Patent Application
20100077990
The present description relates to control of a spark ignited internal combustion engine having a fuel injector which directly injects fuel into the combustion chamber.
To improve efficiency of an internal combustion engine, particularly for an automotive vehicle, it is preferred that the engine have a high geometric compression ratio. However, an increased compression ratio may cause increased combustion chamber temperatures. A higher temperature in the combustion chamber may cause abnormal combustion such as knocking or pre-ignition.
Knocking occurs after spark ignition. In particular, the spark starts combustion of air fuel mixture in the combustion chamber, and the combusted gas expands and compresses the un-combusted air fuel mixture. As the temperature of the compressed air fuel mixture rises, the un-combusted air fuel mixture ignites itself when the gas temperature reaches the ignition temperature.
On the other hand, pre-ignition occurs before spark ignition. Specifically, an air fuel mixture is compressed, and as the mixture is compressed, its temperature rises during the compression stroke. If the gas temperature reaches the ignition temperature before a spark occurs, the whole air fuel mixture ignites. Therefore, since the combusted gas is expanding in the compression stroke, pre-ignition may cause higher pressure in the combustion chamber than knocking.
There is known, and for example described in US2007/0227503A1, a divisional fuel injection method which injects a first stage fuel in the intake stroke and a second stage fuel in the compression stroke during a cylinder cycle. The second stage fuel is injected into the compressed, relatively hot air fuel mixture, and cools the gas mixture by using some of the heat in the gas to vaporize the fuel. Therefore, it can decrease the temperature in the combustion chamber and reduce the possibility of abnormal combustion.
The prior method can reduce the possibility of knocking, but it still does not significantly reduce the possibility of pre-ignition in a spark ignited internal combustion engine having a relatively high compression ratio. Therefore, there is room to improve the prior divisional fuel injection method.
The inventors herein have rigorously studied ways to suppress the possibility of pre-ignition and have developed a unique method to reduce the possibility of pre-ignition.
Accordingly, there is provided, in one aspect of the present description, a method of controlling a spark ignited internal combustion engine having a fuel injector which injects fuel directly into a combustion chamber. The method comprises injecting a total amount of fuel into a combustion chamber during a cylinder cycle, the total amount of fuel injected by early in a compression stroke of the cylinder cycle at a first engine speed. The method further comprises injecting a first stage of fuel in the combustion chamber during a cylinder cycle, the first stage of fuel injected by early in a compression stroke of a cylinder cycle, and injecting a second stage of fuel, after injecting the first stage of fuel, the second stage of fuel injected by late in the compression stroke during the cylinder cycle at a second engine speed that is less than the first engine speed. And, the amount of the second stage fuel is greater than the amount of the first stage fuel.
According to the first aspect, when the engine speed is relatively low, the second stage fuel is injected late in the compression stroke. The second stage fuel receives heat from a combustion chamber wall such as a cylinder wall, a piston top and a cylinder ceiling during a period between the injection and the ignition. The heat receiving period starts late in the compression stroke; therefore, the heat receiving period is shorter, and the heat received by the second stage fuel is reduced. Then, the molecules of the second stage fuel are in less motion such that they do not collide with the oxygen molecules fast enough to ignite. Consequently, the possibility of pre-ignition of the second stage fuel is reduced.
In the same occasion, the first stage fuel is injected much earlier than the second stage fuel is injected, and it receives more heat from the combustion chamber wall and is more likely to cause a pre-ignition. However, the amount of the first stage fuel is less than the amount of the second stage fuel. Therefore, less of the fuel molecules collide with the oxygen molecules and the exothermal reaction caused by the molecular collision causes less heat. As a result, a temperature of the first stage fuel is temporarily decreased at the injection timing of the second stage fuel and does not reach the ignition temperature before the spark ignition. Consequently, there is a reduced possibility of pre-igniting the first and second stage fuels before the spark ignition.
Even though the heat receiving period of the second stage fuel is short enough to suppress the pre-ignition of the second stage fuel, it is long enough to atomize/evaporate the second stage fuel since it is done when the engine speed is relatively low.
On the other hand, when the engine speed is relatively high, the heat receiving period is shorter and the possibility of abnormal combustion is reduced. Therefore, it is possible to complete the fuel injection by early in the compression stroke. This leads to there being enough time to atomize or evaporate the fuel.
