JP4175320B2 - Engine ignition timing control device - Google Patents

Description

  The present invention relates to an ignition timing control device for an engine (internal combustion engine), and particularly to a device that focuses on the turbulence strength of gas flow in a combustion chamber (hereinafter referred to as “in-cylinder turbulence strength”).

Since the in-cylinder turbulence intensity is a factor that greatly affects combustion, in an engine equipped with a variable valve operating system, the in-cylinder turbulence intensity is calculated based on the intake valve operating angle, the intake valve phase (valve opening timing), There is one that estimates based on the pressure in the intake pipe and the engine rotational speed, and controls the engine based on the estimated in-cylinder turbulence intensity (see Patent Document 1).
JP 2002-180894 A

  Incidentally, the in-cylinder turbulence intensity becomes a problem in the actual in-cylinder turbulence intensity at the basic ignition timing. Since the optimal in-cylinder turbulence intensity at the basic ignition timing matches the target value of the in-cylinder turbulence intensity at the basic ignition timing, optimal combustion (combustion speed) can be obtained. If the in-cylinder turbulence intensity deviates from the target value of the in-cylinder turbulence intensity at the basic ignition timing, for example, the actual in-cylinder turbulence intensity at the basic ignition timing becomes larger than the target value of the in-cylinder turbulence intensity at the basic ignition timing, When the combustion speed becomes high and combustion is promoted too much, and conversely, when the actual in-cylinder turbulence intensity at the basic ignition timing is smaller than the target value of the in-cylinder turbulence intensity at the basic ignition timing, the combustion speed becomes slow. Combustion is too slow.

  Considering this, it is necessary to know what the in-cylinder turbulence intensity at the basic ignition timing actually is.

  However, until now, it has not been known what the in-cylinder turbulence intensity at the basic ignition timing is actually.

  Therefore, as a result of experiments conducted by the present inventor, the actual in-cylinder turbulence intensity after the intake valve is closed may be gradually attenuated from the maximum value at the reference crank angle as shown in FIG. Newly found. Therefore, it has been newly found that the characteristics of FIG. 2 can be approximated (set) by a function U (θ) having a crank angle θ from the reference crank angle as a variable. Based on the in-cylinder turbulence intensity function U (θ) shown in FIG. 2, the actual in-cylinder turbulence intensity (UMBT) at the basic ignition timing (θMBT) can be known. Based on the in-cylinder turbulence intensity (UMBT), it is possible to control the actual in-cylinder turbulence intensity at the basic ignition timing to coincide with the target value of the in-cylinder turbulence intensity at the basic ignition timing.

  If the technique of the present invention is compared with the technique of Patent Document 1, the technique of Patent Document 1 ignores a change in the actual in-cylinder turbulence intensity after the intake valve is closed, and a certain crank angle. This value corresponds to the value representing in-cylinder turbulence strength. In this case, even if a certain crank angle estimates the actual in-cylinder turbulence intensity at the basic ignition timing, the intake valve operating angle, intake valve phase (opening timing) ) Based on the pressure in the intake pipe and the engine speed, the actual in-cylinder turbulence intensity at the basic ignition timing cannot be accurately estimated.

  For example, the damping coefficient of the in-cylinder turbulence intensity shown in FIG. 2 is determined by the volume at the reference crank angle of the combustion chamber, the temperature at the reference crank angle of the combustion chamber, and the pressure at the reference crank angle of the combustion chamber, as will be described later. Therefore, based on only the pressure in the intake pipe, the actual in-cylinder turbulence intensity at the basic ignition timing cannot be accurately estimated.

  Accordingly, an object of the present invention is to provide an apparatus capable of appropriately performing ignition timing control using the newly found attenuation curve representing the actual in-cylinder turbulence intensity.

  In the present invention, the actual in-cylinder turbulence intensity after the intake valve is closed is set as a function, the basic ignition timing (θMBT) is calculated based on the operating conditions, and the set in-cylinder turbulence intensity function Is used to calculate the actual in-cylinder turbulence intensity (UMBT) at the basic ignition timing, and the basic ignition timing (θMBT) is corrected based on the actual in-cylinder turbulence intensity (UMBT) at the basic ignition timing. The spark ignition is performed at the corrected basic ignition timing.

  The actual in-cylinder turbulence intensity (UMBT) at the basic ignition timing deviates from the target value (Um) of the in-cylinder turbulence intensity at the basic ignition timing. For example, the actual in-cylinder turbulence intensity (UMBT) at the basic ignition timing is basic. If it exceeds the target value (Um) of the in-cylinder turbulence intensity at the ignition timing, the combustion speed increases and combustion is accelerated too much. On the contrary, the actual in-cylinder turbulence intensity (UMBT) at the basic ignition timing is fundamental. When it is smaller than the target value (Um) of the in-cylinder turbulence intensity at the ignition timing, the combustion speed becomes slow and the combustion becomes too slow.

