JP2015081594A - Engine control device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide an engine control device capable of controlling torque variation with improved efficiency in an operation state, such as idling in particular, where engine vibration is likely to cause a problem.SOLUTION: An engine control device preliminarily sets a correlation coefficient R(θ) representing a correlation between variation Qin a heat generation amount in combustion cycles of an engine 1 and torque variation TF as a function of a predetermined crank angle position (for example, a gravity center of combustion θ) where a heat release rate dQ/dθ is high based on the specifications of an engine. Then, the engine control device: obtains a current value of the predetermined crank angle position while the engine 1 is operated (Step ST5); and performs corrective control of an ignition timing with a value of the correlation coefficient calculated by using the current value as an index (Steps ST6 to ST11).

Description

  The present invention relates to an engine control apparatus that controls ignition timing so that torque fluctuation is reduced, and more particularly to a control technique that focuses on the correlation between variation in heat generation and torque fluctuation.

  In general, in a spark ignition type engine mounted on a vehicle or the like, control of intake air amount, fuel supply amount, ignition timing and the like is performed in order to adjust the output. As an example, in the control device described in Patent Document 1, the amount of heat generated by combustion and the indicated torque are calculated based on the signal from the in-cylinder pressure sensor facing the combustion chamber, and the parameter indicating the misfire tendency from the amount of heat generated. While calculating the value of (misfire ratio), the value of the parameter showing the torque fluctuation between combustion cycles is calculated from the indicated torque. The ignition timing is controlled so that both of these two parameters fall within a predetermined range.

  For example, in an operating state where engine vibration is likely to be a problem, such as idling, the combustion state is stabilized by correcting the ignition timing retardation to reduce torque fluctuations. If becomes stronger, the ignition timing is corrected to the advance side. As a result, the ignition timing can be gradually brought close to MBT while suppressing torque fluctuation, and fuel consumption can be reduced.

JP 2007-285194 A

  However, even if the ignition timing can be kept within a preferable range to some extent by simply correcting to the advance side or retard side according to parameters such as torque fluctuation and misfire ratio as in the conventional example, it is optimal. It is hard to say that it can be controlled. For example, if the ignition timing retardation correction amount increases, the density of the air-fuel mixture at the time of combustion decreases, so there is a concern that torque fluctuations increase as a result of unstable combustion.

  By the way, in the reciprocating engine, the combustion pressure in the cylinder is transmitted to the crankshaft via the piston and the connecting rod. As a feature of this piston-crank mechanism, the combustion pressure is converted into the rotation of the crankshaft. Efficiency varies depending on the position of the piston (crank angle position). For example, when the piston is in the vicinity of the top dead center or the bottom dead center, the connecting rod and the crank web are aligned substantially in a straight line, so that the combustion pressure is difficult to be converted into the rotation of the crankshaft.

  Focusing on the characteristics of such a piston-crank mechanism, the object of the present invention is to set the ignition timing so that torque fluctuations can be more effectively suppressed than in the prior art, particularly in an operating state where engine vibration is likely to cause problems. There is to control.

  The inventor of the present application pays attention to the fact that the efficiency at which the combustion pressure of the engine is reflected in the torque as described above varies depending on the crank angle position. The ignition timing was corrected so that the combustion pressure increased.

  Specifically, the present invention is directed to a control device that controls the ignition timing of an engine so that the torque fluctuation between combustion cycles is equal to or less than a predetermined value. Then, a correlation coefficient representing the correlation between the variation in the heat generation amount between the combustion cycles and the torque fluctuation is set in advance based on the engine specifications as a function of the predetermined crank angle position at which the heat generation rate becomes higher than a predetermined value. In addition, during the operation of the engine, the current value of the predetermined crank angle position is obtained, and the ignition timing correction control is performed using the value of the correlation coefficient calculated using this value as an index.

  That is, first, as described above, in a reciprocating engine, as a characteristic of the piston-crank mechanism, the mechanical conversion efficiency to the torque of the combustion pressure depends on the crank angle position of the piston between the top dead center and the bottom dead center. Change. The conversion efficiency at each crank angle position is mainly determined by predetermined engine specifications such as the length of the connecting rod, the length of the arm of the crankshaft, and the crank offset.

  Therefore, based on the conversion efficiency and heat generation amount at each crank angle position, an instantaneous torque at each crank angle position is calculated and integrated over the entire combustion period, so in one combustion cycle. The engine torque can be obtained. Approximately, it is assumed that the total heat generation amount in the entire combustion period is generated at a predetermined crank angle position where the heat generation rate is high, such as the combustion center of gravity, and the conversion efficiency at the predetermined crank angle position and the total heat generation amount are Based on this, the engine torque can also be obtained.

