Referring to the drawings, a preferred embodiment of the invention will be described below. First, referring to
Hydraulic electromagnetic valve drive units 16, 17 that vertically reciprocate the intake valve 14 and the exhaust valve 15 according to an instruction signal from a valve event control unit 40 and open and close the intake inlet 10 or the exhaust outlet 11 are arranged at respective ends of the intake valve 14 and the exhaust valve 15. That is, timing at which the intake valve 14 and the exhaust valve 15 are opened and closed in the engine 1 and can be freely set by the valve event control unit 40. In the cylinder head 2, a fuel injection system 18 that injects fuel into the combustion chamber 9 according to an instruction signal from an FI control unit 41 is arranged, and fuel injection quantity and timing can be freely set by the FI control unit 41.
In the engine 1, to detect a condition of the engine 1, an intake air temperature sensor 30 that senses intake air temperature, an exhaust gas temperature sensor 31 that senses exhaust gas temperature, an absolute pressure sensor 32 that senses pressure in the combustion chamber 9 (in the cylinder), a crank timing sensor 33 that senses a turning angle of the crankshaft 8 and a throttle angle sensor 34 that senses an angle of a throttle valve 19 are mounted. To determine fuel injection quantity and injection timing, intake air temperature sensed by the intake air temperature sensor 30, a throttle angle sensed by the throttle angle sensor 34 and a turning angle of the crankshaft 8 sensed by the crank timing sensor 33 are input to the FI control unit 41. To determine timing at which the intake valve 14 and the exhaust valve 15 are opened and closed, a turning angle of the crankshaft 8 sensed by the crank timing sensor 33 and air-fuel ratio determined by the FI control unit 41 are input to the valve event control unit 40.
When compression self ignition operation is executed in the engine 1 configured as described above, the intake valve (IV) 14 is opened when the piston 6 passes a compression top dead center (TDC) as shown in
In normal spark ignition operation, the intake valve 14 and the exhaust valve 15 are simultaneously opened (valve overlap) in the vicinity of the compression top dead center between an exhaust stroke and an intake stroke, however, as described above, in the compression self ignition operation, to leave residual gas in the combustion chamber 9, both the intake valve 14 and the exhaust valve 15 are closed in the vicinity of the compression top dead center between the exhaust stroke and the intake stroke and an angle of the crankshaft 8 until the exhaust valve 15 is closed and the intake valve 14 is opened is called negative overlap quantity in the following description.
In the engine 1 in which such compression self ignition operation is enabled, the spark ignition operation and the compression self ignition operation can be switched according to an operational situation. Therefore, though the following ignition plug is not shown in
The reason why compression self ignition occurs in the combustion chamber 9 is that in-cylinder gas temperature reaches the chain initiation temperature of high temperature oxidation and according to this combustion method, the exhaust of nitrogen oxides (NOx) is reduced, however, it is thought that this is that time in which the temperature of a working medium exceeds nitrogen oxides generation initiation temperature (≈1800 K) is short. In-cylinder gas temperature after combustion when the in-cylinder gas temperature reaches 1000 K and fuel of specified quantity is completely combusted isometrically is estimated by thermodynamically estimating the rise of temperature by the combustion of the fuel in the combustion chamber 9. First, in the thermodynamic estimate, it is supposed that in-cylinder gas is completely combusted isometrically, heat by combustion is all used for the rise of in-cylinder gas temperature and no thermal dissociation occurs. Besides, it is supposed that in-cylinder gas is ideal gas, mean specific heat in a temperature raising stroke is applied and the temperature of in-cylinder gas is uniform. On such supposition, when the displacement of the engine is V, the compression ratio is ε, the charging efficiency is ηc, the air density is μ, the EGR rate (a rate at which burned gas after combustion (called EGR) accounts for in-cylinder all gas quantity) is ξ, the equivalence ratio (acquired by dividing the mass mixture ratio of fuel and air (fuel-air ratio) by fuel-air ratio (theoretical mixture ratio) when fuel and air are mixed just enough) is Φ, the low combustion heat of fuel is Hu and the theoretical mixture ratio of fuel is α, the mass Ga of air is expressed in the following expression (1), the mass Gf of fuel is expressed in the following expression (2), the mass Gu of air-fuel mixture is expressed in the following expression (3), the mass Gr of residual gas is expressed in the following expression (4), and the combustion heat Q of fuel is expressed in the following expression (5).
When the results are applied to a first law of thermodynamics, the following expression (6) is acquired and as a result, the rise of temperature after complete combustion since the initiation of self ignition (a deviation between combustion initiation temperature and peak temperature after combustion) ΔT is expressed in the following expression (7). In the expressions (6) and (7), Cvf denotes the isovolumic specific heat of air-fuel mixture and Cvr denotes the isovolumic specific heat of residual gas.