  Therefore, until now, it has not been possible to know what the in-cylinder turbulence intensity (UMBT) actually occurs at the basic ignition timing, but according to the present invention, for the first time, the actual The in-cylinder turbulence intensity can be set by a predetermined function (U (θ)). Based on this in-cylinder turbulence intensity function (U (θ)), the actual in-cylinder turbulence intensity (UMBT) at the basic ignition timing can be known. By correcting the basic ignition timing (θMBT) based on the strength (UMBT), the actual in-cylinder turbulence intensity (UMBT) at the basic ignition timing is changed to the target value (Um Thus, even if the actual in-cylinder turbulence intensity (UMBT) at the basic ignition timing differs depending on the environmental conditions or changes depending on the operating conditions, the in-cylinder at the basic ignition timing The target value (Um) of the turbulence intensity can be realized.

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

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

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

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

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

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

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

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

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

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

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

  Now, the in-cylinder turbulence intensity greatly affects the combustion at the basic ignition timing. This time, it has been newly found through experiments how the actual in-cylinder turbulence intensity changes after the intake valve 15 is closed. This will be explained below. In FIG. 2, when the crank angle θ [deg] from the reference crank angle (for example, intake valve closing timing IVC) is taken on the horizontal axis, the in-cylinder turbulence strength U [m / s] is actually determined. It is the experimental result which showed how it changes. As shown in the figure, it can be seen that the actual in-cylinder turbulence intensity U attenuates more rapidly than the reference turbulence intensity U0 and converges to a certain value (a non-zero value). Such a change in the actual in-cylinder turbulence intensity U is approximated (set) as follows using an exponential function U (θ) with θ as a variable.

U (θ) = U0−GK × {1−e ^ (− θ × MU)} (1)
However, U0: standard disturbance strength,
GK: In-cylinder turbulence attenuation range,
MU: In-cylinder turbulence attenuation coefficient,
The in-cylinder turbulence attenuation range GK basically has a correlation with the rotational speed NRPM, and further enhances in-cylinder turbulence intensity such as a swirl control valve and a tumble control valve. It is considered that it also correlates with the presence or absence (not shown), and a term correlated with the rotational speed NRPM is expressed as a basic turbulence intensity attenuation width GK0, which is expressed by a function f1 (NRPM) of the rotational speed NRPM. A term that correlates with the presence or absence is newly introduced as the in-cylinder gas flow state instruction value d, and the in-cylinder turbulence intensity attenuation width GK is calculated by the product thereof, that is, the following equation.

GK = GK0 × d (2)
GK0 = f1 (NRPM) (3)
Here, since the in-cylinder turbulence intensity attenuation width GK increases as the rotational speed NRPM increases, the basic turbulence intensity attenuation width GK0 is given in proportion to the rotational speed NRPM as shown in FIG.

  The in-cylinder gas flow state instruction value d is an instruction value for the in-cylinder gas flow enhancement device. It is confirmed that the engine with the in-cylinder gas flow strengthening device has a larger attenuation width GK of the in-cylinder turbulence strength than the engine without the in-cylinder gas flow strengthening device. In-cylinder gas flow state instruction value d = 1 when the engine is not provided, and in-cylinder gas flow according to the type of the in-cylinder gas flow enhancement device when the engine is provided with the in-cylinder gas flow enhancement device. The value of the state instruction value d is set to a value exceeding 1.0. Finally, the value of d is determined by adaptation.

  On the other hand, the damping coefficient MU of the in-cylinder turbulence strength is a value corresponding to the time constant, and therefore, the damping curve shown in FIG. 2 becomes gentler as this value increases. The degree of attenuation of in-cylinder turbulence is considered to be correlated with the viscosity of in-cylinder gas, and the in-cylinder gas viscosity coefficient ν is introduced and proportional to this, that is, the in-cylinder turbulence intensity is attenuated by the following equation: The coefficient MU is calculated.

MU = ν × C (4)
Where C: a constant for converting the viscosity coefficient into a damping coefficient,
In this case, it is considered that the temperature TIVC at the reference crank angle of the combustion chamber 5 and the density ρ of the cylinder gas are correlated with the viscosity of the in-cylinder gas, and are correlated with the temperature TIVC at the reference crank angle of the combustion chamber 5. Is given as a function f2 (TIVC) of the temperature TIVC at the reference crank angle of the combustion chamber 5, and this function f2 (TIVC) is divided by the cylinder gas density ρ, that is, the viscosity coefficient ν is calculated by the following equation: To do.

ν = f2 (TIVC) / ρ (5)
Here, since the in-cylinder turbulence intensity increases as the temperature TIVC at the reference crank angle of the combustion chamber 5 increases, the function f2 (TIVC) is expressed as the temperature TIVC at the reference crank angle of the combustion chamber 5 as shown in FIG. Is given in proportion to.

  The in-cylinder gas density ρ is a value obtained by dividing the in-cylinder gas mass m by the volume VIVC at the reference crank angle of the combustion chamber 5, and is thus calculated by the following equation.

ρ = m / VIVC (6)
The mass m of the in-cylinder gas in equation (6) is expressed by the following equation using the volume VIVC at the reference crank angle of the combustion chamber 5, the pressure PIVC at the reference crank angle of the combustion chamber 5, and the temperature TIVC at the reference crank angle of the combustion chamber 5. Calculated by

m = (PIVC × VIVC) / R × TIVC (7)
Where R: gas constant,
Expression (7) is a modification of the equation of state P · V = m · R · T.