There is provided, in a second aspect of the present description, a system comprising a spark ignited internal combustion engine, a fuel injector which injects fuel directly into a combustion chamber, and a controller. The controller is configured to control the fuel injector to inject a total amount of fuel into the combustion chamber during a cylinder cycle, the total amount of fuel injected by early in a compression stroke of said cylinder cycle at a first engine speed. The controller is further configured to control the fuel injector to inject a first stage of fuel into the combustion chamber during a cylinder cycle, the first stage of fuel injected by early in a compression stroke of the cylinder cycle, and injecting a second stage of fuel, after injecting the first stage of fuel, the second stage of fuel injected by late in the compression stroke during the cylinder cycle at a second engine speed, the second engine speed less than the first engine speed. Further, a desired torque of said engine is greater than a predetermined torque when the first stage fuel and second stage fuel are injected. Further still, the amount of the second stage fuel is greater than amount of the first stage fuel.
The system according to the second aspect performs the method according to the first aspect. Therefore, the system can advantageously suppress the possibility of occurrence of pre-ignition.
In embodiments, a ratio of the amount of the second stage fuel relative to the total amount of the first and second stage fuel may be decreased as an engine speed of the engine increases so as to secure enough time to evaporate the injected fuel.
In other embodiments, a ratio of the amount of the second stage fuel relative to the total amount of the first and second stage fuel may be increased as the desired torque of the engine increases or as a temperature in the combustion chamber increases. As such, the method can suppress the possibility of occurrence of the pre-ignition in a situation where more of the fuel molecules may collide with the oxygen molecules or the collision may be faster.
In addition, in other embodiments, a negative valve overlap period may be decreased, for example, by opening the intake valve earlier during a cylinder cycle, as the desired engine torque increases so as to induct more fresh air into the combustion chamber to increase engine torque.
In still other embodiments, the injected fuel may be ignited with a spark after a top dead center of the compression stroke during the cylinder cycle when the second stage fuel is injected so as to suppress the possibility of the occurrence of knocking in addition to suppressing the possibility of pre-ignition. The fuel is injected at a first pressure when the fuel injection is completed by early in the compression stroke during the cylinder cycle, and the fuel is injected at a second pressure when the second stage fuel is injected. The second pressure may be greater than the first pressure so as to enhance the evaporation/atomization of the second stage fuel which is injected in the highly pressurized gas and at a time closer to the ignition timing.
In still other embodiments, the intake valve may be opened after closing the exhaust valve, and the first stage fuel may be injected after the opening the intake valve, thereby enhancing the evaporation/atomization of the first stage fuel. Later intake valve opening increases vacuum in the cylinder before the intake valve opens, and as a result, the flow rate of air entering the cylinder increases when the intake valve opens.
In still other embodiments, the spark ignited internal combustion engine may have a geometric compression ratio of 14 or greater so that an engine operating efficiency at a condition where the abnormal combustion is not likely to occur can be improved.
In still other embodiments, the spark ignited internal combustion engine may have a turbocharger or a mechanical supercharger to increase air charge in the cylinder so that an engine output torque increases without increasing the possibility of the pre-ignition.
The advantages described herein will be more fully understood by reading an example of embodiments in which the above aspects are used to advantage, referred to herein as the Detailed Description, with reference to the drawings wherein:
Embodiments of the present description will now be described with reference to the drawings, starting with
The engine main body 1 is a four-cycle spark-ignited internal combustion engine installed in a vehicle, such as an automobile. An output shaft of the engine main body 1 is coupled to a drive wheel via a transmission in order to drive the vehicle. The engine main body 1 includes a cylinder block 12 and a cylinder head 13 placed thereon. Inside the cylinder block 12 and the cylinder head 13, a plurality of cylinders 11 are formed. The number of cylinders 11 is not limited; however, four cylinders 11 are formed in this embodiment, as one example. Further, in the cylinder block 12, a crankshaft 14 is supported rotatably by a journal, a bearing and the like.
To each of the cylinders 11, a piston 15 is slideably inserted and fitted to connecting rod 16, over which a combustion chamber 17 is laid out.
In this embodiment, a geometric compression ratio of the engine main body 1 is set to approximately 14, which is the ratio of the volume of the combustion chamber 17 when the piston 15 is positioned at the bottom dead center to the volume of the combustion chamber 17 when the piston 15 is positioned at the top dead center. Of course, the value of the geometric compression ratio is not limited to 14. For example, it may be preferable that the geometric compression ratio is higher from the point of view of improving engine efficiency. However, as the geometric compression ratio is set higher, an in-cylinder temperature can become too high in the compression stroke, thereby increasing the possibility of an auto-ignition occurring at an unexpected timing. Therefore, the geometric compression ratio of the engine main body 1 may be, preferably, between 14 and 16, but is not limited to this range, especially if the engine is supercharged or turbocharged.
