceding ignition timing control method applied to the conventional engine. That is, in the equation (15), Vcyl / {(3 · xb + 1) ^ (1 / κ)} is a mass ratio of the burned gas, the combustion time BT is proportional to the mass ratio of the burned gas, and the combustion speed Sb It is an expression that is inversely proportional to. The flame surface area Ab in the equation (15) can be calculated from engine specifications.

  As the combustion mass ratio xb in the equation (15), it is necessary to use a value in a crank angle section (combustion period) from θmb 2% to θmb 10%. In this case, the combustion mass ratio changes from 2% to 10% in the crank angle section from θmb 2% to θmb 10%. Therefore, here, combustion at θmb 6%, which is an average value of θmb 2% and θmb 10% The mass ratio value, ie 6% is used.

  The combustion speed Sb [m / sec] in the above equation (15) is calculated by the following equation.

Sb = SL + [{kVEL · (IVC−θt0) · m ^ c4 · Ne ^ c5}
/ SL] ^ c6 (16)
However, SL: Laminar burning velocity [m / sec],
kVEL: valve lift amount correction coefficient [1 / deg] of tumble strength,
IVC: Intake valve closing timing [degATDC],
θ0t: Tumble formation start angle [degATDC],
c4, c5: fitness coefficient,
c6: conformity factor,
The second term on the right side of Equation (16) is an equation that gives the turbulent combustion speed [m / sec], and the point that approximates the turbulent combustion rate with the second term on the right side of Equation (16) is new. That is, in this embodiment, the turbulence strength u [m / sec] of the combustion chamber gas is calculated by the following equation (16-1), and the following (16− 2) The turbulent combustion speed ST [m / sec] is calculated by the equation (2).

u = kVEL · (IVC−θt0) · m ^ c4 · Nec5 (16-1)
ST = [u / SL] ^ c6 (16-2)
On the other hand, the laminar combustion speed SL [m / sec] in the equation (16) may be the same as the preceding ignition timing control method applied to the conventional engine, and is therefore calculated by the following equation (17). .

SL = SL0 · (T / 298) ^ a · (p / 101.325) ^ b
(1-2.1 × MRESR ^ c) (17)
However, SL0: Laminar burning velocity in standard state [m / sec],
T: unburned gas temperature [K],
p: pressure in the combustion chamber [kPa],
MRESR: Internal inert gas ratio [%]
a, b, c: coefficients,
On the other hand, the ignition delay time τ [sec] in the equation (13) is newly calculated by the following equation (18).

τ = [{{Vcyl · [1- (1-xb) / (3 · xb + 1) ^ (1 / κ)] · 6} / π} ^ (1/3) −0.001]
/ {C7 / SL + c8 · Ne · m ^ (1/2)}
= {0.490 · Vcyl ^ (1/3)}
/ {C7 · SL + c8 · Ne · m ^ (1/2)} (18)
However, xb: combustion mass ratio (= 0.02) at the time of calculating ignition delay time,
κ: polytropic index (= 1.34),
SL: Laminar burning velocity [m / sec],
m: combustion chamber gas mass [g],
Ne: Engine rotation speed [rpm],
Vcyl: combustion chamber volume [m ³ ],
c7, c8: conformity constants,
Note that the value in the crank angle section (ignition delay period) from θmb0% to θmb2% must be used as the combustion mass ratio xb in the equation (18). In this case, the combustion mass ratio changes from 0% to 2% in the crank angle section from θmb 0% to θmb 2%, but here, θmb 0% and θmb 2% are average values of θmb 1%. Instead of using the value of the combustion mass ratio, the value of the combustion mass ratio at the time of θmb 2%, which is the end of the ignition delay period, that is, 2% is used.

  The laminar burning velocity SL [m / sec] in the above equation (18) is calculated by the following equation (19).

SL = SL0 · (T / 298) ^ a · (p / 101.325) ^ b
(1-2.1 × MRESR ^ c) (19)
However, SL0: Laminar burning velocity in standard state [m / sec],
T: unburned gas temperature [K],
p: pressure in the combustion chamber [kPa],
MRESR: Internal inert gas ratio [%]
a, b, c: coefficients,
Here, the laminar combustion speed SL of the equation (19) used to calculate the ignition delay time τ and the laminar combustion speed SL of the above equation (17) used to calculate the combustion time BT are the equations themselves. The values of the laminar combustion speed SL0 and the internal inert gas ratio MRESR in the standard state do not change, but the values to be substituted as the unburned gas temperature T and the combustion chamber pressure p are the expressions (19) and (17 ) And the expression will be different as will be described later.

  The laminar combustion speed SL0 [m / sec] and the coefficients a, b, and c [anonymous number] in the standard state of the above equations (17) and (19) are calculated by the following equations (20) to (23). To do.

SL0 = (0.2632-0.8472 / (φ−1.13) ^ 2)
... (20)
a = 2.18−0.80 · (φ−1) (21)
b = −0.16 + 0.22 · (φ−1) (22)
c = 1 (23)
Where φ: equivalent ratio [anonymous number]
Here, the equations (20) to (23) are known from SAE paper (see SAE 199910175).

  This completes the description of the post-change ignition timing control method.

  Next, a supplementary explanation will be given on how to obtain the volume combustion ratio shown in the above equation (11b) and the method for calculating the unburned gas temperature shown in the above equations (17) and (19).

  First, how to obtain the volume combustion rate shown in the above equation (11b) will be briefly described.

Assuming that the mass combustion ratio of burned gas is x and the volume combustion ratio is y, combustion in the combustion chamber is similar to constant volume combustion,
y = f (x)
= 1 + (x-1) / (1 + x (k1-1) ^ (1 / κ))
... (24)
However, k1≈4-5,
Assuming that the above equation can be written, the above equation (11b) can be easily obtained by substituting 4 for k1 in the equation (24).

  The calculation method of the unburned gas temperature T shown in the above equations (17) and (19) will be described together. Assuming a change in adiabaticity, θmb of the combustion chamber 5 is calculated using the following equation (25) of thermodynamics. The unburned gas temperature Tmb x% at x% is calculated.

Tmb x% = Tivc · (Pmb x% / Pivc) ^ {(κ−1) / κ} (25)
However, Tmb x%: unburned gas temperature [K] at θmb x%,
Tivc: Combustion chamber temperature [K] at IVC,
Pmb x%: pressure in combustion chamber [kPa] at θmb x%,
Pivc: pressure in the combustion chamber [kPa] at IVC,
κ: specific heat ratio (1.3 to 1.4 at a fixed value),
Here, the calculation method of the unburned gas pressure Pmb x% at θmb x% in the combustion chamber 5 of the equation (25) will be described. The unburned gas pressure Pmb x% at θmb x% in the combustion chamber 5 is (26).

Pmb x% = Pivc + Δp (26)
Where Δp is the amount of change in the unburned gas pressure in the combustion chamber,
Here, the amount of change Δp of the unburned gas pressure in the combustion chamber is calculated by dividing into a pressure change Δpv due to volume change and a pressure change Δpc due to combustion. That is, the change amount Δp of the unburned gas pressure in the combustion chamber is calculated by the following equation (27).

Δp = Δpv + Δpc (27)
Among these, the pressure change Δpv due to the volume change in the equation (27) can be calculated by the following equation (28) which is a thermodynamic equation.

Δpv = Pivc · {(Vivc / Vmb x%) ^ κ−1} (28)
Where Vivc: combustion chamber volume at IVC,
Vmb x%: Volume in combustion chamber at θmb x%,
Here, the volume Vivc of the combustion chamber 5 when the intake valve is closed can be obtained from the position of the piston 6. With respect to the unburned gas volume Vmb x% at the time of θmb x% in the combustion chamber 5, as shown in FIG. 12, a characteristic representing the relationship between θmb x% and Vmb x% is created in advance, and from θmb x% Using this characteristic, the unburned gas volume Vmb x% of the combustion chamber 5 at θmb x% may be obtained. FIG. 12 is a characteristic of the combustion chamber volume with respect to the crank angle.

  On the other hand, the pressure change Δpc due to combustion in the above equation (27) may be simply calculated by the following equation (29).

Δpc = (x / 100) × Δptotal (29)
Where Δptotal: total pressure increase due to combustion,
Here, for the total pressure increase Δptotal due to combustion, as shown in FIG. 13, a characteristic representing the relationship between the fuel injection pulse width Ti [msec] (or the fuel injection amount) and Δptotal is prepared in advance. A total pressure increase Δptotal due to combustion is obtained from the fuel injection pulse width Ti (or fuel injection amount) using this characteristic.

  This completes the supplementary explanation.

  Next, the calculation method of the basic ignition timing MBTCAL in the post-change ignition timing control method executed by the engine controller 31 will be described in detail with reference to the flowcharts of FIG. 14, FIG. 16, FIG. 17, FIG.

  FIG. 14 is for calculating various physical quantities necessary for calculating the basic ignition timing MBTCAL, and is executed at regular time intervals (for example, every 10 msec).

  First, in step 1, the intake valve closing timing IVC [degATDC], the valve lift amount Lift [m], the collector internal temperature TCOL [K] detected by the temperature sensor 43, and the collector internal pressure PCOL [Pa] detected by the pressure sensor 44. ], Exhaust temperature TEXH [K] detected by temperature sensor 45, internal inert gas rate MRESR [%], fuel injection pulse width Ti [msec], engine rotation speed detected by crank angle sensors (33, 34) Read Ne [rpm] and total gas mass MGAS [g].

