JP2005133603A - Intake gas temperature inferring device for engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide technolgy for inferring gas temperature in a cylinder accurately without requiring many matching processes irrespective of change of engine specifications by putting physical phenomenon related to behavior of intake gas in a suction system of an engine including the inside of the cylinder into a model. <P>SOLUTION: This engine provided with a fuel supply device for injecting and supplying fuel into the suction system is provided with a means A01 for obtaining suction air temperature, a means A02 for obtaining vaporization characteristic of fuel spray supplied to the suction system, a means A03 for obtaining gas flow speed and pressure in a section for which fuel is vaporized in the suction system, and a gas temperature computing means A04 for inferring temperature of mixed gas in the cylinder using the result of each means. Each means constitutes a model of phenomenon in which vaporization latent heat of floating fuel in suction air changes gas temperature so that gas temperature in the cylinder can be inferred by calculating gas temperature in a suction system part up to the inside of the cylinder on the downstream side more than a fuel supply section in the suction system. In this case, adaptation on desk becomes possible individually per change section in accordance with the model irrespective of change of specifications of the fuel supply device and control characteristics. <P>COPYRIGHT: (C)2005,JPO&NCIPI

Description

  The present invention relates to a technique for estimating the temperature of air or mixed gas sucked into a cylinder of an internal combustion engine.

  In order to optimally control the fuel supply amount and ignition timing of the internal combustion engine, or to accurately estimate the generated torque, it is necessary to accurately determine the gas temperature when the intake valve is closed or when compression starts. As the prior art relating to the estimation of the intake gas temperature, those disclosed in the following Patent Documents 1 to 3 are known.

Patent Document 1 stores a method for estimating the amount of heat transmitted to intake air by a residual heat amount obtained by subtracting a heat release amount from an engine room from a heat generation amount per unit time of an internal combustion engine, and an engine body represented by a coolant temperature. A technique for estimating the intake air temperature by correcting the outside air temperature based on a method for estimating the amount of heat transmitted to the intake system from the amount of heat is disclosed. Patent Document 2 discloses a technique for estimating the intake air temperature using an equation of state from the intake pipe pressure of the engine, the amount of air passing through the throttle, and the intake pipe volume from the throttle to the intake valve. Patent Document 3 discloses a technique for treating intake air as a mixed gas composed of an intake gas component and a residual gas component, and calculating a mixed gas temperature from each temperature and mass of the intake gas and the residual gas using a predetermined arithmetic expression. ing.
JP 09-189256 A Japanese Patent Laid-Open No. 05-180057 JP-A-11-148419

  The intake gas temperature in the cylinder of the engine fluctuates depending on the vaporization state of the fuel supplied during intake, and this tendency is particularly remarkable in a vehicular engine in which fluctuations in the fuel supply amount and the intake air flow rate are large. However, since the conventional method does not sufficiently take into account the temperature variation factor related to the fuel vaporization characteristics, there is a problem that the temperature estimation accuracy is basically low.

  The accuracy can be raised to a practical level by performing a correction using an experimental technique called matching, but in an internal combustion engine where the operating conditions and operating environment vary widely, the number of man-hours for matching is very large. There is a problem of becoming. In addition, once matching is completed, if the part is replaced with a part having a different specification, it is unclear how the characteristics of the part affect the temperature estimation. Therefore, the matching process with many man-hours must be repeated from the beginning.

  By modeling physical phenomena related to the behavior of intake gas in the engine intake system including the inside of the cylinder, the present invention accurately estimates the gas temperature in the cylinder without requiring many matching steps regardless of engine specification changes. Provide technology that can.

  In the present invention, a gas temperature estimation model as shown in FIG. 7 is formed on the premise of an engine provided with a fuel supply device that injects and supplies fuel to the intake system. In the figure, A01 is a means for obtaining the intake air temperature, A02 is a means for obtaining the vaporization characteristics of the fuel spray supplied to the intake system, A03 is a means for obtaining the gas flow velocity and pressure of the fuel vaporization target portion in the intake system, and A04 is the above-mentioned Gas temperature calculation means for estimating the temperature of the air-fuel mixture in the cylinder using the result of each means.

  The above-described configuration is obtained by modeling the phenomenon of changing the gas temperature based on the latent heat of vaporization when the fuel floating in the intake air evaporates according to the laws of physics. In this case, the gas temperature in the cylinder when the target intake valve is closed can be estimated by examining the gas temperature in the intake system portion up to the cylinder downstream of the fuel supply portion of the intake system.

  According to the present invention, regardless of changes in the specifications and control characteristics of the fuel supply device including the fuel injection valve, desktop correction can be made individually for each changed portion according to the model. Therefore, the matching man-hours for obtaining an accurate intake gas temperature can be reduced regardless of changes in engine specifications.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings. FIG. 1 is a schematic diagram for explaining a system according to an embodiment of the present invention applied to an L-Jetronic gasoline injection engine.

  The air metered by the intake throttle valve 23 is stored in the intake collector 2 and then introduced into the combustion chamber 5 of each cylinder via the intake manifold 3. The engine speed is calculated based on the intake air flow rate detected by the air flow meter 32 and the signal from the crank angle sensor (33, 34) from the fuel injection valve 21 arranged in the intake port 4 of each cylinder. Accordingly, the injection is intermittently supplied into the intake port at a predetermined timing, more specifically, toward the intake valve 15 (back of the umbrella) that exists so as to be blocked by the intake port.

  The fuel injected toward the intake valve 15 is mixed with the intake air to form an air-fuel mixture. The air-fuel mixture is confined in the combustion chamber 5 by closing the intake valve 15, compressed by the rise of the piston 6, and ignited. It is ignited by the plug 14 and burns. The gas pressure due to the combustion works to push down the piston 6, and the reciprocating motion of the piston 6 is converted into the rotational motion of the crankshaft 7. The combusted gas (exhaust gas) is discharged into the exhaust passage 8 when the exhaust valve 16 is opened.

  A three-way catalyst 9 is provided in the exhaust passage 8. The three-way catalyst 9 can efficiently remove HC, CO and NOx contained in the exhaust gas simultaneously when the air-fuel ratio of the exhaust gas is in a narrow range centered on the stoichiometric air-fuel ratio. For this reason, the engine controller 31 determines the basic fuel injection amount from the fuel injection valve 21 in accordance with the operating conditions, and sets the air-fuel ratio based on a signal from an oxygen sensor (not shown) provided upstream of the three-way catalyst 9. Feedback control.

  The intake throttle valve 23 is driven by a throttle motor 24. Since the torque required by the driver appears in the amount of depression of the accelerator pedal 41 (accelerator opening), the engine controller 31 determines a target torque based on a signal from the accelerator sensor 42, and a target air for realizing this target torque. The amount is determined, and the opening degree of the intake throttle valve 23 is controlled via the throttle motor 24 so that this target air amount is obtained.