  In this way, when the in-cylinder turbulence intensity U (θ) can be set as an exponential function of the crank angle θ from the reference crank angle (IVC) as in the above equation (1), the basic ignition timing θMBT is set to ( By substituting for θ in equation (1), that is, the actual in-cylinder turbulence intensity UMBT at the basic ignition timing θMBT can be obtained (calculated) by the following equation.

UMBT = θ (θMBT) (8)
However, the unit of the basic ignition timing θMBT is [degBTDC] as will be described later, and the unit is different from the crank angle [deg] from the reference crank angle on the horizontal axis in FIG.

  Thus, the actual in-cylinder turbulence intensity UMBT at the basic ignition timing depends on the current environmental conditions such as the collector temperature TCOL (equal to the atmospheric temperature) and the current operating conditions such as the collector pressure PCOL and the engine speed NRPM. In addition, if it is possible to calculate with high accuracy according to the intake valve closing timing IVC as the reference crank angle controlled by the variable valve operating device (intake VTC mechanism 27) or the presence or absence of the in-cylinder gas flow strengthening device, this is fed back. By using as a value, it is possible to perform feedback control so that the actual in-cylinder turbulence intensity UMBT at the basic ignition timing matches the target value Um (constant value) of the in-cylinder turbulence intensity at the basic ignition timing.

  The contents of this control executed by the engine controller 31 will be described in detail with reference to the block diagram of FIG.

  FIG. 5 is a block diagram for calculating the ignition timing θADV [degBTDC], which is executed at regular intervals (for example, every 10 ms).

  In FIG. 5, a filling efficiency detection unit 51 calculates a filling efficiency detection value ITACr based on the output of the air flow meter 32. Regarding this, since a technique for calculating the cylinder air mass per intake using a collector response model is described in Japanese Patent Application Laid-Open No. 2000-161113, the cylinder air mass per intake is used. Detect the filling efficiency.

  FIG. 6 is a detailed block diagram of the filling efficiency detector 51. In FIG. 6, the charging efficiency detection unit 51 includes a voltage → flow rate calculation unit 81, an air flow meter error correction unit 82, an intake pulsation smoothing processing unit 83, a pressure change amount calculation unit 84, a cylinder intake mass calculation unit 85 per intake air, A filling efficiency detection value calculation unit 86 is provided.

  The voltage that receives the output voltage of the hot wire type air flow meter 32 → the flow rate calculation unit 81 converts the air flow meter output voltage into an air flow rate Q0 (mass flow rate) by searching a predetermined table. The air flow meter error correction unit 82 calculates a value obtained by multiplying the air flow rate Q0 [g / ms] by a correction coefficient as an air flow rate Qafm [g / ms] after error correction. Here, the correction coefficient is a value obtained by searching a predetermined map from the engine speed NRPM and the throttle valve opening detected by the throttle sensor 47. The correction of the air flow rate Q0 is for correcting that the air flow rate Q0 becomes too large due to the influence of the intake pulsation.

  The intake pulsation smoothing processing unit 83 performs a smoothing process on the air flow rate Qafm after error correction by the following equation to calculate a throttle valve passing air flow rate Mt [g / ms] per unit time.

Mt = Qafm × weighted average coefficient + (1−weighted average coefficient) × Mt (previous value)
... (9)
Where Mt (previous value): previous value of Mt,
The collector pressure change amount calculation unit 84 that inputs the collecter 2 pressure PCOL [kPa] detected by the collector pressure sensor 44 calculates the collector pressure change amount ΔP (= PCOL−PCOL (previous value)).

  The cylinder intake mass calculation unit 85 per intake air calculates a cylinder intake mass Mc [g / cycle] per intake air according to the following equation.

Mc0 [g / ms] = Mt−ΔP × VCOL / (R · TCOL)
(10)
Mc [g / cycle] = coefficient × (Mc0 / NRPM) / number of cylinders (11)
However, VCOL: collector volume (constant value),
TCOL: collector temperature,
R: gas constant,
The charging efficiency detection value calculation unit 86 uses the cylinder intake mass Mc per intake and the cylinder intake mass Mcbase [g / cycle] (constant value) per intake in the standard state, that is, the charging efficiency detection value by the following equation. ITACTr [nameless number] is calculated.

ITACr = Mc / Mcbase (12)
Returning to FIG. 5, the basic ignition timing calculation unit 52 retrieves a map having the contents shown in FIG. 7 from the engine rotation speed NRPM [rpm] detected by the crank angle sensor and the charging efficiency detection value ITACr [anonymous number]. Thus, the basic ignition timing θMBT [degBTDC] is calculated. The basic ignition timing θMBT is an ignition timing at which MBT (Minimum Advance for Best Torque) is obtained. As a unit of the basic ignition timing θMBT, a value measured from the compression top dead center to the advance side is adopted.