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

  The valve lift amount Lift of the intake valve 15 is known from the command value given to the VEL mechanism 28. That is, as shown in the lower part of FIG. 4 and the lower part of FIG. 10, the first valve lift amount Lift1 is obtained when the VEL mechanism 28 is not operated, whereas the second valve lift amount Lift2 is the valve lift amount when the VEL mechanism 28 is operated. Lift.

  The fuel injection pulse width Ti is calculated in a fuel injection pulse width calculation routine (not shown). For example, the calculation formula of the fuel injection pulse width Ti [msec] given to the fuel injector 21 at the time of sequential injection is as follows.

Ti = Tp × Tfbya × (α + αm−1) × 2 + Ts (30)
Where Tp: basic injection pulse width [msec],
Tfbya: [anonymous number],
α: Air-fuel ratio feedback correction coefficient [anonymous number]
αm: Air-fuel ratio learning value [anonymous number]
Ts: invalid pulse width [msec],
The target equivalent ratio Tfbya in the equation (30) is an unknown number, and is a value represented by the following equation (31) when the theoretical air-fuel ratio is 14.7.

Tfbya = 14.7 / target air-fuel ratio (31)
For example, from equation (31), when the target air-fuel ratio is the stoichiometric air-fuel ratio (14.7), Tfbya = 1.0, and when the target air-fuel ratio is a lean side value such as 22.0, Tfbya is less than 1.0. Is a positive value.

  The crank angle sensor includes a position sensor 33 that detects the position of the crankshaft 7 and a phase sensor (= cam angle sensor) 34 that detects the position of the intake camshaft 25. The signals from these two sensors 33 and 34 are used as signals. Based on this, the engine rotation speed Ne [rpm] is calculated.

  The internal inert gas ratio MRESFR is a value obtained by dividing the amount of inert gas remaining in the combustion chamber 5 by the total gas amount in the combustion chamber 5, and the calculation will be described later together with the total gas mass MGAS with reference to FIG.

In step 2, the volume (compression start volume) Vivc [m ³ ] of the combustion chamber 5 at the intake valve closing timing IVC 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. 15, consider the 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 (41) to (45).

Vivc = f1 (θivc) = Vc + (π / 4) D ² · Hivc
... (41)
Vc = (π / 4) D ² · Hx / (ε−1) (42)
^{Hivc = {(CND + ST 2} /2) - (CRoff-PISoff) 2} 1/2
− {(ST / 2) · cos (θivc + θoff)}
+ (CND ² −X ² ) ^1/2 (43)
X = (ST / 2) · sin (θivc + θoff) −CRoff + PISoff
... (44)
θoff = arcsin {(CRoff−PISoff) / (CND · (ST / 2))}
... (45)
Where Vc: gap volume [m ³ ],
ε: compression ratio,
D: cylinder bore diameter [m],
ST: Full piston stroke [m],
Hivc: Piston pin 76 at the intake valve closing timing
Distance from TDC [m],
Hx: difference between the maximum value and the minimum value of the distance from the TDC of the piston pin 76 [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: an angle [deg] between a line connecting the piston pin 76 and the crankshaft rotation center 72 and a vertical line in TDC,
X: horizontal distance [m] between the nodal point 75 and the piston pin 76,
As described above, the intake valve closing timing crank angle θivc is known because it is determined by the command signal from the engine controller 31 to the VTC mechanism 27. If the crank angle θivc (= IVC) at this time is substituted into the equations (41) to (45), the volume Vivc of the combustion chamber 5 at the intake valve closing timing can be calculated. Therefore, in practice, the volume Vivc of the combustion chamber 5 at the closing timing of the intake valve is set by a table having the intake valve closing timing IVC as a parameter.

  In step 3, the temperature (compression start timing temperature) Tivc [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. Tivc follows from the fresh air temperature TCOL in the intake collector 2, the exhaust gas temperature TEXH, and the internal inert gas ratio MRESR 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. (46).

Tivc = TEXH × MRESR + TCOL × (1−MRESR)
... (46)
In step 4, the pressure (compression start timing pressure) Pivc [kPa] 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 Pivc at the intake valve closing timing IVC.

  In step 5, a pressure increase Δptotal [Pa] due to total combustion is calculated by searching a table having the contents shown in FIG. 13 from the fuel injection pulse width Ti (or fuel injection amount).

  In step 6, the valve lift amount correction coefficient kVEL [1 / deg] of the tumble strength is calculated by searching a table having the upper part of FIG. 4 from the valve lift amount Lift. That is, when the VEL mechanism 28 is not operated, the first valve lift amount correction coefficient kVEL1 is calculated, and when the VEL mechanism 28 is operated, the second valve lift amount correction coefficient kVEL2 is calculated as the valve lift amount correction coefficient kVEL.

  In step 7, the flow rate coefficient k1VEL [nameless number] is calculated by searching a table having the upper part of FIG. 10 from the valve lift amount Lift. That is, the first flow velocity coefficient k1VEL1 is calculated as the flow velocity coefficient k1VEL when the VEL mechanism 28 is not operated, and the second flow velocity coefficient k1VEL2 is calculated when the VEL mechanism 28 is operated.

  In step 8, the crank angle position when the combustion mass ratio becomes 10% according to the following equation (47) using the flow velocity coefficient k1VEL, the intake valve closing timing IVC, the total gas mass MGAS, and the engine rotational speed Ne. Θmb10% [degATDC] is calculated.

θmb10% = k1VEL · (IVC−θt0) · MGAS ^ c1 · Ne ^ c2
+ C3 (47)
However, k1VEL: flow velocity coefficient [anonymous number],
IVC: Intake valve closing timing [degATDC],
θt0: tumble formation start angle [degATDC],
MGAS: total gas mass [g],
Ne: Engine rotation speed [rpm],
c1 to c3: conformity factors,
The equation (47) uses the total gas mass MGAS as the combustion chamber gas mass m in the equation (14), and is basically the same as the equation (14).

  Here, the crank angle θpmax at which the combustion pressure is maximized is given by the first term on the right side of the equation (47), that is, the following equation (47-1).

θpmax = k1VEL · (IVC−θt0) · MGAS ^ c1 · Ne ^ c2
(47-1)
In step 9, a value obtained by simply subtracting 4 deg in crank angle from θmb 10% which is the crank angle position when the combustion mass ratio becomes 10% is θmb 6% which is the crank angle position when the combustion mass ratio becomes 6% [ degATDC]. This is based on the assumption that the combustion mass ratio changes linearly in the crank angle section from θmb 2% to θmb 10%, and θmb 6% is obtained.

  In step 10, the volume Vmb6% at the time of θmb6% of the combustion chamber 5 is calculated by searching a table having the contents shown in FIG.

  Steps 11 and 12 are the same as steps 9 and 10. In Step 11, a value obtained by simply subtracting 8 deg of the crank angle from θmb10% which is the crank angle position when the combustion mass ratio becomes 10% is θmb2% which is the crank angle position when the combustion mass ratio becomes 2% [ degATDC]. In step 12, the volume Vmb2% at the time of θmb2% of the combustion chamber 5 is calculated by searching a table having the contents of FIG.12 from this θmb2%.

  FIG. 16 is for calculating the internal inert gas ratio MRESR [%] 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 21, the output of the air flow meter 32 and the target equivalent ratio Tfbya are read. In step 22, based on the output of the air flow meter 32, a fresh air amount (cylinder fresh air amount) MACYL [g] 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 23, an internal inert gas amount MRES [g] in the combustion chamber 5 is calculated. A known method may be used for calculating the internal inert gas amount MRES (see Japanese Patent Application Laid-Open No. 2005-171856).

  In step 24, the total gas mass MGAS [g] in the combustion chamber 5 is calculated from the internal inert gas amount MRES, the cylinder fresh air amount MACYL, and the target equivalent ratio Tfbya by the following equation (48).

MGAS = MACYL × (1 + Tfbya / 14.7) + MRES
... (48)
“1” in parentheses on the right side of the equation (48) is a fresh feeling, and “Tfbya / 14.7” is a fuel.

  In step 25, the total gas mass MGAS and the internal inert gas amount MRES are used to calculate the internal inert gas ratio MRESR (the ratio of the internal inert gas amount to the total gas mass in the combustion chamber 5) according to the following equation (49). ) Calculate [%].

MRESR = MRES / MGAS (49)
FIG. 17 is for calculating the combustion time BT [sec], and FIG. 18 is for calculating the ignition delay time τ [sec], which are executed at regular intervals (for example, every 10 msec). 17 and 18 are executed following FIG. Either of FIG. 17 and FIG. 18 may be executed first.

Here, as explained earlier from FIG. 17, in step 31, the intake valve closing timing IVC [degATDC], the volume Vivc [m ³ ] in the intake valve closing timing calculated in step 2 of FIG. 14, the temperature Tivc [K] at the intake valve closing timing of the combustion chamber 5 calculated in step 3 of FIG. 14, the pressure Pivc [Pa] at the intake valve closing timing of the combustion chamber 5 calculated in step 4 of FIG. 14, the pressure increase Δpctotal [kPa] due to the total combustion calculated in step 5 of FIG. 14, the valve lift amount correction coefficient kVEL [1 / deg] of the tumble intensity calculated in step 6 of FIG. 14, and the step of FIG. The volume Vmb 6% [degATDC] at θmb 6% of the combustion chamber 5 calculated in 10 is calculated in step 24 in FIG. Total gas mass MGAS [g], internal inert gas ratio MRESR [%] calculated in step 25 of FIG. 16, fuel injection pulse width Ti [msec], target equivalent ratio Tfbya [anonymous number], engine speed Ne Read [rpm].

