Patent Application
20120055153
1. Field of the Invention
The present invention relates to an exhaust gas energy recovery device by which a part of the energy included in the exhaust gas discharged from an engine body such as a diesel engine for marine use or land generator use is recovered as power output.
2. Background of the Invention
As for the exhaust gas energy recovery device that recovers the exhaust gas energy included in the exhaust gas (combustion gas) as power output, Patent Reference 1 discloses, for instance, the turbocharger and the power turbine.
In the diesel engine provided with an exhaust gas energy recovery device in which, the exhaust gas (combustion gas) discharged from the engine body is mainly guided into a turbocharger and partly fed to a power turbine, a part of the exhaust gas energy is used for driving not the turbocharger but the power turbine. Therefore, according to an engine operating condition, the efficiency of the turbocharger decreases and the scavenging air pressure (charging air pressure) regarding the air into the engine body from the turbocharger becomes in shortage. As a result, the combustion efficiency of the engine decreases and the fuel consumption rate increases.
Contrary to the general aspect as described above, in recent years, it becomes possible that, even when a larger quantity of the exhaust gas discharged from the engine is fed to the power turbine in comparison with the conventional quantity, the turbocharger is able to supply the scavenging (charging) air of sufficient pressure to the engine body, owing to the improvement (as a general trend regarding the research and development of turbochargers) of the turbocharger performance in recent years. It is hereby noted that the turbocharger performance improvement has been brought, for instance, by the improved blade profile regarding the blades of the turbine, the improved impeller profile regarding the impellers of the compressor driven by the driving power generated by the turbine, the reduced clearance between the blades and the housing thereof, the reduced clearance between the impellers and the housing thereof, and the reduced driving resistance in the turbine and the compressor.
In this way, during the high load operation of the diesel engine such as disclosed in Patent Reference 1, when the opening of the by-pass valve is reduced (namely, the power output of the power turbine is reduced), then the exhaust gas that is fed toward the power turbine is supplied to the turbine (exhaust gas turbine) of the turbocharger. Thus, the driving power of the exhaust gas turbine as well as the speed of the exhaust gas turbine increases; and speed of the compressor driven by the exhaust gas turbine also increases.
Consequently, the pressure (the scavenging air pressure: the charging air pressure) of the compressed air supplied to the diesel engine from the compressor exceeds the allowable operation pressure limit inherent to the engine. Thereby, the scavenging air pressure is checked so as to not exceed the allowable limit; however, the check is performed only from an engine safety standpoint. Thus, in a case where the opening of the by-pass valve is reduced, the operation point of the engine goes away from the optimal operation point; and, the increased efficiency of the turbocharger does not translate into the increased thermal efficiency of the engine.
In view of the situations as described above, the present invention aims at providing an engine exhaust gas recovery device in which the flow rate of the exhaust gas that bypasses the turbocharger as well as the power turbine is adjusted so that the optimal scavenging (charging) air pressure is achieved and the optimal operation conditions are continuously achieved, by use of a control device in which the calculation formulae or the databases are incorporated and the optimal scavenging air pressure is computed based on the formulae or the databases to obtain the optimal engine performance (the fuel consumption rate), the formulae or the databases being related to the engine load or the engine speed.
Further, the present invention aims at providing the engine exhaust gas recovery device in which the optimal operating conditions of the engine are maintained even when the pressure of the compression air compressed in the cylinders of the engine decreases due to the aged deterioration.
In order to reach the goals as described above, the present invention discloses an engine exhaust gas energy recovery device comprising:
an exhaust gas turbocharger comprising a turbine and a compressor, the turbine being driven by the exhaust gas discharged from the engine, the compressor compression-transporting outside air to the engine body by driving the turbine;
a first exhaust pipe connecting the exhaust gas turbocharger to an exhaust manifold of the engine;
a power turbine being driven by the exhaust gas;
a second exhaust pipe connecting the exhaust manifold to the power turbine;
a gas flow-rate control valve being provided on the way of the second exhaust pipe so as to control the flow-rate of the exhaust gas fed to the power turbine;
a bypass pipe being connected to the second exhaust pipe which is provided on the upstream side of the gas-flow rate control valve so as to bypass the power turbine;
an exhaust gas bypass control valve which is provided on the way of the bypass pipe and which controls the flow-rate of the exhaust gas bypassing the power turbine;
an engine load detecting sensor which detects the load of the engine;
an engine speed detecting sensor which detects the speed of the engine;
a scavenging (charging) air pressure detecting sensor which detects the scavenging (charging) air pressure of the engine; and
a control device which comprises a database computing the optimal scavenging air pressure at which an fuel consumption rate of the engine becomes the optimum operating state on the basis of each value detected by the engine load detecting sensor and the engine speed detecting sensor,
wherein
the control device controls the exhaust gas bypass control valve so that the scavenging (charging) air pressure agrees with the computed optimal scavenging (charging) air pressure.