  As shown in FIG. 7, under the condition that the rotational speed NRPM is constant, the basic ignition timing θMBT is a value that becomes retarded as the charging efficiency detection value ITACCr increases. This is because as the charging efficiency detection value ITACCr increases, the combustion speed increases and the combustion proceeds too much. Therefore, it is necessary to suppress the combustion too much by setting the ignition timing to the retarded side. Further, under the condition that the charging efficiency detection value ITACr is constant, the basic ignition timing θMBT becomes a value on the advance side as the rotational speed NRPM increases. This is because ignition tends to be delayed as the rotational speed NRPM increases, so that the ignition timing is not delayed.

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

  The reference crank angle detection unit 53 detects the intake valve closing timing IVC [degBTDC] as the reference crank angle. The intake valve closing timing IVC is known from the command value given to the intake VTC mechanism 27. Alternatively, the actual intake valve closing timing may be detected by the phase sensor 34. The reference crank angle is preferably the intake valve closing timing IVC having the greatest in-cylinder turbulence strength, but is not limited to this, and may be a value related to the intake valve closing timing IVC. For example, the intake valve opening timing IVO or the average value of the intake valve opening timing IVO and the intake valve closing timing IVC may be considered.

  The reference turbulence intensity calculation unit 54 searches the map having the contents shown in FIG. 8 from the rotational speed NRPM and the reference crank angle (IVC), thereby determining the in-cylinder turbulence intensity at the reference crank angle as the reference turbulence intensity. It is calculated as U0 [m / s]. As shown in FIG. 8, the reference turbulence intensity U0 is a value that increases as the reference crank angle (IVC) lags behind the intake bottom dead center when the load and the rotational speed NRPM are constant. This is because when the intake valve 15 closes beyond the intake bottom dead center BDC, the in-cylinder gas is pushed upward by the rising piston 6 and the in-cylinder turbulence strength increases. FIG. 8 includes the case where the intake valve 15 is closed before the intake bottom dead center BDC, that is, the time when the intake valve closing timing is on the more advanced side than the intake bottom dead center. The characteristics shown in FIG. 8 are appropriate for each engine model.

In the volume calculation unit 55, the temperature calculation unit 56, and the pressure calculation unit 57, the volume VIVC [m ³ ] at the reference crank angle (IVC) of the combustion chamber 5, the temperature TIVC [K] at the reference crank angle of the combustion chamber 5, and the combustion chamber The pressure PIVC [kPa] at a reference crank angle of 5 is calculated. As these calculation methods, the technique disclosed in Japanese Patent Laid-Open No. 2003-148236 may be used.

  Explaining from the volume VIVC of the combustion chamber 5 at the intake valve closing timing IVC, the volume VIVC of the combustion chamber 5 at the intake valve closing timing 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. 9, 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 (13) to (17).

VIVC = f1 (θivc)
= Vc + (π / 4) D ² · Hivc (13)
Vc = (π / 4) D ² · Hx / (ε−1) (14)
^{Hivc = {(CND + ST 2} /2) - (CRoff-PISoff) 2} 1/2
− {(ST / 2) · cos (θivc + θoff)} + (CND ² −X ² ) ^1/2
... (15)
X = (ST / 2) · sin (θivc + θoff) −CRoff + PISoff
... (16)
θoff = arcsin {(CRoff−PISoff) / (CND · (ST / 2))}
... (17)
Where Vc: gap volume [m ³ ],
ε: compression ratio,
D: cylinder bore diameter [m],
ST: Full piston stroke [m],
Hivc: From the TDC of the piston pin 76 when the intake valve is closed
Distance [m],
Hx: Maximum value and minimum value of the distance from the TDC of the piston pin 76
Difference [m],
CND: length of connecting rod 74 [m],
CRoff: offset distance of the nodal point 75 from the cylinder center axis 73 [m],
PISoff: offset distance [m] of the crankshaft rotation center 72 from the cylinder center axis 73,
θivc: Intake valve closing timing crank angle [degATDC],
θoff: piston pin 76 and crankshaft rotation center 72
The angle [deg] between the line connecting the two and the vertical line in TDC,
X: horizontal distance [m] between the nodal point 75 and the piston pin 76,
As described above, the crank angle θivc at the time of closing the intake valve is known because it is determined by the command signal from the engine controller 31 to the intake VTC mechanism 27. By substituting the crank angle θivc (= IVC) at this time into the equations (13) to (17), the volume VIVC of the combustion chamber 5 at the closing timing of the intake valve can be calculated. Therefore, in practice, the volume VIVC of the combustion chamber 5 at the intake valve closing timing is set by a table using the intake valve closing timing IVC as a parameter. When the intake VTC mechanism 27 is not provided, a constant value can be given.

  The temperature calculation unit 56 calculates the temperature TINI [K] of the combustion chamber 5 at the intake valve closing timing IVC. 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 [K] in the intake collector 2 and remaining in the combustion chamber 5 can be approximated by the exhaust temperature TEXH [K] in the vicinity of the exhaust port portion, the intake air in the combustion chamber 5 The temperature TINI at the valve closing timing IVC is a ratio of the fresh air temperature TCOL, the exhaust gas temperature TEXH in the intake collector 2 and the inert gas remaining in the combustion chamber 5 at the timing when the intake valve closing timing IVC is reached. It can obtain | require by following Formula from active gas rate MRESFR.