  That is, as a function of the predetermined crank angle position, a correlation coefficient representing a correlation between combustion fluctuation (variation in heat generation amount) and torque fluctuation can be defined and set in advance based on the engine specifications. Then, when the value of the predetermined crank angle position in the current operating state is obtained during operation of the engine and the value of the correlation coefficient is calculated using this value, the value of the correlation coefficient is This is a suitable index representing the correlation between the combustion fluctuation and the torque fluctuation in the state.

  Therefore, the torque fluctuation can be effectively suppressed by correcting the ignition timing using the value of the correlation coefficient as an index as in the specific matter. For example, if the correlation coefficient is multiplied by the value of variation in heat generation amount to obtain the magnitude of torque fluctuation, the smaller the correlation coefficient value, the larger the variation in heat generation amount. However, torque fluctuation is unlikely to increase. Therefore, in this case, the ignition timing may be corrected so that the value of the correlation coefficient becomes small.

  The predetermined crank angle position is preferably the center of gravity (combustion center of gravity) of the heat generation rate in the entire combustion period as described above, but is not limited to this, and the initial period of combustion during which the heat generation rate increases after ignition It may be the center of gravity of the heat generation amount, or the time when the heat generation rate reaches a peak. And the present value of such a predetermined crank angle position can be calculated | required based on the output of the various sensors which can measure a combustion state.

  For example, the waveform of the heat generation rate may be obtained from the output of the in-cylinder pressure sensor, and the predetermined crank angle position (combustion center of gravity, heat generation peak, etc.) may be obtained. It is also possible to detect the fluctuation state of the instantaneous rotation speed of the crankshaft from the output of the crank position sensor and obtain the waveform of the heat generation rate based on this.

  In addition, the ignition timing correction control as described above may be performed in accordance with the operating state of the engine. For example, during idling, engine vibration caused by torque fluctuations can easily cause occupants to feel uncomfortable. At this time, the ignition timing is corrected so that the value of the correlation coefficient is less than the smallest threshold, for example, according to a strict index. do it.

  On the other hand, if the engine is in an acceleration operation state, the engine vibration is unlikely to become a problem, but rather an increase in torque is required. In some cases, the ignition timing may be corrected according to an intermediate index between the strict index and the loose index.

  According to the present invention, the correlation coefficient between the variation in the heat generation amount between the combustion cycles and the torque fluctuation is set so that the ignition timing is such that even if the combustion fluctuation of the engine is large, the torque fluctuation due to this is difficult to increase. Since the ignition timing is corrected as an index, torque fluctuation can be more effectively suppressed than in the past. This has an excellent effect particularly in an operating state where engine vibration is likely to be a problem, such as idling.

It is a figure showing a schematic structure of an engine concerning an embodiment. It is explanatory drawing of operation | movement of an engine, (a) shows the time of being in the compression top dead center vicinity, (b) shows the time of being in an expansion stroke, respectively. It is a graph which shows an example of a heat release rate waveform and a combustion gravity center. It is a chart showing a list of variables used in mathematical expressions such as conversion coefficients. It is a graph which shows an example of the change of the value of the conversion coefficient by the change of a combustion gravity center position. It is a flowchart figure which shows the flow of the process of the first half of the correction control of ignition timing. It is a flowchart figure which shows the flow of the process of the latter half of the correction control. It is an image figure of the process which calculates | requires a gravity center threshold value with reference to the correction | amendment map, (a) shows the case of an intermediate | middle index, (b) shows the case of a severer index | indication, respectively. 1 is a skeleton diagram showing a configuration of a piston-crank mechanism of an engine.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings. In the present embodiment, the case where the present invention is applied to a reciprocating (reciprocating) gasoline engine will be described. However, the present invention is not limited to this, and a reciprocating engine that performs spark ignition is an engine using a gas engine or alcohol fuel. May be.

(Schematic configuration of the engine)
FIG. 1 shows a schematic configuration of an engine 1 according to the present embodiment. The engine 1 is a four-cylinder gasoline engine as an example, and a piston 12 is accommodated in each of four cylinders 2 (only one cylinder 2 is shown in FIG. 1) to define a combustion chamber 11. Each of the four pistons 12 is connected to a crankshaft 13 via a connecting rod 14 so that the reciprocating motion of the piston 12 is converted into rotation of the crankshaft 13.

  That is, as shown in an enlarged view in FIG. 2, the small end of the connecting rod 14 is rotatably connected to the piston pin 12a of the piston 12 in each cylinder 2, while the large end of the connecting rod 14 is connected to the crank pin 13a. Is rotatably connected to. Thus, a reciprocating slider-crank mechanism (piston-crank mechanism) in which the rotational motion of the crankshaft 13 and the reciprocating motion of the piston 12 are interlocked is configured.