According to the engine exhaust gas energy recovery device of the present invention as described above, the control device computes the optimal scavenging air pressure, based on the values detected by the engine load detecting sensor and the engine speed detecting sensor, so that the engine is placed under the optimal operating condition; the control device controls the exhaust gas bypass control valve provided on the way of the bypass pipe so that the scavenging air pressure agrees with the computed optimal scavenging air pressure. In this way, the flow rate of the exhaust gas passing through the bypass pipe, and the optimal operating condition of the engine is continuously maintained. Thus, the fuel consumption rate can be restrained, and the running cost of the engine can be reduced; as a result, desirable effects on environmental improvement can be achieved.
A preferable embodiment of the above-described disclosure is the engine exhaust gas energy recovery device, wherein the control device further comprises a map which computes the optimal injection timing of a fuel injection device when the fuel consumption rate of the engine becomes the optimum operating state on the basis of each value detected by the engine load detecting sensor and the engine speed detecting sensor,
and controls a injection timing of the fuel injection device with reference to the map.
In response to the change of the load of the driven device coupled to the power turbine, the flow rate of the exhaust gas fed to the power turbine changes; thus, the flow rate of the exhaust gas passing through the exhaust gas turbocharger also changes. Hence, the scavenging air pressure of the engine changes. In order to compensate the change of scavenging air pressure, by adjusting the fuel injection timing in addition to the scavenging air pressure, the combustion condition (including the commencement of combustion) in the engine cylinder can be controlled so that the fuel consumption rate is always of the optimal, engine operating condition. Thus, the running cost of the engine can be reduced; as a result, desirable effects on environmental improvement can be achieved.
Another preferable embodiment of the above-described disclosure is the engine exhaust gas energy recovery device, wherein the control device further comprises a map which computes an closing timing of the exhaust valve when the fuel consumption rate of the engine becomes the optimum operating state on the basis of each value detected by the engine load detecting sensor and the engine speed detecting sensor,
and controls a closing timing of the exhaust valve with reference to the map.
The compression pressure of the air compressed inside of the cylinder is determined by the scavenging air pressure (the charging air pressure) and the exhaust (gas) valve closing timing. By increasing the scavenging air pressure and delaying the exhaust (gas) valve closing timing at the same time, the compression work during the upward movement of the piston can be reduced, and the fuel consumption rate can be reduced. Further, since the temperature of the air (or gas) inside the cylinder before the commencement of combustion (in a time point near to top dead center) can be reduced, the formation of NOx (nitrogen oxide) can be constrained. Thus, environmental load can be reduced.
Another preferable embodiment of the above-described disclosure is the engine exhaust gas energy recovery device,
the engine comprising either an operating oil accumulation room for an fuel-pump operating oil or a fuel accumulation room for a common-rail type fuel pump,
the control device further comprising a map which computes either an operating oil accumulation pressure or an fuel accumulation pressure when a fuel consumption rate of the engine becomes the optimum operating state on the basis of each value detected by the engine load detecting sensor and the engine speed detecting sensor,
wherein the control device controls either the operating oil accumulation pressure or the fuel accumulation pressure with reference to the map.
The operating oil accumulation pressure or the fuel oil accumulation pressure has an effect directly on the fuel injection pressure. Hence, by maintaining the fuel injection pressure at a high level, the fine atomization of the fuel during the injection duration in the cylinder can be achieved; the mixing of the atomized fuel and the air in the cylinder can be promoted so that the optimal combustion state is achieved; the fuel consumption rate can be reduced thanks to the improved thermal efficiency. Thus, the improved combustion brings about desirable results in cleaning the emissions of the engine.
Another preferable embodiment of the above-described disclosure is the engine exhaust gas energy recovery device, wherein the control device feedbacks an opening of the exhaust gas bypass control valve so that the opening of the exhaust gas bypass control valve agrees with a target bypass control valve opening which is calculated in a manner that a fuel consumption rate becomes the optimum operating state on the basis of a signal given by an exhaust gas bypass control valve opening detecting sensor which detects an opening of the exhaust gas bypass control valve.
In continuously detecting and controlling the opening of the exhaust gas bypass control valve, the flow rate of the exhaust gas fed to the exhaust gas turbocharger can be adjusted. Thus, the pressure of the scavenging air delivered to the engine body can be adjusted so that the deviation of the actual scavenging air pressure from the target (ordered) pressure value can be corrected, the deviation being attributable to aged deterioration. In this way, the engine operating condition can be prevented from deviating from the optimal operating condition.
Another preferable embodiment of the above-described disclosure is the engine exhaust gas energy recovery device, wherein the control device computes a cylinder compression pressure Pcomp and a maximum pressure Pmax on the basis of a cylinder pressure detected by a cylinder pressure detecting sensor;
the control device preliminarily sets, as a map, an optimal compression pressure Pcomp0 and an optimal maximum pressure Pmax0 so that a fuel consumption rate agrees with the optimal value with regard to the engine load and the engine speed;
the control device controls a fuel injection timing so that the computed maximum pressure Pmax agrees with the value obtained from the map,
and the control device controls the exhaust valve closing timing so that the computed compression pressure Pcomp agrees with the value obtained from the map.
