JP2008002328A - Multiple-fuel internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To execute good compression self ignition operation in accordance with an ignitibility of fuel. <P>SOLUTION: A multiple-fuel internal combustion engine is provided with a compression ratio variable means 81 for varying an actual compression ratio to an arbitrary compression ratio, a fuel characteristic detection means (an electronic control unit 1) for detecting the quality of the ignitibility of fuel, in a combustion chamber CC, which is composed of at least two types of fuels F1, F2 with different properties, and a compression ratio control means (an electronic control unit 1) for controlling the compression ratio variable means 81 so as to increase the actual compression ratio when the ignitibility of fuel in the combustion chamber CC is worse than a prescribed ignitibility and to decrease the actual compression ratio when the ignitibility of fuel in the combustion chamber CC is better than a prescribed ignitibility in the compression self ignition operation. <P>COPYRIGHT: (C)2008,JPO&INPIT

Description

  The present invention relates to a multi-fuel internal combustion engine that is operated by introducing at least one of at least two types of fuels having different properties into a combustion chamber or guiding a mixed fuel composed of the at least two types of fuel into the combustion chamber.

  Conventionally, so-called multi-fuel internal combustion engines that are operated using a plurality of types of fuels having different properties are known. For example, Patent Document 1 below describes a multi-fuel internal combustion engine that is operated by spark ignition using a mixed fuel composed of a plurality of types of fuel. This Patent Document 1 discloses a technique for changing the actual compression ratio so as to obtain an optimal compression ratio that does not cause knocking according to the ratio of the plurality of types of fuels.

  The following Patent Document 2 describes a multi-fuel internal combustion engine that is operated by compression autoignition using a mixed fuel of gasoline and dimethyl ether. In the multi-fuel internal combustion engine disclosed in Patent Document 2, the compression ignition region is expanded without increasing the compression ratio by changing the mixing ratio of the respective fuels. Patent Document 3 below describes an internal combustion engine that uses a single fuel and is operated in a spark ignition region or a compression auto-ignition region according to operating conditions (engine speed, engine load).

JP-A-6-248988 JP 2002-357139 A JP 2000-320333 A

  As described above, the technique disclosed in Patent Document 1 eliminates inconvenience at the time of spark ignition operation (here, knocking due to abnormal combustion) by varying the actual compression ratio. On the other hand, there are many inconveniences caused by the ignitability of the fuel introduced into the combustion chamber as a disadvantage during the compression self-ignition operation. For example, as a typical example, there are situations where the fuel has poor ignitability and cannot be operated without being able to self-ignite in the compressed air in the combustion chamber, and even when self-ignition is possible, misfiring can occur. In addition, the use of fuel with poor ignitability prolongs the ignition delay period, and the fuel and compressed air are excessively mixed. In other words, there is a situation in which the in-cylinder pressure is increased and vibration and noise (so-called diesel knock) are generated.

  Therefore, the present invention has an object to provide a multi-fuel internal combustion engine that can improve the disadvantages of the conventional example and perform a good compression self-ignition operation according to the ignitability of the fuel guided into the combustion chamber. To do.

  In order to achieve the above object, according to the first aspect of the present invention, at least one of at least two types of fuels having different properties is led to the combustion chamber or a mixed fuel composed of the at least two types of fuel is supplied to the combustion chamber. In a multi-fuel internal combustion engine that is guided and operated, a compression ratio variable means for changing the actual compression ratio to an arbitrary compression ratio, a fuel characteristic detection means for detecting the quality of the ignitability of the fuel in the combustion chamber, and compression auto-ignition When operating, if the ignitability of the fuel in the combustion chamber is worse than the predetermined ignitability, the actual compression ratio is increased, and if the ignitability of the fuel in the combustion chamber is better than the predetermined ignitability, the actual compression ratio is decreased. And a compression ratio control means for controlling the compression ratio variable means.

  In the multi-fuel internal combustion engine according to the first aspect, since the temperature of the compressed air in the combustion chamber rises by increasing the actual compression ratio, even a fuel with poor ignitability is likely to self-ignite in the compressed air. On the other hand, in this multi-fuel internal combustion engine, the mechanical loss is reduced and the maximum in-cylinder pressure is reduced by lowering the actual compression ratio during the operation of the compression auto-ignition diffusion combustion mode when the ignitability of the fuel led into the combustion chamber is good. Is planned.

  In order to achieve the above object, according to a second aspect of the present invention, in the multifuel internal combustion engine according to the first aspect, an exhaust turbocharger or / and a supercharger with an electric motor and a compression ratio control means are implemented. There is provided a supercharging pressure control means for setting a high upper limit of the supercharging pressure of the exhaust turbocharger and / or the supercharger with electric motor when the compression ratio is lowered.

  In the multi-fuel internal combustion engine according to claim 2, in addition to the effect of the invention according to claim 1, it can be operated at a high boost pressure by increasing the upper limit of the boost pressure. The supercharging effect can be obtained and higher torque and higher output can be achieved. For this reason, in this multi-fuel internal combustion engine, it is possible to improve the engine performance that decreases as the actual compression ratio decreases.

  The multi-fuel internal combustion engine according to the present invention is easily self-ignited in compressed air even by fuel with poor ignitability by increasing the actual compression ratio, so that compression self-ignition diffusion combustion with fuel with poor ignitability is realized, Moreover, misfire and diesel knock during compression self-ignition diffusion combustion are suppressed. On the other hand, this multi-fuel internal combustion engine reduces mechanical loss and lowers the maximum in-cylinder pressure by lowering the actual compression ratio, so the engine performance and fuel efficiency performance during compression auto-ignition diffusion combustion with fuel with good ignitability Can be improved. Therefore, according to this multi-fuel internal combustion engine, it is possible to perform a good compression self-ignition operation by varying the actual compression ratio according to the ignitability of the fuel guided into the combustion chamber.

  Embodiments of a multi-fuel internal combustion engine according to the present invention will be described below in detail with reference to the drawings. The present invention is not limited to the embodiments.

  A multifuel internal combustion engine according to a first embodiment of the present invention will be described with reference to FIGS. This multi-fuel internal combustion engine is an internal combustion engine that is operated by introducing at least one of at least two types of fuels having different properties into the combustion chamber or by introducing a mixed fuel composed of the at least two types of fuel into the combustion chamber. It is. In the first embodiment, the latter multi-fuel internal combustion engine will be described as an example.

  In this multifuel internal combustion engine, various control operations such as combustion control are executed by an electronic control unit (ECU) 1 shown in FIG. The electronic control unit 1 includes a CPU (Central Processing Unit) (not shown), a ROM (Read Only Memory) that stores a predetermined control program and the like, and a RAM (Random Access Memory) that temporarily stores the calculation result of the CPU. , And a backup RAM for storing information prepared in advance.