  In the latter half of the compression stroke shown in FIG. 1, the connecting rod 14 rises as the crankshaft 13 rotates clockwise in the drawing, and the compression top dead center (TDC) as shown in FIG. ) In the vicinity, the connecting rod 14 and the crank web 13b are arranged substantially in a straight line. Thereafter, when the expansion stroke is started, the connecting rod 14 falls down as the piston 12 descends, and when the piston 12 extends in the tangential direction of the crank rotation as shown in FIG.

  In order to detect the rotation angle θ of the crankshaft 13, that is, the crank angle θ (representing the crank angle position of the piston 12), a signal rotor 15 is attached to the end of the crankshaft 13. A plurality of protrusions (teeth) 15 a are formed on the outer periphery of the signal rotor 15, and each time the protrusion 15 a passes as the crankshaft 13 rotates, a pulse signal is output from the crank position sensor 61. It transmits to ECU6 mentioned later.

  On the other hand, a cylinder head 18 is fastened to the upper end of the cylinder block 1a, and the upper end opening of each cylinder 2 is closed. A spark plug 20 is disposed in the cylinder head 18 so as to face the inside of the cylinder 2, and a spark is discharged by receiving a high voltage from an igniter 21 controlled by an ECU 6 described later. Further, an in-cylinder pressure sensor 62 is disposed so as to face the cylinder 2 as with the spark plug 20.

  An intake passage 3 and an exhaust passage 4 are communicated with the combustion chamber 11 of each cylinder 2 to perform intake of fresh air and exhaust of combustion gas. An intake valve 31 is disposed at the downstream end of the intake passage 3 (downstream end of the intake flow) facing the combustion chamber 11, while the intake passage 31 is disposed at the upstream end of the exhaust passage 4 (upstream end of the exhaust flow). An exhaust valve 41 is arranged, and a valve operating system for operating the intake valve 31 and the exhaust valve 41 is provided.

  As an example, the valve operating system of the present embodiment is of the DOHC type including an intake camshaft 31a and an exhaust camshaft 41a that drive the intake valve 31 and the exhaust valve 41, respectively. The camshafts 31a and 41a are rotated in synchronization with the crankshaft 13 by a timing chain (not shown) or the like, so that the intake valve 31 and the exhaust valve 41 are opened and closed at a suitable timing for each cylinder 2.

  In the intake passage 3, an air cleaner 32, an air flow meter 63, an intake air temperature sensor 64 (built in the air flow meter 63), and an electronically controlled throttle valve 33 are disposed. The throttle valve 33 is driven by a throttle motor 34 to adjust the intake air amount of the engine 1 by restricting the flow of intake air, and the opening degree (throttle opening degree) is controlled by an ECU 6 described later.

  In addition, an injector (fuel injection valve) 35 for fuel injection is also provided in the intake passage 3 for each cylinder 2, and this injector 35 is controlled by an ECU 6 described later to supply fuel into the intake passage 3. It comes to inject. Thus, the fuel injected from the injector 35 into the intake passage 3 is sucked into the cylinder 2 while being mixed with the intake air to form an air-fuel mixture. This air-fuel mixture is compressed by the piston 12 rising in the compression stroke of the cylinder 2.

  The air-fuel mixture compressed in this way is usually ignited by the spark plug 20 at the end of the compression stroke, combustion starts, and the piston 12 is compression top dead center (TDC: see FIG. 2 (a)). It goes down after going over. In the expansion stroke (see FIG. 2B), the piston 12 is pushed down by the combustion pressure, and the crankshaft 13 is rotated via the connecting rod 14 (generation of torque).

Such a combustion state of the air-fuel mixture by spark ignition is generally represented by a heat generation rate waveform as shown in FIG. The horizontal axis of the graph of FIG. 3 is the crank angle θ, and the vertical axis is the heat generation rate (dQ / dθ). As shown in the graph of FIG. 3, the heat generation rate dQ / dθ increases due to the combustion of the air-fuel mixture, and after reaching the peak, it decreases to the end of combustion. The area below the graph of the heat generation rate waveform corresponds to the total heat generation amount Q _all (total amount of heat generation by combustion), and the crank angle θ that is the center of the area is the combustion center of gravity θ _{g, ave} .

In FIG. 3, when the combustion becomes intense as shown by the broken line graph with respect to the combustion state shown by the solid line graph, the peak of the heat generation rate dQ / dθ is increased while the combustion period is shortened. Further, when the combustion becomes slow as in the dashed line graph, the peak of the heat generation rate dQ / dθ is lowered while the combustion period is lengthened. And, since the combustion state varies between the combustion cycles as described above, the variation Q _{all, std} of the heat generation amount between the combustion cycles occurs as shown in the figure.