The cylinder head 13 is formed with two intake ports and two exhaust ports communicating with the respective one of the combustion chambers. In
The intake valve driving mechanism 30 and the exhaust valve driving mechanism 40 have an intake camshaft 31 and an exhaust camshaft 41, respectively. The intake camshaft 31 and the exhaust camshaft 41 are coupled to the crankshaft 14 via a power transmission mechanism such as a known chain-sprocket mechanism. The power transmission mechanism is configured such that the camshafts 31 and 41 rotate one time while the crankshaft 14 rotates two times.
Further, in the intake valve driving mechanism 30, there is provided an intake camshaft phase changing mechanism 32 between the power transmission mechanism and the intake camshaft 31. The intake camshaft phase changing mechanism 32 is set to change the valve timing of the intake valve 21, in which a phase difference between the crankshaft 14 and the intake camshaft 31 is changed by changing the phase difference between the driven shaft, which is arranged concentrically with the intake camshaft 31 and is directly driven by the crankshaft 14, and the intake camshaft 31.
The intake camshaft phase changing mechanism 32 includes, for example, a hydraulic pressure mechanism where a plurality of liquid holding chambers are arranged in a circumferential direction between the driven shaft and the intake camshaft 31. A pressure difference between the liquid holding chambers is used to change the phase difference. An electromagnetic mechanism having an electromagnet is provided between the driven shaft and the intake camshaft 31. Current is applied to the electromagnet to change the phase difference between the camshaft and the crankshaft. The intake camshaft phase changing mechanism 32 changes the phase difference based on the valve timing of the intake valve 21 calculated by the engine controller 100, described later.
In this embodiment, the intake camshaft phase changing mechanism 32 changes the valve opening timing IVO and valve closing timing IVC of the intake valve 21 by changing the phase difference. The lift amount (i.e., a valve profile of the intake valve 21) is kept constant. A phase angle of the intake camshaft 31 is detected by a cam phase sensor 35, and a signal θINT
Also, in the exhaust valve driving mechanism 40, there is provided an exhaust camshaft phase changing mechanism 42 between the power transmission mechanism and the intake camshaft 41. The exhaust camshaft phase changing mechanism changes the valve opening timing EVO and valve closing timing EVC of the exhaust valve 22 in the same manner as in the intake camshaft phase changing mechanism.
The intake port 18 communicates with a surge tank 55a via an intake manifold 55b. The air intake passage upstream of the surge tank 55a is provided with the throttle body (throttle valve actuator) 56. A throttle valve 57 is pivotally provided inside the throttle body 56 for adjusting the air flowing from atmosphere to the surge tank 55a. The throttle valve 57 can change the opening area of the air intake passage (i.e., the flow passage area) to change the mass air flow rate, and the pressure in the air intake passage downstream of the throttle valve 57. The throttle valve 57 is actuated by a throttle valve actuator 58. The throttle valve actuator 58 actuates the throttle valve 57 such that the opening TVO of the throttle valve 57 is to be a target throttle valve opening TVOD calculated in the engine controller 100. Here, the air intake passage 55 may include all of the intake port 18, the intake manifold 55b and the surge tank 55a downstream of the throttle valve 57. In this embodiment, an amount of air to be inducted into the cylinder 11, that is, the air charge amount CE inside the cylinder 11 is controlled to have an adequate value by adjusting the opening of the throttle valve 57 and the closing timing of the intake valve 21.
The exhaust port 19 communicates with an exhaust pipe via an exhaust manifold 60. In the exhaust pipe, an exhaust gas treatment system is arranged. A specific constitution of the exhaust gas treatment system is not limited to, but may include those having a catalytic converter 61 of a three-way catalyst, a lean NOx catalyst, an oxidation catalyst and the like.
The surge tank 55a and the exhaust manifold 60 communicate with each other via an EGR pipe 62, constituted such that a part of the exhaust gas may be circulated to an intake side. Provided in the EGR pipe 62 is an EGR valve 63 for adjusting the flow volume of EGR gas circulating to the intake side through the EGR pipe 62. The EGR valve 63 is actuated by an EGR valve actuator 64. The EGR valve actuator 64 actuates the EGR valve 63 such that the opening of the EGR valve 63 becomes an EGR opening EGRopen calculated by the engine controller 100. This makes it possible to adjust the flow volume of the EGR gas to an adequate value.
