JP2011149351A - Egr rate estimation-detecting device - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide an EGR (Exhaust Gas Recirculation) rate estimation-detecting device capable of estimation-detecting an ERG rate in an internal EGR, based on an output signal from a crank pulser rotor. <P>SOLUTION: The EGR rate estimation-detecting device 30 includes an intake new-air mass detecting part 39 for detecting an intake new-air mass of the engine 5, based on an output signal from an air flowmeter 15, an NeA calculating part 33 for calculating an average engine speed NeA of the engine, based on a crank pulse, a ▵ω1 calculating part 34 for calculating the first crank-angular speed ω1 in the first prescribed section τ1 overlapped with a compression top dead center (TDC) of the engine 5, and for calculating the first fluctuation amount ▵ω1 by subtracting the first crank-angular speed ω1 from the average engine speed NeA, and a total cylinder gas mass estimation value drawing-out part 38 for estimating total cylinder gas mass, based on a value of the first fluctuation amount ▵ω1. The EGR rate estimation-detecting device 30 draws out an estimated value of the ERG rate that is a recirculation rate of an exhaust gas, based on the intake new-air mass and the total cylinder gas mass estimation value. <P>COPYRIGHT: (C)2011,JPO&INPIT

Description

  The present invention relates to an EGR rate estimation detection device, and more particularly to an EGR rate estimation detection device that can estimate and detect an exhaust gas recirculation rate (EGR rate) in an engine.

  Conventionally, a part of the engine exhaust gas is led to the intake side, and this is mixed with the air-fuel mixture and then taken into the intake air to reduce nitrogen oxides in the exhaust gas and improve the fuel efficiency. The technique of (EGR: Exhaust Gas Recirculation) is known. The EGR has an “external EGR” in which a bypass pipe for circulating the exhaust gas is provided between the intake pipe and the exhaust pipe, and the exhaust valve is allowed to remain in the cylinder earlier by closing the exhaust valve. Although there is “internal EGR”, in either EGR technique, “EGR rate” indicating how much exhaust gas is contained in the air-fuel mixture sucked into the cylinder is an important control parameter.

  In Patent Document 1, in an engine having an external EGR bypass pipe, a gas concentration sensor is provided in each of an intake pipe, an exhaust pipe, and a cylinder. Based on an output value of the gas concentration sensor, the EGR rate of the external EGR and the internal A gas concentration measuring device that can calculate the EGR rate of EGR is disclosed.

JP 2009-203874 A

  However, the technique described in Patent Document 1 requires a plurality of high-performance sensors that directly detect the gas concentration, and there is a problem that the configuration is complicated and the cost is increased.

  An object of the present invention is to solve the above-described problems of the prior art and to provide an EGR rate estimation detection device that can estimate and detect an EGR rate in an internal EGR based on an output signal of a crank pulser rotor.

  In order to achieve the above object, the present invention provides a pickup for detecting the passage of a plurality of reluctors (4) provided in a crank pulser rotor (2) that rotates in synchronization with a crankshaft (1) of an engine (5). In the EGR rate estimation detection device (30) to which a crank pulse is supplied from the PC), based on the intake fresh air mass detection unit (39) for detecting the intake fresh air mass of the engine (5), and the crank pulse, A NeA calculation unit (33) for calculating an average engine speed (NeA) of the engine (5), and a first crank of a first predetermined section (τ1) overlapping with a compression top dead center (TDC) of the engine (5) Δω1 calculation for calculating the first fluctuation amount (Δω1) by calculating the angular velocity (ω1) and subtracting the first crank angular velocity (ω1) from the average engine speed (NeA). (34) and an in-cylinder gas total mass estimated value deriving unit (38) for estimating the in-cylinder gas total mass (Gtotal) based on the value of the first fluctuation amount (Δω1), and the EGR rate The estimation detection device (30) has a first feature in that an estimation value of an EGR rate that is a recirculation rate of exhaust gas is derived based on an estimation value of the intake fresh air mass and the total mass of the in-cylinder gas. is there.