  In step 32, from the pressure Pivc at the intake valve closing timing of the combustion chamber 5, the volume Vivc at the intake valve closing timing of the combustion chamber 5, and the volume Vmb 6% at θmb of 6% of the combustion chamber 5, the volume is calculated by the following equation (50). A pressure change Δpv [kP] due to the change is calculated.

Δpv = Pivc · {(Vivc / Vmb6%) ^ κ−1} (50)
Where κ: specific heat ratio (fixed value 1.3-1.4),
Equation (50) is the same as equation (28) above. Since the pressure increase Δpv due to the volume change here is the pressure increase in the crank angle section (combustion period) from θmb2% to θmb10%, θmb6% which is an average value of θmb2% and θmb10% is used.

  In step 33, the pressure increase Δpc [kP] due to combustion is calculated from the total pressure increase Δpctotal due to combustion by the following equation (51).

Δpc = 0.06 × Δptotal (51)
The formula (51) uses 6% as the combustion mass ratio x in the above formula (29).

  In step 34, the pressure increase Δpc due to combustion and the pressure increase Δpv due to volume change are added to the pressure Pivc at the intake valve closing timing of the combustion chamber 5, that is, θmb 6% of the combustion chamber 5 by the following equation (52). The unburned gas pressure Pmb 6% [kP] at the time is calculated.

Pmb6% = Pivc + Δpv + Δpc (52)
In step 35, using the unburned gas pressure Pmb 6% when θmb of the combustion chamber 5 is 6%, the pressure Pivc when the combustion chamber 5 is closed, and the temperature Tivc when the combustion chamber 5 is closed, the following ( 53) The unburned gas temperature Tmb 6% [K] at θmb 6% in the combustion chamber 5 is calculated by the equation (53).

Tmb6% = Tivc · (Pmb6% / Pivc) ^ {(κ−1) / κ} (53)
Where κ: specific heat ratio (fixed value 1.3-1.4),
Equation (53) is a value when the combustion mass ratio x is 6% in the above equation (25).

  In step 36, the laminar combustion velocity SL0 [m / sec] in the standard state and coefficients a and b [anonymous number] are obtained from the target equivalent ratio Tfbya by the following equations (54) to (56). , Laminar combustion velocity SL0 in these standard states, coefficients a and b, unburned gas temperature Tmb 6% when combustion chamber 5 is θmb 6%, unburned gas pressure Pmb 6% when combustion chamber 5 is θmb 6%, internal Using the active gas ratio MRESR, the laminar combustion speed SL [m / s] at θmb of 6% is calculated by the following equation (57).

SL0 = (0.2632-0.8472 / (Tfbya-1.13) ^ 2)
... (54)
a = 2.18-0.80 · (Tfbya-1) (55)
b = −0.16 + 0.22 · (Tfbya−1) (56)
SL = SL0 · (Tmb6% / 298) ^ a · (Pmb6% / 101.325) ^ b
(1-2.1 × MRESR) (57)
The formulas (54) to (56) are obtained by using the target equivalent ratio Tfbya as the equivalent ratio φ in the above formulas (20) to (22). Basically, the formulas (20) to (22) does not change. Similarly, equation (57) uses Tmb 6% and Pmb 6% as unburned gas temperature T and combustion chamber pressure p in equation (17), and is basically the same as equation (17).

  In step 38, using the laminar combustion speed SL at the time of θmb 6%, the valve lift amount correction coefficient kVEL for the tumble strength, the intake valve closing timing IVC, the total gas mass MGAS, and the engine speed Ne, To calculate the combustion speed Sb [m / s] when θmb is 6%, which is the crank angle position when the combustion mass ratio becomes 6%.

Sb = SL + [{kVEL · (IVC−θt0) · MGAS ^ c4 · N ^ c5}
/ SL] ^ c6 (58)
Where θt0: tumble formation start angle [degATDC],
c4, c5, c6: fitness coefficients,
The equation (58) uses the total gas mass MGAS as the combustion chamber gas mass m in the equation (16), and is basically the same as the equation (16).

  Here, the equation (58) is obtained by combining the following four equations in detail.

It = kVEL · (IVC−θt0) · MGAS ^ c4 · Ne ^ (c5-1)
... (58-1)
u = It · Ne (58-2)
ST = [u / SL] ^ c6 (58-3)
Sb = SL + ST (58-4)
However, It: Tumble strength [anonymous number],
u: Turbulence intensity of combustion chamber gas [m / sec],
ST: Turbulent combustion speed [m / sec],
As the intake valve closing timing IVC is retarded by the operation of the VTC mechanism 27, that is, as the crank angle section (IVC-θt0) from the tumble start angle to the IVC increases, the actual tumble strength increases. However, according to the equation (58-1), when the intake valve closing timing IVC is delayed (when the IVC value increases), the tumble strength It also increases, and the calculated value It agrees well with the actual tumble strength. Will be.

  Further, the actual tumble strength increases as the valve lift amount of the intake valve increases due to the operation of the VEL mechanism 28. However, according to the upper part of FIG. 4, when the valve lift amount Lift increases, the valve lift amount correction coefficient kVEL increases. According to the equation (58-1), when the valve lift amount correction coefficient kVEL increases, the tumble strength It also increases, and the calculated value It agrees well with the actual tumble strength.

  Furthermore, according to the equation (58-1), the tumble strength It increases as the total gas mass MGAS increases.

  Formula (58-2) is considered that the turbulence of the combustion chamber gas becomes stronger as the tumble is converted into small vortices during the compression process, so that the turbulence of the combustion chamber gas becomes stronger as the tumble strength It increases. The turbulence strength u of the combustion chamber gas is estimated based on the intensity It (gas flow in the combustion chamber).

  In step 39, the combustion speed Sb at the time of θmb6%, which is the crank angle position when the combustion mass ratio becomes 6%, and the volume Vmb6% at θmb6% of the combustion chamber 5 are used by the following equation (59). Burning time BT [s] is calculated.

BT = 0.0707 · Vmb6% / (Ab · Sb) (59)
Where Ab: flame surface area [m ² ],
Formula (59) uses Vmb 6% as the combustion chamber volume Vcyl in Formula (15) above, and is basically the same as Formula (15).

  FIG. 18 is for calculating the ignition delay time τ [sec] and is executed at regular time intervals (for example, every 10 msec).

Steps 41 to 47 are the same as steps 31 to 37 in FIG. That is, in step 41, the intake valve closing timing IVC [degATDC], the volume Vivc [m ³ ] at the intake valve closing timing calculated in step 2 of FIG. 14, and calculated in step 3 of FIG. The temperature Tivc [K] at the closing timing of the intake valve of the combustion chamber 5, the pressure Pivc [Pa] at the closing timing of the intake valve of the combustion chamber 5 calculated at step 4 in FIG. 14, and calculated at step 5 in FIG. 14. Pressure increase Δpctotal [Pa] due to the total combustion, the tumble strength valve lift amount correction coefficient kVEL [1 / deg] calculated in step 6 of FIG. 14, and the combustion chamber calculated in step 12 of FIG. Volume Vmb 2% [degATDC] at θmb 2% of 5; total gas mass MGAS [g] calculated in step 24 of FIG. The internal inert gas ratio MRESR [%], fuel injection pulse width Ti [ms], target equivalent ratio Tfbya [anonymous number], and engine speed Ne [rpm] calculated in step 25 of FIG. 16 are read.

  In step 42, from the pressure Pivc at the intake valve closing timing of the combustion chamber 5, the volume Vivc at the intake valve closing timing of the combustion chamber 5, and the volume Vmb2% at the time of θmb2% of the combustion chamber 5, the volume is calculated by the following equation (60). A pressure change Δpv [kP] due to the change is calculated.

Δpv = Pivc · {(Vivc / Vmb2%) ^ κ−1} (60)
Where κ: specific heat ratio (fixed value 1.3-1.4),
Equation (59) is the same as equation (28) above. Here, the pressure increase Δpv due to the volume change is the pressure increase in the crank angle section (ignition delay period) from θmb0% to θmb2%, but θmb1% which is the average value of θmb0% and θmb2% is not used. The final value of the ignition delay period is θmb2%.

  In step 43, the pressure increase Δpc [kP] due to combustion is calculated from the total pressure increase Δpctotal due to combustion by the following equation (61).

Δpc = 0.02 × Δptotal (61)
The equation (60) uses 2% as the combustion mass ratio x in the above equation (29).

  In step 44, the pressure increase Δpc due to combustion and the pressure increase Δpv due to volume change are added to the pressure Pivc at the closing timing of the intake valve of the combustion chamber 5, that is, θmb2% of the combustion chamber 5 by the following equation (62). The unburned gas pressure Pmb 2% [kP] at the time is calculated.

Pmb2% = Pivc + Δpv + Δpc (62)
In step 45, using the unburned gas pressure Pmb2% at the time of θmb2% of the combustion chamber 5, the pressure Pivc at the intake valve closing timing of the combustion chamber 5, and the temperature Tivc at the intake valve closing timing of the combustion chamber 5, the following ( 63) The unburned gas temperature Tmb2% [K] at the time of θmb2% of the combustion chamber 5 is calculated by the equation 63).

Tmb2% = Tivc · (Pmb2% / Pivc) ^ {(κ−1) / κ} (63)
Where κ: specific heat ratio (fixed value 1.3-1.4),
Equation (63) is a value when the combustion mass ratio x is 2% in the above equation (25).

