JP2005207366A - Control device for internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To compatibly suppress torque fluctuation and suppress HC emission by accurately determining lean-burn limit not to vary a combustion condition. <P>SOLUTION: This control device for an internal combustion engine determines whether combustion is appropriate normal combustion, partial combustion or rapid combustion for each combustion in each cylinder in steady operation (for example cold fast idling). Based on the determination result, frequency of occurrence of partial combustion and rapid combustion to normal combustion is calculated. Based on frequency of occurrence of partial combustion and rapid combustion to normal combustion, target air fuel ratio is corrected. <P>COPYRIGHT: (C)2005,JPO&NCIPI

Description

  The present invention relates to a control technique for an internal combustion engine, and more particularly to a control technique for an internal combustion engine that is suitable for use in suppressing HC emissions during idling.

  2. Description of the Related Art Conventionally, in an internal combustion engine, lean air-fuel ratio and retarded ignition timing are performed in cold first idle immediately after cold start from the viewpoint of exhaust countermeasures. Making the air-fuel ratio lean is effective in suppressing the emission of HC, which is a harmful substance, and retarding the ignition timing enables early activation of the catalyst by raising the exhaust temperature.

  However, lean air-fuel ratio is effective in suppressing HC emission, but also makes combustion unstable and causes the combustion state of the internal combustion engine to fluctuate. The fluctuation of the combustion state not only causes torque fluctuation and decreases drivability, but also may cause deterioration of exhaust gas performance. For this reason, in the air-fuel ratio control at the time of cold first idling, it is desirable to determine the lean limit that does not change the combustion state and to control the air-fuel ratio so as not to exceed the limit.

Conventionally, for example, a technique described in Patent Document 1 is known as a technique for determining the limit of leaning and suppressing the air-fuel ratio. In the technique described in Patent Document 1, the in-cylinder pressure is integrated in a predetermined integration interval to obtain an in-cylinder pressure integrated value, and the standard deviation of the in-cylinder pressure integrated value during a predetermined cycle and the average value of the in-cylinder pressure integrated value during a predetermined cycle. And the value obtained by dividing the standard deviation by the average value is used as an index for determining the limit of output fluctuation. In addition, the air-fuel ratio is made lean as much as possible within a range that does not exceed the output fluctuation limit, thereby achieving both suppression of torque fluctuation and suppression of HC emission.
JP-A-8-232752

  However, as an index for discriminating the limit of leaning, rather than taking into account the individual state quantities of the combustion state as in the prior art described above, it is rather the case that the variation in the combustion state itself is used as an index. It is considered that the limit can be accurately determined. In the above prior art, the standard deviation and average value of the in-cylinder pressure integral values, which are individual state quantities of the combustion state, are used as the index value. In this case, a large increase / decrease in torque is greatly reflected in the index value. It may end up.

  The present invention has been made to solve the above-described problems, and can accurately determine the limit of leaning that does not fluctuate the combustion state so as to achieve both suppression of torque fluctuation and suppression of HC emission. It is an object of the present invention to provide a control device for an internal combustion engine.

In order to achieve the above object, a first invention is a control device for an internal combustion engine,
Combustion state determining means for determining whether the combustion is proper normal combustion, partial combustion, or rapid combustion for each combustion of each cylinder at a constant time;
An occurrence frequency calculating means for calculating an occurrence frequency of partial combustion and rapid combustion with respect to normal combustion from the determination result of the combustion state determining means;
And target air-fuel ratio correcting means for correcting the target air-fuel ratio based on the occurrence frequency.

In a second aspect based on the first aspect, the occurrence frequency calculating means calculates the occurrence frequency for each cylinder,
The target air-fuel ratio correcting means corrects the target air-fuel ratio for each cylinder based on the occurrence frequency obtained for each cylinder.

  According to the first invention, since the occurrence frequency of partial combustion and rapid combustion with respect to normal combustion is used as an index value, the limit of leaning that does not change the combustion state without being affected by the individual state quantities of the combustion state is set. It can be determined accurately. Then, by correcting the target air-fuel ratio based on the frequency of occurrence, it becomes possible to reduce the air-fuel ratio to the limit and suppress the discharge of HC without causing torque fluctuation.

  Further, according to the second invention, the limit of leaning of each cylinder can be individually set and the air-fuel ratio can be controlled for each cylinder, so that it is possible to further suppress HC emission. .