  Further, the relationship between the first fluctuation amount (Δω1) and the in-cylinder gas total mass (Gtotal) is shown, and a plurality of Δω1-Gtotal maps (37) provided for each predetermined engine speed are provided. The in-cylinder gas total mass estimated value deriving unit (38) selects one Δω1-Gtotal map (37) that matches the average engine speed (NeA), and responds to the first fluctuation amount (Δω1). A second feature is that the derived Gtotal value is used as an estimated value of the total in-cylinder gas mass.

  The engine (5) is configured to be capable of premixed compression ignition combustion, and the EGR rate causes the exhaust gas to remain in the cylinder according to the opening / closing timing of the intake valve (IV) and the exhaust valve (EV). A third feature is that the EGR rate of the internal EGR.

  Further, the first predetermined section (τ1) is from a falling point of the crank pulse positioned immediately before the compression top dead center (TDC) to a falling point of the crank pulse positioned immediately after the compression top dead center (TDC). A fourth characteristic is that the period is up to (C2).

  Further, a second crank angular speed (ω2) in a second predetermined section (τ2) overlapping the combustion bottom dead center (BDC) of the engine (5) is calculated, and the first crank angular speed is calculated from the second crank angular speed (ω2). A Δω2 calculating unit that calculates the second variation (Δω2) by subtracting (ω1), calculating an engine load factor based on the value of the second variation (Δω2), and an estimated value of the EGR rate; A fifth feature is that a target value of valve timing is calculated based on the engine load factor.

  According to the first feature, the intake fresh air mass detection unit that detects the intake fresh air mass of the engine, the NeA calculation unit that calculates the average engine rotation speed of the engine based on the crank pulse, and the compression top dead center of the engine Is calculated based on the value of the first variation amount, and a Δω1 calculation unit that calculates the first variation amount by subtracting the first crank angular velocity from the average engine rotation speed. An in-cylinder gas total mass estimated value deriving unit for estimating the in-cylinder gas total mass, and the EGR rate estimation detection device regenerates exhaust gas based on the estimated values of the intake fresh air mass and the in-cylinder gas total mass. Since the estimated value of the EGR rate that is the circulation rate is derived, the EGR rate can be estimated and detected by using the crank pulse signal and the output of an air flow sensor or the like for detecting the intake fresh air mass. Become. This eliminates the need to provide a gas concentration sensor or the like for directly detecting the EGR rate, avoids complication of the engine configuration, and reduces production costs.

  According to the second feature, the relationship between the first fluctuation amount and the in-cylinder gas total mass is shown, and a plurality of Δω1-Gtotal maps provided for each predetermined engine speed are provided to estimate the in-cylinder gas total mass. The value deriving unit selects one Δω1-Gtotal map that matches the average engine rotation speed, and uses the Gtotal value derived from the correspondence with the first fluctuation amount as an estimated value of the total gas in the cylinder. The in-cylinder gas total mass can be estimated and detected using the first fluctuation amount calculated from the crank pulse output and a map derived in advance through experiments or the like.

  According to the third feature, the engine is configured to be capable of premixed compression ignition combustion, and the EGR rate is the EGR rate of the internal EGR that causes the exhaust gas to remain in the cylinder at the opening and closing timing of the intake valve and the exhaust valve. Therefore, in an engine that enables premixed compression ignition (HCCI) combustion using internal EGR, it is possible to accurately obtain an EGR rate that is a control parameter necessary for HCCI combustion control and perform appropriate combustion control. Become.

  According to the fourth feature, the first predetermined section is a period from the falling point of the crank pulse located immediately before the compression top dead center to the falling point of the crank pulse located immediately after the compression top dead center. Therefore, it is possible to accurately detect the crank angular velocity at the position overlapping the compression top dead center.

  According to the fifth feature, the second crank angular speed of the second predetermined section overlapping the combustion bottom dead center of the engine is calculated, and the second fluctuation amount is calculated by subtracting the first crank angular speed from the second crank angular speed. Since the Δω2 calculation unit is provided, the engine load factor is calculated based on the value of the second fluctuation amount, and the target value of the valve timing is calculated based on the estimated value of the EGR rate and the engine load factor. Therefore, based on the crank pulse signal Thus, the engine load factor and the target value of the valve timing can be calculated.