  In step 46, the laminar combustion speed SL0 [m / sec] and coefficients a and b in the standard state are obtained from the target equivalent ratio Tfbya by the following equations (64) to (66). In step 47, these standard states are obtained. Laminar combustion speed SL0, coefficients a and b, unburned gas temperature Tmb2% when combustion chamber 5 is at θmb2%, unburned gas pressure Pmb2% when combustion chamber 5 is at θmb2%, internal inert gas ratio MRESR Is used to calculate the laminar combustion speed SL [m / s] at θmb 2% by the following equation (67).

SL0 = (0.2632-0.8472 / (Tfbya-1.13) ^ 2)
... (64)
a = 2.18-0.80 · (Tfbya-1) (65)
b = −0.16 + 0.22 · (Tfbya−1) (66)
SL = SL0 · (Tmb2% / 298) ^ a · (Pmb2% / 101.325) ^ b
(1-2.1 × MRESR) (67)
Expressions (64) to (66) are obtained by using the target equivalent ratio Tfbya as the equivalent ratio φ in the above expressions (20) to (22). Basically, the expressions (20) to (22) does not change. Similarly, the expression (67) uses Tmb 2% and Pmb 2% as the unburned gas temperature T and the combustion chamber pressure p in the above expression (19), and is basically the same as the expression (19).

  In step 48, using this laminar combustion speed SL at θmb2%, the volume Vmb2% at θmb2% of the combustion chamber 5, the engine speed Ne, and the total gas mass MGAS, the ignition delay time is calculated by the following equation (68). τ [sec] is calculated.

τ = {0.490 · Vmb2% ^ (1/3)}
/ {C7 · SL + c8 · Ne · MGAS ^ (1/2)}
... (68)
However, c7, c8: conformity coefficient,
Equation (68) uses the volume Vmb2% and the total gas mass MGAS at θmb2% of the combustion chamber 5 as the combustion chamber volume Vcyl and combustion chamber gas mass m in the above equation (18), respectively. The same as (18).

  FIG. 19 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 executed later in FIGS.

  In step 51, θmb 10% [degATDC], which is the crank angle position when the combustion mass ratio calculated in step 8 of FIG. 14 becomes 10%, and the combustion time BT [sec] calculated in step 39 of FIG. ], The ignition delay time τ [sec] calculated in step 48 of FIG. 18 and the engine speed Ne [rpm] are read.

  In step 52, using the ignition delay time τ, the combustion period BT, the crank angle position when the combustion mass ratio becomes 10%, θmb 10%, and the engine rotational speed Ne, the basic ignition timing MBTCAL is calculated by the following equation (69). [DegATDC] is calculated, and in step 53, the absolute value is taken by the following equation (70), and the negative sign is again calculated as the basic ignition timing MBTCAL [degBTDC].

MBTCAL = (τ + BT) · Ne · 6-θmb 10% (69)
MBTCAL = − | MBTCAL | (70)
Expressions (69) and (70) are the same as the above expression (13).

  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.

  Here, the effect of this embodiment is demonstrated.

The tumble strength It (gas flow in the combustion chamber) is represented by the sum of the angular momentum of small vortices distributed at various locations in the combustion chamber, so, for example, from the center of the cylinder (combustion chamber 5) to the i-th small vortex. Table and the distance r _i, i-th small swirl of gas mass m _i, the i-th small vortex of gas flow rate and v _i, the (3) tumble intensity it (the combustion chamber of a gas flow) by equation However, it is currently impossible to calculate these values m _i , r _i , v _i for all the small vortices in the combustion chamber online in the engine controller 31.

  Therefore, the present inventor considers that the turbulence of the combustion chamber gas becomes stronger due to the tumble being converted into small vortices during the compression process, so that the turbulence of the combustion chamber gas becomes stronger as the tumble strength It increases. The turbulence intensity u of the combustion chamber gas is estimated based on the strength It (gas flow in the combustion chamber). That is, according to the present embodiment, the turbulence intensity u of the combustion chamber gas is calculated based on the tumble strength It (gas flow in the combustion chamber) (see the above equation (58-2)), and the combustion chamber gas A turbulent combustion speed ST is calculated based on the turbulence intensity u (see the above equation (58-3)), and a basic ignition timing for obtaining MBT is calculated based on the turbulent combustion speed ST (described above (58 -4) and step 38 in FIG. 17 and steps 51 to 53 in FIG. 19), even if the turbulent combustion speed ST fluctuates due to the influence of the tumble intensity It (gas flow in the combustion chamber), the fluctuation Since the turbulent combustion speed ST to be calculated can be calculated with high accuracy, the prediction accuracy of the basic ignition timing MBTCAL (basic ignition timing from which MBT is obtained) can be improved.

As described above, since it is currently impossible to calculate m _i , r _i , and v _i for all small vortices in the combustion chamber online in the engine controller 31, some approximation is required to obtain the tumble strength It. is required. Therefore, the present inventor newly estimates the gas flow in the combustion chamber based on the lift characteristics of the intake valve 15 and the engine speed Ne. Here, the lift characteristic of the intake valve 15 is the valve lift amount of the intake valve 15 (or the valve operating angle of the intake valve 15) or the opening / closing timing of the intake valve 15.

  That is, according to the present embodiment, the tumble strength It (gas flow in the combustion chamber) is estimated based on the valve lift amount Lift of the intake valve 15 and the intake valve closing timing IVC (lift characteristic) and the engine speed Ne. (Refer to the above equation (58-1)), the tumble strength It (gas flow in the combustion chamber) can be accurately calculated with a simple configuration.

  When the VEL mechanism 28 is provided, the direction of gas flow in the combustion chamber when the lift amount of the intake valve 15 is small when the VEL mechanism 28 is operating and when the lift amount of the intake valve 15 is large when the VEL mechanism 28 is not operated. (See FIG. 2 (f) and FIG. 2 (g)), the turbulence intensity u of the combustion chamber gas also changes. According to this embodiment, at least the turbulent combustion speed ST is set to the intake air. Since it is calculated based on the valve lift amount Lift (lift characteristic) of the valve 15 (see step 6 in FIG. 14, equations (58-1) to (58-3) above), the VEL mechanism 28 is activated or deactivated. The estimation accuracy of the turbulent combustion speed ST can be improved regardless of the magnitude of the valve lift amount Lift of the intake valve 15 associated therewith.

According to the present embodiment, the operation of the VTC mechanism 27 corresponds to the fact that the longer the intake valve closing timing IVC is, the longer the period during which the combustion chamber gas turbulence is formed and the turbulence strength u of the combustion chamber gas increases. Since the turbulent combustion speed ST is calculated based on at least the intake valve closing timing IVC (see the above equations (58-1) to (58-3)), the intake valve closing timing IVC is advanced or retarded. Regardless of this, the estimation accuracy of the turbulent combustion speed ST can be improved.

This embodiment corresponds to the fact that as the total gas mass MGAS (gas amount in the combustion chamber) of the combustion chamber 5 increases, the kinetic energy of the gas increases and the turbulence strength u of the combustion chamber gas also increases and the combustion becomes faster. According to this, since the turbulent combustion speed ST is also calculated based on the total gas mass MGAS (gas amount in the combustion chamber) of the combustion chamber 5 (see the above equations (58-1) to (58-3)), The estimation accuracy of the turbulent combustion speed ST can be improved regardless of the total gas mass MGAS (gas amount in the combustion chamber) of the combustion chamber 5.

  Corresponding to the large influence of the laminar combustion speed SL on the turbulent combustion speed ST when the turbulence in the combustion chamber gas is small, according to the present embodiment, the laminar combustion speed calculation for calculating the laminar combustion speed ST is performed. Including the processing procedure (see step 37 in FIG. 17), the turbulent combustion speed ST is also calculated based on the laminar combustion speed SL (see the above equation (58-3)), so the turbulence in the combustion chamber gas is disturbed. Even in a small state, the estimation accuracy of the turbulent combustion speed ST can be improved.

  In the above <2>, the method for setting the reference crank angle has been described. Next, an idea that is the basis of the concept will be described.

  The crank angle θpmax at which the combustion pressure is maximized is considered to be affected by engine loss according to the result of examination on the desk. Engine loss includes engine cooling loss Qcool and engine time loss Qtime. Of these, the crank angle θpmax at which the combustion pressure is maximum is mostly affected by the cooling loss Qcool of the engine, and is slightly affected by the time loss Qtime of the engine. Therefore, if the correction coefficients f1 (Qcool) and g1 (Qtime) using the engine cooling loss Qcool and the engine time loss Qtime as parameters are introduced, the crank angle θpmax at which the combustion pressure is maximized by the following equation (71) is obtained. Can be calculated.

θpmax = TH0 · f1 (Qcool) · g1 (Qtime)
... (71)
TH0 in the equation (71) is a reference value of the crank angle at which the combustion pressure becomes maximum, for example, the compression top dead center [TDC], and is retarded from the compression top dead center by the correction coefficients f1 (Qcool) and g1 (Qtime). The value on the side is calculated.

  Here, since it is considered that the crank angle θpmax at which the combustion pressure becomes maximum is shifted to the retard side as the engine cooling loss Qcool increases, the correction coefficient f1 (Qcool) increases as the engine cooling loss Qcool increases. In addition, since it is considered that the crank angle θpmax at which the combustion pressure becomes maximum shifts to the retard side as the engine time loss Qtime increases, the correction coefficient g1 (Qtime) increases as the engine time loss Qtime increases. Set as follows.

  Since the representative value for estimating the engine cooling loss Qcool is the engine load, this engine load (for example, the combustion chamber gas mass m) can be used as it is as the engine cooling loss. In addition to the engine load, there is a combustion period as a representative value for estimating the engine cooling loss Qcool. If the combustion period becomes longer (the engine speed Ne becomes smaller), the engine cooling loss Qcool becomes larger. Therefore, a rotational speed correction factor f2 (Ne) corresponding to the engine rotational speed Ne that is inversely related to the combustion period is introduced, and the cooling loss compensation coefficient f1 (Qcool) is corrected by this rotational speed correction factor f2 (Ne). .