Embodiment 1 FIG.
Embodiment 1 of the present invention will be described below with reference to the drawings.
FIGS. 1-6 is a figure for demonstrating the control apparatus of the internal combustion engine as Embodiment 1 of this invention. The control device for an internal combustion engine of the present embodiment is configured as an ECU (Electronic Control Unit). The ECU 2 comprehensively controls various actuators related to the operating state of the internal combustion engine based on the output signals of the plurality of sensors. In the present embodiment, as shown in FIG. 2, the A / F sensor 4 and the crank angle sensor 6 are connected to the input side of the ECU 2, and the fuel injection valve 10 is connected to the output side thereof. The A / F sensor 4 is provided in the exhaust passage of the internal combustion engine, and outputs a signal corresponding to the air-fuel ratio of the exhaust gas to the ECU 2. The crank angle sensor 6 is provided in the vicinity of the crankshaft, and outputs a signal to the ECU 2 at a predetermined crank angle position. The ECU 2 receives an output signal from the A / F sensor 4 and the crank angle sensor 6 and also supplies a drive signal to the fuel injection valve 10. The fuel injection valve 10 is provided for each cylinder, but only one is shown here representatively. In addition to the sensors 4 and 6 and the fuel injection valve 10, a plurality of sensors (for example, in-cylinder pressure sensors) and actuators (for example, spark plugs and ISC valves) are connected to the ECU 2. Is omitted.

  The ECU 2 performs combustion stabilization control that stabilizes combustion by controlling the air-fuel ratio and suppresses fluctuations in the combustion state in a situation where the combustion state of the internal combustion engine is fluctuating as shown in FIG. FIG. 1 is a graph showing a change in torque of a specific cylinder that occurs when the air-fuel ratio is excessively leaned in cold first idling. In FIG. 1, the torque region sandwiched between the upper and lower threshold values indicates a normal combustion region (appropriate torque region) during idling, a region higher than the normal combustion region is a rapid combustion region, and a region lower than the normal combustion region is partial combustion. Indicates the area. Here, partial combustion means combustion in which a large torque is not generated, and rapid combustion means combustion in which a large torque is generated. In the figure, cycles that have entered the normal combustion region are indicated by white circles, and cycles that have deviated from the normal combustion region are indicated by black circles. As shown in this figure, when the torque decreases due to partial combustion, in the next cycle, the combustion of that cylinder becomes rapid combustion, and the torque increases rapidly. Conversely, when the torque increases due to rapid combustion, in the next cycle, the combustion in that cylinder becomes partial combustion, and the torque decreases rapidly.

  The reason why rapid combustion occurs in the next cycle of partial combustion is that in partial combustion, the exhaust gas temperature is increased by the amount that fuel energy was not used to generate torque, so the residual gas ratio in the next cycle decreased. This is probably because the temperature of unburned gas increases. The increase in the equivalent ratio in the next cycle due to the generation of a large amount of residual HC is considered to be one of the reasons why rapid combustion occurs in the next cycle. In addition, it is considered that the partial combustion occurs in the next cycle of the rapid combustion due to the opposite reason. In any case, when the air-fuel ratio is made excessively lean in the cold first idling, partial combustion and rapid combustion occur alternately, which induces torque fluctuation and reduces drivability.

  The outline of the combustion stabilization control according to the present embodiment is that the target exhaust air-fuel ratio in the air-fuel ratio F / B control is corrected to the lean side until the occurrence frequency of partial combustion and rapid combustion with respect to normal combustion exceeds a predetermined threshold, When the occurrence frequency exceeds a predetermined threshold, the target exhaust air-fuel ratio is corrected to the rich side. Specifically, the content of the combustion stabilization control performed by the ECU 2 in the present embodiment is shown in the flowchart of FIG. Note that the combustion stabilization control routine of FIG. 3 is executed in units of an expansion stroke (for example, 180 degrees for the four cylinders and 120 degrees for the six cylinders) during the cold first idle where the combustion is particularly likely to become unstable.

  First, in step 100, it is determined whether or not the internal combustion engine has been started. The determination in the first step 100 is executed when the ECU 2 is activated by turning on the ignition switch. Prior to starting the internal combustion engine, the ECU 2 sets a target exhaust air-fuel ratio AFr to a predetermined value (for example, theoretical air-fuel ratio) as an initial value of each parameter related to the combustion stabilization control, and sets count values N1 and N2 described later to 0. Set to.

  If the start of the internal combustion engine is detected in the determination in step 100, it is determined in the next step 102 whether or not air-fuel ratio F / B (feedback) control is started. The air-fuel ratio F / B control is a control for correcting the fuel injection amount in accordance with the deviation between the actual exhaust air-fuel ratio and the target exhaust air-fuel ratio so that the actual exhaust air-fuel ratio becomes the target exhaust air-fuel ratio. The actual exhaust air-fuel ratio is detected by the A / F sensor 4, but the sensor element of the A / F sensor 4 is not activated unless it rises to a predetermined temperature. For this reason, after the internal combustion engine is started, the air-fuel ratio F / B control cannot be performed until the exhaust gas temperature rises and the A / F sensor 4 is activated. The ECU 2 controls the fuel injection amount by open loop control until the air-fuel ratio F / B control is started.