As described above, the pressure in the cylinder 2 (in-cylinder pressure) also increases during the period in which the heat generation rate dQ / dθ increases, and a rotational force (torque) is effectively applied to the crankshaft 13 via the piston 12. be able to. Thus, the magnitude of the torque applied in one combustion cycle can be obtained by integrating the instantaneous torque at each crank angle θ over the entire combustion period. It can also be calculated assuming that Q _all occurs at the combustion center of gravity θ _{g, ave} .

The burned gas that has pushed down the piston 12 by the combustion pressure and generates torque flows out into the exhaust passage 4 by opening the exhaust valve 41 (not shown), and is converted by the catalytic converters 42 and 43 shown in FIG. Purified. In the present embodiment, an air-fuel ratio sensor (A / F sensor) 65 is disposed upstream of the upstream catalyst 42 in the flow of burned gas (exhaust gas) discharged as described above, and the downstream catalyst 43. An oxygen sensor (O ₂ sensor) 66 is disposed on the upstream side.

-ECU-
The ECU 6 includes a known electronic control unit (Electronic Control Unit). Although not shown, the ECU 6 includes a CPU (Central Processing Unit), a ROM (Read Only Memory), a RAM (Random Access Memory), a backup RAM, and the like. . The CPU executes various arithmetic processes based on the control program and map stored in the ROM. In addition, the RAM temporarily stores calculation results in the CPU, data input from each sensor, and the like, and the backup RAM stores data to be saved when the engine 1 is stopped, for example.

As described above, the ECU 6 is connected to the crank position sensor 61, the in-cylinder pressure sensor 62, the air flow meter 63, the intake air temperature sensor 64, the air-fuel ratio sensor 65, the O ₂ sensor 66, and the like. As shown in FIG. 1, the ECU 6 is also connected with an accelerator opening sensor 67 that detects an operation amount (accelerator opening Acc) of an accelerator pedal 7 provided in the vehicle interior.

  The ECU 6 executes various control programs based on signals input from the various sensors and the like. For example, the ignition timing control of the spark plug 20 by the igniter 21 for each cylinder 2 of the engine 1 and the throttle motor 34 are controlled. Control of throttle opening by operation (that is, control of intake air amount), fuel injection control by the injector 35, and the like are executed. In the present embodiment, the ECU 6 also executes ignition timing correction control that is effective in suppressing torque fluctuations as described below.

(Ignition timing correction control)
Hereinafter, the ignition timing correction control, which is a feature of the present embodiment, will be described. This control is characterized by the mechanical efficiency at which the combustion pressure in the cylinder 2 is converted into the rotation of the crankshaft 13 via the piston 12 and the connecting rod 14 as a feature of the piston-crank mechanism of the engine 1 described above. It focuses on the difference between the two.

  Specifically, first, as shown in FIG. 2A, in the vicinity of the compression top dead center (TDC), the connecting rod 14 and the crank web 13b are aligned in a straight line, so that the combustion pressure acting on the top surface of the piston 12 is Even if an attempt is made to push down the piston 12 downward in the figure, most of this force presses the crankshaft 13 in the radial direction via the connecting rod 14 and is hardly converted to the rotational force of the crankshaft 13.

  Thereafter, when the expansion stroke is started, the connecting rod 14 is tilted as the piston 12 is lowered as shown in FIG. 2B, and the combustion pressure is effectively converted into the rotational force of the crankshaft 13. . At this time, the conversion efficiency increases as the crank angle θ increases in the range of ATDC 0 to 30 ° CA, and gradually decreases after the efficiency reaches the maximum, but the conversion efficiency is compared up to ATDC 60 ° CA. High state.

  The change in conversion efficiency due to the change in the crank angle θ is mainly determined by predetermined engine specifications such as the length of the connecting rod 14, the arm length of the crankshaft 13, and the crank offset, and should be set in advance. Can do. As an example, the conversion coefficient R (θ) used in the following (Equation 1) is a coefficient (correlation) indicating the correlation between the amount of heat generated at an arbitrary crank angle θ during the combustion period and the instantaneous torque generated thereby. Number).

The above (Formula 1) represents the correlation between the variation Q _{all, std} of the heat generation amount between combustion cycles and the torque fluctuation caused thereby. Specifically, (Equation 1) indicates that the heat generation amount variation Q _{all, std} has a value R (θ = θ _{g, ave} ) of the conversion coefficient R (θ) at the crank angle θ _{g, ave} of the combustion center of gravity. The torque fluctuation TF of the engine 1 can be calculated by multiplying the predetermined adjustment coefficient k and further adding the torque fluctuation TF ₀ due to motoring.