It is a block diagram which shows the structure of the EGR rate estimation detection apparatus which concerns on one Embodiment of this invention. It is a block diagram which shows the flow of the engine control using an EGR rate estimated value. 3 is a time chart showing a relationship between a crank pulse signal and a change in crank angular velocity ω during one cycle. FIG. 4 is a partially enlarged view of FIG. 3. It is an example of (DELTA) omega1-Gtotal map. It is a graph which shows the combustion characteristic of an HCCI engine. It is the graph which showed the valve timing changed with a VVT mechanism. It is a graph which shows the relationship between an EGR rate and an engine load factor.

  Hereinafter, preferred embodiments of the present invention will be described in detail with reference to the drawings. FIG. 1 is a block diagram showing a configuration of an EGR rate estimation detection device 30 according to an embodiment of the present invention. FIG. 2 is a block diagram showing a flow of engine control using the estimated value of the EGR rate. Further, FIG. 3 is a time chart showing the relationship between the crank pulse signal and the fluctuation of the crank angular velocity ω during one cycle, and FIG. 4 is a partially enlarged view of FIG. The EGR rate estimation detection device 30 is built in the ECU 50 that controls the engine 5.

  The EGR rate estimation detection device 30 according to the present embodiment makes it possible to obtain an estimated value of the EGR rate based on the output signal of the crank pulsar rotor 2 for detecting the rotational position of the crankshaft 1 of the engine 5. Is. The EGR rate is obtained from the total in-cylinder gas mass (Gtotal) and the intake fresh air mass (Gf). That is, the EGR rate can be calculated by a calculation formula of EGR rate (%) = (Gtotal−Gf) ÷ Gtotal × 100. In the present invention, the in-cylinder gas total mass (Gtotal) is estimated and detected based on the output signal of the crank pulsar rotor 2, while the intake fresh air mass (Gf) is attached to the intake pipe 11 to measure the intake air amount. Is obtained based on the output signal of the air flow meter 15.

  A crank pulsar rotor 2 that rotates synchronously with the crankshaft 1 is attached to the crankshaft 1 of the engine 5. The crank pulsar rotor 2 according to the present embodiment has a configuration in which a total of eleven retractors 4 are provided at intervals of 30 degrees, except for a tooth missing portion H in a rotor 3 that rotates synchronously with the crankshaft 1. .

  The crank pulse detector 31 in the EGR rate estimation detection device 30 detects the rotational position and the rotational speed of the crankshaft 1 by detecting the passing state of the reluctator 4 as a pulse signal by the magnetic pickup type pulse generator PC. . The crank pulse detector 31 detects the reference position of the crank pulsar rotor 2 by detecting the passage of the tooth missing portion H, and the crankshaft makes one rotation of the crankshaft # 0-10 in total 11 crank stages based on the arrangement of the retractor 4. Divide by. Thereafter, when the stroke determination based on the intake pressure fluctuation generated in the intake pipe 11 is determined, the front / back determination of the stage (determination of whether the crankshaft is the first rotation or the second rotation in one cycle) is determined. One cycle (720 degrees) of the engine is divided into 22 cycle stages of # 0 to 21 in total. The stroke determination based on the intake pressure change is executed by, for example, comparing the detected intake pressure fluctuation pattern with the intake pressure fluctuation pattern obtained through experiments or the like. The variation pattern obtained by experiments or the like is associated with the cycle stage.

  The EGR rate estimation detection device 30 includes a NeA calculation unit 33 that calculates an average engine speed NeA in a predetermined detection section based on output signals of the crank pulse detection unit 31 and the timer 32. In addition, the Δω1 calculation unit 34 is based on the average engine speed NeA calculated by the NeA calculation unit 33 and the first crank angular speed ω1 detected in the first predetermined section overlapping the top dead center position of the crankshaft 1. The first variation amount Δω1 of the crank angular speed is calculated. In the present embodiment, the first crank angular velocity ω1 is detected in a predetermined section overlapping the compression top dead center (TDC).