One of the conditions under which the optimal engine operating state is maintained is the fuel combustion condition; namely, the fuel combustion state has an effect on the optimal engine operating condition. The fuel combustion state is influenced by the factors such as the engine speed, the engine scavenging air pressure (the engine charging air pressure), the used fuel properties (e.g. cetane number, viscosity, contained impurities) and so on. Thus, the combustion state changes in response to the fuel ignition timing and the fuel atomization level. Accordingly, the fuel combustion state can be judged by directly detecting the cylinder pressure (the compression pressure Pcomp and the maximum firing pressure Pmax).
Therefore, when the exhaust gas bypass control valve is controlled so that the engine is operated under the condition that the fuel consumption rate in the map corresponding to the engine load, the scavenging air pressure can be controlled; and, the engine operating condition under which the fuel consumption is optimal can be maintained.
Further, the present invention discloses an engine exhaust gas energy recovery device comprising:
an exhaust gas turbocharger comprising a turbine and a compressor, the turbine being driven by the exhaust gas discharged from the engine, the compressor compression-transporting outside air to the engine body by driving the turbine;
a first exhaust pipe connecting the exhaust gas turbocharger to an exhaust manifold of the engine;
a power turbine being driven by the exhaust gas;
a second exhaust pipe connecting the exhaust manifold to the power turbine;
a gas flow-rate control valve being provided on the way of the second exhaust pipe so as to control the flow-rate of the exhaust gas fed to the power turbine;
a bypass pipe being connected to the second exhaust pipe which is provided on the upstream side of the gas-flow rate control valve so as to bypass the power turbine;
an exhaust gas bypass control valve which is provided on the way of the bypass pipe and which controls the flow-rate of the exhaust gas bypassing the power turbine;
an engine load detecting sensor which detects the load of the engine;
an engine speed detecting sensor which detects the speed of the engine;
a cylinder pressure detecting sensor which detects a cylinder pressure of the engine;
and a control device which comprises a database computing the optimal compression pressure and the optimal cylinder maximum pressure at which a fuel consumption rate of the engine becomes the optimum operating state on the basis of each value detected by the engine load detecting sensor and the engine speed detecting sensor,
wherein the control device controls an exhaust valve closing timing so that a compression pressure agrees with the computed optimal compression pressure, and controls a fuel injection timing so that a cylinder maximum pressure agrees with the computed optimal cylinder maximum pressure.
When the actual opening of the exhaust gas bypass control valve deviates from the order value, for instance, because of aged deterioration, and the scavenging air pressure drops, or when the seat wear regarding the exhaust valve takes place, then the engine performance becomes worse due to the lowered level of the scavenging air pressure. Thus, when the exhaust gas bypass control valve is controlled so that the cylinder compression pressure Pcomp stays at a prescribed level, the scavenging air pressure can be controlled. In this way, the actual engine operation condition can be prevented from deviating from the optimal operating condition.
By use of the exhaust gas energy recovery device according to the present invention, the scavenging air pressure can be determined based on the value detected by each of the engine load detecting sensor, the speed detecting sensor and the scavenging air pressure detecting sensor, so that the engine performance (the fuel consumption rate) is optimal; thus, the scavenging air pressure can be controlled so as to continuously maintain the engine performance at the optimal condition.
Hereafter, the present invention will be described in detail with reference to the modes shown in the figures. However, the dimensions, materials, shape, the relative placement and so on of a component described in these modes shall not be construed as limiting the scope of the invention thereto, unless especially specific mention is made.
As shown in
The numeral 3 denotes an exhaust gas turbocharger that includes, but not limited to, an exhaust gas turbine 3a that is driven by the exhaust gas discharged from the exhaust gas manifold 7; and a compressor 3b that is driven by the exhaust gas turbine 3a via a shaft connecting the exhaust gas turbine 3a and the compressor 3b, the compressor pressurizing the outside air and supplying the pressurized air as scavenging (charging) air to the engine body 2.
The alphanumeric character L1 denotes a first exhaust pipe that connects the exhaust gas manifold 7 to the exhaust gas turbine 3a, the exhaust gas being fed to the exhaust gas turbine 3a through the first exhaust pipe.
The numeral 18 denotes an intercooler that cools the scavenging air (the charging air) that is the outside air compressed at the compressor 3b so as to increase the density of the scavenging air (the charging air).
The alphanumeric character K1 denotes a first charging air pipe that connects the compressor 3b to the intercooler 18; the alphanumeric character K2 denotes a second charging air pipe that connects the intercooler 18 to a charging air manifold 8 which is included in the engine body 2 or is a component of the engine, the scavenging (charging) air cooled at the intercooler being fed into the charging air manifold 8 of the engine body 2.
The numeral 4 denotes a power turbine that is driven by the diverging exhaust gas flow branched from the exhaust gas manifold 7 (toward the exhaust gas turbine), the power turbine 4 being the power source of the generator as described later. The numeral 9 denotes a heat exchanger that generates steam from water by heating the water by use of the heat included in the exhaust gas that has been used for driving the power turbine 4 as well as the exhaust gas turbine 3a of the turbocharger 3. The alphanumeric character L10 denotes a feed water pipe through which water is supplied to the heat exchanger 9. The heat exchanger 9 is a device in which the water supplied through the feed water pipe L10 is guided into the steam generating tubes (not shown) inside the heat exchanger 9 and the water is transformed into steam by use of the heat included in the exhaust gas.