  As a result of the determination in step 102, when the air-fuel ratio F / B is started, the torque of the internal combustion engine is calculated (step 104). For example, as described below, the torque can be calculated from an output signal (crank angle signal) supplied from the crank angle sensor 6 according to an equation of motion.

The following formulas (1) and (2) are formulas for calculating torque from the output signal supplied from the crank angle sensor 6.
Ti = J × (dω / dt) + Tf + Tl (1)
Ti = Tgas + Tinertia (2)
In the above equations (1) and (2), Ti is the indicated torque generated on the crankshaft by engine combustion. Here, the right side of equation (2) indicates the torque that generates the indicated torque Ti, and the right side of equation (1) indicates the torque that consumes the indicated torque Ti.

  In the right side of equation (1), J is the moment of inertia of the drive member driven by the combustion of the air-fuel mixture, dω / dt is the angular acceleration of the crankshaft, Tf is the friction torque of the drive unit, and Tl is the load received from the road surface during travel Torque. Here, J × (dω / dt) is a dynamic loss torque (= Tac) resulting from the angular acceleration of the crankshaft. The friction torque Tf is a torque due to mechanical friction of each fitting portion such as friction between the piston and the inner wall of the cylinder, and includes torque due to mechanical friction of auxiliary machinery. The load torque Tl is a torque due to a disturbance such as a road surface condition during traveling. Since the shift gear is in the neutral state during cold first idling, Tl = 0 is assumed in the following description.

  In the right side of the equation (2), Tgas represents torque due to cylinder cylinder gas pressure, and Tinertia represents inertial torque due to reciprocating inertial mass such as a piston. The torque Tgas due to the in-cylinder gas pressure is a torque generated by the combustion of the air-fuel mixture in the cylinder. In order to accurately estimate the combustion state, it is necessary to obtain the torque Tgas by the cylinder gas pressure.

  As shown in the equation (1), the indicated torque Ti can be obtained as the sum of dynamic loss torque J × (dω / dt) due to angular acceleration, friction torque Tf, and load torque Tl. However, as shown in the equation (2), the indicated torque Ti and the torque Tgas due to the in-cylinder gas pressure do not match, so the combustion state cannot be accurately estimated from the indicated torque Ti.

  FIG. 4 is a characteristic diagram showing the relationship between each torque and crank angle in equation (2). In FIG. 4, the vertical axis indicates the magnitude of each torque, the horizontal axis indicates the crank angle, the one-dot chain line in FIG. 4 indicates the indicated torque Ti, the solid line indicates the torque Tgas due to the in-cylinder gas pressure, and the broken line reciprocates. Inertia torque Tinertia due to inertial mass is shown. Here, FIG. 4 shows the characteristics in the case of four cylinders. TDC and BDC in FIG. 4 are the top dead center (TDC) or the bottom dead center of the piston 34 of one of the four cylinders. The crank angle (0 °, 180 °) in the (BDC) position is shown. When the internal combustion engine has four cylinders, an explosion stroke is performed for each cylinder every time the crankshaft rotates 180 °, and torque characteristics from TDC to BDC in FIG.

  As shown by the solid line in FIG. 4, the torque Tgas due to the in-cylinder gas pressure rapidly increases and decreases between TDC and BDC. Here, the rapid increase in Tgas is due to the explosion of the air-fuel mixture in the combustion chamber during the explosion stroke. After the explosion, Tgas decreases and becomes a negative value due to the influence of the cylinders in other compression strokes or exhaust strokes. Thereafter, when the crank angle reaches BDC, the change in volume of the cylinder becomes zero and Tgas becomes zero.

  On the other hand, the inertia torque Tinertia due to the reciprocating inertial mass is an inertial torque generated by the inertial mass of a reciprocating member such as a piston, regardless of the torque Tgas due to the in-cylinder gas pressure. The reciprocating member repeatedly accelerates and decelerates, and Tinertia always occurs as long as the crank rotates, even if the angular velocity is constant. As shown by the broken line in FIG. 4, the member that reciprocates is stopped at the position where the crank angle is TDC, and Tinertia = 0. When the crank angle advances from TDC toward BDC, the reciprocating member starts to move from the stopped state. At this time, Tinertia increases in the negative direction due to the inertia of these members. When the crank angle reaches around 90 °, the reciprocating members are moving at a predetermined speed, so that the crankshaft rotates due to the inertia of these members. Therefore, Tinertia changes from a negative value to a positive value between TDC and BDC. After that, when the crank angle reaches BDC, the reciprocating member stops and Tinertia = 0.