The cylinder head 13 has spark plugs 51 attached thereto such that a tip of each spark plug faces the combustion chamber 17. The spark plug 51 generates a spark in the combustion chamber 17 when supplied with current by an ignition system 52, based on an ignition timing signal SA output from the engine controller 100, described later in detail.
Further, the cylinder head 13 has fuel injectors 53 attached thereto for injecting fuel directly into the respective combustion chambers 17 such that a tip of each of the fuel injectors faces the combustion chamber 17. In more detail, the fuel injector 53 is arranged such that the tip thereof is positioned below the two intake ports 18 in a vertical direction, and midway between the two intake ports 18 in a horizontal direction. The fuel injector 53 injects a predetermined amount of fuel into the combustion chamber 17 when a solenoid coupled to the fuel injector 53 is supplied with current by a fuel system 54 for a predetermined period of time based on a fuel pulse signal FP calculated by and output from the engine controller 100. The fuel system 54 includes a pressure regulator and supplies fuel to the fuel injectors 53 at a fuel pressure in accordance with a fuel pressure signal PFUEL calculated by and output from the engine controller 100.
The engine controller 100 is a controller having a known microcomputer as a base and includes a CPU for executing a program, a memory such as RAM and ROM for storing a program and data, and an I/O bus for inputting and outputting various signals.
The engine controller 100 receives inputs via the I/O bus, with various information such as an intake airflow AF detected by an air flow meter 71, an air pressure MAP inside the surge tank 55a detected by an intake pressure sensor 72, a crank angle pulse signal detected by a crank angle sensor 73, an oxygen concentration EGO of the exhaust gas detected by an oxygen concentration sensor 74, an amount a of depression of an accelerator pedal by a driver of the automobile detected by an accelerator pedal position sensor 75, a vehicle speed VSP detected by a vehicle speed sensor 76, an engine temperature TENG detected by an engine coolant temperature sensor 77, and an intake air temperature TAIR detected by an intake air temperature sensor which detects a temperature inside the surge tank 55a or a temperature upstream of the throttle body 56. Then, the engine controller 100 calculates control parameters for various actuators such that the air charge amount, ignition timing and the like in cylinder 11 may be an appropriate value according to the operating conditions based on the input information. For example, control parameters such as a throttle valve opening TVO, the fuel injection amount FP, the fuel pressure PFUEL, the ignition timing SA, a target value of the intake valve timing θINT
Control routines the engine controller 100 executes will be described with reference to flowcharts illustrated in
Referring to
Then, the first routine R1 proceeds to a step S4 and determines a target angular phase θINT
On the other hand, as the target air charge CED increases, the target angular phase θINT
After step S4, the first routine R1 proceeds to a step S5 and determines a target angular phase θEXH
In another exemplary state where the target air charge CED and engine speed NENG is not in the Positive region, but in a region labeled “Negative” in the map of
Referring back to
After the step S7, the first routine R1 proceeds to step S8 and drives the respective actuators according to the computed control parameters such as the fuel injection amount FP, the fuel pressure PFUEL, the ignition timing SA, the target intake camshaft phase θINT
Referring to
The second routine R2 chooses one of the fuel injection modes in accordance with one of areas A1 through A4 of
Referring back to
When it is determined at step S22 that the engine operating condition is not in the area A1 or A4 (NO), the second routine R2 proceeds to a step S25 and determines whether or not the engine operating condition is in the area A3 of
FP3=FP×1/2
FP4=FP×1/2.
Then, since the FP3+FP4=FP, the other fuel pulses FP0, FP1 and FP2 remain zero. After the step S26, the second routine R2 proceeds to the step S24 and determines the fuel pressure PFUEL to be the lower pressure P1 as described above. At step S8 of the first routine R1, the fuel system 54 is controlled to set fuel pressure at the lower pressure P1, and the fuel injector 53 is driven to open its nozzle at a predefined timing after the intake valve 21 opens and to close the nozzle when the pulse width FP3 has passed. The injector nozzle is again opened at a predefined timing and closed when the pulse width FP4 has passed as is illustrated in the bottom diagram of
When it is determined at the step S25 that the engine operating condition is not in the area A3 of
When it is determined at the step S27 that the engine temperature TENG detected by the engine coolant temperature sensor 77 is higher than the threshold temperature TENG1, the second routine proceeds to a step S28 and determines whether or not the intake air temperature TAIR detected by the intake air temperature sensor 78 is higher than a threshold temperature TAIR1. If it is determined NO at the step S28, the second routine R2 proceeds to the step S26 and the system takes the divided fuel injection mode as described above.