That is, as described above with reference to FIG. 3, heat generation by combustion is approximately performed at time zero at the combustion center of gravity θ _{g, ave} , and the combustion pressure acts on the piston 12 only at the combustion center of gravity θ _{g, ave.} Then it can be considered. Therefore, as shown in the above (Equation 1) _, the heat generation amount variation Q _{all, std} at the combustion center of gravity θ _{g, ave} is multiplied by the conversion coefficient R (θ = θ _{g, ave} ) at the combustion center of gravity θ _{g, ave} . Thus, the torque fluctuation TF-TF ₀ due to combustion can be calculated.

In other words, the conversion coefficient R (θ = θ _{g, ave} ) at the combustion center of gravity θ _{g, ave} is a suitable index of how much variation in the amount of heat generated by combustion is reflected in the torque fluctuation of the engine 1. However, when the value of the conversion coefficient R (θ = θ _{g, ave} ) is small, the variation in heat generation is large, and even if the combustion fluctuation is large, the torque fluctuation is difficult to increase. it can.

Here, as shown in (Formula 1), the conversion coefficient R (θ) is set based on the engine specifications. Functions such as v (θ _a ), V (θ _a ), and L (n) in the conversion coefficient R (θ) of (Equation 1) are expressed by the following equations (Equation 2) to (Equation 9). It is defined by a variable that represents the specifications. The definition of each variable is shown in the list of FIG. Note that the variable n (n = 1, 2,...) Represents a calculation step of numerical calculation. A method of deriving (Equation 1) using these variables will be described later with reference to FIG.

The inventor of the present invention shows how the value of the conversion coefficient R (θ = θ _{g, ave} ) changes when the position of the combustion center of gravity (crank angle θ _{g, ave} ) changes depending on the operating state of the engine 1. As a result, the characteristics as shown in FIG. 5 were found. As is apparent from the graph of FIG. 5, when the combustion center of gravity θ _{g, ave} is in the vicinity of the compression top dead center (TDC) of the cylinder 2 (indicated by a broken line S in FIG. 5), refer to FIG. As described above, since the combustion pressure is difficult to be converted into the rotational force of the crankshaft 13, the value of the conversion coefficient R (θ = θ _{g, ave} ) becomes very small.

As the combustion center of gravity θ _{g, ave} moves away from the compression top dead center (TDC), the combustion pressure is effectively converted into the rotational force of the crankshaft 13. In the graph of FIG. 5, the value of the conversion efficiency R (θ = θ _{g, ave} ) rises rapidly with the retardation of the combustion center of gravity θ _{g, ave in} the range of ATDC 0-30 ° CA. The value of the conversion efficiency R (θ = θ _{g, ave} ) gradually decreases after peaking at about ATDC 30 ° CA, but is relatively high up to about ATDC 60 ° CA.

A change in the value of the conversion coefficient R (θ = θ _{g, ave} ) due to such a change in the combustion center of gravity θ _{g, ave} is examined in advance, and in this embodiment, a map for correcting the ignition timing (see FIG. 8) is used. Created and electronically stored in the ROM of the ECU 6. Then, the current combustion center of gravity θ _{g, ave} is calculated from the output of the in-cylinder pressure sensor 62 during operation of the engine 1, and the conversion coefficient R (θ = θ _{g, The} ignition timing is corrected so that the value of _ave ) becomes smaller.

-Flow of correction control processing-
Next, a specific processing flow of the ignition timing correction control will be described with reference to the flowcharts of FIGS. This correction routine is executed at a predetermined crank angle for each cylinder 2 in the ECU 6 during operation of the engine 1.

  First, in step ST1 after the start shown in FIG. 6, a necessary degree of torque fluctuation suppression is set based on the operating state of the engine 1 (for example, accelerator opening Acc, throttle opening, engine speed, etc.). As an example, if the engine 1 is in an accelerating operation state, the degree of suppression becomes moderate (suppression degree = 0), and if the engine is idling, the degree of suppression becomes severe (suppression degree = 2). This is because during idling, the vibration of the engine due to torque fluctuation tends to cause discomfort to the occupant, while the engine vibration hardly becomes a problem in the acceleration operation state.

  If the engine 1 is in one of a steady operation state and a deceleration operation state excluding idling, the degree of suppression of torque fluctuation is set to the middle between the loose time and the severe time, that is, a medium degree of suppression = 1. To do. In addition, what the engine 1 is in an acceleration driving state may be determined by, for example, the accelerator opening degree Acc, the throttle opening degree, or the speed at which they change being equal to or greater than a predetermined threshold value.

  Subsequently, in step ST2, it is determined whether or not the suppression degree set as described above is “0”. If the engine 1 is in the acceleration operation state and the suppression degree = 0, an affirmative determination (YES) is made and the process is terminated. (End) On the other hand, if the suppression degree is not 0 and the determination is negative (NO), the process proceeds to step ST3, where it is determined whether or not the suppression degree is “1”. Then, if the suppression degree = 2 during idling of the engine 1, a negative determination (NO) is made, and the process proceeds to step ST12 of FIG.