  Further, mainly for improving fuel efficiency, an EGR device (comprising an EGR passage 25, an EGR valve 26, and an actuator 27) and a variable valve mechanism 29 that variably controls the intake valve operating angle or the lift amount are provided.

  The intake throttle valve 23 is provided with a hot water heater 61 as a heating means for preventing freezing in the throttle chamber 60 during cold weather. Cooling water is supplied to the hot water heater 61 from the engine body via the hot water passage 62. Is done. A blow-by gas passage 63 is connected between the upstream portion of the throttle chamber and the intake collector, and the gas containing unburned fuel components blown from the cylinder gap to the crankcase is replaced with fresh air taken from the upstream portion of the throttle. The fuel is supplied to the intake collector 2 via the cylinder head for combustion treatment. A purge gas passage 64 of a fuel evaporative gas processing device is connected to the intake collector 2. The fuel evaporative gas in the fuel tank 65 is temporarily adsorbed by the canister 66. When the purge valve 67 is opened, the fuel evaporative gas is desorbed from the atmosphere into the air introduced into the canister at that time. Through the intake collector 2.

  The feature of the present invention is to estimate the intake gas temperature in the cylinder by paying attention to the temperature change caused by the vaporization of the fuel supplied to the intake system. In this embodiment, FIG. As shown in the conceptual diagram, the temperature change due to the heat transfer effect between the wall surface of the specific part of the intake system and the intake air, and the mixing with the gas supplied to the intake system mainly from the outside such as purge gas and EGR gas By building models that take into account temperature changes, boost changes during sudden acceleration / deceleration, and temperature changes associated with adiabatic compression or expansion of gas before and after the intake throttle valve, it is possible to obtain more accurate estimation results. Yes.

  The initial temperature that serves as a reference for the intake gas temperature estimation is a portion that is separated from the heat source for intake air and that does not need to consider heat transfer from inside and outside, for example, in the engine system of FIG. The intake air temperature in the vicinity of the air flow meter 32 is set. When there is no heat source as described above, the intake air temperature at a portion closer to the intake port can be used as a reference for temperature estimation. In an engine mounted on a vehicle, it is possible to estimate the intake air temperature as the initial value in consideration of various conditions such as the atmosphere in the engine room, the outside air temperature, and the radiant heat of the radiator. Since the air in the room has a complicated and unstable temperature distribution, it is difficult to accurately estimate the temperature from these elements. Therefore, it is preferable to provide a temperature sensor 43 in the air flow meter or the like to directly control the intake air temperature. To detect.

Next, details of the mixture temperature estimation will be described. In this specification or the accompanying drawings, “*” is used as a multiplication symbol.
1. Initial intake air temperature Ta0 to intake air temperature Ta1 after passing through the throttle chamber
First, the estimated value or detected value of the intake air temperature of the air flow meter 32 is Ta0. Next, the intake air temperature Ta1 after passing through the intake throttle valve is obtained using the initial intake air temperature Ta0 as a reference. At this time, when the throttle valve opening is large, it can be considered that Ta1 = Ta0. When the throttle valve opening is small, that is, under the condition that the temperature drops due to adiabatic expansion of the intake air after passing through the intake throttle valve 23, Ta1 is calculated using the following equation (1).

Pa0: Atmospheric pressure
Pa1: Pressure at the collector 2 part
Pa0 and Pa1 may be measured or estimated by the pressure sensors 44 and 45, respectively. κ is the specific heat ratio of air, and 1.4 is used in the standard state.
2. Intake air temperature Ta2 after passing through the heater
Next, the intake air temperature Ta2 after passing through the icing prevention heater 61 of the throttle chamber 60 is calculated by the following equation (2).

Tw: Engine coolant temperature
Ta1: Intake air temperature after passing through the throttle chamber
K: Constant determined by heat capacity and heat transfer coefficient of cooling water In this case, the cooling water temperature Tw is applied as a representative value of the wall surface temperature, and the engine rotational speed Ne is applied as a representative value of the gas flow rate and the cooling water flow rate. Although explanation is omitted, temperature estimation by heat transfer from the port wall surface can be performed for the intake port 4 in the same manner as described above.
3. Intake air temperature after mixing external gas Ta3
Next, an air-fuel mixture temperature Ta3 after mixing with the purge gas from the canister 66, external EGR gas, and internal EGR gas (residual gas in the cylinder after the intake stroke) is calculated. In the engine system of FIG. 1, blow-by gas is introduced into the intake collector 2, but here, gas introduced from the outside in addition to EGR gas will be representatively described as purge gas.
3-1. Estimation of internal EGR gas temperature First, the internal EGR gas temperature Tevc5 is calculated. Based on the exhaust temperature detected by the exhaust temperature sensor 46, the in-cylinder temperature Tevc0 when the exhaust valve is closed is calculated. Since the in-cylinder temperature Tevc0 when the exhaust valve is closed varies depending on the amount of heat corresponding to the difference between the fuel injection amount and the amount of work at that time, it may be obtained from a table using such characteristics.

  Next, the in-cylinder pressure Pecv when the exhaust valve is closed is estimated from the exhaust pressure detected by the exhaust pressure sensor 47. The in-cylinder pressure Pecv when the exhaust valve is closed is determined by the mixture volume and the in-pipe resistance of the exhaust system, and may be obtained from a table corresponding to the mixture volume flow rate.

Next, the fresh air when the intake valve is opened and the exhaust pressure Peivc before mixing are calculated. here,
(a) When the intake valve opening timing is before the exhaust valve closing timing, that is, when there is an overlap period, Peivc / Pevc = 1.0,
(b) When the intake valve opening timing is after the exhaust valve closing timing, that is, when there is no overlap period, Peivc / Pevc is calculated from a table corresponding to the valve timing.

At this time, the above-described table sets values by calculating in advance from the relationship among exhaust valve closing, intake valve opening, and top dead center. In particular,
(a) When the exhaust valve closing timing is before the top dead center, adiabatic compression is performed from the exhaust valve closing to the top dead center, so Peicc / Pevc> 1.0. Therefore, Peivc / Pevc <1.0.
(b) When the exhaust valve closing timing is after top dead center, since the adiabatic expansion is performed from the exhaust valve closing to the intake valve opening, Peivc / Pevc <1.0.