  As shown in the equation (2), the indicated torque Ti is the sum of the torque Tgas caused by the in-cylinder gas pressure and the inertia torque Tinertia caused by the reciprocating inertial mass. For this reason, as shown by the one-dot chain line in FIG. 4, between TDC and BDC, the indicated torque Ti increases due to an increase in Tgas due to the explosion of the mixture, once decreases, and then increases again due to Tinertia. The behavior is shown.

  However, when attention is paid to a section with a crank angle of 180 ° from TDC to BDC, the average value of inertia torque Tinertia due to reciprocating inertia mass in this section is zero. This is because a member having a reciprocating inertia mass moves in the opposite direction at a crank angle of 0 ° to 90 ° and a crank angle of 90 ° to 180 °. Therefore, when the torques in the equations (1) and (2) are calculated as average values from TDC to BDC, the inertia torque Tinertia = 0 by the reciprocating inertia mass can be calculated. As a result, the influence of the inertia torque Tinertia due to the reciprocating inertia mass on the indicated torque Ti can be eliminated, and an accurate combustion state can be easily estimated.

  Then, when the average value of each torque is obtained in the section from TDC to BDC, the average value of Tinertia becomes 0. From the equation (2), the average value of the indicated torque Ti and the average of the torque Tgas due to the in-cylinder gas pressure are obtained. The value becomes equal. Therefore, the combustion state can be accurately estimated based on the indicated torque Ti.

  Further, when the average value of the angular acceleration of the crankshaft in the section from TDC to BDC is obtained, the average value of Tinertia in this section is 0, so that the influence of the reciprocating inertia mass on the angular acceleration is eliminated. Can be requested. Therefore, it is possible to calculate the angular acceleration caused only by the combustion state, and it is possible to accurately estimate the combustion state based on the angular acceleration.

  Next, a method for calculating each torque on the right side of the equation (1) will be described. First, a method of calculating dynamic loss torque Tac = J × (dω / dt) resulting from angular acceleration will be described. FIG. 5 is a schematic diagram illustrating a method for obtaining the angular acceleration of the crankshaft, and is a diagram illustrating a crank angle signal and torque calculation timing. As shown in FIG. 5, in this embodiment, a crank angle signal is supplied from the crank angle sensor 6 every 10 ° of rotation of the crankshaft.

The ECU 2 calculates the dynamic loss torque Tac caused by the angular acceleration as an average value from TDC to BDC. For this purpose, the apparatus of this embodiment obtains angular velocities ω ₀ (k) and ω ₀ (k + 1) at two crank angle positions of TDC and BDC, and the crankshaft rotates from TDC to BDC at the same time. Time Δt (k) is obtained.

When obtaining the angular velocity ω ₀ (k), for example, as shown in FIG. 5, the times Δt ₀ (k) and Δt ₁₀ (k) during which the crank angle is rotated by 10 ° forward and backward from the TDC position are calculated. It is detected from the crank angle sensor 6. Since the crankshaft rotates 20 ° during the time Δt ₀ (k) + Δt ₁₀ (k), ω ₀ (k) = (20 / (Δt ₀ (k) + Δt ₁₀ (k)) × ( π / 180) can be calculated to calculate ω ₀ (k) [rad / s] Similarly, when calculating ω ₀ (k + 1), the crank angle is 10 ° forward and backward from the BDC position. Times Δt ₀ (k + 1) and Δt ₁₀ (k + 1) during rotation are detected, and ω ₀ (k + 1) = (20 / (Δt ₀ (k + 1) + Δt ₁₀ ( k + 1))) × (π / 180) can be calculated to calculate ω ₀ (k + 1) [rad / s], and angular velocities ω ₀ (k) and ω ₀ (k + 1) were obtained. After that, (ω ₀ (k + 1) −ω ₀ (k)) / Δt (k) is calculated, and the average value of the angular acceleration during the rotation of the crankshaft from TDC to BDC is calculated.

  After the average value of angular acceleration is obtained, the average value of angular acceleration and the moment of inertia J are multiplied according to the right side of equation (1). As a result, an average value of dynamic loss torque J × (dω / dt) while the crankshaft rotates from TDC to BDC can be calculated. The inertia moment J of the drive unit is obtained in advance from the inertia mass of the drive component.

  Next, a method for calculating the friction torque Tf will be described. FIG. 6 is a map showing the relationship between the friction torque Tf, the engine speed Ne of the internal combustion engine, and the coolant temperature thw. In FIG. 6, the friction torque Tf, the engine speed Ne, and the coolant temperature thw are average values when the crankshaft rotates from TDC to BDC. Further, the cooling water temperature becomes higher in the order of thw1 → thw2 → thw3. As shown in FIG. 6, the friction torque Tf increases as the engine speed (Ne) increases, and increases as the cooling water temperature thw decreases. In the map of FIG. 6, the engine rotational speed Ne and the coolant temperature thw are varied as parameters, the friction torque Tf generated when the crankshaft is rotated from TDC to BDC is measured, and the average value is calculated in advance. Create it. When estimating the combustion state, the average value of the cooling water temperature and the average value of the engine speed in the section from TDC to BDC are applied to the map of FIG. 6 to obtain the average value of the friction torque Tf. At this time, the coolant temperature is detected from a water temperature sensor (not shown), and the engine speed is detected from a crank angle sensor 6.