  First, the configuration of the multi-fuel internal combustion engine exemplified here will be described with reference to FIG. Although only one cylinder is shown in FIG. 1, the present invention is not limited to this, and can be applied to a multi-cylinder multifuel internal combustion engine. In the first embodiment, description will be made assuming that a plurality of cylinders are provided.

  The multifuel internal combustion engine is provided with a cylinder head 11, a cylinder block 12, and a piston 13 that form a combustion chamber CC. Here, the cylinder head 11 and the cylinder block 12 are fastened with bolts or the like via the head gasket 14 shown in FIG. 1, and the recess 11a on the lower surface of the cylinder head 11 and the cylinder bore 12a of the cylinder block 12 formed thereby. In the space, the piston 13 is disposed so as to be able to reciprocate via the connecting rod 15. And the combustion chamber CC mentioned above is comprised by the space enclosed by the wall surface of the recessed part 11a of the cylinder head 11, the wall surface of the cylinder bore 12a, and the top surface 13a of the piston 13. FIG.

  The multifuel internal combustion engine of the first embodiment sends air and fuel into the combustion chamber CC according to the operating conditions such as the engine speed and engine load and the combustion mode, and executes combustion control according to the operating conditions. The air is sucked from the outside through the intake passage 21 and the intake port 11b of the cylinder head 11 shown in FIG. On the other hand, the fuel is supplied using the fuel supply device 50 shown in FIG.

  First, the air supply path will be described. On the intake passage 21 of the first embodiment, an air cleaner 22 that removes foreign matters such as dust contained in air introduced from the outside, and an air flow meter 23 that detects the amount of intake air from the outside are provided. . In this multi-fuel internal combustion engine, the detection signal of the air flow meter 23 is sent to the electronic control unit 1, and the electronic control unit 1 calculates the intake air amount, the engine load and the like based on the detection signal.

  A throttle valve 24 that adjusts the amount of intake air into the combustion chamber CC and a throttle valve actuator 25 that opens and closes the throttle valve 24 are disposed downstream of the air flow meter 23 on the intake passage 21. Is provided. The electronic control device 1 according to the first embodiment controls the throttle valve actuator 25 according to the operating conditions and the combustion mode so that the valve opening degree (in other words, the intake air amount) according to the operating conditions is obtained. The valve opening angle of the throttle valve 24 is adjusted. For example, the throttle valve 24 is adjusted so that the intake air amount necessary for achieving an air-fuel ratio corresponding to the operating conditions and the combustion mode is sucked into the combustion chamber CC. The multifuel internal combustion engine is provided with a throttle opening sensor 26 that detects the valve opening of the throttle valve 24 and transmits the detection signal to the electronic control unit 1.

  Further, one end of the intake port 11b opens to the combustion chamber CC, and an intake valve 31 for opening and closing the opening is disposed at the opening portion. The number of openings may be one or more, and an intake valve 31 is provided for each opening. Therefore, in this multi-fuel internal combustion engine, air is sucked into the combustion chamber CC from the intake port 11b by opening the intake valve 31 and closed in the combustion chamber CC by closing the intake valve 31. Inflow of air to is blocked.

  Here, as the intake valve 31, for example, there is a valve that is driven to open and close in accordance with the rotation of an intake camshaft (not shown) and the elastic force of an elastic member (string spring). In this type of intake valve 31, by interposing a power transmission mechanism such as a chain or a sprocket between the intake side camshaft and the crankshaft 16, the intake side camshaft is interlocked with the rotation of the crankshaft 16, Open / close drive is performed at a preset opening / closing timing. In the multifuel internal combustion engine of the first embodiment, the intake valve 31 that is opened and closed in synchronization with the rotation of the crankshaft 16 is applied.

  However, this multi-fuel internal combustion engine may be provided with a variable valve mechanism such as a so-called variable valve timing & lift mechanism that can change the opening / closing timing and lift amount of the intake valve 31. The opening / closing timing and the lift amount can be changed to suitable ones according to the operating conditions and the combustion mode. Furthermore, in this multi-fuel internal combustion engine, a so-called electromagnetically driven valve that opens and closes the intake valve 31 using electromagnetic force may be used in order to obtain the same effect as the variable valve mechanism.

  Next, the fuel supply device 50 will be described. The fuel supply device 50 guides a plurality of types of fuel having different properties to the combustion chamber CC. In the first embodiment, two types of fuels having different properties (a first fuel F1 stored in the first fuel tank 41A and a second fuel F2 stored in the second fuel tank 41B) are preliminarily stored in a predetermined fuel. An example is shown in which the mixture is mixed at a mixing ratio and the mixed fuel is directly injected into the combustion chamber CC.

  Specifically, the fuel supply device 50 includes a first feed pump 52A that sucks up the first fuel F1 from the first fuel tank 41A and sends it to the first fuel passage 51A, and a second fuel F2 from the second fuel tank 41B. Fuel mixing for mixing the second feed pump 52B sucked up and sent to the second fuel passage 51B, and the first and second fuels F1 and F2 sent from the first and second fuel passages 51A and 51B, respectively. Means 53, a high-pressure fuel pump 55 for pressurizing and feeding the mixed fuel produced by the fuel mixing means 53 to the high-pressure fuel passage 54, and a delivery passage for distributing the mixed fuel in the high-pressure fuel passage 54 to the respective cylinders 56 and a fuel injection valve 57 for each cylinder for injecting the mixed fuel supplied from the delivery passage 56 into the combustion chamber CC.

  In this fuel supply device 50, the first feed pump 52A, the second feed pump 52B and the fuel mixing means 53 are driven and controlled by the electronic control device 1, so that the mixed fuel of a predetermined fuel mixing ratio is supplied to the fuel mixing means. 53 to be generated. For example, the fuel supply device 50 may adjust the fuel mixing ratio of the mixed fuel by causing the electronic control device 1 to adjust the discharge amounts of the first feed pump 52A and the second feed pump 52B. The fuel mixing ratio of the mixed fuel may be adjusted by adjusting the mixing ratio of the first and second fuels F1 and F2 in the fuel mixing means 53 in accordance with an instruction from the control device 1. Here, the fuel mixture ratio may be a constant value set in advance, or may be a variable value that varies depending on the operating conditions and the combustion mode.

  In addition, the fuel supply device 50 causes the electronic control device 1 to drive and control the high-pressure fuel pump 55 and the fuel injection valve 57 in accordance with the operating conditions and the combustion mode, whereby the fuel injection amount corresponding to the operating conditions, The generated mixed fuel is injected under fuel injection conditions such as fuel injection timing and fuel injection period. For example, the electronic control unit 1 causes the mixed fuel to be pumped from the high-pressure fuel pump 55 and causes the fuel injection valve 57 to perform injection under fuel injection conditions corresponding to operating conditions and the like.