On the other hand, if the engine 1 is in a steady operation state or a decelerating operation state and an affirmative determination (YES) is made with the suppression degree = 1 in step ST3, the process proceeds to step ST4. Then, with reference to the ignition timing correction map described above, the absolute value | R (θ = θ _{g, ave} ) | of the conversion coefficient is an index value corresponding to the moderate suppression degree (as an example, 0.5 ) _Is calculated as a center-of-gravity threshold value. As an example, as shown in FIG. 8A, in the present embodiment, the first to third center-of-gravity threshold values θ _{g, ave} 1, θ _{g, ave} 2, θ _{g, ave} 3 are set in order from the advance side. Is done.

In the next step ST5, the current combustion center of gravity (the current value of the crank angle θ _{g, ave} ) is calculated by a known method based on the signal from the in-cylinder pressure sensor 62, and the current combustion center of gravity θ _{g, ave} is determined whether the first advance side of the center of gravity threshold theta _{g, ave} 1 of the (step ST6). If the result is affirmative (YES), the ignition timing is corrected to the retard side by the advance angle of the current combustion center of gravity θ _{g, ave} from the first center of gravity threshold θ _{g, ave} 1 (step ST7). ) End processing (END).

That is, when the value of the conversion coefficient R (θ = θ _{g, ave} ) calculated based on the current combustion center of gravity θ _{g, ave} during operation of the engine 1 is larger than the index value, the combustion center of gravity θ _{g, ave} Is the closest centroid threshold value (in this case, the first centroid threshold value θ _{g, ave} 1), that is, the value of the conversion coefficient R (θ = θ _{g, ave} ) is smaller than the current index value. Next, correct the ignition timing.

On the other hand, if the current combustion center of gravity θ _{g, ave} is equal to or more retarded than the first center of gravity threshold θ _{g, ave} 1 and a negative determination (NO) is made in step ST6, It is determined whether or not the combustion center of gravity θ _{g, ave} is on the more retarded side than the second center of gravity threshold θ _{g, ave} 2 (step ST8). If the result is negative (YES), the absolute value | R (θ = θ _{g, ave} ) | of the conversion coefficient at the current combustion center of gravity θ _{g, ave} is equal to or less than the index value. The process ends without correcting the time (end).

If the current combustion center of gravity θ _{g, ave} is on the retard side with respect to the second center of gravity threshold θ _{g, ave} 2 and an affirmative determination (YES) is made in step ST8, the current combustion center of gravity θ _{g, ave} Is closer to the third centroid threshold θ _{g, ave} 3 than the second centroid threshold θ _{g, ave} 2 (step ST9). If this result is affirmative determination (YES), so that the second center of gravity threshold theta _{g, ave} 2 towards closer, the combustion center theta _g, second centroid threshold theta _g of _ave, from _ave 2 Current The ignition timing is corrected to the advance side by the retarded amount (step ST10), and the process ends (END).

Conversely, if the current combustion center of gravity θ _{g, ave} is closer to the second center of gravity threshold θ _{g, ave} 2 than the third center of gravity threshold θ _{g, ave} 3 and a negative determination (NO) is made in step ST9, The ignition timing is advanced by the amount of delay from the second center-of-gravity threshold θ _{g, ave} 2 of the current combustion center-of-gravity θ _{g, ave} so that the closer second center-of-gravity threshold θ _{g, ave} 2 is _reached. It correct | amends (step ST11) and complete | finishes a process (end). That is, the combustion center of gravity θ _{g, ave} becomes the closer of the second and third center of gravity thresholds θ _{g, ave} 2, θ _{g, ave} 3, and the value of the conversion coefficient R (θ = θ _{g, ave} ) The ignition timing is corrected so that becomes smaller.

In this way, when the degree of suppression of torque fluctuation is set to a medium level in accordance with the steady or decelerating operation state of the engine 1, an index value (for example, 0) of the conversion coefficient R (θ = θ _{g, ave} ) corresponding thereto. .5) Correct the ignition timing so that it is as follows. Thereby, even if the variation in the heat generation amount is large, the ignition timing can be controlled so that the torque fluctuation does not become so large.

On the other hand, if it is determined in step ST3 that the torque fluctuation suppression degree is not “1” (NO), the suppression degree is 2 while the engine 1 is idling, so the absolute value of the conversion coefficient | R The ignition timing is corrected so that (θ = θ _{g, ave} ) | is less than a stricter index value (for example, 0.2). That is, in step ST12 shown in FIG. 7, the ignition timing correction map is referred to in the same manner as in step ST4, and the absolute value of the conversion coefficient | R (θ = θg _{, ave} ) | The threshold value of the combustion center of gravity (crank angle θ _{g, ave} ) is calculated.