TINI = TEXH × MRESFR + TCOL × (1−MRESFR)
... (18)
The collector temperature TCOL in the equation (18) is detected by the collector temperature sensor 43, and the exhaust temperature TEXH is detected by the exhaust temperature sensor 45. 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 amount of gas in the combustion chamber, and is calculated using a known method. 5 is calculated at the intake valve closing timing IVC. That is, the collector 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.

  The collector pressure PCOL [Pa] is detected by the collector pressure sensor 44.

The in-cylinder turbulence strength attenuation coefficient calculation unit 58 has a volume VIVC [m ³ ] at the reference crank angle of the combustion chamber 5, a temperature TIVC [K] at the reference crank angle of the combustion chamber 5, and a pressure at the reference crank angle of the combustion chamber 5. The in-cylinder gas mass m [g] is calculated by the above equation (7) using PIVC [kPa], and the in-cylinder gas mass m and the volume VIVC [m ³ ] at the reference crank angle of the combustion chamber 5 are used. In-cylinder gas density ρ [g / m ³ ] is calculated by the above equation (6). On the other hand, a function f2 (TIVC) is obtained by searching a table having the contents shown in FIG. 4 from the temperature TIVC at the reference crank angle of the combustion chamber 5, and the function f2 (TIVC) and the in-cylinder gas density ρ are used to obtain the above function f2. The viscosity coefficient ν is calculated by the equation (5), and this is multiplied by the constant C, that is, the in-cylinder turbulence strength attenuation coefficient MU is calculated by the above equation (4).

  The in-cylinder turbulence intensity attenuation width calculation unit 59 calculates a basic turbulence intensity attenuation width GK0 [m / s] by searching a table having the contents shown in FIG. 3 from the engine speed NRPM. The in-cylinder turbulence intensity attenuation width GK [m / s] is calculated by multiplying the depth attenuation width GK0 and the in-cylinder gas flow state instruction value d, that is, the above equation (2).

  The in-cylinder turbulence intensity setting unit 60 uses the in-cylinder turbulence intensity attenuation width GK [m / s], the above-described in-cylinder turbulence intensity attenuation coefficient MU, and the reference turbulence intensity U0 to calculate the reference crank angle. Is set to a function of the crank angle θ [deg], that is, a function U (θ) [m / s] expressed by the above equation (1).

  The basic ignition timing in-cylinder turbulence intensity calculation unit 61 substitutes the basic ignition timing θMBT [degBTDC] into θ [deg] of the function set in the in-cylinder turbulence intensity setting unit 60, that is, The actual in-cylinder turbulence intensity UMBT [m / s] at the basic ignition timing θMBT is calculated from the equation (8). In this case, since the unit of the reference ignition timing θMBT is [degBTDC], a value [deg] converted from the reference crank angle (IVC) to the crank angle must be substituted as θ [deg]. If the unit of the reference crank angle (intake valve closing timing IVC) is also [degBTDC], it may be calculated to θ by the following equation.

θ = reference crank angle−θMBT (19)
Accordingly, θ obtained by the equation (19) is substituted into the function U (θ) [m / s] set as described above, and the actual in-cylinder turbulence intensity UMBT [m / s] at the basic ignition timing θMBT. s] is calculated.

  The ignition timing correction amount calculation unit 62 uses FIG. 10 as the content from the value obtained by subtracting the target value Um (constant value) of the in-cylinder turbulence intensity at the basic ignition timing from the actual in-cylinder turbulence intensity UMBT at the basic ignition timing. The ignition timing correction amount Uhos [deg] is calculated by searching the table to be executed.

  The ignition timing calculation unit 64 calculates the ignition timing θADV [degBTDC] from the basic ignition timing θMBT [degBTDC] and the ignition timing correction value Uhos [deg] by the following equation.

θADV = θMBT−Uhos (18)
Here, since the unit of the basic ignition timing θMBT is [degBTDC], the ignition timing correction amount Uhos is retarded from the basic ignition timing θMBT when the ignition timing correction amount Uhos is a positive value, and the basic ignition is performed when the ignition timing correction amount Uhos is a negative value. It is advanced from time θMBT.

  As shown in FIG. 10 above, when the actual in-cylinder turbulence intensity UMBT at the basic ignition timing is larger than the target value Um of the in-cylinder turbulence intensity at the basic ignition timing, the ignition timing correction amount Uhos is a positive value (that is, the delay is delayed). Conversely, when the actual in-cylinder turbulence intensity UMBT at the basic ignition timing is smaller than the target value Um of the in-cylinder turbulence intensity at the basic ignition timing, the ignition timing correction amount Uhos is a negative value (that is, the advance amount). ). When the actual in-cylinder turbulence intensity UMBT at the basic ignition timing is larger than the target value Um of the in-cylinder turbulence intensity at the basic ignition timing, the ignition timing correction amount Uhos is set as the retard amount. If the actual in-cylinder turbulence intensity UMBT is larger than the target value Um of the turbulence intensity, the combustion speed becomes higher and the combustion is accelerated too much. Therefore, the ignition timing is retarded and the in-cylinder turbulence intensity target at the basic ignition timing is increased. This is to return the combustion speed to the value Um. On the contrary, when the actual in-cylinder turbulence intensity UMBT at the basic ignition timing is smaller than the target value Um of the in-cylinder turbulence intensity at the basic ignition timing, the ignition timing correction amount Uhos is used as the advance amount. If the actual in-cylinder turbulence intensity UMBT is smaller than the target value Um of the in-cylinder turbulence at the timing, the combustion speed becomes slow and the combustion becomes too slow, so the ignition timing is advanced and the in-cylinder turbulence at the basic ignition timing This is for returning to the combustion speed at the target value Um of strength.