  The mixed fuel thus supplied to the combustion chamber CC is burned by the ignition operation in the ignition mode corresponding to the combustion mode in combination with the air described above. The in-cylinder gas after the combustion is discharged from the combustion chamber CC to the exhaust port 11c shown in FIG. Here, an exhaust valve 61 that opens and closes an opening between the exhaust port 11c and the combustion chamber CC is disposed. The number of openings may be one or more, and the exhaust valve 61 described above is provided for each opening. Therefore, in this multi-fuel internal combustion engine, the in-cylinder gas after combustion is discharged from the combustion chamber CC to the exhaust port 11c by opening the exhaust valve 61, and the exhaust valve 61 is closed to close the cylinder. The discharge of the internal gas to the exhaust port 11c is blocked.

  Here, as the exhaust valve 61, as in the intake valve 31 described above, a valve with a power transmission mechanism, a valve with a variable valve mechanism such as a so-called variable valve timing & lift mechanism, or a so-called electromagnetically driven valve can be used. Can be applied.

  By the way, in an internal combustion engine, combustion modes are generally divided into a diffusion combustion mode and a flame propagation combustion mode, and a compression auto-ignition mode and a premixed spark ignition mode are prepared as ignition modes corresponding to the combustion modes. Hereinafter, they are collectively referred to as a combustion mode, and are respectively referred to as a compression autoignition diffusion combustion mode and a premixed spark ignition flame propagation combustion mode.

  First, in the compression auto-ignition diffusion combustion mode, a part of the fuel is self-ignited by injecting high-pressure fuel into the high-temperature compressed air formed in the combustion chamber CC, and the fuel and air are diffusely mixed. This is a combustion mode in which combustion is advanced while being performed. Here, since the compressed air in the combustion chamber CC and the fuel are difficult to be mixed instantaneously, immediately after the start of fuel injection, the air-fuel ratio varies in some places. On the other hand, it is generally preferable to use a fuel having excellent ignitability as described below when performing diffusion combustion, and such a fuel with good ignitability does not have to wait for the entire injection amount to be injected. It will ignite by itself at the air fuel ratio suitable for combustion. For this reason, in this compression self-ignition diffusion combustion mode, the fuel of the air-fuel ratio part suitable for combustion self-ignites first, and the flame formed thereby gradually advances the combustion while entraining the remaining fuel and air. Let

  In order to operate in this compressed self-ignition diffusion combustion mode, a fuel with good ignitability whose ignition point is lower than the compression heat of compressed air is usually required. For example, light oil or dimethyl ether can be considered as the fuel with good ignitability. Further, in recent years, GTL (Gas To Liquids) fuel has attracted attention as an alternative fuel for light oil, and this GTL fuel is easily produced in a desired property. For this reason, the GTL fuel produced | generated in order to improve ignitability can also be used for fuel with good ignitability. Such a fuel with good ignitability not only enables compression auto-ignition diffusion combustion, but also reduces the amount of NOx generated when operating in the compression auto-ignition diffusion combustion mode. Vibration can be suppressed.

  On the other hand, in the premixed spark ignition flame propagation combustion mode, the premixed gas in the combustion chamber CC in which fuel and air are mixed in advance is given a spark by spark ignition, and combustion is performed while propagating the flame around the fire type. It is a combustion form that advances the. In this premixed spark ignition flame propagation combustion mode, homogeneous combustion for igniting a homogeneously mixed premixed gas or a highly concentrated premixed gas is formed around the ignition means, and further, a lean mixture is formed around the premixed spark ignition flame propagation combustion mode. It includes a combustion mode such as stratified combustion in which a premixed gas is formed and ignition is performed on the rich premixed gas.

  As a fuel suitable for the premixed spark ignition flame propagation combustion mode, a highly evaporative fuel represented by gasoline is generally considered. Here, since highly evaporable fuel is easily mixed with air, it reduces the fuel rich region and contributes to suppression of PM, smoke, NOx and unburned hydrocarbons (unburned HC). As such highly evaporable fuel, in addition to gasoline, GTL fuel produced as a highly evaporable property, dimethyl ether, and the like are known.

  The multifuel internal combustion engine of the first embodiment is configured to enable operation in both combustion modes. Therefore, the multifuel internal combustion engine of the first embodiment is provided with the spark plug 71 shown in FIG. 1 for spark ignition of the premixed gas in order to enable operation in the premixed spark ignition flame propagation combustion mode. Set up. The spark plug 71 executes spark ignition when the ignition timing according to the operating condition in the premixed spark ignition flame propagation combustion mode is reached in accordance with the instruction of the electronic control unit 1.

  In addition, the electronic control device 1 of the first embodiment is provided with combustion mode setting means for setting the combustion mode. The combustion mode setting means exemplified here uses the combustion mode map data as shown in FIG. 2 with the operating conditions (engine speed and engine load) as parameters, and selects the optimal combustion mode according to the operating conditions. Let For example, this combustion mode map data can be operated in the compression auto-ignition diffusion combustion mode when operating conditions such as medium / high load / low rotation, high load / high rotation, etc., and low load / low rotation, low medium load / high rotation, etc. This is set in advance based on experiments and simulations so as to operate in the premixed spark ignition flame propagation combustion mode under the above operating conditions. The engine speed can be grasped from the detection signal of the crank angle sensor 17 shown in FIG. The crank angle sensor 17 is a sensor that detects the rotation angle of the crankshaft 16. On the other hand, the engine load can be grasped from the detection signal of the air flow meter 23 described above.

  Furthermore, the electronic control unit 1 is configured to set the intake air amount, the fuel mixture ratio, and the fuel injection condition based on the operation condition and the combustion mode. These are set, for example, using map data for each combustion mode using operating conditions (engine speed and engine load) as parameters.

  Here, whether or not the operation is possible in the compression auto-ignition diffusion combustion mode is affected by the fuel characteristics (ignitability) of the fuel introduced into the combustion chamber CC as described above. That is, when the ignitability of the fuel introduced into the combustion chamber CC is good, self-ignition is performed in compressed air, so that diffusion combustion can be performed. However, when the ignitability of the fuel is poor, self-ignition is not possible, so combustion itself is performed. There is a possibility of disappearing.

  Depending on the degree of poor ignitability of the fuel, self-ignition may occur. However, even in such a case, a misfire may occur after self-ignition, which increases vibration and noise. In addition, fuel with poor ignitability causes so-called diesel knock, which increases vibration and noise during combustion.

  By the way, in order to self-ignite the fuel in the compressed air, not only the ignitability of the fuel but also the temperature of the compressed air becomes an important factor. That is, self-ignition of fuel can be performed by high-temperature compressed air and fuel whose ignition point is lower than the compression heat of the compressed air. In compressed air whose temperature is lower than the ignition point, fuel self-ignition. Not.