  The behavior of the friction torque Tf accompanying the variation of the crank angle is very complicated and has a large variation. However, since the behavior of the friction torque Tf mainly depends on the speed of the piston, the average value of the friction torque Tf for each section where the average value of the inertia torque Tinertia due to the reciprocating inertia mass is zero is substantially constant. Therefore, by obtaining the average value of the friction torque Tf for each section (TDC → BDC) where the average value of the inertia torque Tinertia due to the reciprocating inertia mass is 0, the friction torque Tf showing a complex instantaneous behavior can be obtained with high accuracy. it can. Further, by making the friction torque Tf an average value for each section, the map shown in FIG. 6 can be created accurately.

  Further, as described above, the friction torque Tf includes torque due to friction of auxiliary machinery. Here, the value of the torque due to the friction of the auxiliary machines varies depending on whether or not the auxiliary machines are operating. For example, the rotation of the engine is transmitted to a compressor of an air conditioner, which is one of the auxiliary machines, by a belt or the like, and torque due to friction is generated even when the air conditioner is not actually operating.

  On the other hand, when the auxiliary machines are operated, for example, when the air conditioner switch is turned on, the torque consumed by the compressor becomes larger than in the state where the air conditioner is not operated. For this reason, the torque due to the friction of the auxiliary machinery increases, and the value of the friction torque Tf also increases. Therefore, in order to accurately determine the friction torque Tf, the operation state of the auxiliary machinery is detected, and when the auxiliary machinery is switched on, the value of the friction torque Tf obtained from the map of FIG. It is desirable to correct.

  It is more preferable to correct the friction torque Tf in consideration of the difference between the temperature of the portion where the friction torque Tf is actually generated and the cooling water temperature at the time of extremely cold start. In this case, it is desirable to perform correction in consideration of the engine start time after the cold start, the in-cylinder inflow fuel amount, and the like.

  In the present embodiment, the indicated torque (also referred to as estimated indicated torque) Ti is used as a representative value of the engine torque, and the ECU 2 calculates the torque (indicated torque Ti) for each cylinder. In the next step 106, it is determined from the calculated torque whether the combustion state of the internal combustion engine is rapid combustion, partial combustion, or appropriate normal combustion. Specifically, the calculated torque is first compared with a predetermined determination value T1, and when the torque is smaller than the determination value T1, it is determined that the combustion state in the cycle is partial combustion. Determination value T1 is a lower limit value of the appropriate torque region. When it is determined that the partial combustion is not performed, the torque is further compared with a predetermined determination value T2, and when the torque is larger than the determination value T2, it is determined that the combustion state in the cycle is rapid combustion. The The determination value T2 is an upper limit value of the appropriate torque region, and is naturally a value larger than the determination value T1. If neither partial combustion nor rapid combustion is determined, that is, if the torque is not less than the determination value T1 and not more than T2, it is determined that the combustion state is normal combustion.

  If the result of determination in step 106 is normal combustion, 1 is added to the count value N1 (step 108). If rapid combustion or partial combustion is determined, 1 is added to the count value N2. (Step 110). In the next step 112, the total value of the two count values N1 and N2 is compared with a predetermined threshold value. The total value of the count values N1 and N2 represents the number of torque calculations since the start of the air-fuel ratio F / B control, and the threshold value is a torque calculation that can collect enough data to perform the processing after step 114. Is set to the number of times.

  If the total value of the count values N1 and N2 exceeds the threshold value as a result of the determination in step 112, first, in step 114, a ratio R (R = N2 / N1) of rapid combustion and partial combustion to normal combustion is calculated. In step 116, the rapid / partial combustion ratio R is compared with a predetermined threshold value Ro. The rapid / partial combustion ratio R changes depending on the air-fuel ratio, and the HC emission amount changes as the air-fuel ratio changes. The threshold value Ro is a value obtained by experiment or the like for the ratio of rapid combustion and partial combustion to normal combustion when both drivability and HC emission are at satisfactory levels.