The alphanumeric character L2 denotes a second exhaust pipe that connects the exhaust gas manifold 7 to the power turbine 4; the exhaust gas is fed to the power turbine 4 through the second exhaust pipe L2. The alphanumeric character L3 denotes a third exhaust pipe that connects the power turbine 4 to the heat exchanger 9; the exhaust gas discharged from the power turbine 4 is fed to the heat exchanger 9 through the third exhaust pipe L3. The alphanumeric character L4 denotes a fourth exhaust pipe that connects the exhaust gas turbine 3a of the exhaust gas turbocharger 3 to the third exhaust pipe L3; the exhaust gas discharged from exhaust gas turbine 3a of the turbocharger 3 is fed to the third exhaust pipe L3 so that the exhaust gas streams into the heat exchanger 9. The alphanumeric character L5 denotes a fifth pipe that connects the heat exchanger 9 to a steam turbine 5 that, together with the above-described power turbine 4, becomes a power source of a generator 11; the steam generated at the heat exchanger 9 is fed to the steam turbine 5 through the fifth pipe L5. The alphanumeric character L6 denotes a sixth pipe that returns the steam having driven the steam generator 5 to a condenser (not shown). The alphanumeric character L7 denotes a seventh exhaust pipe that discharges the exhaust gas having been used for generating the steam at the heat exchanger, toward a funnel (not shown), so that the exhaust gas is emitted outboard (or outside.)
In addition, the condensed water condensed in the condenser is fed to the heat exchanger 9 through the feed water pipe L10
The alphanumeric character V1 denotes an exhaust gas flow rate control valve that is provided on a part way of the second exhaust pipe L2; and, a controller (not shown) regulates the opening of the exhaust gas flow rate control valve V1. The alphanumeric character V2 denotes a steam flow rate control valve that is provided on a part way of the fifth pipe L5; and, the controller regulates the opening of steam flow rate control valve V2 so as to control the flow rate of the steam streaming into the steam turbine 5. The alphanumeric character V3 denotes an exhaust gas bypass valve provided on a part way of a first exhaust gas bypass pipe L8 that connects a location of the second exhaust pipe L2 on the upstream of the exhaust gas flow rate control valve V1 to a location on a part way of the third exhaust pipe L3. The alphanumeric character V4 denotes a steam bypass flow rate control valve provided on a part way of a second (steam) bypass pipe L11 that connects a location of the fifth pipe L5 on the upstream of the steam flow rate control valve V2 to a location on a partway of the sixth pipe L6.
The alphanumeric character V5 denotes an emergency control valve provided between the location of the second exhaust pipe L2 from which the first (exhaust gas) bypass pipe L8 is branched and the location of the exhaust gas flow rate control valve V1; the emergency control valve V5 is actuated in a case where the exhaust gas flow into the power turbine 4 is needed to be stopped for some reason.
Further, an orifice is provided between the exhaust gas bypass control valve V3 and the third exhaust pipe L3, namely, at the downstream side of the exhaust gas bypass control valve V3 on a part way of the first (exhaust gas) bypass pipe. The orifice 19 is arranged so that the exhaust gas flow rate through the orifice 19 in a case where the exhaust gas flow rate control valve V1 is fully closed (namely, in a case where the power turbine is at standstill) is the same as the exhaust gas flow rate through the exhaust gas flow rate control valve V1 in a case where the engine body 2 is under a high load condition (e.g. a normal service operation) and the power turbine 4 is at the full load operation condition (namely, the exhaust gas bypass control valve V3 is fully closed and the exhaust gas flow rate control valve V1 is fully opened).
Accordingly, when the engine body 2 is under a high load condition and the power turbine 4 is at standstill, the flow rate of the exhaust gas supplied to the exhaust gas turbocharger increases; as a result, the scavenging air pressure increases. The orifice 19 aims at preventing the scavenging air pressure from exceeding a predetermined pressure level so that the engine operation is not infected.
As described above, the orifice is arranged so that the exhaust gas flow rate (in a case where the exhaust gas flow rate control valve V1 is fully closed) through the orifice 19 is the same as the exhaust gas flow rate through the exhaust gas flow rate control valve V1 in a case where the engine body 2 is under a high load condition (e.g. a normal service operation) and the power turbine 4 is at the full load operation condition. On the other hand, when the exhaust gas flow rate control valve V1 is fully closed, the flow rate of the exhaust gas through the orifice is appropriately small so that the flow rate of the exhaust gas supplied toward the exhaust gas turbocharger 3 is not reduced less than a prescribed level; thus, the pressure of the scavenging air delivered from the exhaust gas turbocharger 3 is prevented from decreasing less than a predetermined level. Hence, the optimal operating conditions regarding the engine body 2 are achieved.