  Thus, the self-ignition of the fuel is established by the correlation between the ignitability of the fuel and the temperature of the compressed air. Therefore, in order to realize the self-ignition of the fuel in a situation where self-ignition is impossible, at least the improvement of the ignitability of the fuel guided into the combustion chamber CC (that is, the reduction of the ignition point) or the increase of the temperature of the compressed air Is required.

  The “fuel introduced into the combustion chamber CC” shown here means that the mixed fuel of the fuels F1 and F2 mixed by the fuel mixing means 53 is sent to the combustion chamber CC as in the multi-fuel internal combustion engine shown in FIG. The fuel mixture is referred to as the mixed fuel, and the fuels F1 and F2 are individually supplied to the combustion chamber CC as in the multi-fuel internal combustion engine shown in FIG. When it is, it means the whole of the supplied fuels F1 and F2. For example, here, the first fuel tank 41A stores fuel with good ignitability and poor evaporation (first fuel F1), and the second fuel tank 41B has fuel with low ignitability and good evaporation (second fuel). The case where the fuel F2) is stored will be exemplified. In such a case, if only the first fuel F1 is supplied into the combustion chamber CC, the fuel introduced into the combustion chamber CC has good ignitability and poor evaporation characteristics, and only the second fuel F2 is the combustion chamber CC. If supplied inside, the fuel characteristics are poor in ignitability and good in evaporability. On the other hand, the fuel characteristics of the fuel guided into the combustion chamber CC vary depending on the fuel mixing ratio of the fuels F1 and F2. For this reason, the fuel introduced into the combustion chamber CC in this case has, for example, a fuel characteristic with good ignitability and poor evaporation when the fuel mixing ratio of the first fuel F1 is large, and the fuel mixing ratio of the second fuel F2. If the amount is too large, the flammability is poor and the fuel characteristics are good for evaporation.

  In the first embodiment, the fuel is poorly ignitable by raising the temperature of the compressed air so that the fuel is self-ignited.

  Here, the temperature of the compressed air usually increases as the piston 13 moves toward the top dead center, and reaches the maximum temperature at the top dead center. Further, although the temperature of the compressed air moves up and down depending on the temperature of the intake air and the wall surface temperature of the combustion chamber CC, this does not cause a large temperature difference and cannot be controlled to an arbitrary temperature. Therefore, the compressed air has a substantially constant value for each crank angle, and the maximum temperature is determined to be a substantially constant temperature by the actual compression ratio.

  On the other hand, the temperature of the compressed air has a correlation with the actual compression ratio, and the maximum temperature increases as the actual compression ratio increases, and the maximum temperature decreases as the actual compression ratio decreases. In recent years, the existence of a variable compression ratio engine that changes the actual compression ratio is known, and this variable compression ratio engine is provided with a compression ratio variable means that changes the actual compression ratio to an arbitrary compression ratio. ing. The compression ratio variable means may have any structure or form. For example, the top dead center position of the piston 13 is arbitrarily changed in accordance with the drive control command of the electronic control unit 1, and the actual compression is thereby performed. A structure that changes the ratio is considered as an example. For example, as this type of compression ratio variable means, a variable compression provided with a connecting rod having a structure capable of expanding and contracting the length, a link mechanism for expanding and contracting the connecting rod, and an electric motor for driving and controlling the link mechanism. A ratio mechanism (Patent Document 2 described above) is known.

  Accordingly, the actual compression ratio can be increased by providing such a compression ratio variable means, and thereby, the maximum temperature of the compressed air and the temperature for each crank angle are increased more than before the change of the actual compression ratio. be able to. For fuel that has a high ignition point and was not capable of self-ignition before the change in the actual compression ratio, if the temperature of the compressed air that has risen after the change is higher than the ignition point of the fuel, To be able to self-ignite. In addition, fuel that was originally self-ignitable but inferior in ignitability can be subjected to normal diffusion combustion without causing misfire or diesel knock.

  Therefore, in the multi-fuel internal combustion engine of the first embodiment, the above-described compression ratio variable means 81 is provided, the fuel characteristic detecting means for detecting the quality of the ignitability of the fuel guided into the combustion chamber CC, and its ignition. The electronic control device 1 is provided with a compression ratio control means for driving and controlling the compression ratio variable means 81 in accordance with the quality.

  First, the fuel characteristic detecting means will be described.

  The ignitability of the fuel can be expressed using an index value obtained by indexing the quality (hereinafter referred to as “ignitability index value”) Pc. For this reason, the fuel characteristic detecting means of the first embodiment detects the ignitability index value Pc of the fuel introduced into the combustion chamber CC, and a threshold value (hereinafter referred to as “ignitability determination criterion”) as a reference ignitability determination. The value is referred to as “value”.) The ignition quality is judged by comparison with Pc0. Here, as the difference between the ignitability index value Pc and the ignitability determination reference value Pc0 increases, the ignitability of the fuel becomes better or worse.

  Here, as the ignitability determination reference value Pc0, for example, an ignitability index value of fuel that can be self-ignited in compressed air at the reference compression ratio is set. The reference compression ratio refers to a compression ratio that can be determined as a reference, such as a compression ratio with a high setting frequency or an intermediate compression ratio within a variable range. In addition, when there are multiple types of such self-ignitable fuels, for example, considering the frequency of use of the fuel in the engine, availability of the fuel, stability of properties, etc. The ignitability index value may be set to the ignitability determination reference value Pc0.

  Specifically, as the ignitability index value Pc, a cetane number of fuel or an ignition delay period during diffusion combustion can be used.

  For example, the cetane number of the fuel can be grasped from the properties of the respective fuels F1 and F2 recognized by the fuel characteristic detecting means at the time of refueling. However, in the first embodiment, the respective fuels F1 and F2 are mixed in the fuel mixing means 53 at a predetermined fuel mixing ratio and then sent to the combustion chamber CC. Therefore, the fuel mixing ratio is also taken into consideration. Without it, it is impossible to grasp the exact cetane number of the fuel (mixed fuel) introduced into the combustion chamber CC. For this reason, the fuel characteristic detecting means of the first embodiment calculates the cetane number of the fuel (mixed fuel) introduced into the combustion chamber CC based on the cetane numbers of the respective fuels F1 and F2 and the fuel mixture ratio thereof. Let The cetane number of each of the fuels F1 and F2 acquired by the fuel characteristic detecting means at the time of refueling may be recognized by, for example, providing a vehicle with an input device for inputting the properties of the fuels F1 and F2. It is also possible to recognize fuel supply information such as the type and properties of the fuel supply fuel, the amount of fuel supply, and the like by transmitting and receiving the information from the fuel supply facility to the vehicle via each communication device. In this case, if the ignitability index value Pc is larger than the ignitability determination reference value Pc0, the cetane number of the fuel is larger than the reference cetane number. It shows that it is better than the ignitability. On the other hand, when the ignitability index value Pc is smaller than the ignitability determination reference value Pc0, the cetane number of the fuel is smaller than the reference cetane number, so that the ignitability of the fuel is higher than the reference ignitability. It also shows that it is getting worse.