  Next, the exhaust gas specific heat ratio SHEATR is calculated. FIG. 3 is a calculation table of the exhaust gas specific heat ratio, where the horizontal axis indicates the target equivalent ratio TFBYA during combustion and the vertical axis indicates the exhaust gas specific heat ratio SHEATR. TFBYA is the fuel-air ratio value when the fuel-air ratio at stoichiometric is 1. The dotted line in the figure indicates the stoichiometric position. When the combustion equivalent ratio TFBYA is close to the stoichiometric ratio, the exhaust gas specific heat ratio SHEATR decreases, and when it becomes rich or lean, the exhaust gas specific heat ratio = SHEATR increases. . Also, the thick line arrow in the figure shows the case where the in-cylinder temperature TEVC when the exhaust valve is closed changes.

  The internal EGR gas temperature Tevc is calculated from the thus calculated Peivc / Pevc, exhaust gas ratio SHEATR, and exhaust valve closing cylinder temperature Tevc0 using the following equation (3).

3-2. Estimation of external EGR gas temperature Next, the external EGR gas temperature Tegr is calculated. First, based on the exhaust gas temperature detected by the exhaust gas temperature sensor 46, the EGR gas temperature Tegr0 upstream of the EGR valve is calculated. Next, the EGR gas pressure Pegr0 upstream of the EGR valve is calculated from the exhaust pressure detected by the exhaust pressure sensor 47. Next, the EGR gas pressure Pm downstream of the EGR valve is calculated from the pressure sensor 45 or based on the pressure in the collector estimated from the intake air amount, the intake temperature, and the collector volume. Next, the exhaust gas specific heat ratio SHEATR1 is read out from the same table as in FIG. In this case, the thick arrow in FIG. 3 corresponds to the case where the EGR gas temperature Tegr0 upstream of the EGR valve changes.

  From the EGR gas pressure Pegr0 upstream of the EGR valve, the EGR gas pressure Pm downstream of the EGR valve, the EGR gas temperature Tegr0 upstream of the EGR valve, and the exhaust gas specific heat ratio SHEATR1 calculated in this way, the internal EGR gas is calculated using the following equation (4). Calculate the temperature Tegr.

In the present embodiment, the temperature when the external EGR returns to the collector 2 part is obtained, but there is a system that returns to the intake port instead of the collector part, and in that case, it can be calculated by the same method.
3-3. Temperature Estimation at the Time of Mixing External Gas Next, the gas mixture temperature Ta3 after mixing with the purge gas from the canister 66, the external EGR gas, and the internal EGR gas is calculated from the following equation (5).

Ca: Specific heat of intake air
Ma: Intake air volume
Cegr: Specific heat of external EGR gas
Megr: External EGR gas amount
Cevp: Specific heat of purge gas
Mevp: Purge gas amount
Tevp: purge gas temperature
Cegr is calculated from combustion TFBYA and external EGR gas temperature Tegr. Megr has a correlation with the EGR valve opening area obtained from the EGR valve opening and the ratio between the intake negative pressure and the exhaust pressure, and is obtained using the correlation.

  Tevp is directly detected by the temperature sensor 48, and Mevp can be obtained as a function of the opening degree of the purge valve 67 and the pressure of the purge passage outlet (intake collector 2) detected by the pressure sensor 73. The purge gas amount calculated here is the total mass of the air introduced from the canister 66 and the fuel evaporative gas desorbed from the canister during the purge, and the ratio of the total gas flow rate to the desorption amount of the fuel evaporative gas, that is, the fuel of the purge gas The specific heat of the purge gas can be accurately determined according to the concentration.

Various methods are conceivable for calculating the amount of desorbed fuel held by the purge gas, and here, as an example, the physics proposed by the present applicant in Japanese Patent Application No. 2001-71562 (Japanese Patent Laid-Open No. 2002-276436). The model is shown in FIG. In the figure, B41 is an adsorption amount calculation unit, which calculates the current adsorption amount from the previous value of the adsorption amount of the evaporated fuel and the previous value of the desorption amount. B42 is a reference desorption amount calculation unit, which calculates the desorption amount at the reference purge flow rate from the activated carbon temperature, adsorption amount, predetermined desorption constant and desorption index from the activated carbon temperature calculation unit B44. B43 is a flow rate equivalent desorption amount calculation unit, which calculates a desorption amount corresponding to the purge flow rate using the purge flow rate (purge rate * engine intake air amount) and the calculation result of the reference desorption amount calculation unit B42. In this canister model, as described above, the temperature of the activated carbon that adsorbs the evaporated fuel is used in the desorption amount calculation, so that the influence of the change in the activated carbon temperature according to the desorption amount is compensated to obtain an accurate desorption amount. be able to.
4). Mixture temperature Ta4 when passing through intake valve
Next, an air-fuel mixture temperature Ta4 when passing through the intake valve is calculated. First, the intake air temperature Ta41 after passing through the intake port is calculated from the following equation (6).

Tw: Engine coolant temperature
Ne: Engine speed as a representative value of gas flow rate or cooling water flow rate
K: Constant determined by heat capacity and heat transfer coefficient of cooling water Next, the air-fuel mixture temperature Ta42 in the case of adiabatic compression during sudden acceleration or adiabatic expansion during sudden deceleration is calculated from the following equation (7).

Pm: Pressure in the intake manifold
Pc: In-cylinder pressure
MIXAIRSHR: Mixture specific heat ratio PC = Pm under operating conditions other than rapid acceleration or rapid deceleration. At the time of sudden acceleration or sudden deceleration, in the experiment by the present applicant, Pc <Pm at the time of sudden acceleration, and Pc> Pm at the time of sudden deceleration. The determination of sudden acceleration or sudden deceleration is made using a signal from the accelerator sensor 42, for example. FIG. 5 is a calculation table for obtaining the mixture specific heat ratio MIXAIRSHR. The horizontal axis represents the combustion equivalent ratio TFBYA, and the vertical axis represents the mixture specific heat ratio MIXAIRSHR. The dotted line in the figure indicates the stoichiometric ratio, and the specific heat ratio MIXAIRSHR is large on the lean side and small on the rich side.

  Next, an air-fuel mixture temperature Ta43 when the air-fuel mixture causes choke at the intake valve is calculated. In a variable valve mechanism that controls the intake air amount by variably controlling the intake valve lift amount, it is known that the air-fuel mixture causes choke in the intake valve particularly at a low lift amount. In that case, the mixture temperature = Ta43 is calculated by the following equation (8).

Pc: In-cylinder pressure when intake valve is closed
Pport: Intake port pressure (= collector pressure Pm)
MIXAIRSHR: Mixture specific heat ratio (see Fig. 5)
Finally, assuming that Ta4 = Ta43, the mixture temperature calculation process when the intake valve passes is terminated.
5). Mixture temperature Tivc when intake valve is closed
5-1. Temperature estimation considering the effect of latent heat of vaporization of fuel (corresponding to the invention of each claim of the present application)
Next, when calculating the mixture temperature Tivc when the intake valve is closed, first, the mixture temperature Ta5 when the fuel injected from the injection valve 21 is affected by the latent heat of vaporization of the droplets in the intake port and the combustion chamber is calculated. To do. Here, a method for calculating the vaporization amount Mx0 ′ with respect to the spray particle size distribution (mass ratio), which is necessary for calculating the mixture temperature Ta5, will be described.