  On the other hand, since the representative value for estimating the engine time loss Qtime is the reciprocal of the combustion speed, the reciprocal of the combustion speed can be simply used as the engine time loss Qtime. In addition to the reciprocal of the combustion speed, the representative value for estimating the engine time loss Qtime includes a combustion period. If the combustion period becomes longer (the engine speed Ne becomes smaller), the time loss Qtime becomes larger. Therefore, a rotational speed correction factor g2 (Ne) corresponding to the engine rotational speed Ne that is inversely related to the combustion period is introduced, and the time loss correction coefficient g1 (Qtime) is corrected by this rotational speed correction factor g2 (Ne). .

  Accordingly, at this time, the crank angle θpmax at which the combustion pressure becomes maximum is calculated by the following equation (72) instead of the above equation (71).

θpmax = TH0 · f1 (Qcool) · f2 (Ne)
・ G1 (Qtime) ・ g2 (Ne) (72)
The flow chart of FIG. 20 is for calculating θmb10%, which is the crank angle position when the combustion mass ratio becomes 10% by incorporating the above basic idea (second embodiment), and at regular intervals (for example, Every 10 msec). In relation to the specific example (first embodiment) described above with reference to FIGS. 14 to 19, in FIG. 19, the calculation is performed in FIG. 20 instead of θmb 10% of the first embodiment calculated in step 8 of FIG. The θmb of 10% of the second embodiment is used.

  Referring to the flow, in step 61, as in step 1 of FIG. 14, the intake valve closing timing IVC [degATDC], the valve lift amount Lift [m], the rotational speed Ne [rpm], and the combustion chamber gas mass m [ g].

  Here, as the gas mass m in the combustion chamber, the intake air amount MA [g / sec] detected by the air flow meter 32 is used.

  In step 62, the valve lift amount correction coefficient kVEL [1 / deg] of the tumble strength is calculated by searching a table having the upper part of FIG. 4 from the valve lift amount Lift as in step 6 of FIG. That is, when the VEL mechanism 28 is not operated, the first valve lift amount correction coefficient kVEL1 is calculated, and when the VEL mechanism 28 is operated, the second valve lift amount correction coefficient kVEL2 is calculated as the valve lift amount correction coefficient kVEL.

  In step 63, the combustion chamber gas mass m is directly used as the engine cooling loss Qcool [J]. In step 64, the basic cooling loss correction coefficient is obtained by searching a table having the contents shown in FIG. 21 from the engine cooling loss Qcool. f1 (Qcool) [anonymous number] is calculated.

  As shown in FIG. 21, the basic cooling loss correction coefficient f1 (Qcool) is a value that increases as the cooling loss Qcool of the engine increases. That is, the basic cooling loss correction coefficient f1 (Qcool) is a value that is determined so that the crank angle θpmax at which the combustion pressure becomes maximum is directed to the retard side from the compression top dead center as the engine cooling loss Qcol increases.

  In step 65, a table having the contents shown in FIG. 22 is retrieved from the engine speed Ne to calculate a rotation speed correction factor f2 (Ne) [anonymous number] of the cooling loss correction coefficient. The value obtained by multiplying the basic cooling loss correction coefficient f1 (Qcool) by the correction factor f2 (Ne) is used as the cooling loss correction coefficient f (Qcool), that is, the cooling loss correction coefficient f (Qcool) according to the following equation (73). [Anonymous] is calculated.

f (Qcool) = f1 (Qcool) · f2 (Ne) (73)
As shown in FIG. 22, the rotational speed correction factor f2 (Ne) is a value that decreases as the rotational speed Ne increases. That is, the rotational speed correction factor f2 (Ne) is a value that is determined such that the crank angle θpmax at which the combustion pressure becomes maximum becomes more retarded as the combustion period becomes longer (the engine rotational speed becomes smaller).

  In step 67, the combustion speed [m / sec] is calculated by the following equation (74) using the valve lift amount correction coefficient kVEL, the intake valve closing timing IVC, the combustion chamber gas mass m, and the engine speed Ne.

Combustion rate = kVEL, IVC, m ^ c4, Ne ^ c5 (74)
However, c4, c5: conformity constant,
As can be seen by comparing the equation (74) with the equation (16), the equation (74) is an equation for accurately obtaining the turbulence intensity u, but here the turbulence intensity u is simply burned as it is. The speed is set.

  In step 68, the reciprocal of this combustion speed is directly used as the engine time loss Qtime [J]. In step 69, the basic time loss correction coefficient is retrieved from the engine time loss Qtime by searching a table having the contents shown in FIG. g1 (Qtime) [nameless number] is calculated.

  As shown in FIG. 23, the basic time loss correction coefficient g1 (Qtime) is a value that increases as the engine time loss Qtime increases. That is, the basic time loss correction coefficient g1 (Qtime) is a value that is determined so that the crank angle θpmax at which the combustion pressure becomes maximum is directed toward the retard side as the engine time loss Qtime increases.

  In step 70, a table having the contents shown in FIG. 24 is searched from the engine speed Ne to calculate a rotational speed correction factor g2 (Ne) [anonymous number] of the time loss correction coefficient. In step 71, the rotational speed is corrected. A value obtained by multiplying the basic time loss correction coefficient g1 (Qtime) by the correction factor g2 (Ne) is used as the time loss correction coefficient g (Qtime), that is, the time loss correction coefficient g (Qtime) according to the following equation (75). [Anonymous] is calculated.

g (Qtime) = g1 (Qtime) · g2 (Ne) (75)
As shown in FIG. 24, the rotational speed correction factor g2 (Ne) is a value that decreases as the rotational speed Ne increases. That is, the rotational speed correction factor f2 (Ne) is a value that is determined such that the crank angle θpmax at which the combustion pressure becomes maximum becomes more retarded as the combustion period becomes longer (the engine rotational speed becomes smaller).

  In step 72, a crank angle reference value TH0 [TDC] at which the combustion pressure, which is the compression top dead center, is maximized by the time loss correction coefficient g (Qtime) and the cooling loss correction coefficient f (Qcool). That is, the crank angle θpmax [degATDC] at which the combustion pressure is maximized is calculated by the following equation (76).

θpmax = TH0 · f (Qcool) · g (Qtime) (76)
Here, the cooling loss of the engine, the engine time loss, the engine speed, the basic cooling loss correction coefficient, and the basic time loss complementary coefficient in the reference operating state are respectively Qcool0, Qtime0, Ne0, f1 (Qcool0), g1 ( If Qtime0), the basic cooling loss correction coefficient f1 (Qcool0) in the reference operation state and the basic time loss complement coefficient g1 (Qtime1) in the reference operation state can be determined by adaptation. If θpmax in the reference operating state is θpmax0, from equation (76), θpmax0 = TH0 · f (Qcool0) · g (Qtime0)
... (77)
And θpmax0 is determined at a predetermined crank angle position after compression top dead center.

  Now, assuming that the engine time loss is constant, when the engine cooling loss Qcool becomes larger than the engine cooling loss Tcool0 in the reference operating state, the crank angle θpmax at which the combustion pressure becomes maximum is retarded from θpmax0. In contrast, when the engine cooling loss Qcool becomes smaller than the engine cooling loss Tcool0 in the reference operating state, the crank angle θpmax at which the combustion pressure becomes maximum moves from θpmax0 to the advance side. FIG. 21 described above reflects the difference of the cooling loss Qcool of the engine from Qcool0 in θpmax.

  Similarly, assuming that the engine cooling loss Qcool is constant and the engine time loss Qtime becomes larger than the engine time loss Qtime0 in the reference operating state, the crank angle θpmax at which the combustion pressure becomes maximum is slower than θpmax0. When the engine time loss Qtime becomes smaller than the engine time loss Qtime0 in the reference operating state, the crank angle θpmax at which the combustion pressure becomes maximum moves from θpmax0 to the advance side. FIG. 23 described above reflects the difference in time loss Qtime of the engine from Qtime0 in θpmax.

  On the other hand, the rotational speed correction factors f2 (Ne) and g2 (Ne) are for reflecting the difference in the combustion period in the crank angle θpmax at which the combustion pressure becomes maximum. Specifically, the rotational speed correction factors f2 (Ne) and g2 (Ne) are both 1.0 at the rotational speed Ne0 in the reference operating state. When the rotational speed Ne becomes higher than the rotational speed Ne0 in the reference operating state, the crank angle θpmax at which the combustion pressure becomes maximum moves to the advance side from θpmax0, and conversely, from the rotational speed Ne0 in the reference operating state. When the rotational speed Ne decreases, the crank angle θpmax at which the combustion pressure becomes maximum moves to the retard side from θpmax0. 22 and 24, the difference between the engine rotational speed Ne and Ne0 (the difference from the combustion period at Ne0 in the combustion period having a reciprocal relationship with the engine rotational speed) is expressed as a cooling loss correction coefficient f (Qtime). ) And the time loss correction coefficient g (Qtime).

  In step 73, θmb10% [degATDC], which is the crank angle position when the combustion mass ratio becomes 10%, is calculated according to the following equation (78) using the crank angle θpmax that maximizes the combustion pressure calculated in this way. calculate.