  Further, the ignition delay period during diffusion combustion can be detected from the combustion pressure sensor 18 shown in FIG. 1 during diffusion combustion. In this case, the meaning of the magnitude relationship between the ignitability index value Pc and the ignitability judgment reference value Pc0 (when Pc> Pc0, the ignitability is good, and when Pc <Pc0, the ignitability is deteriorated) is the above cetane number for convenience. Therefore, the reciprocal of the detected ignition delay period is set as an ignitability index value Pc. In such a case, if the ignitability index value Pc is larger than the ignitability determination reference value Pc0, the ignition delay period is shorter than the reference ignition delay period, so that the ignitability of the fuel is the reference. It shows that it is better than the ignitability. On the other hand, when the ignitability index value Pc is smaller than the ignitability determination reference value Pc0, since the ignition delay period is longer than the reference ignition delay period, the fuel ignitability is higher than the reference ignitability. Indicates that it is getting worse. The ignition delay period may be detected from a detection signal of an ignition timing sensor (not shown) that optically detects the ignition timing, or a detection signal of the crank angle sensor 17 when a change in crank angular velocity is detected. .

  Subsequently, the compression ratio control means will be described.

  The compression ratio control means of the first embodiment is a control function of the electronic control unit 1 that drives and controls the compression ratio variable means 81 so that the actual compression ratio becomes an optimum compression ratio according to the ignitability of the fuel.

  The compression ratio control means is configured to increase the actual compression ratio if the ignitability of the fuel introduced into the combustion chamber CC is deteriorated based on the relationship between the actual compression ratio and the temperature of the compressed air. Then, the compression ratio variable means 81 is driven and controlled to increase the temperature of the compressed air. Here, as the ignitability index value Pc is smaller than the ignitability determination reference value Pc0 and the difference between them increases, the ignitability deteriorates. For example, the ignitability of the fuel (Pc−Pc0) is increased. Using the compression ratio setting map data in which the corresponding target compression ratio is determined in advance, drive control is performed so that the actual compression ratio becomes higher than the reference compression ratio as the ignitability deteriorates.

  On the other hand, in a multi-fuel internal combustion engine, the mechanical loss and the maximum in-cylinder pressure Pmax increase with the increase in the compression ratio, which may deteriorate engine performance such as engine output and fuel efficiency. Therefore, the compression ratio control means of the first embodiment drives the compression ratio variable means 81 so that the actual compression ratio is low if the ignitability of the fuel guided into the combustion chamber CC is good, and mechanical loss is caused. And increase in the maximum in-cylinder pressure Pmax can be prevented. Here, as the ignitability index value Pc is larger than the ignitability determination reference value Pc0 and the difference between them increases, the ignitability becomes better. For example, the above compression ratio setting map data is used. Thus, the drive control is performed so that the actual compression ratio becomes lower than the reference compression ratio as the ignitability improves.

  For example, when the compression ratio variable means 81 is the variable compression ratio mechanism described above, this compression ratio control means drives the link mechanism until the top dead center position of the piston 13 that realizes a desired compression ratio is reached. The connecting rod 15 is expanded and contracted by force. In this case, the compression ratio control means uses, for example, map data with parameters such as the compression ratio and the operation amount of the link mechanism, the driving direction and the driving time of the electric motor, and the like so as to obtain a desired compression ratio. 13 Move the top dead center position.

  Hereinafter, an example of the compression ratio change control operation of the electronic control unit 1 according to the first embodiment will be described with reference to the flowchart of FIG.

  First, the electronic control unit 1 according to the first embodiment detects the ignitability index value Pc of the fuel introduced into the combustion chamber CC as described above by the fuel characteristic detecting means (step ST5), and further in the RAM or the like. The stored ignitability index value Pc1 is called (step ST10).

  The ignitability index value Pc1 called in step ST10 means the ignitability index value Pc detected in step ST5 when the compression ratio change control was performed last time. Therefore, after detecting the ignitability index value Pc in step ST5, the electronic control unit 1 according to the first embodiment stores the ignitability index value Pc in the RAM or the like as the ignitability index value Pc1. The ignitability index value Pc1 is rewritten to the newly detected ignitability index value Pc every time the calculation process of step ST5 is performed.

  The electronic control unit 1 according to the first embodiment compares the current ignitability index value Pc with the previous ignitability index value Pc1, and determines whether or not they are the same value (step ST15).

  Here, the fact that they indicate the same value indicates that the ignitability of the fuel guided into the combustion chamber CC has not changed since the previous compression ratio change control. Therefore, when the current ignitability index value Pc and the previous ignitability index value Pc1 are the same value, there is no need to change from the current compression ratio. The compression ratio change control is once terminated.

  On the other hand, when the current ignitability index value Pc and the previous ignitability index value Pc1 are different from each other, the ignitability of the fuel guided into the combustion chamber CC has changed since the previous compression ratio change control. It represents that. For this reason, the electronic control unit 1 according to the first embodiment first determines the engine speed Ne and the engine load Kl of this multi-fuel internal combustion engine in order to change the compression ratio to the optimum according to the ignitability of the fuel. It detects (step ST20). Then, the combustion mode setting means of the electronic control unit 1 compares the engine rotational speed Ne and the engine load Kl (current operating conditions) with, for example, the combustion mode map data shown in FIG. A combustion mode is selected (step ST25).

  Here, when the compression auto-ignition diffusion combustion mode is selected in step ST25, the electronic control unit 1 detects the ignitability index value Pc and the ignitability determination reference value Pc0 detected in step ST5 as fuel characteristic detection. The means is compared (step ST30).

  As a result, if the ignitability index value Pc is the same value as the ignitability determination reference value Pc0, the ignitability of the fuel led into the combustion chamber CC is the same as the ignitability determined as a reference. Since it is determined by the characteristic detection means, the compression ratio control means of the electronic control unit 1 sends a drive control command to the compression ratio variable means 81 to change the actual compression ratio to the reference compression ratio (step ST35).

  If the ignitability index value Pc is smaller than the ignitability determination reference value Pc0 in step ST30, the fuel characteristic detection means determines that the ignitability of the fuel guided into the combustion chamber CC is worse than the reference ignitability. Therefore, the compression ratio control means sends a drive control command to the compression ratio variable means 81 to change the actual compression ratio to a compression ratio higher than the reference compression ratio (step ST40). Here, as the difference between the ignitability index value Pc and the ignitability determination reference value Pc0 at this time increases, the ignitability of the fuel deteriorates. Therefore, as described above, the compression ratio control means has a large difference. As the actual compression ratio increases, the compression ratio variable means 81 is driven and controlled.