  First, the branch model of the injection valve spray is set as shown in FIG. In other words, the spray particle size distribution calculating means C41, the injection-time vaporization ratio calculating means C42, the direct injection ratio calculating means C43, the intake system floating ratio calculating means C44, the combustion chamber floating ratio calculating means C45, the intake system adhesion ratio. An allocation unit C46, a combustion chamber adhesion rate allocation unit C47, and a vaporization / floating rate calculation unit C48 are included.

  First, the spray particle size distribution calculating means C41 reads the spray particle size distribution stored in advance in the ROM in the engine controller 31. Here, the particle size distribution of the spray is a matrix of the mass ratio of the spray for each small particle size classification (for each particle size), and the calculation of the spray particle size distribution is performed from the ROM in the engine controller 31. This is an operation to read out a matrix of the mass ratio of the spray for each small section of the particle size.

  The injection vaporization ratio calculating means C42 calculates the spray vaporization rate for each of the small particle size classifications from the signals such as temperature, pressure, flow velocity, etc., and sums these for all the particle size classifications to obtain the total during injection. Vaporization during injection X0 ′ [%], which is the amount of vaporization from the spray, is calculated. As a result, the spray XB [%] of 100−X0 ′ remains in the intake port 4 without being vaporized.

  The direct injection ratio calculation means C43 receives this residual spray amount XB (= 100−X0 ′) from the injection-time vaporization ratio calculation means C42, and this, the injection timing I / T, and the sandwiching between the injection valve 21 and the intake valve 15 Using the angle β, a spray amount XD [%] that is directly injected into the combustion chamber 5 without colliding with the intake valve 15 or the intake port 4 is calculated. As a result, XB-XD spray XC [%] remains in the intake port 4. The spray XC remaining in the intake port 4 is output to the intake system floating ratio calculating means C44, and the spray XD directly injected into the combustion chamber 5 is output to the combustion chamber floating ratio calculating means C45.

  In the intake system floating ratio calculating means C44, the vaporization rate of the spray for each small particle size classification is calculated, and these are summed up for all the particle size classifications, so that the floating portion X0 ″ [%] at the intake port 4 is calculated. And the remainder of the spray that adheres to the intake port wall 4a and the intake valve wall 15a (hereinafter, the spray that adheres to the intake port wall 4a and the spray that adheres to the intake valve wall 15a are collectively referred to as “intake system”. Calculated as “attachment” XE (= XC−X0 ″) [%].

  Similarly, the combustion chamber floating ratio calculating means C45 calculates the vaporization rate of the spray for each of the small particle size divisions, and sums up these for all the particle size divisions, whereby the floating portion X0 '' in the combustion chamber 5 is calculated. '[%], And the remainder adheres to the combustion chamber wall (excluding the cylinder surface wall as described above) and the cylinder surface wall 52 (hereinafter referred to as the spray adhering to the combustion chamber wall and the cylinder surface wall 52). The spray is collectively called “combustion chamber deposit”.) Calculated as XF (= XD−X0 ′ ″) [%].

  The vaporization / floating ratio calculation means C48 adds up the three of the vaporization amount X0 ′ during injection, the floating portion X0 ″ at the intake port 4 and the floating portion X0 ′ ″ at the combustion chamber 5 thus obtained. Calculate vaporization and floating part X0 in one injection total.

  On the other hand, in the intake system adhesion ratio allocation means C46, the intake system adhesion part XE is divided into the part X1 [%] attached to the intake valve wall 15a and the part X2 [%] attached to the port wall 4a, and the combustion chamber adhesion ratio. The allocating means C47 allocates the combustion chamber adhering portion XF to the portion X3 [%] adhering to the combustion chamber wall and the portion X4 [%] adhering to the cylinder surface wall 52, respectively.

Next, a description will be given for the model identification of spray branching.
<1> Spray branch model identification (spray branch overall process)
Fig. 6-2 shows the process of the entire spray branch used for estimation (identification) of each spray branch (X0, X1, X2, X3, X4) as a model. As shown in the figure, it is decomposed into six in time series.

1) Vaporization during injection:
An injection spray is a collection of fuel sprays having different particle sizes. Therefore, if the particle diameter D [μm] is taken on the horizontal axis and the mass ratio [%] of the spray is taken on the vertical axis, as shown in the upper left corner of FIG. (See thick solid line), the area surrounded by the chevron curve is 100%, which is the total sum of the total spray during injection. A part of the fuel spray having a mountain-shaped distribution is vaporized at the time of injection, and the rest stays in the spray. Since the spray with a smaller particle size is more easily vaporized, the distribution of the spray that remains without being vaporized (see the thin solid line) is smaller on the smaller particle size side than the spray distribution during spraying (XA). The area between these two distributions is the spray X0 ′ [%] that is vaporized at the time of injection, and 100−X0 ′ is the spray XB [%] that remains without being vaporized.

2) Injection spray directly into the combustion chamber:
In the second characteristic from the upper left of FIG. 6-2, a large mountain (see thick solid line) is a distribution of the spray that remains in the intake port 4 without being vaporized, and of this, the spray directly injected into the combustion chamber 5 The distribution is drawn over small mountains (see thin solid lines). This small mountain area is the spray fraction XD [%] that is directly injected into the combustion chamber 5, and XB-XD, that is, the spray fraction XC [%] where the area between the large mountain and the small mountain remains in the intake system. It is.

3) Inhalation system spray adhesion floating:
Part of the spray that is not directly injected into the combustion chamber 5 and remains in the intake port (intake system) floats as it is (including vaporization), and the rest is the wall surface of the intake system (port wall 4a and intake air). It adheres to the valve wall 15a). The smaller the particle size, the more likely it is to float as it is sprayed. Therefore, the distribution of the spray adhering to the wall surface of the intake system (see the thin solid line) in the second characteristic from the upper right of Fig. 6-2 is the distribution of the spray remaining in the intake system. The side with a smaller particle diameter is smaller than that (see thick solid line). The area between these two distributions is the amount that floats in the inhalation system as a spray (the air floating rate in the inhalation system) X0 '' [%], and this floating from the spray portion XB that remains in the inhalation system The value obtained by subtracting the minute X0 ″ is the intake system attachment XE (intake system attachment ratio) [%].