The rotating shaft (not shown) of the power turbine 4 is connected to the rotating shaft (not shown) of the steam turbine 5 via a reduction gear (not shown) and a coupling 10; further, the rotating shaft (not shown) of the steam turbine 5 is connected to the rotating shaft (not shown) of a generator 11 via a reduction gear (not shown) and a coupling 12.
In addition, the generator 11 is electrically connected to a distribution panel 14 provided inboard (in an engine room in this mode) apart from other components, via a resistor 13 for control use (a control resistor 13). Thus, the electric power produced by the generator 11 can be used as the inboard electric power source.
In the next place, with reference to
The control use database as described in
The lateral axis shows the compression pressure (in the cylinders); the compression pressure becomes greater in the rightward area in
The control factor may be the scavenging (charging) air pressure, instead of the cylinder compression pressure Pcomp; thereby, there is a relation that the greater scavenging pressure corresponds to the greater the cylinder compression pressure. Further, the control factor may be the exhaust valve closing timing regarding the engine body, instead of the cylinder compression pressure Pcomp; thereby, there is a relation that the earlier closing timing corresponds to the greater cylinder compression pressure. Thus, in a case where the scavenging (charging) air pressure or the exhaust valve closing timing is used as the control factor instead of the cylinder compression pressure Pcomp, the control maps like the map of
In
Further, in
Accordingly, the control target point at which the best fuel consumption rate is achieved exists in the area on the left side of the upper limit boundary line as well as on the lower right side along a contour line regarding a best fuel consumption rate; more concretely, the control target point exists at a point in the neighborhood of (as well as on left side of) the upper limit boundary line regarding the maximum firing pressure, the fuel consumption rate being minimal on the upper limit boundary line (whereby the contour curve passing the optimal point is close to the upper limit boundary line so as to come in contact with the upper limit boundary line).
On the basis of this control target point, the scavenging air pressure, the exhaust valve closing timing and the fuel injection timing are controlled so as to maintain the optimal engine operation condition.
Further, as the load becomes lower, the scavenging pressure becomes lower; accordingly, the cylinder compression pressure decreases. Thus, the fuel injection timing can be advanced. Therefore, the lower the load of the engine is, the lower and the more leftward the optimal operation point is shifted on the upper limit boundary (thick line) regarding the maximum firing pressure in
Thereby, the contour curves regarding the fuel consumption rate as well as the centers of the contour curves are shifted toward the lower leftward direction along the upper limit boundary (thick line) regarding the maximum firing pressure.
Based on
In
As shown in
In the next place, in the step S3, the controller 23 calculates the difference ΔPs between the scavenging air pressure Ps and the optimal scavenging air pressure Ps0 computed in step S2 so as to determine a to-be-adjusted value ΔA regarding the opening of the exhaust gas bypass control valve V3.
In the following step S4, based on the to-be-adjusted value ΔA and the current opening order value A′, an updated opening order value A regarding the opening of the exhaust gas bypass control valve V3 is determined. In the following step S5, the controller 23 outputs an order regarding the updated opening order value A to regulate the exhaust gas bypass control valve V3. Then, the step S5 is followed by the step S1, and the steps S1 to S5 are repeated. While the steps S1 to S5 are repeated, it is judged whether or not the scavenging air pressure Ps reaches an optimal pressure level Ps0 to maintain the optimal engine operating condition (e.g. minimal fuel consumption rate condition); thereby, when there is a difference between the scavenging air pressure Ps and the optimal pressure level Ps0, the controller works so that the difference is eliminated.
According to the first mode, the control device (the controller) 23 computes the optimal scavenging air pressure Ps0 based on the detected engine load and the detected engine speed so as to achieve the optimal engine operating condition; and, the controller controls the flow rate of the exhaust gas branched toward the power turbine by controlling the opening of the exhaust gas bypass control valve V3, so that the scavenging air pressure Ps agrees with the optimal scavenging air pressure. Thus, the engine is always placed under the optimal operating condition, and the fuel consumption rate is constrained. In this way, the running cost regarding the engine is reduced; consequently, a desirable effect on the reduction of the environmental loads can be achieved.
In the next place, based on
In
In the flow chart of
Further, in response to the engine load L and the engine speed Ne, a map regarding the optimal scavenging air pressure Ps0, a map regarding the fuel injection timing θinj, a map regarding the exhaust valve closing timing θevc, and a map regarding the operating-oil/fuel accumulation pressure are previously prepared in the controller 24. In the step S12, based on each map, the optimal value regarding each parameter (scavenging air pressure, fuel injection timing, exhaust valve closing timing, and operating-oil/fuel accumulation pressure) corresponding to the map is computed.