  On the other hand, if the ignitability index value Pc is larger than the ignitability determination reference value Pc0 in step ST30, the fuel characteristic detecting means indicates that the ignitability of the fuel guided into the combustion chamber CC is better than the reference ignitability. Therefore, the compression ratio control means sends a drive control command to the compression ratio variable means 81 to change the actual compression ratio to a compression ratio lower than the reference compression ratio (step ST45). Here, the larger the difference between the ignitability index value Pc and the ignitability determination reference value Pc0 at this time, the better the ignitability of the fuel. Therefore, as described above, the compression ratio control means has a large difference. As the actual compression ratio decreases, the compression ratio variable means 81 is driven and controlled so that the actual compression ratio becomes lower.

  By the way, when the electronic control unit 1 according to the first embodiment selects a combustion mode other than the compression auto-ignition diffusion combustion mode in step ST25, the electronic control unit 1 selects a compression ratio corresponding to the combustion mode and the compression. The compression ratio variable means 81 is driven and controlled so that the ratio becomes equal (step ST50).

  For example, in the multifuel internal combustion engine of the first embodiment, a premixed spark ignition flame propagation combustion mode is prepared as a combustion mode other than the compression autoignition diffusion combustion mode. However, when operating in this premixed spark ignition flame propagation combustion mode, knocking due to abnormal combustion is likely to occur, particularly in the high load region, as the compression ratio increases. On the other hand, such knocking can be suppressed by using a fuel excellent in knock resistance (for example, high-octane gasoline relative to regular gasoline).

  Therefore, when the premixed spark ignition flame propagation combustion mode is selected in step ST25, for example, the compression ratio is changed so that the actual compression ratio is changed according to the knock resistance of the fuel guided into the combustion chamber CC. The control means is configured. In such a case, if the knock resistance of the fuel is superior to the standard knock resistance, the actual compression ratio is increased, and if the knock resistance of the fuel is inferior to the standard knock resistance, the actual compression ratio is increased. make low. As a result, when the knock resistance is good, the actual compression ratio can be increased while suppressing the occurrence of knocking. Therefore, as the compression ratio is increased, the thermal efficiency is improved to achieve higher torque and higher output. Can do. When knock resistance is poor, the occurrence of knocking can be suppressed by reducing the actual compression ratio.

  Here, the knock resistance can be grasped from the octane number of the fuel. In this case, the octane number becomes a knock resistance index value, and is detected in the same manner as in the case of the cetane number described above. The knock resistance can also be grasped by using information on the trace knock ignition timing at the time of knock control performed based on a detection signal of a knock sensor (not shown). In this case, the relationship between the trace knock ignition timing at the time of knock control and the reference ignition timing becomes the knock resistance index value, for example, good premixed spark ignition flame propagation with fuel exhibiting good knock resistance as an index. The above relationship when combustion is being performed is set as a knock resistance criterion value.

  In the multi-fuel internal combustion engine of the first embodiment shown above, when operating in the compression self-ignition diffusion combustion mode when the ignitability of the fuel guided into the combustion chamber CC is poor, the actual compression is so bad that the ignitability is poor. Increase the ratio. For this reason, according to this multi-fuel internal combustion engine, the temperature of the compressed air in the combustion chamber CC rises as the ignitability of the fuel becomes worse, so even the fuel with poor ignitability self-ignites in the compressed air. It becomes easy. Therefore, this multi-fuel internal combustion engine can be operated in the compression auto-ignition diffusion combustion mode, and misfires and diesel knocks are not caused at the time of compression auto-ignition diffusion combustion, thus increasing noise and vibration during combustion. Will be able to prevent.

  Further, in this multi-fuel internal combustion engine, when operating in the compression auto-ignition diffusion combustion mode when the ignitability of the fuel led into the combustion chamber CC is good, the actual compression ratio is lowered so that the ignitability is good. Go. For this reason, according to the multifuel internal combustion engine, mechanical loss is reduced and the maximum in-cylinder pressure Pmax is reduced, so that the engine performance and fuel efficiency can be improved.

  As described above, according to the multifuel internal combustion engine of the first embodiment, it is possible to perform a good compression ignition operation by changing the actual compression ratio according to the ignitability of the fuel guided into the combustion chamber. .

  By the way, in the multifuel internal combustion engine of the first embodiment, the compression ratio variable means 81 is provided to change the actual compression ratio. For example, when the variable valve mechanism described above is provided, the compression ratio variable means 81 is compressed. The ratio variable means 81 may be used. In this case, the compression ratio control means changes the actual compression ratio by changing the valve closing timing of the intake valve 31 in accordance with the ignitability of the fuel guided into the combustion chamber CC. Here, if the intake valve 31 is closed early, the actual compression ratio becomes low, and if the intake valve 31 is closed late, the actual compression ratio becomes high.

  Second Embodiment A multifuel internal combustion engine according to a second embodiment of the present invention will be described with reference to FIGS.

  In general, the internal combustion engine has improved thermal efficiency by increasing the compression ratio, and therefore can achieve higher torque and higher output. For this reason, in the multi-fuel internal combustion engine of the first embodiment described above, when operating in the compression auto-ignition diffusion combustion mode when the ignitability of the fuel is poor, the actual compression ratio is changed to a high value, so high torque and high output are achieved. Can be achieved.

  On the other hand, when the multi-fuel internal combustion engine of the first embodiment is operated in the compression auto-ignition diffusion combustion mode when the ignitability of the fuel is good, the actual compression ratio is changed to a low value to reduce the mechanical loss and the maximum in-cylinder pressure Pmax. Improvements in engine performance such as engine output due to the decrease are made, but in most cases, the effect of the decrease in engine performance due to the decrease in thermal efficiency becomes stronger than the improvement cost.

  Therefore, the multifuel internal combustion engine of the second embodiment is provided with a supercharger such as an exhaust turbocharger or a supercharger with an electric motor (so-called supercharger) in the multifuel internal combustion engine of the first embodiment. To improve. Here, the case where the exhaust turbocharger 90 is provided as shown in FIG. 4 is illustrated, but a supercharger with an electric motor may be provided instead of the exhaust turbocharger 90, and the exhaust turbocharger is provided. A supercharger with an electric motor may be provided together with 90.

  The exhaust turbocharger 90 according to the second embodiment is disposed between the intake passage 21 and the exhaust passage 101, a turbine 91 disposed on the exhaust passage 101, and a rotating shaft that rotates in synchronization with the turbine 91. 92 and a compressor 93 on the intake passage 21 that rotates together with the turbine 91 via the rotating shaft 92. As a result, the exhaust turbocharger 90 uses the energy of the exhaust gas to rotate the turbine 91 and the compressor 93 and forcibly supplies a large amount of air supercharged by the compressor 93 into the combustion chamber CC. Supply high torque and high output.