  As a result of the comparison in step 116, when the rapid / partial combustion ratio R exceeds the threshold value Ro, the target exhaust air / fuel ratio AFr for the air / fuel F / B control is corrected to the rich side (step 118). That is, a value obtained by subtracting the predetermined value α from the current target exhaust air-fuel ratio AFr is set as a new target exhaust air-fuel ratio AFr. Thereby, in the next fuel injection, the fuel injection amount is increased by the air-fuel ratio F / B control, and the combustion is stabilized by enrichment of the actual air-fuel ratio. Moreover, the ratio R with respect to normal combustion of rapid combustion and partial combustion will fall because combustion becomes stable.

  On the other hand, if it is determined in step 116 that the rapid / partial combustion ratio R is equal to or less than the threshold value Ro, the target exhaust air / fuel ratio AFr for air / fuel F / B control is corrected to the lean side (step 120). ). That is, a value obtained by adding the predetermined value α to the current target exhaust air-fuel ratio AFr is set as a new target exhaust air-fuel ratio AFr. Thereby, in the next fuel injection, the fuel injection amount is reduced by the air-fuel ratio F / B control, and the discharge of HC is suppressed. In addition, since the combustion becomes unstable due to the lean air-fuel ratio, the ratio R of normal combustion to rapid combustion or partial combustion increases. In this embodiment, the target exhaust air-fuel ratio AFr is corrected by adding / subtracting the predetermined value α. However, the target exhaust air-fuel ratio AFr is obtained by multiplying the reference air-fuel ratio by a coefficient, and the coefficient is increased or decreased. Thus, the target exhaust air-fuel ratio AFr may be corrected.

  By executing the combustion stabilization control routine described above, the target exhaust air-fuel ratio AFr at the time of air-fuel ratio F / B control is set so that the ratio R of rapid combustion and partial combustion to normal combustion converges to the threshold value Ro. As a result, the exhaust air-fuel ratio can be leaned to the limit without impairing drivability, and an ideal combustion state that satisfies both drivability and HC emissions can be obtained.

  Further, according to the control device of the present embodiment, since the rapid / partial combustion ratio R indicating the frequency of occurrence of partial combustion and rapid combustion with respect to normal combustion is used as an index value, it is influenced by the individual state quantities of the combustion state. There is also an advantage that the limit of leaning can be accurately determined.

  In the above-described embodiment, the “combustion state determining means” of the first invention is realized by the execution of the processing of steps 104 and 106 by the ECU 2. Further, the “occurrence frequency calculation means” of the first invention is realized by the execution of the processing of steps 108, 110, 112 and 114 by the ECU 2. Furthermore, the “target air-fuel ratio correcting means” according to the first aspect of the present invention is realized by executing the processing of steps 116, 118, and 120 by the ECU 2.

Embodiment 2.
The second embodiment of the present invention will be described below with reference to FIG.
The control apparatus for an internal combustion engine of the present embodiment can be realized by causing the ECU 2 to execute the routine of FIG. 7 instead of the routine of FIG. 3 in the first embodiment.

  In the first embodiment, the combustion stabilization control is performed after the start of the air-fuel ratio F / B control. However, in order to further suppress the emission of HC, the air-fuel ratio is reduced even before the start of the air-fuel ratio F / B control. I want to make lean. In this case, however, the air / fuel ratio is not simply made lean, but it is desired to make the air / fuel ratio lean to the limit without impairing drivability. Therefore, in the present embodiment, the air-fuel ratio can be leaned to the limit even before the start of the air-fuel ratio F / B control by the combustion stabilization control described below.

  FIG. 7 is a flowchart for explaining the flow of the combustion stabilization control executed by the ECU 2 as the control device for the internal combustion engine in the present embodiment. The combustion stabilization control routine of FIG. 7 is executed in units of an expansion stroke (for example, 180 degrees for the four cylinders and 120 degrees for the six cylinders) during the cold first idle where the combustion is likely to be particularly unstable.

  First, at step 200, it is determined whether or not the internal combustion engine has been started. The determination in the first step 200 is executed when the ECU 2 is activated by turning on the ignition switch. When the ECU 2 is activated by turning on the ignition switch, the target in-cylinder air-fuel ratio AFc is set to a predetermined value (for example, theoretical air-fuel ratio) as an initial value of each parameter relating to the present combustion stabilization control prior to starting the internal combustion engine. The count values N1 and N2 are set to 0.

  If the start of the internal combustion engine is detected in the determination in step 200, the in-cylinder air amount mc (k) is calculated in the next step 202. k indicates the number of calculations since the start. The in-cylinder air amount mc (k) can be calculated from, for example, a detection signal of an air flow meter that detects the intake air amount in the intake passage. However, since there is a time lag until the air passes through the air flow meter and flows into the cylinder, the ECU 2 does not use the detection signal of the air flow meter as it is, but uses the physical model of the intake system and considers the time lag. The cylinder air amount mc (k) is estimated.