4) Combustion chamber spray adhesion floating:
A part of the spray directly injected into the combustion chamber 5 floats in the combustion chamber 5 as a spray (including the portion to be vaporized), and the rest adheres to the combustion chamber wall and the cylinder surface wall 52. As the spray having a smaller particle size is more likely to float as it is, the distribution of the spray adhering to the combustion chamber wall and the cylinder surface wall 52 (see the thin solid line) in the second characteristic from the lower right of FIG. The side with a smaller particle size is smaller than the distribution of spray sprayed directly (see thick solid line). The area between these two distributions is the amount that floats in the combustion chamber 5 while being sprayed (the air floating ratio in the combustion chamber 5) X0 '''[%], and is directly injected into the combustion chamber 5 The value obtained by subtracting the floating portion X0 ′ ″ from the sprayed portion XD is the combustion chamber wall deposit (combustion chamber deposit ratio) XF [%].

5) Intake system spray attachment location:
In the upper right characteristic of FIG. 6-2, the large mountain (see thick solid line) is the distribution of XE adhering to the intake system, and the small mountain (see thin solid line) is the distribution of spray adhering to the intake valve wall 15a. . The area of this small mountain is the spray X1 [%] adhering to the intake valve wall 15a, and the value obtained by subtracting this intake valve wall X1 from the intake system adhering XE is the port wall adhering X2 [%]. is there.

6) Combustion chamber spray deposit location:
In the characteristic at the lower right end of FIG. 6B, the large mountain (see thick solid line) is the distribution of the above-mentioned combustion chamber adhering portion XF, and the small mountain (see thin solid line) is the distribution of the spray adhering to the combustion chamber wall. The area of this small mountain is the combustion chamber wall deposit X3 [%], and the value obtained by subtracting this combustion chamber wall deposit X3 from the combustion chamber deposit XF is the cylinder surface wall deposit X4 [%].

  In this way, intake system residuals XB, XC, directly injected spray XD, intake system adhering XE, combustion chamber wall adhering XF, injection vaporization X0 ', floating X0' ', X0' '' Is the same unit [%], but only XA represents the distribution itself.

Hereinafter, the calculation method of the particle size distribution XA of the spray during spraying and the branched portions XB, XC, XD, XF, X0 ′, X0 ″, and X0 ′ ″ will be described in detail.
<2-1> Model identification of spray branching (vaporization)
1) XA: Particle size distribution of spray during spraying:
As the particle size distribution XA for the mass ratio of the spray during injection, the spray measurement result of the injection valve 21 is used.

The particle size classification of the spray may be equally spaced (for example, every 10 μm) (see (a) of FIG. 6-3) or may be divided every 2 ⁿ (see (b) of FIG. 6-3). . The greater the number of particle size categories, the better the accuracy. However, on the other hand, the memory capacity and calculation time increase, so it may be designed according to the CPU capacity.

  For simplicity, only one particle size classification may be used. This means that the average particle size of the total spray during injection is used. In this case, the evaporation rate and the retention rate of the spray are approximately determined from the particle diameter, and when the particle diameters are similar, the evaporation and retention characteristics can be approximated by experimental values. However, since the spraying method and the injection valve in which the spray particle size distribution changes greatly do not match, the spray particle size distribution may be used at this time.

2) X0 ': Evaporation during injection:
As for the vaporization of the spray during injection, as shown in FIG. 6-4, the mass of the spray is m, the surface area is A, the diameter is D, the amount of vaporization of the spray is Δm, the flow velocity of the intake port 4 is V, and the temperature of the intake port 4 Is T, and the pressure of the intake port 4 (this pressure is lower than the atmospheric pressure and becomes negative if the atmospheric pressure is used as a reference) is P, the vaporization rate X0 ′ and the vaporization amount Δm are expressed by the following equations. Is done.
X0 '= Δm / m… (a1)
Δm = f (V, T, P) * A * t… (a2)
Here, f (V, T, P) in the equation (a1) is a unit surface area and an evaporation amount per unit time (this value is hereinafter referred to as “vaporization characteristic”), and the vaporization characteristic f (V, T, P). Represents a function of flow velocity V, temperature T, and pressure P. In the formula (a2), t is a unit time.

In this case, since A = D2 * K1 # and m = D3 * K2 # (K1 # and K2 # are constants), substituting these into the equations (a1) and (a2) and deleting Δm The formula is obtained.
X0 '= ΣXAk * f (V, T, P) * A * t * KA # / Dk… (a3)
Where XAk is the mass ratio to the particle size of the kth segment, Dk is the particle size of the kth segment, and Σ represents the summation over all particle size categories (from 1 to the maximum number of categories for k) ing. KA # is the effective utilization rate (constant smaller than 1) at the surface area of the gas flow velocity V.

  The vaporization characteristic f (V, T, P) is obtained by searching a characteristic map having the contents shown in FIG. 6-5 from the temperature T and the flow velocity V. As shown in FIG. 6-5, the vaporization characteristic f (V, T, P) increases as the temperature T increases and the flow velocity V increases. In FIG. 6-5, the temperature on the horizontal axis is widely set from −40 ° C. to 300 ° C., but in reality, the vaporization and evaporation of the spray are performed in the region indicated as “temperature range”.

  The second term (Pa-P) / Pa · # KPT on the horizontal axis is the temperature correction due to the pressure P. This is because the volatility difference due to pressure P, that is, when the pressure P is lower than the atmospheric pressure Pa as at low load, the amount of evaporation is larger than when the pressure P is higher than at low load as at high load. It is taken into consideration.

By the way, among the parameters of the vaporization characteristics f (V, T, P), the flow velocity V includes a relative flow velocity due to the penetration force of the spray and a flow velocity due to the intake of the combustion chamber. The sum of the spray penetration and the intake airflow is obtained by the following equation instead of the above equation (a3).
X0 '= ΣXAk * f (V1, T, P) * A * t1 * KA # / Dk + ΣXAk * f (V2, T, P) * A * t2 * KA # / Dk… (a4)
V1: Spraying speed by spray penetration force,
t1: Time required for spray penetration,
V2: speed of intake airflow,
t2: The time that the spray is exposed to the intake airflow,
Here, the spray velocity V1 due to the spray penetration force and the time t1 required for the spray penetration are constant values if the fuel pressure Pf acting on the injection valve 21 is determined. The values of V1 and t1 are determined when the specifications of the injection valve 21 are determined. In an engine in which the fuel pressure Pf is variably controlled, V1 and t1 vary depending on the fuel pressure Pf, and therefore are set as a function of the fuel pressure Pf.

Since the intake of air into the combustion chamber 5 is intermittent, the speed of the intake airflow (flow velocity of the intake port 4) V2 is proportional to the engine speed Ne, that is, V2 can be calculated by the following equation.
V2 = Ne * # KV… (a5)
#KV: Flow velocity index,
The flow velocity index #KV in the equation (a5) is a value determined by a value obtained by dividing the flow path area (flow path area of the intake port 4) by the cylinder volume. This index also includes unit alignment. Here, the flow path area and the cylinder volume can be obtained from the drawings.