  Here, also in the second embodiment, as in the multifuel internal combustion engine of the first embodiment, the actual compression ratio is changed according to the ignitability of the fuel guided into the combustion chamber CC. For this reason, when the supercharging pressure of the exhaust turbocharger 90 is set so as to achieve good engine performance during the compression self-ignition diffusion combustion mode operation when the fuel ignitability is good, the ignitability of the fuel that can increase the actual compression ratio. When the compression autoignition diffusion combustion mode operation is performed when the engine is bad, the engine body may be damaged due to an excessive supercharging effect. In other words, supplying the supercharged air into the combustion chamber CC with the exhaust turbocharger 90 has the same meaning as increasing the actual compression ratio, so the engine body is overloaded and damaged by excessively high compression ratio. there's a possibility that.

  Therefore, in the second embodiment, the supercharging pressure is changed to the optimum supercharging pressure in accordance with the compression ratio of the multi-fuel internal combustion engine, in other words, in accordance with the ignitability of the fuel introduced into the combustion chamber CC. Supply control means is provided in the electronic control unit 1. Specifically, for example, at the reference compression ratio {when the reference ignitability (Pc = Pc0)}, an upper limit value of the supercharging pressure (hereinafter referred to as a reference) that can balance engine performance and durability. , "Reference boost pressure upper limit value") is set, and when the compression ratio is higher than the reference compression ratio {when ignitability is poor (Pc <Pc0))}, the boost pressure upper limit value is used as a reference. When the compression pressure is set lower than the upper limit of the supercharging pressure and the compression ratio is lower than the reference compression ratio {when ignitability is good (Pc> Pc0)}}, the supercharging pressure upper limit is set to the reference supercharging pressure upper limit. Higher than that.

  Here, in a general exhaust turbocharger, the upper limit value of the supercharging pressure is set by a control device for the supercharging limit such as a wastegate valve, so that a supercharging pressure larger than the upper limit value is not applied. It has become. For this reason, in order to change the supercharging pressure, a supercharging limit control device that arbitrarily changes the upper limit value of the supercharging pressure is required. In the second embodiment, such a supercharging limit control device 94 is provided in the exhaust turbocharger 90. For example, the supercharging limit control device 94 includes a variable valve whose valve opening angle can be changed in accordance with an instruction from the electronic control device 1, and the upper limit value of the supercharging pressure is varied by operating this variable valve.

  Hereinafter, an example of the compression ratio change control operation and the supercharging pressure change control operation of the electronic control device 1 according to the second embodiment will be described with reference to the flowchart of FIG. In the second embodiment, the compression ratio change control operation is the same as that in the first embodiment, and the difference will be described below.

  When the ignitability of the fuel guided into the combustion chamber CC is the same as the ignitability determined as a reference (Pc = Pc0), the electronic control unit 1 according to the second embodiment sets the actual compression ratio to the reference compression ratio in step ST35. change. In such a case, the supercharging pressure control means of the electronic control unit 1 sets the upper limit value of the supercharging pressure of the exhaust turbo supercharger 90 as the reference supercharging pressure upper limit value (step ST36).

  In addition, when the ignitability of the fuel introduced into the combustion chamber CC is worse than the reference ignitability (Pc <Pc0), the electronic control unit 1 according to the second embodiment sets the actual compression ratio to the reference compression ratio in step ST40. Change to a higher compression ratio. In such a case, the supercharging pressure control means sets the upper limit value of the supercharging pressure of the exhaust turbocharger 90 to be lower than the reference supercharging pressure upper limit value (step ST41). Here, the greater the difference between the ignitability index value Pc and the ignitability determination reference value Pc0 at this time, the worse the ignitability of the fuel. As the difference increases, the actual compression ratio is increased in step ST40. ing. For this reason, in step ST41, the upper limit value of the supercharging pressure is lowered as the difference between the ignitability index value Pc and the ignitability determination reference value Pc0 increases (the actual compression ratio is increased).

  On the other hand, when the ignitability of the fuel guided into the combustion chamber CC is better than the reference ignitability (Pc> Pc0), the electronic control unit 1 according to the second embodiment uses the actual compression ratio as the reference compression ratio in step ST45. Change to a lower compression ratio. In such a case, the supercharging pressure control means sets the upper limit value of the supercharging pressure of the exhaust turbo supercharger 90 to be higher than the reference supercharging pressure upper limit value (step ST46). Here, the greater the difference between the ignitability index value Pc and the ignitability determination reference value Pc0 at this time, the better the ignitability of the fuel. Therefore, in step ST45, the actual compression ratio is lowered as the difference increases. ing. For this reason, in step ST46, the upper limit value of the supercharging pressure is increased as the difference between the ignitability index value Pc and the ignitability determination reference value Pc0 increases (lowers the actual compression ratio).

  Furthermore, when the electronic control unit 1 according to the second embodiment selects a combustion mode other than the compression auto-ignition diffusion combustion mode, the actual compression ratio is changed to a compression ratio corresponding to the combustion mode in step ST50. In such a case, the supercharging pressure control means lowers the upper limit value of the supercharging pressure and lowers the actual compression ratio if the actual compression ratio is to be increased, based on the same idea as in, for example, the compression auto-ignition diffusion combustion mode operation. If so, the upper limit value of the supercharging pressure is increased (step ST51). Accordingly, in this multi-fuel internal combustion engine, as in the case of the above-described operation in the compression auto-ignition diffusion combustion mode, it is possible to achieve high torque and high output while ensuring the durability of the engine body.

  In the multi-fuel internal combustion engine of the second embodiment as described above, when the fuel guided into the combustion chamber CC is poor in ignitability, when it is operated in the compression autoignition diffusion combustion mode, the actual compression is so bad that the ignitability is poor. While increasing the ratio, the upper limit value of the supercharging pressure of the exhaust turbocharger 90 is decreased. Therefore, according to this multi-fuel internal combustion engine, it becomes possible to perform compression auto-ignition diffusion combustion even with fuel with poor ignitability as in the case of the first embodiment. Since diesel knock is suppressed, increase in noise and vibration during combustion can be prevented. Furthermore, according to the multifuel internal combustion engine, even if the actual compression ratio is increased, the upper limit value of the supercharging pressure can be reduced to suppress the increase in the maximum in-cylinder pressure Pmax. A fatal burden is not applied, and the actual compression ratio can be increased without deteriorating the durability of the engine body. Therefore, in this multi-fuel internal combustion engine, not only good compression self-ignition diffusion combustion is realized, but also high torque and high output can be achieved while ensuring the durability of the engine body.