  The ignition timing advance limiter calculation unit 63 calculates an ignition timing advance limiter θMAX [degBTDC]. If the ignition timing is advanced too much, misfires or combustion noises become serious, so a limit value that should not be advanced any more is determined as the advance angle limiter θMAX. That is, the advance angle limiter θMAX is a limiter that takes into account the effects of misfire limit and sound vibration.

  A method of setting the advance angle limiter θMAX will be described with reference to FIG. 11. FIG. 11A shows an attenuation characteristic of the in-cylinder gas turbulence strength in the case of an engine provided with an in-cylinder gas flow strengthening device. In this case, as shown in the figure, since there is a tendency to settle to the convergence value before the basic ignition timing θMBT, the actual in-cylinder turbulence intensity UMBT at the basic ignition timing θMBT is obtained from the basic ignition timing θMBT. The θ when the in-cylinder turbulence intensity becomes the actual in-cylinder turbulence intensity UMBT for the first time is obtained as the advance angle limiter θMAX. Specifically, after obtaining the actual in-cylinder turbulence intensity UMBT at the basic ignition timing from the basic ignition timing θMBT (see a and b), the actual in-cylinder turbulence intensity UMB is set to the left side of the above (1). Substituting and solving the substituted equation for θ, the obtained θ may be used as an advance limiter θMAX (see c and d). The advance limiter θMAX thus obtained is always on the advance side of the basic ignition timing θMBT as shown in the upper part of FIG.

  On the other hand, FIG. 11B shows the damping characteristic of the in-cylinder turbulence strength in the case of an engine that does not include the in-cylinder gas flow strengthening device. In this case, as shown in FIG. It is decaying. Accordingly, at this time, the advance angle limiter θMAX coincides with the basic ignition timing θMBT.

  Accordingly, the ignition timing advance limiter calculation unit 63 looks at the in-cylinder gas flow state instruction value d. When d is a value exceeding 1.0, the advance angle limiter θMAX is set as shown in FIG. In addition, when d = 1, the basic ignition timing θMBT is directly calculated as the advance angle limiter θMAX.

  The ignition timing calculation unit 64 compares the ignition timing θADV [degBTDC] calculated by the above equation (18) with the advance angle limiter θMAX [degBTDC]. When the ignition timing θADV is larger than the advance angle limiter θMAX, Limited to the advance angle limiter θMAX.

  The ignition timing θADV 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 the timing when the actual crank angle coincides with the ignition timing command value is 13 is output.

  Here, the operation of the present embodiment will be described.

  The actual in-cylinder turbulence intensity UMBT at the basic ignition timing θMBT deviates from the target value Um of the in-cylinder turbulence intensity at the basic ignition timing. For example, the actual in-cylinder turbulence intensity UMBT at the basic ignition timing If the in-cylinder turbulence intensity becomes larger than the target value Um, the combustion speed increases and combustion is promoted too much. Conversely, the actual in-cylinder turbulence intensity UMBT at the basic ignition timing becomes in-cylinder turbulence intensity at the basic ignition timing. When it is smaller than the target value Um, the combustion speed becomes slow and the combustion becomes too slow.

  Therefore, until now, it has not been possible to know what the in-cylinder turbulence intensity UMBT at the basic ignition timing θMBT is actually, but according to this embodiment (the invention described in claim 1), The actual in-cylinder turbulence intensity after the intake valve is closed can be set by the function U (θ). Based on this in-cylinder turbulence intensity function U (θ), the actual in-cylinder turbulence intensity UMBT at the basic ignition timing θMBT can be known, and the actual in-cylinder turbulence intensity UMBT at the obtained basic ignition timing can be obtained. By correcting the basic ignition timing θMBT based on the above, it is possible to control the actual in-cylinder turbulence intensity UMBT at the basic ignition timing to the target value Um of the in-cylinder turbulence intensity at the basic ignition timing. Even if the actual in-cylinder turbulence intensity UMBT at the basic ignition timing differs depending on the environmental conditions or changes depending on the operating conditions, the target value Um of the in-cylinder turbulence intensity at the basic ignition timing can be realized.

Corresponding to the fact that the in-cylinder turbulence intensity attenuates from the timing when the intake valve 15 is closed, according to the present embodiment (the invention according to claim 2 ), the reference crank angle is the intake valve closing timing IVC. In addition, the setting of the in-cylinder turbulence strength function U (θ) is facilitated.