θmb10% = a · θpmax−b (78)
Where a and b are conforming values,
In FIG. 19 of the first embodiment, instead of θmb 10% calculated in step 8 of FIG. 14 of the first embodiment, the basic ignition timing MBTCAL [degBTDC] is calculated using θmb 10% calculated by the equation (78). calculate. That is, the remaining part excluding step 8 in FIG. 14, FIG. 16, FIG. 17, FIG. 18, and FIG.

  Here, the effect of 2nd Embodiment is demonstrated.

  In view of the fact that the crank angle θpmax at which the combustion pressure becomes maximum varies depending on the operating conditions, but the value is determined by the loss of the engine, according to the second embodiment (the invention according to claims 1 and 13), the combustion pressure Is calculated based on the engine loss (see step 72 of FIG. 20), and the basic ignition timing at which MBT is obtained is calculated based on the crank angle θpmax at which this combustion pressure is maximized. 20 (see step 73 in FIG. 20 and steps 51 to 53 in FIG. 19), even if the crank angle θpmax at which the combustion pressure becomes maximum varies depending on the operating conditions, the varying θpmax can be accurately estimated. Thus, the prediction accuracy of the obtained basic ignition timing can be improved.

  In particular, in an engine equipped with the VEL mechanism 28 and the VTC mechanism 27 (variable valve actuating device), the crank angle θpmax at which the combustion pressure becomes maximum fluctuates more greatly than in an engine not equipped with the VEL mechanism 28 and the VTC mechanism 27. Especially effective for engines. However, it goes without saying that the present invention is also applicable to an engine that does not include the VEL mechanism 28 and the VTC mechanism 27 (variable valve operating device).

  According to the second embodiment (the invention described in claims 2 and 14), the combustion mass ratio representing the ratio of the fuel mass to the fuel supplied to the combustion chamber is based on the crank angle θpmax at which the combustion pressure is maximum. The crank angle position θmb10% reaching 10% (predetermined reference ratio) is calculated (see step 73 in FIG. 20), and the basic ignition timing MBTCAL (MBT) is calculated based on the θmb10% (crank angle position reaching the reference ratio). Is obtained (see steps 51, 52 and 53 in FIG. 19). That is, the crank angle position that reaches the predetermined reference ratio is the crank angle position that is most correlated between the crank angle θpmax at which the combustion pressure is maximized and the crank angle position at which the combustion mass ratio reaches a certain value (specifically, Is the crank angle position when the combustion mass ratio reaches 10% (10mb), and the basic ignition timing MBTCAL (basic ignition timing from which MBT is obtained) is calculated based on the most correlated crank angle position. Thus, the calculation accuracy of the basic ignition timing MBTCAL can be improved.

  In view of the fact that the crank angle θpmax at which the combustion pressure becomes maximum fluctuates mostly due to the engine cooling loss Qcool, according to the second embodiment (the invention according to claims 3 and 15), the engine Since the crank angle θpmax at which the combustion pressure is maximized is calculated based on the cooling loss Qcool (see steps 64, 66, and 72 in FIG. 20), the crank angle θpmax at which the combustion pressure is maximized can be easily calculated. it can.

  In view of the fact that the fluctuation of the crank angle θpmax at which the combustion pressure becomes maximum is affected by the engine time loss Qtime in addition to the engine cooling loss Qcool, the second embodiment (the inventions according to claims 4 and 16) ), The crank angle θpmax at which the combustion pressure is maximized is calculated based on the engine time loss Qtime in addition to the engine cooling loss Qcool (see steps 69, 71 and 72 in FIG. 20). It is possible to further improve the estimation accuracy of the crank angle θpmax that maximizes.

  Next, the flowchart of FIG. 25 is for calculating θmb10%, which is the crank angle position when the combustion mass ratio of the third embodiment reaches 10%, and replaces FIG. 20 of the second embodiment. . The same steps as those in FIG. 20 are given the same step numbers.

  In the second embodiment, the combustion chamber gas mass m is simply set as the engine cooling loss Qcool and the reciprocal of the combustion speed is set as the engine time loss Qtime as it is (see steps 63 and 68 in FIG. 20). In the embodiment, the engine cooling loss Qcool is accurately calculated by the Wassini equation and the combustion speed is also accurately calculated.

  The difference from the second embodiment will be mainly described. In step 82, an engine cooling loss Qcool [J] is calculated. The calculation of the engine cooling loss Qcool will be described with reference to the flowchart of FIG.

  In FIG. 26 (subroutine of step 82 in FIG. 25), in step 91, as in step 1 in FIG. 14, the intake valve closing timing IVC [degATDC] and the valve lift amount correction coefficient calculated in step 62 in FIG. kVEL [1 / deg], collector temperature TCOL [K] detected by the temperature sensor 43, collector pressure PCOL [Pa] detected by the pressure sensor 44, exhaust temperature TEXH [K] detected by the temperature sensor 45 Then, the internal inert gas rate MRESR [%], the engine rotation speed Ne [rpm] detected by the crank angle sensor, and the intake air amount MA [g / sec] detected by the air flow meter 32 are read.

In steps 92, 93, and 94, the volume Vivc [m ³ ] at the intake valve closing timing IVC of the combustion chamber 5 and the temperature Tivc of the combustion chamber 5 at the intake valve closing timing IVC are the same as steps 2, 3, and 4 in FIG. [K], the pressure Pivc [kPa] at the intake valve closing timing IVC of the combustion chamber 5 is calculated.

  Steps 95 to 101 are parts for calculating values (gas flow rate VS, coefficient c1, combustion chamber pressure P0, motoring pressure PM, W, unburned gas temperature TUGAS0, HC) necessary for the Wassini equation.

  First, in step 95, the combustion chamber gas flow rate VS [m / sec] is calculated by the following equation (81) using the valve lift amount correction coefficient kVEL, the intake valve closing timing IVC, and the rotational speed Ne.

VS = kVEL · IVC · Ne (81)
Here, the valve lift amount correction coefficient kVEL is used from the first embodiment. However, since the unit of the gas flow rate VS is [m / sec], the unit of the valve lift amount correction coefficient kVEL is [m / deg]. Must. In order to clarify this point, the second valve lift amount correction coefficient k2VEL [m / deg] is calculated by searching a table similar to the table including the upper part of FIG. 4 from the valve lift amount Lift, The second valve lift amount correction coefficient k2VEL may be used instead of the valve lift amount correction coefficient kVEL of the equation (81).

  In the first embodiment, the valve lift amount correction coefficient kVEL is a tumble valve lift amount correction coefficient. However, in the third embodiment, the combustion chamber gas flow is not necessarily limited to tumble. Applicable when flow is swirl.

  Further, according to the equation (81), since the valve lift amount correction coefficient kVEL and the intake valve closing timing IVC are included, it seems that the engine including the VEL mechanism 28 and the VTC mechanism 27 is also targeted in the third embodiment. However, the third embodiment is also intended for an engine that does not include the VEL mechanism 28 and the VTC mechanism 27. When the engine including the VEL mechanism 28 and the VTC mechanism 27 is targeted, the valve lift amount correction coefficient kVEL is a binary value to be switched and the intake valve closing timing VTC is a variable value. However, the VEL mechanism 28 and the VTC mechanism 27 are When an engine that is not provided is targeted, both the valve lift amount correction coefficient kVEL and the intake valve closing timing IVC are only fixed values.

  In step 96, the coefficient c1 is calculated by the following equation (82) using the combustion chamber gas flow velocity VS and the engine rotational speed Ne.

c1 = 2.28 + 0.308 · VS / 2 / ST / Ne (82)
Where ST: full piston stroke [m]
In steps 97 and 98, combustion is performed by the following equations (83) and (84) using the pressure Pivc at the intake valve closing timing of the combustion chamber 5, the volume Vivc at the intake valve closing timing of the combustion chamber 5, and the intake air amount MA. The indoor pressure P0 and the pressure PM during motoring of the combustion chamber 5 are calculated.

P0 = Pivc · (V / Vivc) ^ κ + MA · MBAVE · DPC
... (83)
PM = Pivc · (Vivc / V) ^ κ (84)
V: combustion chamber volume [m ³ ],
κ: specific heat ratio (1.3 to 1.4 at a fixed value),
MBAVE: Standard combustion mass ratio [%]
DPC: Pressure increase due to combustion per gram of intake air
[KPa / g],
The first term on the right side of equation (83) is the pressure change due to the volume change (that is, the pressure during motoring of the combustion chamber 5), and the second term is the combustion up to the crank angle position (θmb 10%) where the combustion mass ratio reaches 10%. The pressure rise due to. Therefore, the reference combustion mass ratio MBAVE in the equation (83) is 10%.

  The combustion chamber volume V in the above equations (83) and (84) is set to an appropriate value.

  In step 99, the coefficient c1, the rotational speed Ne, the temperature Tivc at the intake valve closing timing of the combustion chamber 5, the pressure Pivc at the intake valve closing timing of the combustion chamber 5, the volume Vivc at the intake valve closing timing of the combustion chamber 5, and the combustion chamber pressure W is calculated by the following equation (85) using P0 and the pressure PM when the combustion chamber 5 is motored.

W = c1 ・ 2 ・ ST ・ Ne
+ C2 · Vc · Tivc / Pivc / Vivc · (P0-PM)
... (85)
Where ST: full piston stroke [m],
Vc: gap volume [m ³ ],
c1, c2: coefficients,
In step 100, the unburned gas temperature TUGAS0 [K] is calculated by the following equation (86) using the temperature Tivc of the combustion chamber 5 when the intake valve is closed, the pressure Pivc of the combustion chamber 5 when the intake valve is closed, and the pressure P0 of the combustion chamber. Is calculated.