  Reference is now made to FIGS. The crank angular speed ω is cyclically varied in accordance with one cycle of the engine, that is, four strokes of compression, combustion / expansion, exhaust, and intake, due to fluctuations in the cylinder pressure even when the average engine speed NeA is constant. ing. Specifically, in the compression stroke, the crank angular speed ω is reduced due to the compression resistance due to the increase in the cylinder internal pressure. Further, in the section of the combustion / expansion stroke, crank rotational energy is generated due to an increase in the cylinder internal pressure due to combustion, and an increase due to this is generated. In addition, the crank angular speed ω reaches a peak at the end of the combustion / expansion stroke, and then decreases due to pump work such as mechanical friction resistance in the engine, discharge resistance of burned gas in the exhaust stroke, and suction resistance in the intake stroke. Then, the fluctuation of reaching the intake stroke / compression stroke again is repeated.

  According to the fluctuation of the crank angular speed ω, the first crank angular speed ω1 detected in the vicinity of the compression top dead center is smaller than the average engine speed NeA. Note that the second crank angular speed ω2 detected in the vicinity of the combustion bottom dead center is higher than the average engine rotational speed NeA. For example, when the average engine rotational speed NeA is 3000 rpm, the first crank angular speed ω1 = 2900 rpm, The second crank angular speed ω2 = 3100 rpm.

  The fluctuation peak of the crank angular speed ω increases in accordance with the generated torque of the engine, and the subsequent decrease amount increases as the intake air amount increases. Therefore, the variation in the crank angular speed ω increases as the engine has a larger generated torque and a larger intake air amount. Further, this fluctuation becomes larger as the inertial force of the crankshaft is smaller and the engine is smaller in the low rotation range, and becomes larger as the engine has a smaller number of cylinders and a larger explosion interval. In other words, an engine having a relatively small crankshaft moment of inertia, such as a single-cylinder engine of a motorcycle, tends to have a large variation in crank angular velocity ω.

  The Δω1 calculating unit 34 calculates a first variation amount Δω1 (a variation amount with respect to the average engine rotational speed NeA) of the first crank angular velocity ω1 in the vicinity of the compression top dead center by an equation of Δω1 = NeA−ω1. 3 and 4 also show the second fluctuation amount Δω2 (the fluctuation amount with respect to the crank angular speed ω1) of the second crank angular speed ω2 near the combustion bottom dead center. The second variation amount Δω2 is calculated by the equation: Δω2 = ω2−ω1.

  A procedure for calculating Δω1 in the Δω1 calculating unit 34 will be described. The crank angular velocity ω is smallest when the crankshaft 1 is at the position of compression top dead center (TDC), that is, when the crank angle is 0 degree. Therefore, the degree of deceleration of the crankshaft 1 caused by the compression stroke is represented by the first variation amount Δω1 of the crank angular speed (average engine rotational speed NeA−first crank angular speed ω1).

  Further, the crank angular velocity ω becomes the highest when the crankshaft 1 is at the combustion bottom dead center position, that is, when the crank angle is 180 degrees. Accordingly, the degree of acceleration of the crankshaft 1 due to the combustion / expansion stroke is represented by the second variation amount Δω2 of the crank angular speed from the compression top dead center to the combustion bottom dead center (second crank angular speed ω2−first crank angular speed ω1). The

  Referring to FIG. 4, in the present embodiment, the first crank angular velocity ω1 is set from the falling point C1 of the crank pulse P1 located immediately before the compression top dead center to the rising edge of the crank pulse P2 located immediately after the compression top dead center. It is calculated by the passage time τ1 of the 30 degree section (first predetermined section) up to the falling point C2. The second crank angular velocity ω2 is 30 degrees from the falling point C3 of the crank pulse P3 located immediately before the combustion bottom dead center (BDC) to the falling point C4 of the crank pulse P4 located immediately after the combustion bottom dead center. It is calculated by the passage time τ2 of the section (second predetermined section).

  The first variation Δω1 is calculated by subtracting the crank angular velocity ω1 from the average engine speed NeA, and the second variation Δω2 is calculated by subtracting the first crank angular velocity ω1 from the second crank angular velocity ω2. .

  The first fluctuation amount Δω1 is calculated by subtracting the crank angular speed ω1 from the average engine rotational speed NeA. Next, the relationship between the first variation Δω1 and the in-cylinder gas total mass Gtotal will be described.