  Since the exposure time t2 of the intake airflow that represents the degree of exposure to the flow rate of the spray is affected by the injection timing I / T and the engine rotational speed Ne, FIG. Find by searching the map to be.

  Of the parameters of the vaporization characteristics f (V, T, P), the intake air temperature is used as the temperature T. However, when considering the residual gas (external EGR gas or internal EGR gas), the gas temperature mixed with the residual gas is used. This gas temperature is estimated from the intake air temperature and the water temperature. Simply, a simple average value or a weighted average value of the intake air temperature and the water temperature may be used as the estimated value of the gas temperature. The intake air temperature and the water temperature detect the respective sensor values. Ignore the heat of vaporization and respond with adaptation. Of the parameters of the vaporization characteristic f (V, T, P), the intake pressure is used as the pressure P. The intake pressure is detected by a pressure sensor 45 provided in the intake collector 2.

3) XB: Spray remaining in the intake port:
When the vaporization amount X0 ′ during injection is obtained in this way, the spray amount XB remaining in the intake port 4 while being sprayed is given by the following equation.
XB = XA−X0 '(a6)
<2-2> Model identification of spray branching (direct injection)
1) XD: Spray amount directly injected into the combustion chamber 5:
If the spray from the injection valve 21 is an injection during the exhaust stroke, the intake valve 15 is fully closed, so it only hits the intake valve 15 and the intake port 4, but it aims at the back of the intake valve umbrella. When injecting in the intake stroke, a part thereof is directly injected into the combustion chamber 5 through the gap between the intake valve 15 and the valve seat without colliding with the intake valve 15 or the intake port as shown in FIG. This direct injection rate is defined as KXD, and the spray fraction XD that is directly injected into the combustion chamber 5 is calculated by the following equation.
XD = XB * KXD… (a7)
In addition to the injection timing, the direct injection rate KXD is also affected by the injection direction (the direction of the injection valve 21 and the direction of the intake valve 15). Therefore, the direct injection rate KXD is obtained by searching a map having the contents shown in FIGS. 6-8 from the injection timing I / T and the sandwich angle β between the axis of the injection valve 21 and the axis of the intake valve 15. The sandwiching angle β is obtained from the drawing. The characteristics shown in FIGS. 6-8 are obtained by matching.

Further, in an engine having a variable valve mechanism, the valve lift and profile of the intake valve also directly affect the injection rate KXD. Therefore, in the engine, the direct injection rate KXD is calculated by the following equation.
KXD = KXD0 * H / H0… (a8)
H: Maximum intake valve lift,
H0: Standard maximum lift,
H0 in the equation (a8) is the maximum lift of the intake valve when the variable valve mechanism is not operated. When the variable valve mechanism is operated, the maximum lift H of the intake valve is usually smaller than H0, so the direct injection rate is reduced accordingly. Therefore, the amount of decrease correction is performed by equation (a8).

2) XC: Inhalation system residual spray:
When the spray amount XD directly injected is thus obtained, the spray amount XC remaining in the intake system is given by the following equation.
XC = XB-XD (a9)
<2-3> Model identification of spray branch (floating)
1) X0 ″; Float in the intake system:
It is assumed that the spray is distributed throughout the intake port 4 and that each spray falls against the air by gravity acceleration as shown in FIGS. 6-9. In such a natural fall model, the spray that falls and does not reach the port wall 4a floats, and the spray that reaches the port wall 4a is considered to adhere to the port wall 4a.

  However, in the case of natural fall, the drop speed of the spray is not related to the particle diameter D (∝ mass), including the case where there is an air resistance proportional to the speed or the square of the speed. V is a function of the particle size D, and is assumed to increase as the particle size D increases as shown in FIG. 6-10.

The spray drop distance L is a value obtained by multiplying the drop velocity V by the spray floating time (or arrival limit time) t. When the maximum distance #L (port height #LP) to the wall surface is taken in FIG. 6-10, all the sprays whose spray drop distance L is greater than this maximum distance #L adhere to the port wall 4a. As shown in Fig. 6-10, the characteristics of each floating part are downward-sloping, and the sum of the particle diameters of the area where the floating part for each particle size is 0 or more is the floating part X0 "in the intake system. Become. This can be obtained by the following equation.
X0 '' = Σ (1-Lk / # LP)… (a10)
Here, Lk is the spray reach distance in the particle size category k. This Lk is
Lk = Vk * tp… (a11)
(Vk is the spray drop velocity in the particle size category k, tp is the time from the injection timing I / T as the floating time (or the arrival limit time) to the start of the compression stroke), and this is expressed as (a10 Substituting into the formula gives the following formula:
X0 '' = Σ (1-Vk * tp / # LP)… (a12)
As a result, a table (see FIG. 6-10) of the spray drop speed V for each of the small sections using the particle diameter D as a parameter is prepared, and the expression (a12) is applied until the particle diameter section k becomes 1 from D0. By summing up, the floating part X0 '' in the intake system can be obtained. D0 is the particle size when the floating portion for each particle size becomes 0 in FIG. The tp may be measured by a timer built in the engine controller 31. #LP is a constant value and is obtained from the drawing.

2) X0 ''': Floating matter in the combustion chamber:
The idea is the same as the floating part X0 '' in the intake system. That is, it is assumed that sprays are distributed throughout the combustion chamber 5 and that each spray falls against the air by gravity acceleration as shown in FIGS. 6-9. In such a natural fall model, the spray that falls and does not reach the piston crown surface 6a floats, and the spray that reaches the piston crown surface 6a is considered to adhere to the combustion chamber (fuel chamber wall or cylinder surface wall 52).

  The spray drop speed V is a function of the particle size D, and is assumed to increase as the particle size D increases as shown in FIG. 6-10.

The spray drop distance L is a value obtained by multiplying the drop velocity V by the spray floating time (or arrival limit time) t. If the combustion chamber height #LC (for example, represented by the piston center point), which is the maximum distance #L to the wall surface, is taken in FIG. 6-10, the fuel whose spray fall distance L is equal to or greater than the combustion chamber height #LC Since all of the spray adheres to the combustion chamber, the characteristics of the floating part for each particle size become a downward-sloping characteristic as shown in FIG. 6-10. The value becomes the float X0 '''in the combustion chamber. This can be obtained by the following equation.
X0 '''= Σ (1-Lk / # LC)… (a13)
Here, Lk is the spray reach distance in the particle size category k, and this Lk is
Lk = Vk * tc… (a14)
(Vk is the spray drop velocity in the particle size category k, tc is the end of the compression stroke (or combustion) from the injection timing I / T (or the start of the intake stroke) as the floating time (or arrival limit time) Substituting this into the equation (a13) gives the following equation:
X0 '''= Σ (1-Vk * tc / # LC)… (a15)
As a result, a table (see FIG. 6-10) of the spray drop speed V for each of the small sections with the particle diameter D as a parameter is prepared. Until the particle diameter section becomes D0 from 1, the equation (a15) is used. Summing up, it is possible to obtain the float X0 '''in the combustion chamber. D0 is the particle size when the floating portion for each particle size becomes 0 in FIG. tc may be measured by a timer built in the engine controller 31. #LC is a constant value and is obtained from the drawing.