  In step 204, the target in-cylinder fuel amount fc (k) is calculated by dividing the in-cylinder air amount mc (k) calculated in step 202 by the current target in-cylinder air-fuel ratio AFc. The target in-cylinder fuel amount fc (k) is an in-cylinder fuel amount necessary to realize the target in-cylinder air-fuel ratio AFc. However, the target in-cylinder fuel amount fc (k) The amount of fuel injected from the injection valve 10 (fuel injection amount) does not necessarily match the in-cylinder fuel amount. Therefore, in the present embodiment, the relationship between the fuel injection amount and the in-cylinder fuel amount is set in advance as a fuel model through experiments or the like. In step 206, the target in-cylinder fuel amount fc (k) is used using the fuel model. A fuel injection amount for realizing the above is calculated. The calculated fuel injection amount is output to the fuel injection valve 10 of the cylinder in which the current fuel injection stroke is performed, and fuel injection is executed in the cylinder.

  After the fuel injection in step 206, the same processing as the processing in steps 104 to 114 according to the first embodiment is performed (step 208). By this process, it is determined whether the combustion state of the internal combustion engine obtained by the fuel injection in step 206 is proper normal combustion, partial combustion, or rapid combustion. A rapid / partial combustion ratio R indicating the frequency of occurrence of combustion is calculated.

  In step 210, the rapid / partial combustion ratio R is compared with a predetermined threshold value Ro. As a result of the comparison in step 210, when the rapid / partial combustion ratio R exceeds the threshold value Ro, the target in-cylinder air-fuel ratio AFc is corrected to the rich side (step 212). That is, a value obtained by subtracting the predetermined value α from the current target cylinder air-fuel ratio AFc is set as a new target cylinder air-fuel ratio AFc. As a result, the next target in-cylinder fuel amount fc (k + 1) is increased from the current target in-cylinder fuel amount fc (k), and the stabilization of combustion is achieved by the enrichment of the in-cylinder air-fuel ratio. Figured. Moreover, the ratio R with respect to normal combustion of rapid combustion and partial combustion will fall because combustion becomes stable.

  On the other hand, if it is determined in step 210 that the rapid / partial combustion ratio R is equal to or less than the threshold value Ro, the target in-cylinder air-fuel ratio AFc is corrected to the lean side (step 214). That is, a value obtained by adding the predetermined value α to the current target cylinder air-fuel ratio AFc is set as a new target cylinder air-fuel ratio AFc. As a result, the next target in-cylinder fuel amount fc (k + 1) is reduced from the current target in-cylinder fuel amount fc (k), and the discharge of HC is suppressed. In addition, since the combustion becomes unstable due to the lean air-fuel ratio, the ratio R of normal combustion to rapid combustion or partial combustion increases. In the present embodiment, the target in-cylinder air-fuel ratio AFc is corrected by adding / subtracting the predetermined value α, but the target in-cylinder air-fuel ratio AFc is obtained by multiplying the reference air-fuel ratio by a coefficient, and the coefficient is increased or decreased. By doing so, the target in-cylinder air-fuel ratio AFc may be corrected.

  By executing the combustion stabilization control routine described above, the target in-cylinder air / fuel ratio AFc is set so that the ratio R of the rapid combustion and the partial combustion to the normal combustion converges to the threshold value Ro. Since the air-fuel ratio control performed in the present embodiment is open loop control, the actual in-cylinder air-fuel ratio does not necessarily match the target in-cylinder air-fuel ratio AFc, but the above-described combustion stabilization control As a result, the in-cylinder air-fuel ratio is realized such that the rapid / partial combustion ratio R converges to the threshold value Ro. As a result, the in-cylinder air-fuel ratio can be made lean without impairing drivability, and an ideal combustion state that satisfies both drivability and HC emissions even before the start of air-fuel ratio F / B control. Can be obtained.

  Further, according to the control device of the present embodiment, as in the first embodiment, the rapid / partial combustion ratio R indicating the frequency of occurrence of partial combustion and rapid combustion with respect to normal combustion is used as an index value. There is also an advantage that the lean limit can be accurately determined without being affected by the state quantity. Furthermore, according to the control device of the present embodiment, torque fluctuation can be positively suppressed before the start of air-fuel ratio F / B control, so that the robustness of drivability immediately after starting can be significantly improved. There is also an advantage of being able to.

  In the above-described embodiment, the “combustion state determination means” and the “occurrence frequency calculation means” of the first invention are realized by the execution of the processing of step 208 by the ECU 2. Furthermore, the “target air-fuel ratio correcting means” according to the first aspect of the present invention is realized by the execution of the processing of steps 210, 212, and 214 by the ECU 2.

