KJEMTRUP, Niels (Carinaparken 20, Brikerød, DK-3460, DK)
| CLAIMS :
1. A large turbocharged diesel engine comprising:
a plurality of cylinders that are each connected to an exhaust gas receiver via respective manifold pipes,
an upstream exhaust gas conduit for leading the exhaust gases from the exhaust gas receiver to the inlet of the turbine of the turbocharger,
a downstream exhaust gas conduit for leading the exhaust gases from the outlet of the turbine of the turbocharger to the atmosphere,
one or more exhaust gas heated boilers or heat exchangers for recovering heat energy from the exhaust gases,
characterized in that
at least one of said boilers or heat exchangers is disposed within said exhaust gas receiver.
2. An engine according to claim 1, further comprising a preheating boiler at the low-pressure side of the turbocharger and wherein the boiler that is disposed within the exhaust gas receiver is used to overheat steam produced by the boiler at the low-pressure side of the turbocharger.
3. An engine according to claim 1 or 2, further comprising a steam turbine driven by steam produced by the boiler or boilers .
4. An engine according to claim 3, wherein said power turbine drives an electric generator.
5. An engine according to any of claims 1 to 4, wherein the exhaust gas receiver houses a plurality of boilers.
6. An engine according to claim 5, wherein said plurality of boilers forms a multistage steam superheated steam production system including preheating and superheating boilers.
7. An engine according to claim 1, wherein said exhaust gas receiver is transversely divided into an exhaust gas collection channel and a heat exchanging channel
8. An engine according to claim 7, wherein the heat exchanging channel has a substantially ring shaped cross section in which substantially ring segment shaped boiler sections are received.
9. A large turbocharged diesel engine comprising:
a turbocharger with an exhaust gas driven turbine that is connected to a charging air compressor,
a first exhaust gas heated boiler on the high pressure side of the turbocharger,
a power turbine driven by a portion of the exhaust gases that is branched off from the high pressure side of the turbocharger.
10. An engine according to claim 9, further comprising a second exhaust gas heated boiler on the low-pressure side of the turbocharger .
11. An engine according to claim 10 or 11, wherein the first boiler is flowed through by all of the exhaust gasses and the exhaust gas portion for the power turbine is branched off downstreams of the first exhaust gas heated boiler.
12. An engine according to claim 10 or 11, wherein the first boiler is flowed through by the branched off portion of the exhaust gases only.
13. An engine according to any of claim 9 to 12, wherein the exhaust gases leaving the power turbine are reintroduced to the main exhaust gas stream at the low- pressure side of the turbocharger.
14. An engine according to any of claims 9 to 13, wherein said power turbine drives an electric generator.
15. An engine according to any of claims 9 to 14, wherein the second exhaust gas driven boiler acts as a preheating boiler and wherein the first exhaust gas heated boiler is used to overheat steam produced by the second exhaust gas heated boiler.
16. An engine according to claim 15, further comprising a steam turbine driven by superheated steam produced by the first and second exhaust gas heated boilers.
17. An engine according to claim 15, wherein said engine is operated to recover a substantial amount of energy in the first exhaust gas boiler for obtaining highly superheated steam and thereby improve the efficiency of the steam turbine .
18. An engine according to any of claims 9 to 17, wherein the scavenging air is humidified and cooled to a relatively high temperature so that scavenging air entering the cylinders has a high absolute water vapor content, thus increasing the energy content of the exhaust gases for subsequent recovery in the boiler and/or power turbine.
19. An engine according to any of claims 9 to 18, said engine comprising a plurality of cylinders that are each connected to an exhaust gas receiver via respective manifold pipes, and wherein the first exhaust gas heated boiler and/or the second exhaust^ gas heated boiler are disposed inside the exhaust gas receiver.
20. An engine according to any of claims 9 to 19, wherein the cooling capacity of the first and/or second boiler is selected to result in an exhaust gas temperature below ambient .
21. An engine according to any of claims 9 to 20, wherein a portion of the exhaust gas flow is recirculated.
22. An engine according to claim 21, wherein the exhaust gas portion to be recirculated is branched off from the exhaust gas stream downstreams of the first boiler.
23. A large charged two-stroke diesel engine comprising:
an exhaust gas driven turbine driving an electric generator, a charging air compressor driven by an electric drive motor, and
a heat exchanger on the high pressure side of the turbine for extracting heat from the exhaust gas.
24. A large charged two-stroke diesel engine according to claim 23 that does not comprise a turbo charger.
25. A large charged two-stroke diesel engine according to claim 23 or 24, wherein the heat exchanger is used to produce steam.
26. A large charged two-stroke diesel engine according to claim 25, further comprising means for accumulating a portion of the electrical energy produced by the generator, and means for supplying the stored electrical energy to the electric drive motor.
27. A large charged two-stroke diesel engine according to claim 26, further comprising means to control the distribution of the electric energy produced by the electric generator and the stored electric energy.
28. A large two-stroke diesel engine according to any of claims 25 to 27, further comprising a steam turbine driven by steam generated with the help of heat from the heat exchanger.
29. A large two-stroke diesel engine according to any of claims 23 to 28, wherein the heat exchanger is configured to lower the temperature of the exhaust gasses leaving the heat exchanger to an extend that results in the exhaust gases leaving the turbine downstreams of the heat exchanger to have a temperature below ambient.