As a result, as shown in FIG. 8B, the first to third center-of-gravity threshold values θ _{g, ave} 1, θ _{g, ave} 2, θ _{g, ave} 3 are set in order from the advance side. Therefore, in subsequent steps ST13 to ST19, the ignition timing is corrected in the same manner as steps ST5 to ST11, respectively. That is, first, the current combustion gravity center θ _{g, ave} is calculated based on the signal from the in-cylinder pressure sensor 62 (step ST13), and if this is on the advance side of the first gravity center threshold θ _{g, ave} 1 (step ST13). (YES in ST14), the ignition timing is corrected to the retard side (step ST15), and the process is terminated (end).

Further, the current combustion center of gravity θ _{g, ave} is equal to or more retarded than the first center of gravity threshold θ _{g, ave} 1 (NO in step ST14), and the second center of gravity threshold θ _{g, ave} 2 If it is the same or more advanced (NO in step ST16), the absolute value | R (θ = θ _{g, ave} ) | of the conversion coefficient at the current combustion center of gravity θ _{g, ave} is equal to or less than the index value. Therefore, the process is terminated without correcting the ignition timing (end).

Furthermore, if the current combustion center of gravity θ _{g, ave} is on the more retarded side than the second center of gravity threshold θ _{g, ave} 2 (YES in step ST16), the second center of gravity threshold θ _{g, ave} 2 and the third center of gravity threshold θ _The ignition timing is corrected so as to be closer to _{g, ave} 3 (steps ST17 to ST19). That is, the ignition timing is corrected to the retard side so that the combustion center of gravity θ _{g, ave} becomes the third center of gravity threshold θ _{g, ave} 3 (step ST18) or becomes the second center of gravity threshold θ _{g, ave} 2 Thus, the ignition timing is corrected to the advance side (step ST19), and the process is terminated (end).

As described above, when the degree of torque fluctuation suppression is set more severely corresponding to the idling of the engine 1, the index value (for example, 0.2) of the conversion coefficient R (θ = θ _{g, ave} ) corresponding thereto. The ignition timing is corrected to be as follows. Thereby, even if the variation in the amount of heat generation is large, it is possible to control the ignition timing so that the torque fluctuation is as small as possible.

As described above, in the engine 1 of the present embodiment, the ignition timing is corrected using the correlation coefficient R (θ = θ _{g, ave} ) between the variation in the heat generation amount due to the combustion of the air-fuel mixture and the torque fluctuation as an index. By doing so, it is possible to control to an ignition timing at which the torque fluctuation hardly increases even if the combustion fluctuation is large, and the torque fluctuation can be more effectively suppressed than before.

In addition, depending on the operating state of the engine 1, for example, when idling, the value of the correlation coefficient R (θ = θ _{g, ave} ) is less than the minimum threshold value (for example, 0.2) according to a stricter index value. Thus, the ignition timing is corrected as described above, so that the engine vibration does not cause discomfort to the occupant even during idling.

(Derivation of conversion coefficient formula)
Next, a method for deriving (Equation 1) related to the conversion coefficient R (θ = θ _{g, ave} ) described above will be described. First, in a spark ignition engine, the relationship between the heat generation rate due to combustion of the air-fuel mixture and the in-cylinder pressure is generally given by the first law of thermodynamics and the gas state equation. That is, for example, assuming that a discrete numerical calculation is performed in increments of a crank angle set in advance from the time when the intake valve 31 is closed, and the amount of heat given to the cylinder in the nth step of this calculation is dQ (n), the cylinder Using the internal pressure P (n), the in-cylinder volume V (n), and the specific heat ratio κ (n) of the working gas, the heat quantity dQ (n) is given by the following (Equation 10). Then, when (Equation 10) is discretized with a backward difference, the following (Equation 11) is obtained.

  Here, when (Equation 11) is rearranged using functions A (n) and B (n) like the following (Equation 12) and (Equation 13), the following (Equation 14) is obtained.

  Then, by transforming (Equation 14) as a recurrence formula, the following (Equation 15) is obtained.

-Correlation between in-cylinder pressure and torque-
Next, regarding the relationship between the in-cylinder pressure and the torque of the reciprocating engine, the following theoretical formula is established as in (Equation 16). (Equation 16) is an equation of motion of the piston-crank mechanism schematically shown in FIG. 9, and the term on the left side and the fifth term on the right side are the torque (dynamic torque) due to the moment of inertia of the rotating part such as the crankshaft ). The first term on the right side represents torque due to in-cylinder pressure, and the second term on the right side represents torque due to inertial mass that reciprocates such as a piston and a connecting rod. Further, the third term on the right side represents torque loss due to friction, and the fourth term on the right side represents load torque, that is, engine torque output from the crankshaft.