TUGAS0 = Tivc · (P0 / Pivc) ^ ((κ−1) / κ)
... (86)
Where κ: specific heat ratio (fixed value 1.3-1.4),
In step 101, HC is calculated by the following equation (87) using the W, the combustion chamber pressure P0, and the unburned gas temperature TUGAS0.

HC = 3.26 ・ D ^ -0.2 ・ P0 ^ 0.8 ・ TUGAS0 ^ -0.55
・ W ^ 0.8
... (87)
Where D: cylinder bore diameter [m],
In step 102, the engine cooling loss Qcool [J] is calculated from the following equation (88), which is a cooling loss equation (Wossini equation), using the HC, the unburned gas temperature TUGAS0, and the rotational speed Ne.

Qcool = ACYL · HC · (TUGAS0-Tw) · dθ · 6 / Ne
... (88)
Where ACYL: combustion chamber wall area [m ² ],
Tw: combustion chamber wall temperature [K],
dθ: average combustion period [deg],
The combustion chamber wall area ACYL, the combustion chamber wall temperature Tw, and the average combustion period dθ are suitable values.

  When the engine cooling loss Qcool is calculated in this way, the process returns to FIG. 25, and the operations of Steps 64-66 are executed in the same manner as Steps 64-66 of FIG. 20 of the second embodiment.

  In step 83 of FIG. 25, engine time loss Qtime [J] is calculated. The calculation of the engine time loss Qtime will be described with reference to the flowchart of FIG.

  In FIG. 27 (subroutine of step 83 in FIG. 25), in step 111, the intake valve closing timing IVC [degATDC], the internal inert gas rate MRESR [%], the target equivalent ratio Tfbya, and the engine rotation detected by the crank angle sensor. The speed Ne [rpm], the intake air amount MA [g / sec] detected by the air flow meter 32, the valve lift amount correction coefficient kVEL calculated in step 62 in FIG. 25, and calculated in step 97 in FIG. The combustion chamber pressure P0 [kPa] and the unburned gas temperature TUGAS0 [K] calculated in step 100 of FIG. 26 are read.

  In step 112, a laminar combustion speed SL0 [m / sec] in the standard state and coefficients a and b [anonymous number] are obtained from the target equivalent ratio Tfbya by the following equations (89) to (91). Using the laminar combustion speed SL0, coefficients a and b, the unburned gas temperature TUGAS0, the combustion chamber pressure P0, and the internal inert gas ratio MRESR in the standard state, the laminar combustion speed SL [ m / sec] is calculated.

SL0 = 0.305-0.549 (Tfbya-1.21) ^ 2
... (89)
a = 2.4-0.271 · Tfbya ^ 3.51 (90)
b = −0.357 + 0.14 · Tfbya ^ 2.77 (91)
SL = SL0 · (TUGAS0 / 298) ^ a · (P0 / 101) ^ b
・ (1-2.06 ・ MRESR ^ 0.77)
... (92)
Expressions (89) to (92) are similar to the expressions (54) to (57).

  In step 114, the combustion speed Sb [m / sec] is calculated by the following equation (93) using the laminar combustion speed SL, the valve lift amount correction coefficient kVEL, the intake valve closing timing IVC, the intake air amount MA, and the engine speed Ne. ] Is calculated.

Sb = SL + (kVEL, IVC, MA ^ e, Ne ^ f / SL) ^ g
... (93)
Where f and g are conforming values,
Equation (93) is the same as equation (58) above.

  Thus, also in the third embodiment, the valve lift amount correction coefficient kVEL, the intake valve closing timing IVC, and the intake air amount MA are newly added to the turbulent combustion speed calculation formula (the second term on the right side of the formula (93)). Has been introduced. However, as described above, the engine includes not only the engine including the VEL mechanism 28 and the VTC mechanism 27 but also the engine not including the VEL mechanism 28 and the VTC mechanism 27.

  In step 115, the engine time loss Qtime [J] is calculated by the following equation (94) using the combustion speed Sb and the rotational speed Ne.

Qtime = d / (Sb · Ne) (94)
Where d: conforming value,
Expression (94) is an expression in which the engine time loss Qtime is inversely proportional to the combustion speed Sb and the rotational speed Ne.

  When the engine time loss Qtime is calculated in this way, returning to FIG. 25, the operations of steps 69 to 73 are executed in the same manner as steps 69 to 73 of FIG. 20 of the second embodiment, so that the combustion mass ratio becomes 10%. Calculate θmb 10%, which is the crank angle position at which it reaches.

  Then, in FIG. 19 of the first embodiment, instead of θmb 10% calculated in step 8 of FIG. 14 of the first embodiment, the basic ignition timing MBTCAL [degBTDC] is used by using θmb 10% calculated in step 73 of FIG. Is calculated. That is, the remaining part excluding step 8 in FIG. 14, FIG. 16, FIG. 17, FIG. 18, and FIG.

  Here, the function and effect of the third embodiment will be described.

  The engine cooling loss Qcool is determined from the unburned gas temperature TUGAS0 (temperature in the combustion chamber), the pressure in the combustion chamber P0 (pressure in the combustion chamber), the gas flow velocity VS (gas flow velocity in the combustion chamber) and the combustion period (rotational speed Ne). The impact of conversion is large. According to the third embodiment (the invention described in claims 5 and 17), based on these, the engine cooling loss Qcool is calculated by the Wassini equation (see steps 95 to 102 in FIG. 26). The cooling loss Qcool can be accurately calculated.

  Although it is very difficult to accurately calculate the combustion chamber gas flow velocity VS, in view of the fact that the combustion chamber gas flow velocity VS is substantially proportional to the engine speed Ne, the third embodiment (described in claims 6 and 18). Since the combustion chamber gas flow rate VS is calculated based on at least the engine speed Ne (see step 95 in FIG. 26), the combustion chamber gas flow rate VS can be simply estimated.

  When the VEL mechanism 28 is operated and the valve lift amount is smaller than that when the VEL mechanism 28 is not operated and the opening area of the intake valve 15 is small, the combustion chamber gas flow rate VS passes through the intake valve 15 with the opening area. Since the flow velocity of the intake air flowing into the combustion chamber 5 is limited, the gas flow velocity VS in the combustion chamber is different from that when the VEL mechanism 28 is inactive and the opening area of the intake valve 15 is large. In other words, the valve lift amount (lift characteristic of the intake valve 15) differs depending on the operating conditions depending on whether the VEL mechanism 28 (variable valve operating device) is activated or deactivated, and the valve lift amount (intake air) according to the operating conditions. Due to the difference in the lift characteristics of the valve 15, the gas flow velocity VS in the combustion chamber changes more greatly than in the case of an engine that does not include the VEL mechanism 28 (variable valve operating device). According to the third embodiment (the invention described in claims 7 and 19), when the VEL mechanism 28 (variable valve operating device) is provided, the gas flow velocity VS in the combustion chamber is added to the engine rotational speed Ne, and the valve lift amount. Since the calculation is also based on (the lift characteristic of the intake valve) (see step 95 in FIG. 26), the opening area of the intake valve 15 changes depending on the valve lift amount (the lift characteristic of the intake valve) according to the operating conditions. Even in this case, the gas flow velocity VS in the combustion chamber can be calculated with high accuracy.

  Although it is difficult to calculate the engine time loss Qtime itself by the engine controller 31, according to the third embodiment (the inventions described in claims 8 and 20), the engine time loss Qtime decreases as the combustion speed decreases. Since the time loss Qtime of the engine is calculated based on the combustion speed Sb of the combustion chamber gas (see step 115 in FIG. 27), the time loss Qtime of the engine is easily calculated.

  In view of the fact that the combustion speed Sb of the combustion chamber gas is substantially determined by the engine speed Ne during the period when the combustion has progressed, according to the third embodiment (the invention according to claims 9 and 21), the combustion of the combustion chamber gas Since the speed Sb is calculated based on at least the engine rotational speed Ne (see step 114 in FIG. 27), the engine time loss Qtime can be easily calculated without relatively reducing the estimation accuracy.

  When the VEL mechanism 28 is activated and the valve lift amount Lift is changed from a large state to a small state, or when the VEL mechanism 28 is not operated, the valve lift amount Lift is changed from a small state to a large state. 27, when the opening / closing timing of the intake valve 15 is changed to the retard side or the advance side (when the variable valve mechanism is activated and the lift characteristic of the intake valve 15 is changed), the gas flow in the combustion chamber is changed. The gas combustion speed Sb changes more greatly than in the case of an engine that does not include the VEL mechanism 28 and the VTC mechanism 27. In this case, according to the third embodiment (the invention described in claims 10 and 22), in addition to the engine rotational speed Ne, the combustion chamber gas is also determined based on the valve lift amount (lift characteristic of the intake valve 15). Since the combustion speed Sb is calculated (see kVEL and IVC in step 114 of FIG. 27), even if the valve lift amount of the intake valve 15 and the opening / closing timing of the intake valve 15 (lift characteristic of the intake valve) are different, The combustion speed Sb can be calculated with high accuracy.

  According to the third embodiment (the inventions according to claims 11 and 23), the combustion speed Sb of the combustion chamber gas is greatly influenced by the temperature, pressure, and inert gas ratio of the combustion chamber gas at the early stage of combustion. In addition to the engine rotation speed Ne, the valve lift amount and opening / closing timing of the intake valve 15 (intake valve lift characteristics), the unburned gas temperature TUGAS0 (combustion chamber gas temperature), the combustion chamber pressure P0 (combustion chamber pressure) Since the laminar combustion speed SL (combustion speed Sb of the combustion chamber gas) is also calculated based on the inert gas ratio MRESR (see step 113 in FIG. 27), particularly the laminar combustion speed SL (combustion chamber gas in the initial stage of combustion) The estimation accuracy of the combustion speed Sb) is improved.