As shown in the expression (6), when in-cylinder gas is completely combusted, temperature which the in-cylinder gas reaches is univocally determined by the heat capacity (the product of the isovolumic specific heat Cvf, Cvr and the mass Gf, Gr of working media) of the working medium to the combustion heat Q of fuel. That is, difference in heat capacity between the same quantity of heat has an effect on burning velocity. As shown in the expression (7), as the rise of temperature from self ignition initiation temperature (hereinafter called a deviation) ΔT is determined by air-fuel ratio Φ/α and an EGR rate ξ, the deviation ΔT is determined independent of the displacement of the engine and an operational load if the air-fuel ratio Φ/α and the EGR rate ξ are known. As equivalence ratio Φ according to an operational condition is already known in the engine 1 where fuel injection quantity is controlled by the FI control unit 41 as described above, the deviation ΔT can be calculated if only an EGR rate ξ is sensed. As described above, the deviation ΔT between self ignition initiation temperature and peak temperature after combustion is a parameter substituted for combustion heat Q and burning velocity itself can be controlled by this parameter. Therefore, if respective adjustment elements are controlled so that the deviation ΔT is always in a range of suitable values, combustion temperature is kept between stable high temperature oxidation sustaining minimum temperature and nitrogen oxide generation temperature and stable compression self ignition combustion is enabled.
As described above, control elements for realizing compression self ignition combustion are the following two. First, it is control over timing and in-cylinder gas temperature at the compression top dead center is required to exceed 1000 K. Second, it is control over heat release after ignition and the deviation ΔT is required to be in a predetermined range. When in-cylinder gas temperature at the compression top dead center is lower than 1000 K, no compression self ignition combustion can be caused and in the meantime, when the deviation ΔT is not in the predetermined range, misfire after ignition or knocking is caused.
Further, as clear from the expression (7) and others, as an EGR rate ξ can be substituted for the deviation ΔT, combustion heat Q is controlled by controlling the EGR rate ξ, ideal compression self ignition combustion can be realized, and as a control parameter over the deviation ΔT can be possibly reduced by the EGR rate ξ, control algorithm can be simplified.
As described above, the deviation ΔT can be unequivocally determined by the EGR rate independent of an operational load of the engine 1 and velocity. Circles having a void inside shown in
Generally, when the EGR rate is in the vicinity of 20%, knocking is caused to be a limit because a nitrogen oxide increases. When the EGR rate is in the vicinity of 80%, misfire is caused to be a limit because hydrocarbon (HC) increases. As clear from
EGR includes outside EGR returned to an intake system through an EGR line (not shown in
To calculate the inside EGR rate, first, the mass Gr of residual gas is calculated according to an enthalpy preservation law of air-fuel mixture and residual gas and an equation of a state when the intake stroke is finished. To calculate the mass Gr of the residual gas, intake air temperature Tu sensed by the intake air temperature sensor 30, exhaust gas temperature Tr sensed by the exhaust gas temperature sensor 31, in-cylinder pressure Pivc sensed by the absolute pressure sensor 32 when the intake valve 14 is closed, the volume Vivc of the combustion chamber 9 at this time, the mass Gu of air-fuel mixture, the mass Ga of air and a constant R of gas are used. When the EGR line is provided in the engine 1, the mass Gegre of outside EGR gas is further used.
The mass Gf of air-fuel mixture is an already known value by fuel injection quantity controlled by the FI control unit 41 or a fuel consumption meter and the mass Ga of air can be calculated based upon air-fuel ratio and the mass Gf of fuel. The volume Vivc of the combustion chamber 9 when the intake valve 14 is closed can be calculated based upon the turning angle of the crankshaft 8 sensed by the crank timing sensor 33 and the data of the engine 1 (a bore×a stroke, the length of the connecting rod, compression ratio, offset quantity and others). The mass Gegre of outside EGR gas can be calculated based upon an outside EGR rate EGRe [%] shown in the following expression (8).
An inside EGR rate yegr can be calculated in the following expression (9) based upon mass Ga of fresh air, the mass Gf of air-fuel mixture and the mass Gr of residual gas respectively calculated as described above.
As shown in
In this embodiment, a case that negative overlap quantity is varied by the valve event control unit 40 and the deviation ΔT is controlled will be described as feedback control using the EGR rate (the inside EGR rate yegr). First, referring to
As shown in
As described above, as compression self ignition combustion in the next cycle is stably generated and combustion after ignition can be smoothed when the engine is configured so that the negative overlap quantity for example is varied based upon the EGR rate and the deviation ΔT is converged on a desired value, satisfactory compression self ignition combustion can be made. Besides, control over the compression self ignition combustion engine 1 is greatly simplified by controlling the deviation ΔT by the EGR rate and the security of control over an unexpected event can be greatly enhanced. At this time, as the data of the engine used for calculating the EGR rate are already known values at the time of the design of the engine 1 and injection quantity, air-fuel ratio, intake air temperature, exhaust gas temperature and in-cylinder pressure can be acquired from sensors normally used in the engine 1, no dedicated sensor is required to be provided and the cost of the engine 1 can be reduced.
Although a specific form of embodiment of the instant invention has been described above and illustrated in the accompanying drawings in order to be more clearly understood, the above description is made by way of example and not as a limitation to the scope of the instant invention. It is contemplated that various modifications apparent to one of ordinary skill in the art could be made without departing from the scope of the invention which is to be determined by the following claims.
Number | Date | Country | Kind |
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2006-181142 | Jun 2006 | JP | national |