3) XE, XF: The amount adhering to the intake system and combustion chamber:
When the floating part X0 ″ in the intake system and the floating part X0 ′ ″ in the combustion chamber are obtained in this way, the intake system adhering part XE and the combustion chamber adhering part XF are given by the following equations.
XE = XC-X0 ''… (a16)
XF = XD-X0 '''… (a17)
In an engine equipped with a variable valve mechanism, secondary atomization of the spray directly injected is promoted, so that the spray XD directly injected and the floating X0 ′ ″ in the combustion chamber are corrected. Here, secondary atomization means that when the intake valve operating angle variable mechanism works, the maximum lift of the intake valve becomes small and the airflow flowing through the gap between the intake valve and the valve seat does not work for the intake valve operating angle variable mechanism. It is faster than that, and it means that atomization of the spray that is directly injected is promoted.

As shown in the second characteristic from the right in the lower part of Fig. 6-2, the distribution of the floating part for each particle size and the deposit in the combustion chamber for each particle size by the secondary atomization are the characteristics from the solid line to the broken line. Migrate to In order to obtain each distribution of the broken line characteristic, each distribution of the spray component XD directly injected and the floating component X0 '''in the combustion chamber is corrected by shifting by two grids in the direction of decreasing the particle size. Using these new corrected distributions, the spray fraction XD directly injected as described above and the floating fraction X0 ′ ″ in the combustion chamber are obtained, and the obtained XD and X0 ′ ″ are calculated as (32 ) Used in the formula.
<2-4> Model identification of spray branch (attachment site)
1) X1, X2: Intake valve wall deposit, port wall deposit:
The distribution of the intake system adhering component XE is a lower solid line in FIG. 6-11, and the distribution of the intake valve wall adhering component X1 is as shown by the lower broken line in FIG. This is the distribution of port wall adhesion X2. Therefore, the intake system adhesion part XE is allocated to the intake valve wall adhesion part X1 and the port wall adhesion part X2 according to the intake valve direct hit rate #DVR as in the following equation.
X1 = XE * KX1… (a18)
X2 = XE-X1… (a19)
However, KX1: intake valve direct hit rate coefficient,
Here, the intake valve direct hit rate coefficient KX1 is obtained by searching a map having the contents shown in FIG. 6-12 from the intake valve direct hit rate #DVR and the pressure P. As shown in FIG. 6-12, the intake valve direct hit rate coefficient KX1 increases as the intake valve direct hit rate #DVR increases. Further, even when the intake valve direct hit rate #DVR is the same, the value of the intake valve direct hit rate coefficient KX1 is smaller at the time of low load where the pressure P is small. In FIG. 6-12, “no negative pressure” means a high load when the pressure P approaches atmospheric pressure, and “high negative pressure” means a low load when the pressure P becomes smaller than the atmospheric pressure. The intake valve direct hit rate #DVR is a rate at which the spray from the injection valve 21 collides with the intake valve 15 and can be calculated from the drawings of the intake port 4 and the injection valve spray.

2) X3, X4: Combustion chamber wall deposit, cylinder surface wall deposit:
The distribution of the spray adhering to the combustion chamber wall and the cylinder face wall 52 is shown superimposed on FIG. The combustion chamber deposit XF is allocated to the combustion chamber wall deposit X3 and the cylinder surface wall deposit X4 as shown in the following equation at an allocation rate KX4.
X4 = XF * KX4… (a20)
X3 = XF-X4 (a21)
Here, a cylinder adhesion index is determined based on the spray inflow layout, and an allocation rate KX4 is obtained by searching a table having the contents shown in FIG. Here, the cylinder index represents the ratio of the fuel that sprays from the injection valve 21 passes through the gap between the intake valve 15 and the valve seat and enters the combustion chamber 5 and adheres to each wall toward the cylinder wall. For example, if the spray shape is a cone and the ratio of passing through the gap between the intake valve 15 and the valve seat is B, and the ratio of B toward the cylinder wall is A, A / B may be used as the cylinder index. As shown in FIG. 6-13, the allocation rate KX4 is a value that increases as the cylinder adhesion index increases. The cylinder adhesion index can be set from the result of a flow simulation model, a part-by-part wall flow recovery experiment in a unit test, or the like.

  In this way, the branch model of the injection valve spray shown in FIG. 6A can calculate the branch ratios X0, X1, X2, X3, and X4 of the spray during injection from the injection valve 21, Compared to the conventional method that uses a map or table directly from the operating conditions such as temperature, rotation speed, load signal, etc., the physical model is promoted, so it is possible to eliminate the adaptation by individual engine experiments. It is possible to reduce the adaptation man-hours and shorten the adaptation period.

  Of the spray branching ratio X0, X1, X2, X3, X4 obtained from the injection valve as described above, the spray branching ratio X0 from the injection valve and the fuel injection amount Mfin injected from the injection valve 21 are used. Vaporization amount Mx0 ′ = Mfin * X0 is calculated with respect to the spray particle size distribution (mass ratio).

  Using the vaporization amount Mx0 'for this spray particle size distribution (mass ratio), the mixture temperature decrease allowance ΔTBvap when vaporized as droplets correlates with the vaporization amount Mx0' for the spray particle size distribution (mass ratio) Therefore, it can be expressed as the following equation (9) using the constant = Kbvap #.

  Therefore, the mixture temperature Ta5 when the fuel injected from the injection valve 21 is affected by the latent heat of vaporization of the droplets in the intake port and the combustion chamber is calculated using the following equation (10).