That is, the function a (θ _a ) of the second term on the right side of (Expression 16) is expressed by the following (Expression 17), and the function b (θ _a ) of the term on the left side is expressed by the following (Expression 18). Is done.

The function c (θ _a ) of the fifth term on the right side of (Equation 16) is represented by the following (Equation 19), and the function v (θ _a ) of the first term on the right side is represented by the following (Equation 20). The Note that the v P to be multiplied by (θ _a) (θ _a) is a cylinder pressure. Further, the function f (θ _a ) of the third term on the right side is represented by the following (formula 21), and θ _a is represented by the following (formula 22). Note that one point above θ in (Equation 16) represents a first-order differentiation, and two points represent a second-order differentiation.

Then, from the above (Expression 15) and (Expression 16), when the in-cylinder pressure P (n) = P (θ _a ) is eliminated and the load torque T _l is arranged, (Expression 23) is obtained as follows. .

Here, if it is assumed that the heat generation due to the combustion of the air-fuel mixture is performed at time zero at a predetermined crank angle (for example, the combustion center of gravity θ _{g, ave} described above) at which the heat generation rate becomes high during the combustion period, the predetermined crank angle as calculation step n = n _d corresponding to the following (equation 24) is obtained, whereby the following (expression 25) is obtained.

From the above, using (Equation 25), the relational expression between the variation (standard deviation) between the combustion cycles of the total heat generation amount Q _all and the torque fluctuation is derived. First, paying attention to the first term on the right side of (Expression 25), substituting the above (Expression 12) and (Expression 13), the following (Expression 26) is obtained. Assuming that the specific heat ratio κ does not change even if the composition of the working gas changes with combustion, the following (Equation 27) is obtained.

In this (Expression 27), n _d is replaced with the combustion center of gravity θ _{g, ave,} and the second and subsequent terms on the right side of (Expression 25) are replaced with a constant term TF ₀ for convenience, and an error caused by this replacement is adjusted. (Equation 1) can be obtained by introducing an adjustment coefficient k (proportional constant adapted by experiments and simulations).

(Other embodiments)
The description of the above-described embodiment is merely an example, and is not intended to limit the configuration or application of the present invention. That, is regarded as a total heat generation amount Q _all by combustion occurs in the combustion center of gravity theta _{g, ave} in the above embodiments, the correlation coefficient in the combustion center theta _{g, ave} as shown in Equation (1) R (θ = θ _{g, ave} ) and the variation Q _{all, std} of the total heat generation amount are obtained _, but the torque variation TF is not limited to this.

For example, in addition to the combustion center of gravity θ _{g, ave} , as a predetermined crank angle position having a high heat generation rate dQ / dθ, the center of heat generation amount during the initial period of combustion in which the heat generation rate dQ / dθ increases after ignition, Examples include the time (crank angle) at which the rate dQ / dθ peaks. Further, the predetermined crank angle position may not be one, and the torque fluctuation is based on the correlation coefficient R (θ) at two or more predetermined crank angle positions (crank angle θ) and the variation in heat generation amount. TF may be obtained.

In the above-described embodiment, the current value of the predetermined crank angle position such as the combustion center of gravity θ _{g, ave} is calculated based on the signal from the in-cylinder pressure sensor 62, but is not limited thereto. For example, it can be calculated based on the output of any sensor that can measure the combustion state, such as a crank position sensor 61 or an ion current sensor.

Further, in the above embodiment, the degree of suppression of torque fluctuation is changed according to the operating state of the engine 1, and the index value of the conversion coefficient R (θ = θ _{g, ave} ) is also changed accordingly. However, the present invention is not limited to this, and the degree of suppression of torque fluctuation, that is, the index value may be constant regardless of the operating state of the engine 1.

  The present invention can suppress torque fluctuation more effectively than before by characteristic correction control of the ignition timing of the engine. For example, the present invention has an excellent effect when applied to a spark ignition engine mounted on a passenger car. .

1 engine 6 ECU (control device)
TF Torque variation between combustion cycles Q _{all, std} Variation in heat generation between combustion cycles R (θ _{g, ave} ) Conversion coefficient (correlation coefficient)
θ _{g, ave} combustion center of gravity (predetermined crank angle position)

Claims (1)

An engine control device that controls the ignition timing of an engine so that torque fluctuation between combustion cycles is less than or equal to a predetermined value,
A correlation coefficient representing the correlation between the variation in heat generation amount between combustion cycles and torque fluctuation is preset based on engine specifications as a function of a predetermined crank angle position at which the heat generation rate becomes higher than a predetermined value. ,
The present invention is characterized in that a current value of the predetermined crank angle position is obtained during engine operation, and ignition timing correction control is performed using the value of the correlation coefficient calculated using this value as an index. Engine control device.

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