More concretely, the map previously prepared in the controller 24 means a map in which the coordinate axes relate to the compression pressure Pcomp and the injection timing θing with regard to various engine loads L and various engine speed Ne as is the case with
Further, the control factor (parameter) may be the scavenging (charging) air pressure, the exhaust valve closing timing, the operating-oil accumulating pressure (the accumulation pressure regarding the operating oil that drives the fuel pump in an electronic control engine), or the fuel accumulating pressure (the accumulation pressure regarding the fuel in the accumulator of the common rail type fuel injection device), instead of the compression pressure Pcomp of the lateral axis in
In the step S13, the controller 24 solves the difference ΔPs between the scavenging air pressure Ps and the optimal scavenging air pressure Ps0 that is computed in step S12; based on the difference ΔPs, the controller 24 determines a to-be-adjusted value ΔA regarding the opening of the exhaust gas bypass control valve V3. In the following step S14, based on the to-be-adjusted value ΔA and the current opening order value A′, an updated opening order value A regarding the opening of the exhaust gas bypass control valve V3 is determined. In the following step S15, the controller 24 outputs an order regarding the updated opening order value A to regulate the exhaust gas bypass control valve V3. The step S16 is followed by the step S17 where the difference between the newly detected opening value B regarding the exhaust gas bypass control valve V3, and the updated opening order value A is computed. When there remains a difference (an error) between the value B and the value A, the step S17 is returned to the step S14; based on the error, the correction value is computed, and the correction regarding the opening of the exhaust gas bypass control valve V3 is repeated.
When the current opening value B regarding the exhaust gas bypass control valve V3 agrees with the updated opening order value A, the process of the control returns to the step S11 so that the scavenging air pressure Ps remains at the optimal scavenging air pressure Ps0; thus, the control is repeated.
On the other hand, the step S12 is also followed by the step S18 where the optimal values as the order signals regarding the fuel injection timing θinj, the exhaust valve closing timing θevc and the operating-oil/fuel accumulation pressure are retrieved from the corresponding map; and, the signals are transmitted into an engine controller 25. Thus, the control regarding the engine body 2 is also performed.
As described above, instead of the compression pressure Pcomp (the lateral axis parameter) in
In the electronic control engine, the accumulation pressure regarding the operating oil that drives the fuel pump, the accumulation pressure regarding the fuel in the accumulator of the common rail type fuel injection device, or the exhaust valve closing timing has an effect directly on the fuel injection pressure. Therefore, the operating-oil accumulating pressure or the fuel injection pressure is maintained at a high level so that, in the cylinders, the fine fuel atomization and the prompt air-fuel mixing are promoted. In this way, the optimal combustion can be achieved; the fuel consumption rate can be reduced owing to the improved thermal efficiency, and the cleaned-up emissions owing to the improved combustion can be achieved. Thus, the desirable results such as the reduction of the environmental loads owing to the cleaned-up emissions can be obtained.
As shown in the steps 14 to 17 of
In the next place, based on
In the control block diagram of
In the flow chart of
Further, in response to the engine load L and the engine speed Ne, a map regarding the optimal scavenging air pressure Ps0, a map regarding the optimal compression pressure Pcomp0 and a map regarding the optimal maximum firing pressure Pmax are previously prepared in the controller 28. In the step S23, based on each map, the optimal value regarding each parameter (scavenging air pressure, compression pressure and maximum firing pressure) corresponding to the map is computed.
In the following step S24, the controller 28 calculates the difference ΔPs between the scavenging air pressure Ps and the optimal scavenging air pressure Ps0 that is computed in step S23 so as to determine a to-be-adjusted value ΔA regarding the opening of the exhaust gas bypass control valve V3, based on the difference ΔPs. In the following step S25, based on the to-be-adjusted value ΔA and the current opening order value A′, an updated opening order value A regarding the opening of the exhaust gas bypass control valve V3 is determined.
In the step S26, the controller 28 outputs an order regarding the updated opening order value A that is determined in the step S25, to regulate the exhaust gas bypass control valve V3. In the following step S27, the difference between the newly detected opening value B regarding the exhaust gas bypass control valve V3 and the updated opening order value A is computed. In step S28, it is judged whether or not there remains a difference (an error) between the value B and the value. When there remains the difference, the step S28 is followed by the step S30 where the correction value is computed based on the error. The step S30 is returned to the step S25 so that the corrections regarding the opening of the exhaust gas bypass control valve V3 is repeated. When the current opening value B regarding the exhaust gas bypass control valve V3 agrees with the updated opening order value A, the process of the control returns to the step S21 and the control flow is repeated so that the scavenging air pressure Ps remains at the optimal scavenging air pressure Ps0.
On the other hand, in the step S31, based on the difference ΔPcomp between the optimal compression pressure Pcomp0 computed in the step S23 and the cylinder compression pressure Pcomp, a correction value Δθevc regarding the exhaust valve closing timing θevc is determined.
In parallel to the determination of the correction value Δθevc regarding the exhaust valve closing timing in the above-described step S31, in the step S32, based on the optimal maximum firing pressure Pmax0 computed in the step S23 and the maximum firing pressure Pmax computed in the step S22, a correction value Δθinj regarding the fuel injection timing θinj is determined.