Others.
Although the embodiments of the present invention have been described above, the present invention is not limited to the above-described embodiments, and various modifications can be made without departing from the spirit of the present invention. For example, the combustion stabilization control according to the first embodiment may be combined with the combustion stabilization control according to the second embodiment. That is, until the air-fuel ratio F / B control becomes possible, the combustion stabilization control according to the second embodiment is executed, and after the start of the air-fuel ratio F / B control, the combustion stabilization control according to the first embodiment. Execute. According to this, it becomes possible to further suppress the discharge of HC.

  Further, both the combustion stabilization control according to the first embodiment and the combustion stabilization control according to the second embodiment may be performed for each cylinder. That is, the frequency of occurrence of partial combustion and rapid combustion with respect to normal combustion is calculated for each cylinder, and the target air-fuel ratio (target exhaust air-fuel ratio, target in-cylinder air-fuel ratio) is corrected for each cylinder based on the occurrence frequency obtained for each cylinder. You may make it do. Intake pipe shape irregularity, non-uniform intake volume due to intake interference, etc., variation in combustion speed due to slight difference in in-cylinder gas flow, slight difference in combustion temperature between cylinders caused by cooling water path, combustion chamber of each cylinder Occurrence of partial combustion and rapid combustion even with the same air-fuel ratio due to various differences among cylinders, such as differences in fuel injection amount due to manufacturing variations such as volume and piston shape, manufacturing errors of fuel injection valves, etc. The frequency is different for each cylinder, and the limit of leaning is also different for each cylinder. By performing the above-described combustion stabilization control for each cylinder, the limit of leaning of each cylinder can be individually set to control the air-fuel ratio for each cylinder, and HC emissions can be further suppressed. Is possible.

  When the above-described combustion stabilization control is performed for each cylinder, a torque difference (average torque difference) may occur between the cylinders due to the difference in the air-fuel ratio between the cylinders. The torque difference between the cylinders causes torque fluctuation and deteriorates drivability. Therefore, preferably, the torque difference between the cylinders (or the torque variation of all cylinders) is determined, and if the torque difference or variation exceeds the allowable range, the target air-fuel ratio (target exhaust air-fuel ratio, target A lean upper limit value is provided for the in-cylinder air-fuel ratio so that the torque difference falls within an allowable range.

  In the above-described embodiment, the combustion state is determined based on the torque. However, the combustion state may be determined based on an output parameter other than the torque. For example, the combustion ratio, heat generation amount, combustion speed, etc. can be used as output parameters for determining the combustion state. Any of the internal combustion engines equipped with the in-cylinder pressure sensor can be calculated based on the output signal. Each acceleration of the crankshaft can also be used as an output parameter for determining the combustion state.

  In the above-described embodiment, the routines of FIGS. 3 and 7 are executed at the time of the cold first idle. However, the control by these routines is not effective only at the time of the cold first idle. That is, if the combustion state is fluctuating, the fluctuation of the combustion state can be suppressed by executing the control according to the above routine. However, in a situation where the load fluctuates, such as when accelerating or climbing, the combustion state is controlled according to the load fluctuation, and the fluctuation of the combustion state at this time is the result of active control. Therefore, in such a situation, it is not necessary to execute control by the above routine. The control by the above routine is effective when the combustion state fluctuates in a steady state where there is no load fluctuation. Therefore, the routine shown in FIG. 3 or FIG.

It is the graph which showed about the torque change of the specific cylinder which arises at the time of cold first idling. It is a figure for demonstrating the structure of the control apparatus of the internal combustion engine as Embodiment 1 of this invention. It is a flowchart of the combustion stabilization control routine performed in Embodiment 1 of this invention. It is a characteristic view showing the relationship between the indicated torque, the torque due to in-cylinder gas pressure, the inertia torque due to reciprocating inertia mass, and the crank angle. It is a schematic diagram which shows a crank angle signal and a torque calculation timing. It is a schematic diagram which shows the map showing the relationship between friction torque, engine speed, and cooling water temperature. It is a flowchart of the combustion stabilization control routine classified by cylinder performed in Embodiment 2 of this invention.

Explanation of symbols

2 ECU
4 A / F sensor 6 Crank angle sensor 8 Spark plug 10 Fuel injection valve

Claims (2)

Combustion state determining means for determining whether the combustion is proper normal combustion, partial combustion, or rapid combustion for each combustion of each cylinder at a constant time;
An occurrence frequency calculating means for calculating an occurrence frequency of partial combustion and rapid combustion with respect to normal combustion from the determination result of the combustion state determining means;
Target air-fuel ratio correcting means for correcting the target air-fuel ratio based on the occurrence frequency;
A control device for an internal combustion engine, comprising: The occurrence frequency calculating means calculates the occurrence frequency for each cylinder,
2. The control apparatus for an internal combustion engine according to claim 1, wherein the target air-fuel ratio correcting means corrects the target air-fuel ratio for each cylinder based on the occurrence frequency obtained for each cylinder.

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