  Further, in this multi-fuel internal combustion engine, when operating in the compression auto-ignition diffusion combustion mode when the ignitability of the fuel introduced into the combustion chamber CC is good, the actual compression ratio is lowered as the ignitability is improved. Thus, the upper limit value of the supercharging pressure of the exhaust turbocharger 90 is increased. For this reason, according to this multi-fuel internal combustion engine, as in the case of the first embodiment, the engine performance and the fuel efficiency can be improved as the mechanical loss is reduced and the maximum in-cylinder pressure Pmax is reduced. Further, according to this multi-fuel internal combustion engine, the exhaust turbo supercharger 90 is operated at a higher supercharging pressure as the ignitability of the fuel introduced into the combustion chamber CC becomes better (as the actual compression ratio decreases). As a result, it is possible to obtain more supercharging effects, increase torque and output, and improve engine performance that decreases as the actual compression ratio decreases. . Further, when the upper limit value of the supercharging pressure is increased in this way, the maximum in-cylinder pressure Pmax is increased and the durability of the engine body is deteriorated. However, as the upper limit value of the supercharging pressure is increased, the actual compression ratio is increased. Since the pressure decreases, an increase in the maximum in-cylinder pressure Pmax can be suppressed. Therefore, in this multi-fuel internal combustion engine, since a fatal burden is not applied to the engine body, the engine performance and fuel consumption performance are improved by reducing the compression ratio and the boost pressure is reduced without deteriorating the durability of the engine body. High torque and high output can be achieved by increasing the upper limit value.

  Thus, according to the multifuel internal combustion engine of the second embodiment, a good compression autoignition operation can be performed by varying the actual compression ratio and the supercharging pressure in accordance with the ignitability of the fuel guided into the combustion chamber. become able to.

  By the way, the compression ratio variable means 81 by the variable valve mechanism described in the first embodiment lowers the actual compression ratio by the early closing of the intake valve 31, but at that time, the actual displacement is reduced and the output is reduced. It becomes a factor. However, when the variable valve mechanism used as the compression ratio variable means 81 is applied to the multifuel internal combustion engine of the second embodiment, the supercharging pressure of the exhaust turbocharger 90 is reduced when the actual compression ratio is reduced. Therefore, the engine performance can be improved only by compensating for the decrease in engine performance due to the decrease in the actual displacement.

  Here, in each of the first and second embodiments described above, the multi-fuel internal combustion engine operated in either the compression auto-ignition diffusion combustion mode or the premixed spark ignition flame propagation combustion mode is illustrated, but the multi-fuel according to the present invention is illustrated. The internal combustion engine only needs to have at least the compression auto-ignition diffusion combustion mode prepared as the combustion mode.

  In the first and second embodiments, the so-called in-cylinder direct injection type multi-fuel internal combustion engine in which the mixed fuel of the first fuel F1 and the second fuel F2 is directly injected into the combustion chamber CC is illustrated. Each invention in the first and second embodiments may be applied to a multi-fuel internal combustion engine in which the mixed fuel is injected not only into the combustion chamber CC but also into the intake port 11b. In each of the first and second embodiments, the fuel supply device 50 is configured to inject the mixed fuel that has been mixed in advance by the fuel mixing means 53 from the fuel injection valve 57 into the combustion chamber CC. The respective fuels (first fuel F1 and second fuel F2) may be supplied individually into the combustion chamber CC without using the fuel mixing means 53. In such a multi-fuel internal combustion engine, each fuel injection valve is driven and controlled so as to achieve a set fuel mixture ratio.

  For example, this type of multi-fuel internal combustion engine is configured by replacing the fuel supply device 50 with the fuel supply device 150 shown in FIG. The fuel supply device 150 shown in FIG. 6 has a first fuel supply path for directly injecting the first fuel F1 (fuel having excellent ignitability) into the combustion chamber CC, and a second fuel F2 (evaporable) to the intake port 11b. And a second fuel supply path for injecting excellent fuel). The first fuel supply path includes a first feed pump 52A that sucks up the first fuel F1 from the first fuel tank 41A and sends it to the first fuel passage 51A, and a high-pressure fuel for the first fuel F1 in the first fuel passage 51A. A high pressure fuel pump 55A that pumps the fuel into the passage 54A, a delivery passage 56A that distributes the first fuel F1 in the high pressure fuel passage 54A to the respective cylinders, and the first fuel F1 supplied from the delivery passage 56A in the combustion chamber CC. And a fuel injection valve 57A for each cylinder that injects into the cylinder. On the other hand, the second fuel supply path draws the second fuel F2 from the second fuel tank 41B and sends it to the second fuel passage 51B, and the second fuel F2 in the second fuel passage 51B is second. A high pressure fuel pump 55B that pumps the fuel to the third fuel passage 54B, a delivery passage 56B that distributes the second fuel F2 in the third fuel passage 54B to the respective cylinders, and an intake of the second fuel F2 supplied from the delivery passage 56B. And a fuel injection valve 57B for each cylinder that injects into the port 11b.

  As described above, the multi-fuel internal combustion engine according to the present invention is useful as a technique for realizing a good compression self-ignition operation according to the ignitability of the fuel guided into the combustion chamber.

It is a figure shown about the structure of Example 1 of the multi-fuel internal combustion engine which concerns on this invention. It is a figure which shows an example of the combustion mode map data used when setting a combustion mode. 3 is a flowchart illustrating an operation of changing a compression ratio in the multifuel internal combustion engine of the first embodiment. It is a figure shown about the structure of Example 2 of the multi-fuel internal combustion engine which concerns on this invention. 6 is a flowchart illustrating an operation of changing a compression ratio and a supercharging pressure in the multifuel internal combustion engine of the second embodiment. It is a figure shown about the structure of the modification of the multi-fuel internal combustion engine which concerns on this invention.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Electronic controller 41A 1st fuel tank 41B 2nd fuel tank 81 Compression ratio variable means 90 Exhaust turbo supercharger 94 Supercharging limit control device CC Combustion chamber F1 1st fuel F2 2nd fuel Pc Ignition index value Pc0 Ignition property Judgment standard value

Claims (2)

In a multi-fuel internal combustion engine in which at least one of at least two types of fuels having different properties is led to a combustion chamber or a mixed fuel composed of the at least two types of fuel is led to a combustion chamber and operated.
Compression ratio variable means for changing the actual compression ratio to an arbitrary compression ratio;
Fuel characteristic detection means for detecting the ignitability of the fuel in the combustion chamber;
In the compression self-ignition operation, if the ignitability of the fuel in the combustion chamber is worse than the predetermined ignitability, the actual compression ratio is increased, and if the ignitability of the fuel in the combustion chamber is better than the predetermined ignitability. Compression ratio control means for controlling the compression ratio variable means so as to lower the actual compression ratio;
A multi-fuel internal combustion engine characterized by comprising: An exhaust turbocharger or / and a supercharger with electric motor;
A supercharging pressure control means for setting a high upper limit of a supercharging pressure of the exhaust turbocharger or / and the supercharger with electric motor when the compression ratio control means lowers the actual compression ratio;
The multifuel internal combustion engine according to claim 1, further comprising:

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