The step S31 is followed by the step S33 where the exhaust valve closing timing is determined based on the correction value Δθevc regarding the exhaust valve closing timing. The step S32 is followed by the step S34 where the fuel injection timing θinj is determined. In the step S35, the controller 28 issues the correction orders to correct the exhaust valve closing timing θev and the fuel injection timing θinj, toward the engine controller 25. In the step S36, the difference (error) between the target optimal maximum firing pressure Pmax0 and the detected maximum firing pressure Pmax, as well as, the difference (error) between the target optimal compression pressure Pcomp0 and the detected compression pressure Pcomp is computed. In the step S37, when there remains the error regarding the compression pressure as well as the maximum firing pressure, the correction value regarding the exhaust valve closing timing θevc as well as the fuel injection timing θinj is computed; the correction value regarding θevc is returned to the step S 33 so as to perform a feedback control, and the correction value regarding θinj is returned to the step S 35 so as to perform a feedback control. Thus, the correction processes are repeated.
In this way, the engine can be optimally operated under further precise optimal operation conditions (e.g. under a minimal fuel consumption condition); namely, further improved fuel consumption rate and further reduced environmental load can be achieved.
Further, the compression pressure is the pressure of the air compressed in the cylinder of the engine; the compression pressure is determined by two factors, namely, the scavenging air pressure and the exhaust valve closing timing. Thus, by enhancing the scavenging (charging) air pressure as well as by delaying the exhaust valve closing timing, the piston up-rising work during the compression stroke can be reduced. Thus, the effect on the reduction of the fuel consumption rate can be obtained. Further, the air temperature in the cylinder at the top dead center of the compression stroke can be reduced. As a result, the NOx (nitrogen oxide) formation through the fuel combustion in the cylinder can be restrained. Hence, the environmental load can be reduced.
Further, as described in relation to the steps from S24 to S27, the change of the scavenging air pressure is successively detected so that the scavenging air pressure becomes the optimal level and the opening regarding the exhaust gas bypass control valve is controlled; and, the flow rate of the exhaust gas fed toward the exhaust gas turbocharger is adjusted. Thus, it becomes possible to adjust the pressure of the scavenging air fed toward the engine body. Thus, the deviation of the actual opening level from the ordered opening value can be corrected, the deviation being caused by the aged deterioration or the like; and, the engine operating condition is prevented from deviating from the optimal operating condition.
Further, the fuel combustion condition becomes a factor for maintaining the optimal operating state of the engine; the fuel combustion condition is influenced by the factors such as the engine speed, the scavenging (charging) air pressure, and the fuel properties (e.g. cetane number, viscosity, contained impurities) and so on. Further, the combustion velocity changes in response to the fuel ignition timing and the fuel atomization level. Thus, as shown in the processes of the steps S31 to S37, when the cylinder pressures (the compression pressure Pcomp and the maximum firing pressure (the maximum combustion pressure) Pmax) are directly detected, as well as, when each of the detected compression pressure Pcomp and the detected maximum firing pressure Pmax agrees with the corresponding optimal value of the optimal point established in the corresponding map in response to the engine operating condition and the engine load so that the control factors such as the opening of the exhaust gas bypass control valve, the fuel injection timing and the exhaust valve closing timing are controlled, then the optimal engine operating condition (e.g. the minimal fuel consumption condition) can be maintained even when the cylinder pressure is reduced due to the change of fuel properties or the aged deterioration.
In addition, instead of the approach that the compression pressure Pcomp and the maximum firing pressure Pmax are controlled so as to achieve the optimal compression pressure and the optimal maximum firing pressure, there can be an approach that the ratio of Pmax/Pcomp is taken as a control factor; thereby, a map for realizing the optimal ratio of Pmax/Pcomp can be prepared in response to the engine load so that the opening of the exhaust gas bypass control valve is controlled.
In the next place, based on
In
In the following step S43, in reference to the maps (regarding the optimal compression pressure Pcomp0 and the optimal maximum firing pressure Pmax0 with regard to the engine load L and the engine speed Ne) previously prepared in the controller 29 the controller 28 calculate, the optimal value of each parameter is computed.
In the following step S44, the controller 29 calculates the difference
ΔPcomp between the cylinder compression pressure Pcomp and the optimal cylinder compression pressure Pcomp0; based on the difference ΔPcomp, a to-be-adjusted value ΔA regarding the opening of the exhaust gas bypass control valve V3 is determined. In the following step S45, based on the to-be-adjusted value ΔA and the current opening order value A′, an updated opening order value A regarding the opening of the exhaust gas bypass control valve V3 is determined. In the step S46, the controller 29 outputs an order regarding the updated opening order value A that is determined in the step S45, to regulate the exhaust gas bypass control valve V3.
In the step S47, the optimal compression pressure Pcomp0 and the newly detected compression pressure Pcomp are compared; and, the difference (error) between the pressures Pcomp0 and Pcomp is computed. In the following step S48, it is judged whether or not the opening A of the exhaust gas bypass control valve V3 is 0. When the opening A of the exhaust gas bypass control valve V3 is not 0, namely, when the valve is opened, the correction value regarding the opening of the exhaust gas bypass control valve V3 is computed in the step S49, based on the computed error. The result obtained in the step S49 is returned to the step S 45 so as to perform the control regarding the opening A of the exhaust gas bypass control valve V3. On the other hand, when the opening A is equal to 0 in the step S48, namely, when the valve is closed, the correction value Δθevc regarding the exhaust valve closing timing is computed in the step S50, based on the error regarding the cylinder compression pressure Pcomp. Further, in the step S51, the exhaust valve closing timing θevc is determined.