The torque fluctuation ΔN of the internal combustion engine (engine) is the difference between the net torque of the internal combustion engine and the running resistance torque. The output torque of the internal combustion engine due to the cylinder pressure is N _{cylindere.work,} and the friction resistance torque of the internal combustion engine is N _friction. When the running resistance torque is N _load , the relationship with the equivalent moment of inertia I of the crankshaft 1 can be expressed by the following equation of motion.
ΔN = (N _{cylindere.work} −N _friction ) −N _load = I · (dω / dt) (1)
Here, when the pressure in the _cylinder is P _cylinder , the cylinder inner diameter is B, the gas constant is R, the gas absolute temperature is T, the cylinder internal volume is V, and the effective radius for torque calculation is r,
N _{cylindere.work} = P _cylinder · (π / 4) B ² · r ... (2)
P _cylinder = Gtotal · R · T / V (3)
And substituting the above equations (2) and (3) into the above equation (1) ignoring the frictional resistance torque N _friction and the running resistance torque N _load ,
dω / dt = (1 / I) · (Gtotal · R · T / V) · (π / 4) B ² · r (4)
It becomes.

As shown in FIGS. 3 and 4, the crank angular speed ω is decelerated before the compression top dead center, and the deceleration gradient (dω / dt) before the compression top dead center is It is possible to approximate between two points, and assuming that the time between the two points is Δτ, the average crank angular velocity, that is, the angular velocity fluctuation amount (first fluctuation amount) from the rotational speed Ne of the internal combustion engine is Δω. ,
dω / dt = Δω / Δτ (5)
It is.

Thus, the angular velocity fluctuation amount Δω before the compression top dead center is based on the average angular velocity ωtdc obtained from the pulse output from the pulser pickup (pulse generator) PC that detects the reluctator 4 (Δω = Ne−ωtdc). And the above equation (4) is
Δω / Δτ = Gtotal · T · (1 / I) · (R / V) · (π / 4) B ² · r (6)
It becomes.

Here, it is assumed that {(1 / I) · (R / V) · (π / 4) B ² · r} is constant and Δτ is constant when the rotational speed Ne of the internal combustion engine is the same. , Δω∝Gtotal · T, and when the suction temperature T is constant, ΔωａｌGtotal, which is Δω∝Gtotal, and the amount of angular velocity fluctuation Δω obtained based on the pulse output from the pulser pickup PC that detects the reluctator 4 The total gas mass Gtotal can be easily estimated. Therefore, the in-cylinder gas total mass Gtotal can be predicted based on Δω1.

  Returning to FIG. 1, in the Δω1-Gtotal map group 36 for each engine speed, a plurality of Δω1-Gtotal maps 37 corresponding to a predetermined engine speed (for example, 10 maps every 1000 rpm up to 1000-10000 rpm). It is stored. The Δω1-Gtotal map 37 is created based on experimental data performed in advance, and shows the relationship between the first fluctuation amount Δω1 and the in-cylinder gas total mass (Gtotal).

  As shown in FIG. 5, the Δω1-Gtotal map shows the relationship between the first fluctuation amount Δω1 and the in-cylinder gas total mass (Gtotal). For example, an experimental engine in which a gas concentration sensor is attached in the cylinder. Is determined in advance. A substantially proportional relationship is established between the first fluctuation amount Δω1 and the total mass of the in-cylinder gas.

  Based on the value of the average engine speed NeA calculated by the NeA calculator 33, the map collating unit 35 sets a Δω1-Gtotal map 37 that matches the engine speed from the Δω1-Gtotal map group 36 for each engine speed. Select one. Then, the in-cylinder gas total mass estimated value deriving unit 38 derives the Gtotal value shown in the Δω1-Gtotal map 37 as an estimated value of the in-cylinder gas total mass.

  On the other hand, the intake fresh air mass detection unit 39 detects the mass of the intake fresh air, that is, the intake fresh air mass (Gf), based on the intake air amount detected by the air flow sensor 15. Then, the EGR rate estimated value calculation unit 41 calculates an estimated value of the EGR rate by a calculation formula of EGR rate (%) = (Gtotal−Gf) ÷ Gtotal × 100.