5-2. Estimation of In-Cylinder Gas Temperature when the Intake Valve is Closed Next, an air-fuel mixture temperature Ta6 is calculated by heat transfer from the intake valve, exhaust valve, combustion chamber wall / piston crown, and cylinder wall. First, the wall temperature of each part is estimated by the following method. This is a temperature estimation method proposed by the present applicant in Japanese Patent Application No. 2003-185133. The outline of this method is as follows. First, in this method, as a component constituting the combustion chamber, for example, the cooling water temperature of the engine for each intake valve. Then, the equilibrium temperature is calculated from the intake air amount, the combustion equivalent ratio, and the rotation speed, and then the umbrella direction temperature of the intake valve is calculated using the temperature increase amount that changes with a first-order lag with respect to the coolant temperature. On the other hand, for the wall surface temperature of other cylinder components such as the combustion chamber wall and cylinder wall, the equilibrium temperature in the intake valve umbrella direction and the temperature change rate given as a time constant are multiplied by a predetermined coefficient. The equilibrium temperature and temperature change rate of the cylinder component are obtained, and the temperature of the cylinder component is estimated from this and the previous calculated value. This method can simplify the configuration of the calculation unit by estimating the temperature of other cylinder components from the temperature of a specific part such as an intake valve, and the intake air amount, combustion equivalent ratio, rotation speed can be calculated for the temperature calculation of the specific part. Therefore, there is a feature that a relatively accurate wall surface temperature reflecting the influence of a change in heat generation amount and a change in the flow rate of cooling water according to the engine operation state can be obtained. However, it goes without saying that the method for determining the wall surface temperature of each part in the cylinder is not limited to this, and depending on the target part, the wall surface temperature is represented by the detection value of the water temperature sensor, or it is measured using the temperature sensor. May be.

  Using the intake valve / exhaust valve, combustion chamber wall, cylinder wall, piston crown surface temperature Twalli calculated in this way, and the heat transfer equation of claim 3, the following equation (11) In-cylinder gas temperature Ta6 (= Tivc) under the influence of heat transfer from is calculated.

  Here, Twalli is the wall temperature of the cylinder component, and represents the temperature Twall1, the cylinder wall temperature Twall2, the intake valve umbrella temperature Twall3, and the exhaust valve umbrella temperature Twall4 of the cylinder chamber combustion chamber wall and piston crown. Considering in more detail about the heat receiving / dissipating model between the wall surface of each part of the cylinder and the intake gas, the reception between each part of i = 1 to 4 in the cylinder (the meaning of i is as described above) and the intake gas. The heat quantity Q of heat release can be expressed as the following equation (12) from the law of conservation of heat as a model for heating a gas of mass M and specific heat Cgas by ΔT (K).

hi: Heat transfer coefficient of each heat transfer section
Ai: Heat transfer area
Tgasi: Gas temperature before receiving heat (= intake gas temperature in cylinder when intake valve is open)
Here, the heat transfer coefficient h is calculated using the following equation (13), which is an improved equation of Woschni.

d: Cylinder bore diameter
C1: Constant
Cm: Average piston speed The dimensional values and constants of each known heat transfer target are summarized, and Ki is set to a coefficient that is modified according to the engine speed representing the piston speed (gas flow velocity). After all, the above formula (12) can be expressed as follows.

  When the initial temperature of the intake gas is T0 in the engine intake system, the temperature change due to heat transfer at any n subsequent heat transfer target parts is added to this and the intake valve is closed. This means that the gas temperature in the cylinder can be estimated.

  The order of estimating the temperature change amount is arbitrary when there are a plurality of heat transfer target parts, but the gas temperature is obtained sequentially from upstream to downstream according to the flow of intake air as in the above embodiment. The difference in engine system specifications, for example, when there is an intake air temperature sensor in the vicinity of the air flow meter, when there is an intake air temperature sensor in the intake collector part, the presence or absence of the anti-icing heater in the throttle chamber part, the difference in the layout, etc. On the other hand, the arithmetic expression can be separately corrected according to the flow of the intake air heated or cooled, the estimation accuracy can be improved, and the number of adaptation man-hours can be reduced.

1 is a system diagram of an automobile engine showing an embodiment of the present invention. The conceptual diagram regarding the intake gas temperature behavior which represented the various factors which influence the intake gas temperature change of an engine along the arrangement | positioning of an intake system. Explanatory drawing of the table which gives exhaust gas specific heat ratio from target combustion equivalent ratio and exhaust temperature. The model figure for estimating the desorption amount of the fuel evaporative gas from a canister. Explanatory drawing of the table which gives air-fuel | gaseous mixture specific heat ratio from target combustion equivalent ratio. The data flow figure of the branch model of injection valve spray. The model figure which shows the process of the whole spray branch. The characteristic view of the particle size distribution about the mass ratio of the spray at the time of injection. The model figure for demonstrating the vaporization rate of spraying. The characteristic view of the vaporization characteristic f (V, T, P). The characteristic figure of the exposure time of intake airflow. The model figure for demonstrating direct injection to the combustion chamber of spray. The characteristic figure of the direct injection rate with respect to injection timing and (beta). The model figure for demonstrating the floating in the intake system of spray, and the floating in a combustion chamber. The characteristic figure of the spray fall speed and the floating ratio for every particle size. The characteristic view which shows spraying particle size distribution. The characteristic diagram of the intake valve direct hit rate coefficient with respect to the intake valve direct hit rate and the ratio X1 / X2. The characteristic figure of the allocation rate with respect to ratio X3 / X4. The model figure for the intake gas temperature estimation by this invention.

Explanation of symbols

2 intake collector 4 intake port 5 combustion chamber 6 piston 8 exhaust passage 15 intake valve 16 exhaust valve 21 fuel injection valve 23 intake throttle valve 25 EGR passage 26 EGR valve 29 variable valve mechanism 31 engine controller 32 air flow meter 42 accelerator sensor 43, 46, 48 Temperature sensor 44, 45, 47 Pressure sensor 60 Throttle chamber 61 Hot water heater 63 Blow-by gas passage 64 Purge gas passage 65 Fuel tank 66 Canister 67 Purge valve

Claims (4)

In an engine equipped with a fuel supply device that injects and supplies fuel to the engine intake system,
Means for determining the intake air temperature;
Means for determining the vaporization characteristics of the fuel spray;
Means for determining the gas flow velocity and pressure of the fuel vaporization target portion in the intake system or cylinder;
Gas temperature calculation means for estimating the mixture temperature in the cylinder using the result of each means;
An intake gas temperature estimation device for an engine comprising: The fuel spray vaporization characteristics are:
The engine intake gas temperature estimation device according to claim 1, which is a change characteristic of a fuel evaporation amount with respect to at least one of a gas flow rate, a temperature, and a pressure. The gas temperature calculation means includes
Means for determining the amount of floating fuel in the intake system using the intake air temperature, the vaporization characteristics of the fuel spray, the gas flow velocity and pressure of the fuel vaporization target part in the intake system;
Temperature estimation means for estimating the temperature of the mixture in the cylinder using the floating fuel amount and the latent heat of vaporization of the fuel;
The intake gas temperature estimation device for an engine according to claim 1, comprising: The means for obtaining the amount of floating fuel is:
Means for determining a particle size distribution for the mass of fuel spray injected from the fuel injector;
Means for calculating the floating fuel amount from a model for calculating the floating fuel amount based on the particle size distribution;
An engine intake gas temperature estimation device according to claim 3.

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