  In contrast to the third embodiment, it is conceivable to provide a variable intake air flow device that can change the intake air flow into the combustion chamber 5, such as a swirl control valve and a tumble control valve. For example, in the state where the swirl control valve is operated, the turbulent combustion speed ST of the combustion chamber gas is larger than the state where the swirl control valve is not operated. Therefore, when the swirl control valve is operated, What is necessary is just to correct | amend to the side where the turbulent combustion speed ST becomes large by the part which the turbulent combustion speed ST becomes large. That is, the combustion speed Sb of the combustion chamber gas is also calculated according to the operating state (operating state, non-operating state) of the swirl control valve (the state of the variable intake air flow device) (the invention according to claims 12 and 24). Thereby, even when the variable intake flow device is provided, the combustion speed Sb of the combustion chamber gas can be calculated with high accuracy.

  In the embodiment, the case where the VEL mechanism 28 and the VTC mechanism 27 are provided has been described. However, the present invention can be applied to the case where any one of the VEL mechanism 28 and the VTC mechanism 27 is provided or none.

  In the first aspect of the present invention, the combustion pressure maximum crank angle calculation processing procedure is executed by step 72 of FIG. 20, and the basic ignition timing calculation processing procedure is executed by steps 51, 52, and 53 of FIG.

  In the invention described in claim 13, the function of the combustion pressure maximum crank angle calculating means is performed by step 72 of FIG. 20, and the function of the basic ignition timing calculating means is performed by steps 51, 52, and 53 of FIG. .

The schematic block diagram of the ignition timing control apparatus of the engine of 1st Embodiment of this invention. The figure which represents from the production | generation of tumble in a combustion chamber to extinction according to the motion of a piston. The characteristic figure of the tumble strength with respect to the intake valve closing timing. The characteristic figure of the valve lift amount correction coefficient with respect to the valve lift amount. The characteristic view showing the relationship between the crank angle θpmax at which the combustion pressure is maximum and θmb 0%, θmb 10%, and θmb 60%. The characteristic view showing the relationship between the crank angle θpmax at which the combustion pressure becomes maximum and the engine speed. The combustion mass ratio and the crank angle position when the combustion mass ratio is 0%, 10%, 20%, 30%, 40%, 50%, 60%, θmb 0%, θmb 10%, θmb 20%, θmb 30%, The characteristic view showing the relationship with the square value of R when θmb40%, θmb50%, and θmb60% are linearly approximated by θpmax, respectively. The combustion mass ratio characteristic view for explaining the difference between the reference crank angle in the preceding ignition timing control method applied to the conventional engine and the reference crank angle in the post-change ignition timing control method of the present embodiment. FIG. 6 is a characteristic diagram of a crank angle θpmax at which the combustion pressure is maximum with respect to the intake air amount and the engine rotation speed. The characteristic diagram of the flow velocity coefficient with respect to the valve lift amount. The characteristic view for demonstrating the calculation method of the basic ignition timing of 1st Embodiment. The characteristic view which shows the relationship between (theta) mb x% which is a crank angle position when a combustion mass ratio will be x%, and unburned gas volume Vmb x% at the time of (theta) mb x% of the combustion chamber 5. FIG. The characteristic view which shows the relationship between a fuel-injection pulse width and the total pressure rise by combustion. 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 flowchart for demonstrating calculation of an internal inert gas rate. The flowchart for demonstrating calculation of a combustion period. The flowchart for demonstrating calculation of ignition delay time. The flowchart for demonstrating calculation of basic ignition timing. The flowchart for demonstrating calculation of (theta) mb10% which is a crank angle position when the combustion mass ratio of 2nd Embodiment becomes 10%. The characteristic figure of a basic cooling loss correction coefficient. The characteristic figure of the rotational speed correction factor of a cooling loss correction coefficient. The characteristic figure of a basic time loss correction coefficient. The characteristic figure of the rotational speed correction factor of a time loss correction coefficient. The flowchart for demonstrating calculation of (theta) mb10% which is a crank angle position when the combustion mass ratio of 3rd Embodiment becomes 10%. The flowchart for demonstrating calculation of the cooling loss of 3rd Embodiment. The flowchart for demonstrating calculation of the time loss of 3rd Embodiment.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Engine 5 Combustion chamber 11 Ignition device 15 Intake valve 21 Fuel injector 27 VTC mechanism (variable valve operating device)
28 VEL mechanism (variable valve operating device)
31 Engine controller 33, 34 Crank angle sensor

Claims (24)

Combustion pressure maximum crank angle calculation processing procedure for calculating the crank angle at which the combustion pressure is maximum based on engine loss,
And a basic ignition timing calculation processing procedure for calculating a basic ignition timing at which MBT is obtained based on a crank angle at which the combustion pressure is maximum.   Based on the crank angle at which the combustion pressure becomes maximum, the crank angle position at which the combustion mass ratio representing the ratio of the fuel mass to the fuel supplied to the combustion chamber reaches a predetermined reference ratio is calculated, and the crank angle at which the reference ratio is reached is calculated. 2. The engine ignition timing control method according to claim 1, wherein a basic ignition timing at which the MBT is obtained is calculated based on an angular position.   The engine ignition timing control method according to claim 1, wherein the engine loss includes at least an engine cooling loss.   The engine ignition timing control method according to claim 1, wherein the engine loss includes at least an engine cooling loss and an engine time loss.   5. The engine ignition timing control according to claim 3, wherein the engine cooling loss is calculated based on a combustion chamber temperature, a combustion chamber pressure, a combustion chamber gas flow rate, and an engine rotation speed. Method.   6. The engine ignition timing control method according to claim 5, wherein the combustion chamber gas flow velocity is calculated based on at least the engine rotation speed. With a variable valve gear that can change the lift characteristics of the intake valve,
6. The engine ignition timing control method according to claim 5, wherein the combustion chamber gas flow rate is calculated based on at least an engine speed and a lift characteristic of the intake valve.   5. The engine ignition timing control method according to claim 4, wherein the time loss of the engine is calculated based on a combustion speed of a combustion chamber gas.   9. The engine ignition timing control method according to claim 8, wherein a combustion speed of the combustion chamber gas is calculated based on at least an engine rotation speed. With a variable valve gear that can change the lift characteristics of the intake valve,
The engine ignition timing control method according to claim 8, wherein the combustion speed of the combustion chamber gas is calculated based on at least an engine rotation speed and lift characteristics of the intake valve. With a variable valve gear that can change the lift characteristics of the intake valve,
9. The combustion speed of the combustion chamber gas is calculated based on the temperature of the combustion chamber gas, the pressure of the combustion chamber gas, the inert gas rate, the engine speed, and the lift characteristics of the intake valve. The engine ignition timing control method described. A variable intake air flow device that can change the intake air flow to the combustion chamber,
12. The engine ignition timing control method according to claim 10, wherein the combustion speed of the combustion chamber gas is also calculated according to the state of the variable intake air flow device. Combustion pressure maximum crank angle calculating means for calculating a crank angle at which the combustion pressure is maximum based on engine loss;
An ignition timing control device for an engine, comprising: basic ignition timing calculation means for calculating a basic ignition timing at which MBT is obtained based on a crank angle at which the combustion pressure becomes maximum.   Based on the crank angle at which the combustion pressure becomes maximum, the crank angle position at which the combustion mass ratio representing the ratio of the fuel mass to the fuel supplied to the combustion chamber reaches a predetermined reference ratio is calculated, and the crank angle at which the reference ratio is reached is calculated. The engine ignition timing control device according to claim 13, wherein a basic ignition timing at which the MBT is obtained is calculated based on an angular position.   14. The engine ignition timing control device according to claim 13, wherein the engine loss includes at least an engine cooling loss.   14. The engine ignition timing control device according to claim 13, wherein the engine loss includes at least an engine cooling loss and an engine time loss.   The engine ignition timing control according to claim 15 or 16, wherein the engine cooling loss is calculated based on a combustion chamber temperature, a combustion chamber pressure, a combustion chamber gas flow rate, and an engine rotation speed. apparatus.   18. The engine ignition timing control device according to claim 17, wherein the combustion chamber gas flow velocity is calculated based on at least an engine rotation speed. With a variable valve gear that can change the lift characteristics of the intake valve,
18. The engine ignition timing control device according to claim 17, wherein the combustion chamber gas flow rate is calculated based on at least an engine rotation speed and a lift characteristic of the intake valve.   The engine ignition timing control device according to claim 16, wherein the time loss of the engine is calculated based on a combustion speed of a combustion chamber gas.   21. The engine ignition timing control device according to claim 20, wherein a combustion speed of the combustion chamber gas is calculated based on at least an engine rotation speed. With a variable valve gear that can change the lift characteristics of the intake valve,
21. The engine ignition timing control device according to claim 20, wherein a combustion speed of the combustion chamber gas is calculated based on at least an engine rotation speed and a lift characteristic of the intake valve. With a variable valve gear that can change the lift characteristics of the intake valve,
21. The combustion speed of the combustion chamber gas is calculated based on the temperature of the combustion chamber gas, the pressure of the combustion chamber gas, the inert gas rate, the engine rotation speed, and the lift characteristics of the intake valve. The engine ignition timing control device described. A variable intake air flow device that can change the intake air flow to the combustion chamber,
The engine ignition timing control device according to claim 22 or 23, wherein the combustion speed of the combustion chamber gas is also calculated according to the state of the variable intake air flow device.

Images