  A flow of driving and controlling a variable valve timing (VVT) mechanism of the engine 5 using the estimated EGR rate will be described with reference to the block diagram of FIG. A cylinder head 8 having a VVT mechanism for arbitrarily changing the valve timing of the intake and exhaust valves is attached to the upper part of the cylinder 10 of the four-cycle single cylinder engine 5. The VVT mechanism moves the control motor based on the drive command of the ECU 50, thereby changing the valve timing of the intake valve IV and the exhaust valve EV. Note that the amount of valve lift also changes with this change in valve timing. The variable state of the valve timing by the VVT mechanism is transmitted to the ECU 50 by the sensor 19 that detects the rotation angle of the control motor.

  An air cleaner box 16 that filters fresh air is attached to one end of the intake pipe 11. An intake air temperature sensor 17 and an atmospheric pressure sensor 18 are provided inside the air cleaner box 16. In addition, an air flow sensor 15 that measures the amount of intake air, a throttle valve opening sensor 14 that detects the rotation angle of the throttle valve 13, and an intake pressure sensor 20 that detects the intake pressure are attached to the intake pipe 11. An ignition device 9 is provided in the upper part of the combustion chamber, and a fuel injection valve 12 is provided in the intake pipe 11 on the downstream side of the throttle valve 13. An oxygen concentration sensor 7 is attached to the exhaust pipe 6.

  Note that the air flow sensor 15 includes a hot wire type that utilizes a change in electrical resistance when heat is taken away from a heated platinum wire, and a Karman vortex that measures the number of Karman vortices generated in a flow path with ultrasonic waves. A sensor element of the formula can be used.

  The ECU 50 executes various calculations using the output signals of various sensors, and drives and controls the fuel injection device 12, the ignition device 9, and the VVT mechanism. The estimated value of the EGR rate described above is mainly used for controlling the VVT mechanism. In this embodiment, first, the valve timing target value deriving unit 51 derives a valve timing target value using the EGR rate estimated value and the engine load factor. Then, the VVT mechanism target position deriving unit 52 derives the drive amount of the control motor of the VVT mechanism for realizing the target value of the valve timing, and based on this motor drive amount, the VVT mechanism control unit 53 becomes the control motor. A drive signal is output. The engine load factor can be calculated based on the magnitude of Δω2 (see FIG. 4) described above.

  When the engine speed is constant, the ignition timing is set to MBT (Minimum Advance for Best Torque), and the air-fuel ratio (A / F) in the combustion gas is constant, the IMEP (the average effective in the figure) It is known that a proportional relationship is established between (pressure) and ηc (filling efficiency). Thereby, IMEP can be obtained from the air-fuel ratio detected by the oxygen concentration sensor 7 and the above-described second fluctuation amount Δω2. The MBT at the ignition timing is an ignition timing at which the throttle opening is constant, the engine rotational speed is constant, and the generated torque is maximum, and is a value derived in advance through experiments or the like (for example, 3000 rpm) Time is 0 degrees). The charging efficiency ηc is an efficiency related to the mass of fresh intake air that can be taken into the combustion chamber in the intake stroke at a predetermined pressure and a predetermined temperature. Further, IMEP (the indicated mean effective pressure) is a value (for example, 500 kPa) obtained by dividing the work amount generated in the cylinder by the stroke volume (for example, 500 kPa). Regardless of the displacement, the engine performance depends on the work generation degree. It is one of the indicators for expressing.

Here, the relationship between the second variation amount Δω2 and the above-described IMEP and the like will be described. As described above, the angular speed ω of the crankshaft is caused by the fluctuation torque of the crankshaft and pulsates around the average engine rotational speed NeA. Here, the rotational energy increase ΔE (rotational energy increase from the compression top dead center to the combustion bottom dead center) in the combustion / expansion stroke (see FIGS. 3 and 4) is expressed as follows: From the compression top dead center and ω at the end of the combustion / expansion stroke, that is, ω1 and ω2, the following equation is obtained.
ΔE = 1/2 × I × (ω2 ^ 2-ω1 ^ 2) (7)
Since ΔE is a work caused by combustion of the engine, it can be obtained by the following equation.
ΔE = IMIP × displacement Vs (8)