According to the present embodiment (the invention described in claim 3 ), the intake VTC mechanism 27 (variable valve operating device) is provided, and the intake valve closing timing is the intake valve closing timing IVC realized by the intake VTC mechanism 27. Therefore, even in an engine equipped with the intake VTC mechanism 27, the function U (θ) of the in-cylinder turbulence strength can be set easily.

According to the present embodiment (the invention described in claim 1 ), the function U (θ) for setting the in-cylinder turbulence intensity is defined as a reference turbulence intensity U0 (in-cylinder turbulence intensity at a reference crank angle), Since it is set based on the in-cylinder turbulence intensity attenuation width GK and the in-cylinder turbulence intensity attenuation coefficient MU, the standard turbulence intensity U0, the in-cylinder turbulence intensity attenuation width GK, and the in-cylinder turbulence intensity attenuation are set depending on the engine model. Even if the coefficient MU may be different, the function U (θ) for setting the in-cylinder turbulence intensity can be appropriately given.

According to the present embodiment (the invention described in claim 4 ), the in-cylinder turbulence strength attenuation coefficient MU is determined based on the volume VIVC at the reference crank angle of the combustion chamber 5, the temperature TIVC at the reference crank angle of the combustion chamber 5, and the combustion chamber. Since the temperature TIVC at the reference crank angle of the combustion chamber 5 may vary depending on environmental conditions, the volume at the reference crank angle of the combustion chamber 5 depends on the operating conditions. Even if the pressure PIVC at the reference crank angle of the VIVC or the combustion chamber 5 is different, the in-cylinder turbulence intensity attenuation coefficient MU can be accurately calculated.

According to the invention of this embodiment (claim 5 ), the in-cylinder turbulence intensity attenuation width GK is calculated based on the engine rotation speed NRPM. Therefore, even if the engine rotation speed NRPM is different, the in-cylinder disturbance The strength attenuation width GK can be appropriately given.

According to the present embodiment (the invention described in claim 6 ), the reference turbulence intensity U0 is calculated based on the reference crank angle and the engine rotational speed NRPM. Therefore, the reference crank angle and the engine rotational speed NRPM are different. However, the reference disturbance strength U0 can be appropriately given.

According to the present embodiment (the invention described in claim 8 ), as shown in FIG. 5, the ignition timing correction means is configured such that the actual in-cylinder turbulence intensity UMBT at the basic ignition timing θMBT is the cylinder at the basic ignition timing θMBT. An ignition timing correction amount calculating unit 62 as an ignition timing correction amount calculating means for calculating the ignition timing correction amount Uhos so as to coincide with the target value Um of the internal disturbance strength, and the basic ignition timing θMBT with the ignition timing correction amount Uhos. Since the ignition timing calculation unit 64 is provided as the ignition timing correction means for correcting, the actual in-cylinder turbulence intensity UMBT at the basic ignition timing θMBT may deviate from the target value Um of the in-cylinder turbulence intensity at the basic ignition timing θMBT. Even in this case, the actual in-cylinder turbulence intensity UMBT at the basic ignition timing θMBT quickly returns to the target value Um of the in-cylinder turbulence intensity at the basic ignition timing θMBT. From being, it can be realized in the turbulence intensity cylinder in always unchanged basic ignition timing ShitaMBT.

In the present embodiment (when the in-cylinder gas flow strengthening device such as the swirl control valve and the tumble control valve is provided in the engine, the in-cylinder turbulence strength is enhanced as compared with the case where they are not provided. According to the eleventh aspect of the present invention, the in-cylinder turbulence intensity attenuation width GK is also calculated based on the in-cylinder gas flow state indication value d representing the in-cylinder gas flow strength by the in-cylinder gas flow strengthening device. Therefore, even when the in-cylinder gas flow strengthening device is provided in the engine, the in-cylinder turbulence intensity attenuation width GK can be appropriately given.

The in-cylinder turbulence intensity basically decreases as the latter stage of the compression stroke, but when the in-cylinder gas flow strengthening device is provided in the engine, it tends to settle to the convergence value on the advance side of the basic ignition timing. It has been found through experiments. By utilizing this tendency, according to the present embodiment (the invention described in claim 12 ), the in-cylinder gas flow state indication value d indicating the in-cylinder gas flow strength by the in-cylinder gas flow strengthening device, and the basic ignition timing Based on the actual in-cylinder turbulence intensity UMBT and a function U (θ) for setting the in-cylinder turbulence intensity, the advance angle limiter θMAX is calculated, and the ignition that is the basic ignition timing corrected by the ignition timing correction value Uhos When the timing θADV advances beyond the advance angle limiter θMAX, the advance angle limiter θMAX is limited to the advance angle limiter θMAX, so that the advance timing limiter θMAX of the ignition timing can be easily calculated.

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

  In the embodiment, the basic ignition timing θMBT is calculated based on the charging efficiency detection value ITACCr and the engine rotational speed NRPM. However, the present invention is not limited to this, and instead of the charging efficiency detection value ITACCr, the engine load is calculated. Can be used. Moreover, although the case where the basic ignition timing is an ignition timing at which MBT is obtained has been described, the present invention is not limited to this.