Further, in the step S52, based on the difference between the maximum firing pressure Pmax and the optimal maximum firing pressure Pmax0 that is computed in the step S43, the correction value Δθinj regarding the fuel injection timing θinj is determined. In the step S54, the controller 29 outputs the control orders regarding the exhaust valve closing timing θevc determined in the step S51 and the fuel injection timing θinj determined in the step S53, toward the engine controller 25.
In the step S55, the difference (error) between the target optimal maximum firing pressure Pmax0 and the detected maximum firing pressure Pmax, as well as, the difference (error) between the target optimal compression pressure Pcomp0 and the detected compression pressure Pcomp is computed.
When the computed result shows that there is a difference between the detected maximum firing pressure Pmax and the target optimal maximum firing pressure Pmax0, then the correction value Δθinj regarding the fuel injection timing θinj is computed in the step S56, based on the error (difference) regarding the maximum firing pressure Pmax. Consequently, the step S56 is returned to the step S 53 where the fuel injection timing θinj is again determined on the basis of the above-described error regarding the maximum firing pressure Pmax, and the signal regarding the determined fuel injection timing is outputted to the engine controller 25.
On the other hand, when there is the difference (error) between the detected compression pressure Pcomp and the target optimal compression pressure Pcomp0, the step S55 is returned to the step S50 where the correction value regarding the exhaust valve closing timing θevc is computed again, based on the error (difference) regarding the optimal compression pressure Pcomp0.
In a case where the actual opening level of the exhaust gas bypass control valve is deviated from the order value, so that the scavenging air pressure is reduced, for instance, owing to aged deterioration, or in a case where of the seat surfaces of the exhaust valve wear, the compression pressure is reduced and the engine performance is deteriorated. As shown in the processes of the steps S44 to S48, when the cylinder pressures are directly detected and the exhaust gas bypass control valve is regulated so that the compression pressure Pcomp stays at a predetermined level, the actual engine operating conditions can be prevented from deviating from the optimal operation conditions due to aged deterioration.
In this way, the engine can be optimally operated under further precise optimal operation conditions (e.g. under a minimal fuel consumption condition); namely, further improved fuel consumption rate and further reduced environmental load can be achieved.
Further, as described above, the cylinder pressures (the compression pressure Pcomp and the maximum firing pressure (the maximum combustion pressure) Pmax) are directly detected; the control factors such as the opening of the exhaust gas bypass control valve, the fuel injection timing and the exhaust valve closing timing are controlled so that the detected compression pressure Pcomp agrees with the target optimal value regarding the compression pressure and the detected maximum firing pressure Pmax agrees with the target optimal value regarding the maximum firing pressure, each optimal value being described in the corresponding map with an optimal engine operation point, each map being related to the engine load. In this way, the optimal operation (minimal fuel consumption operation) of the engine can be maintained, even when the properties of fuel change.
Further, even in a case where the exhaust gas bypass control valve V3 is fully closed and the opening is 0, the compression pressure Pcomp can be controlled toward the optimal compression pressure Pcomp0, by regulating the exhaust valve closing timing. Thus, the optimal control regarding the control parameter (factor) can be surly performed.
Incidentally, the present invention is not limited to the above-described modes; there can be the modified or varied modes of the present invention where necessary, within the scope of the technical concept.
Further, in the above-described modes of the present invention, the exhaust gas energy recovery device as a concrete example is provided with one turbocharger 3, one power turbine 4, and one steam turbine 5. However, the present invention is not limited to the configuration of the example. For instance, two turbochargers 3 may be provided; the power turbine may be provided with, for instance, a variable geometry nozzle device that is of a rotation speed control type and adjusts the flow rate of the exhaust gas streaming into the power turbine on the inlet side of the exhaust gas.
Further, according to the modes of the present invention, each of the power turbine 4 and the steam turbine can be operated, not in stages, by the manner in which the exhaust gas flow rate control valve V1, the steam flow rate control valve V2, the exhaust gas bypass control valve V3 and the steam bypass flow rate control valve V4 are precisely adjusted. Accordingly, the span of adjustable range regarding the power output of the generator 13 can be extended. Thus, the capacity of the control resistor 13 (a resistor for control use) can be restrained and the control resistor of a compact size can be adopted even in a case where the electric power demand inboard greatly changes. Thus, the cost effective control-resistor can be selected.
According to the present invention, the energy included in the exhaust gas emitted from the engine is recovered via the power turbine and the steam turbine, the power turbine and the steam turbine being connected in series via a coupling and a reduction gear, the power turbine and the steam turbine driving a generator so as to supply electricity inboard. The recovered energy can be used for driving a compressor so as to accumulate compressed air in air tanks, the air tank being able to be a power source for air actuators.
| Number | Date | Country | Kind |
|---|---|---|---|
| 2009-150571 | Jun 2009 | JP | national |
| Filing Document | Filing Date | Country | Kind | 371c Date |
|---|---|---|---|---|
| PCT/JP2010/060781 | 6/24/2010 | WO | 00 | 11/17/2011 |