Here, ½ × (ω2 ^ 2-ω1 ^ 2) on the right side of the equation (7) can be converted into the following equation.
1/2 × (ω2 ^ 2-ω1 ^ 2) = (ω2-ω1) × 1/2 × (ω2 + ω1) (9)
From the above, the acceleration amount Δω2 of the crank angular velocity in the combustion / expansion stroke section is defined as the following equation.
Δω2 = ω2−ω1 (10)
Further, 1/2 × (ω2 + ω1) on the right side of the equation (9) is the cycle average ω, and therefore coincides with the average engine speed NeA.
1/2 × (ω2 + ω1) = NeA (11)

From the above equations (7) to (11), Δω2 becomes the following equation.
Δω2 = (IMEP × Vs) / (I × NeA) (12)
That is, Δω2 is proportional to IMEP (the illustrated mean effective pressure) and the displacement Vs, and is inversely proportional to the average engine speed NeA and the crankshaft inertia moment I. Since IMEP is proportional to the engine load factor, the engine load factor can be calculated based on the value of Δω2.

  FIG. 6 is a graph showing combustion characteristics of an HCCI engine that enables premixed compression ignition (HCCI). HCCI is a combustion system in which gasoline is compressed and heated to self-ignite like a diesel engine. In this embodiment, HCCI combustion is realized by using internal EGR. Specifically, since the gasoline engine has a low compression ratio and is difficult to self-ignite, exhaust gas is left in the cylinder by closing the exhaust valve EV quickly, and self-ignition is performed using the thermal energy of the residual exhaust gas.

  This graph shows the relationship between engine torque and engine speed. The region where HCCI combustion is possible is limited to a relatively narrow region where the engine torque and the engine speed are low compared to the region where normal SI (spark ignition) combustion is possible (the shaded area in the figure). Therefore, in order to realize HCCI combustion reliably, accurate detection of the EGR rate is necessary. The “WOT” line indicates the engine characteristics when the throttle is fully open (Wide open throttle), that is, when the load is full.

  FIG. 7 is a graph showing the valve timing varied by the VVT mechanism. In normal SI combustion, the exhaust valve EV starts to close and the intake valve IV starts to open at close valve timing. In SI combustion of the broken line A, TDC (overlap top, exhaust bottom dead center) A state occurs in which both the exhaust valve EV and the intake valve IV are open.

  On the other hand, in the HCCI combustion indicated by the solid lines B and C, by applying the “minus overlap method” in which the timing of closing the exhaust valve is advanced and the timing of opening the intake valve IV is delayed by the VVT mechanism. A large amount of exhaust gas is left inside, and the mixture is self-ignited.

  FIG. 8 is a graph showing the relationship between the EGR rate and the engine load factor. As shown in the figure, HCCI combustion can be realized only in a region where the engine load factor is relatively low, and further has a characteristic that a higher EGR rate is required as the engine load factor decreases. On the other hand, in the SI combustion region, a lower EGR rate is required as the engine load factor increases. A, B, and C in the figure correspond to the valve timings A, B, and C shown in FIG. 7, and the ECU 50 controls the VVT mechanism so that the EGR rate corresponding to the engine load factor is realized. .

  As described above, according to the EGR rate estimation detection apparatus according to the present invention, it is possible to estimate and detect the EGR rate based on the crank angular velocity detected from the crank pulser signal. This makes it possible to appropriately control the VVT mechanism, the ignition device, the fuel injection device, and the like without using a gas concentration sensor or the like for directly detecting the EGR rate.

  The configuration and shape of the crank pulsar rotor and the pulse generator, the form and number of the Δω1-Gtotal map, the configuration in the ECU, and the like are not limited to the above-described embodiments, and various changes can be made. For example, in the above-described embodiment, Δω1 and Δω2 are calculated in a period spanning the compression top dead center and the combustion bottom dead center, respectively, but the calculation positions are advanced in accordance with the engine displacement, the form, and the like. It may be shifted by a predetermined angle in the angular direction or the retarded direction. Further, the length of the period for calculating ω1 and ω2 can also be arbitrarily changed according to the shape of the reluctator of the crank pulsar rotor. Further, the relucters of the crank pulsar rotor can be formed so as to straddle the compression top dead center and the combustion bottom dead center, and Δω1 and Δω2 can be calculated based on the passage times of the respective reluctors.