  In the first aspect of the invention, the function of the in-cylinder turbulence intensity setting means is performed by the in-cylinder turbulence intensity setting unit 60 of FIG. 5, and the function of the basic ignition timing calculation means is performed by the basic ignition timing calculation unit 52 of FIG. The function of the basic ignition timing in-cylinder turbulence intensity calculating means is based on the basic ignition timing in-cylinder turbulence intensity calculating section 61 of FIG. 5, and the function of the ignition timing correcting means is the ignition timing correction amount calculating section 62 and the ignition timing in FIG. Each calculation is performed by the calculation unit 64.

The engine control system figure of one Embodiment. The wave form diagram which shows the change of the turbulence intensity in a cylinder after an intake valve closes. The characteristic diagram of basic turbulence intensity attenuation width. The characteristic view of the function f2 (TIVC). The block diagram for demonstrating calculation of ignition timing. The detailed block diagram of a filling efficiency detection part. The characteristic diagram of basic ignition timing. The characteristic diagram of standard turbulence strength. Diagram explaining the positional relationship between the crankshaft of the engine and the connecting rod. The characteristic diagram of ignition timing correction amount. The characteristic view for demonstrating the calculation method of an advance angle limiter.

Explanation of symbols

1 Engine 5 Combustion chamber 11 Ignition device (spark ignition means)
15 Intake valve 27 Intake VTC mechanism (variable valve operating device)
31 Engine controller

Claims (12)

In-cylinder turbulence intensity setting means for setting an actual in-cylinder turbulence intensity after the intake valve is closed by a function;
Basic ignition timing calculation means for calculating basic ignition timing based on operating conditions;
A basic ignition timing in-cylinder turbulence intensity calculating means for calculating an actual in-cylinder turbulence intensity at the basic ignition timing using the function of the set in-cylinder turbulence intensity;
Ignition timing correction means for correcting the basic ignition timing based on the actual in-cylinder turbulence intensity at the basic ignition timing;
Spark ignition means for performing spark ignition at the corrected basic ignition timing ,
The function for setting the in-cylinder turbulence strength uses the crank angle from the reference crank angle as a variable,
The function for setting the in-cylinder turbulence intensity is set based on the in-cylinder turbulence intensity at the reference crank angle, the in-cylinder turbulence intensity attenuation width, and the in-cylinder turbulence intensity attenuation coefficient. Engine ignition timing control device. 2. The engine ignition timing control device according to claim 1, wherein the reference crank angle is an intake valve closing timing. Equipped with a variable valve device capable of variably controlling the intake valve closing timing, the intake valve closing timing is according to claim 1, characterized in that the intake valve closing timing which is implemented by the variable valve device Engine ignition timing control device. The in-cylinder turbulence intensity attenuation coefficient is calculated based on a volume of the combustion chamber at the reference crank angle, a temperature of the combustion chamber at the reference crank angle, and a pressure of the combustion chamber at the reference crank angle. Item 2. The ignition timing control device for an engine according to Item 1 . The engine ignition timing control device according to claim 1 , wherein the in-cylinder turbulence intensity attenuation width is calculated based on an engine rotation speed . The reference in-cylinder turbulence intensity of the crank angle, an ignition timing control device for an engine according to claim 1, characterized in that calculated on the basis of said reference crank angle and the engine rotational speed. 2. The engine ignition timing control device according to claim 1 , wherein the basic ignition timing is an ignition timing at which MBT is obtained . The ignition timing correction means includes
Ignition timing correction amount calculating means for calculating the ignition timing correction amount so that the actual in-cylinder turbulence intensity at the ignition timing at which the MBT is obtained matches the target value of the in-cylinder turbulence intensity at the ignition timing at which the MBT is obtained. When,
Ignition timing correction means for correcting the ignition timing obtained by the MBT with this ignition timing correction amount;
Ignition timing control device for an engine according to claim 7, characterized in that it comprises a. When the actual in-cylinder turbulence intensity at the ignition timing at which the MBT is obtained is larger than the target value of the in-cylinder turbulence intensity at the ignition timing at which the MBT is obtained, the ignition timing correction amount is a retard amount. The engine ignition timing control device according to claim 8 . When the actual in-cylinder turbulence intensity at the ignition timing at which the MBT is obtained is smaller than the target value of the in-cylinder turbulence intensity at the ignition timing at which the MBT is obtained, the ignition timing correction amount is an advance amount. The engine ignition timing control device according to claim 8 . An in-cylinder gas flow strengthening device is provided, and the in-cylinder turbulence intensity attenuation width is also calculated based on an in-cylinder gas flow state indication value representing the in-cylinder gas flow strength by the in-cylinder gas flow strengthening device. The engine ignition timing control device according to claim 1 . An in-cylinder gas flow strengthening device is provided, the in-cylinder gas flow state indicating value indicating the in-cylinder gas flow strength by the in-cylinder gas flow strengthening device, the actual in-cylinder turbulence intensity at the basic ignition timing, and the in-cylinder turbulence The advance angle limiter is calculated based on a function for setting the strength, and when the corrected basic ignition timing advances beyond the advance angle limiter, the advance angle limiter is limited to the advance angle limiter. ignition timing control system for an engine according to 1.

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