When it is determined at the step S28 that the intake air temperature TAIR detected by the intake air temperature sensor 78 is higher than the threshold temperature TAIR1, the second routine R2 proceeds to a step S29 and determines fuel division ratios DR1 and DR2 based on the target air charge CED, the engine speed NENG and the intake air temperature TAIR. The determination is made by reading values stored in tables, one example of which is illustrated in
Further, in one example, the tables are made to decrease the ratio DR1 and increase the ratio DR2 as the intake air temperature TAIR is increases, or, in other words, as a temperature in the combustion chamber 17 increases, as shown in
Referring back to
FP1=FP×P1/P2×DR1
FP2=FP×P1/P2×DR2.
Then, since DR1<DR2, the second fuel pulse width FP2 is greater than the first fuel width FP1. And, since DR1+DR2=1 and FP1+FP2=FP, the other fuel pulse widths FP0, FP3 and FP4 remain zero.
After the step S31, the second routine R2 returns. At step S8 of the first routine R1, the fuel system 54 is controlled to set fuel pressure at the higher pressure P2, and the fuel injector 53 is driven to open its nozzle at a predefined timing after the intake valve 21 opens. The nozzle is closed when the first pulse width FP1 has passed, for example, at 20° CA after the bottom dead center of the intake stroke, and again the nozzle is opened at a predetermined timing in a late compression stroke and closes when the second pulse width FP2 has passed, for example, at 20° CA before the top dead center of compression stroke, as is illustrated in the second bottom diagram of
The increased amount of the second stage fuel contributes to suppress the possibility of occurrence of pre-ignition as is described in greater detail below. It should also be noted that the ratio of the second stage fuel relative to the total of first and second stage fuel decreases as engine speed increases.
A graph of
On the other hand, a certain amount of the first stage fuel needs to be injected in consideration of the promotion of the vaporization or atomization of the fuel and uniform air fuel mixture inside of the cylinder. More specifically, the percentage of the second stage fuel amount may be preferably selected from 60% to 85%. In particular, the percentage of the second stage amount may be more preferably set to 75% or more (the percentage of the first stage fuel amount may be preferably 25% or less, for example, approximately ¼ or less, wherein 15% or more of the first stage injection quantity can be desirably secured).
The graph of
As can be seen from the graph of
In the embodiment described above, when the engine operating condition is in the area A2 illustrated in
As shown in
The higher fuel pressure in the intake and compression stroke injection mode contributes to enhance the atomization or evaporation of the first stage fuel. It also allows the second stage fuel to be injected in a shorter duration and secures the period for the atomization or evaporation of the second stage fuel until the spark ignition.
The spark ignition after the top dead center during a cylinder cycle in the intake and compression stroke injection mode contributes to suppress the possibility of occurrence of knocking.
Thus, the present description provides for a method of controlling a spark ignited internal combustion engine having a fuel injector which injects fuel directly into a combustion into a combustion chamber comprising; injecting a first stage of fuel in the combustion chamber during a cylinder cycle, the first stage of fuel injected by early in a compression stroke during the cylinder cycle, and injecting a second stage of fuel, after injecting the first stage of fuel, the second stage of fuel injected by late in the compression stroke during the cylinder cycle, the amount of the second stage fuel being greater than the amount of the first stage fuel. Wherein said second amount of fuel is increased in relation to an air charge amount, engine speed, and intake air temperature. In one example, the second fuel amount increases as intake temperature increases.
It is needless to say that this description is not limited to the illustrated embodiments, nor is the description limited to the various improvements described herein. Therefore, alternative designs are possible without departing from the substance of the description and claims.
For example, in the embodiment described above, when the engine operating condition is in the area A2 illustrated in
It should be understood that the embodiments herein are illustrative and not restrictive, since the scope of the invention is defined by the appended claims rather than by the description preceding them, and all changes that fall within metes and bounds of the claims, or equivalence of such metes and bounds thereof are therefore intended to be embraced by the claims.
| Number | Date | Country | Kind |
|---|---|---|---|
| 2008-247541 | Sep 2008 | JP | national |