  Further, a first fluctuation amount (Δω1) is calculated from the difference between the angular velocity at the start of the compression stroke of the engine and the angular velocity near the compression top dead center, while the second predetermined section overlaps the combustion bottom dead center (BDC) of the engine. Δω1, Δω2 calculation unit for calculating the second crank angular velocity (ω2) of (τ2) and calculating the second variation amount (Δω2) by subtracting the first crank angular velocity (ω1) from the second crank angular velocity (ω2). You may make it comprise.

  The EGR rate estimation detection apparatus according to the present invention is used in combination with a load detection apparatus that estimates and detects an engine load based on fluctuations in crank angular speed, and is not limited to a motorcycle engine, but can be applied to various types of engines. Is possible.

  DESCRIPTION OF SYMBOLS 1 ... Crankshaft, 2 ... Crank pulsar rotor, 4 ... Retractor, 15 ... Air flow sensor, 30 ... EGR rate estimation detection apparatus, 31 ... Crank pulse detection part, 32 ... Timer, 33 ... NeA calculation part, 34 ... (DELTA) omega1 calculation part 35... Map collating unit 36... .DELTA..omega.1-Gtotal map group by engine speed 37. .DELTA..omega.1-Gtotal map 38 .. In-cylinder gas mass estimated value deriving unit 39 39. Intake fresh air mass detecting unit 41 ... EGR estimating Value calculation unit, PC ... pulse detector

Claims (5)

Estimating the EGR rate to which a crank pulse is supplied from a pickup (PC) that detects the passage of a plurality of reluctors (4) provided in a crank pulser rotor (2) that rotates in synchronization with the crankshaft (1) of the engine (5) In the detection device (30),
An intake fresh air mass detector (39) for detecting the intake fresh air mass of the engine (5);
A NeA calculator (33) that calculates an average engine speed (NeA) of the engine (5) based on the crank pulse;
A first crank angular speed (ω1) of a first predetermined section (τ1) overlapping the compression top dead center (TDC) of the engine (5) is calculated, and the first crank angular speed (NeA) is calculated from the average engine speed (NeA). a Δω1 calculating unit (34) that calculates the first variation (Δω1) by subtracting ω1);
An in-cylinder gas total mass estimated value deriving unit (38) for estimating the in-cylinder gas total mass (Gtotal) based on the value of the first fluctuation amount (Δω1),
The EGR rate estimation detecting device (30) derives an estimated value of an EGR rate that is a recirculation rate of exhaust gas based on the estimated value of the intake fresh air mass and the total mass of the in-cylinder gas. EGR rate estimation detection device. A relationship between the first fluctuation amount (Δω1) and the in-cylinder gas total mass (Gtotal) is shown, and a plurality of Δω1-Gtotal maps (37) provided for each predetermined engine speed are provided.
The in-cylinder gas total mass estimated value deriving unit (38) selects one Δω1-Gtotal map (37) that matches the average engine speed (NeA), and the correspondence with the first fluctuation amount (Δω1). 2. The EGR rate estimation detection apparatus according to claim 1, wherein the Gtotal value derived by the above equation is used as an estimation value of the total mass of the in-cylinder gas. The engine (5) is configured to be capable of premixed compression ignition combustion,
3. The EGR rate according to claim 1, wherein the EGR rate is an EGR rate of an internal EGR that causes exhaust gas to remain in the cylinder according to opening / closing timings of the intake valve (IV) and the exhaust valve (EV). Inference detection device.   The first predetermined interval (τ1) is from the falling point of the crank pulse located immediately before the compression top dead center (TDC) to the falling point (C2) of the crank pulse located immediately after the compression top dead center (TDC). The EGR rate estimation detection apparatus according to claim 1, wherein the EGR rate estimation detection apparatus is a period up to A second crank angular speed (ω2) in a second predetermined section (τ2) overlapping the combustion bottom dead center (BDC) of the engine (5) is calculated, and the first crank angular speed (ω1) is calculated from the second crank angular speed (ω2). ), A second variation amount (Δω2) is calculated, and a Δω2 calculation unit is provided.
An engine load factor is calculated based on the value of the second fluctuation amount (Δω2),
5. The EGR rate estimation detection apparatus according to claim 1, wherein a target value of valve timing is calculated based on the estimated value of the EGR rate and the engine load factor.

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