30. A large charged two-stroke diesel engine comprising:
an exhaust gas driven turbine driving a hydraulic pump,
a charging air compressor driven by an hydraulic drive motor, and
a heat exchanger on the high pressure side of the turbine for extracting heat from the exhaust gas.
31. A charged internal combustion engine for use in a combined heating and power plant, said engine comprising:
an intake system for taking in air at ambient pressure and temperature,
the intake system including a compressor for delivering charging air with a pressure above ambient to the cylinders of the internal combustion engine,
a turbine driven by exhaust gas, and
a heat exchanger on the high pressure side of the turbine for extracting heat from the exhaust gas,
the heat exchanger and the turbine being configured to obtain an exhaust gas temperature at the low pressure side of the turbine below ambient.
32. A charged internal combustion engine according to claim 31, wherein an exhaust gas temperature below ambient is obtained by a large capacity of the heat exchanger for increasing the temperature fall of the exhaust gas passing though the heat exchanger and a low effective turbine area for increasing the fall in the temperature of the exhaust gas during its expansion in the turbine.
33. A charged internal combustion engine according to claim 31, wherein the temperature of the exhaust gases leaving the cylinders is between 400 and 500 °C, the temperature of the exhaust gases leaving the exhaust gas heated boiler is below 110 °C, and the pressure of the exhaust gas leaving the boiler is above 2 bar.
34. A charged internal combustion engine according to any of claims 31 to 34, wherein the turbine and the compressor are connected by a shaft to form a turbocharger .
35. A charged internal combustion engine according to claim 34, further comprising an auxiliary blower for assisting the turbine to deliver charging air to the cylinders of the combustion engine, preferably also when the engine is operating at its maximum continuous rating.
36. A charged internal combustion engine according to claim 34 or 35, wherein the engine further comprises a power turbine driven by exhaust gas branched off from the exhaust gas flow to the turbocharger turbine downstream of the boiler.
37. A charged internal combustion engine according to any of claims 31 to 36, further comprising a steam turbine powered by steam produced with heat extracted from the exhaust gas by the heat exchanger.
38. A charged internal combustion engine according to any of claims 31 to 37, further comprising a charging air humidification unit at the high pressure side of the compressor .
39. A charged internal combustion engine according to any of claims 31 to 38, wherein the pressure of the exhaust gases leaving the turbine is equal to or slightly above ambient pressure.
40. A charged internal combustion engine according to any of claims 31 to 39, wherein the temperature of the exhaust gases leaving the turbine is below ambient at least when the engine is running at its maximum continuous rating.
41. A charged internal combustion engine according to any of claims 31 to 40, wherein the temperature of the exhaust gases leaving the turbine is between -5 and -40 "C for at least when the engine is running at its maximum continuous rating.
42. A charged internal combustion engine according to any of claims 31 to 41, further comprising a further turbine, used instead of or in combination with the turbine, to change the effective turbine area for operation of the engine with an exhaust gas temperature at the low pressure side of the turbine or turbines above ambient.
43. A charged internal combustion engine according to any of claims 31 to 41, wherein said turbine is of a type with a variable effective turbine area for operation of the engine with an exhaust gas temperature at different temperatures .
44. A charged combustion engine comprising an intake system for taking in air at ambient pressure and temperature, the intake system including a compressor for delivering charging air with a pressure above ambient to the cylinders of the internal combustion engine, a first turbine with a given effective turbine area driven by exhaust gas, a second turbine with a given effective turbine area driven by exhaust gas and a heat exchanger on the high pressure side of the turbine for extracting heat from the exhaust gas, and means for selectively using either or both turbines in order to operate the engine with different exhaust gas temperatures at the low pressure side of the turbine.
45. A charged combustion engine comprising an intake system for taking in air at ambient pressure and temperature, the intake system including a compressor for delivering charging air with a pressure above ambient to the cylinders of the internal combustion engine, a turbine with a variable effective turbine area driven by exhaust gas and a heat exchanger on the high pressure side of the turbine for extracting heat from the exhaust gas.
46. A method of operating a charged combustion engine, said charged combustion engine comprising an intake system for taking in air at ambient pressure and temperature, the intake system including a compressor for delivering charging air with a pressure above ambient to the cylinders of the internal combustion engine, a first turbine with a given effective turbine area driven by exhaust gas, a second turbine with a given effective turbine area driven by exhaust gas and a heat exchanger on the high pressure side of the turbine for extracting heat from the exhaust gas, comprising the steps of selectively using turbines to obtain different exhaust gas temperatures at the low pressure side of the turbine or turbines. |
A LARGE TURBOCHARGED DIESEL ENGINE WITH ENERGY RECOVERY ARRANGEMENT
FIELD OF THE INVENTION
The present invention relates to a large turbocharged diesel engine with one or more exhaust gas heated boilers, and in particular to large turbocharged diesel engines provided with a power turbine driven by exhaust gas branched off upstreams of the turbocharger turbine.
BACKGROUND OF THE INVENTION
EP 0 434 419 discloses a large two-stroke turbocharged diesel engine in which heat energy from the exhaust gases is recovered by a combination of a boiler on the low- pressure side of the turbocharger and a boiler on the high pressure side of the turbocharger. At lower engine loads, recovery of heat energy from the exhaust gases before leading them to the turbocharger is reduced by leading a proportion of the exhaust gases directly into the turbocharger by-passing the upstream boiler. However, by placing a boiler between the exhaust gas receiver and the turbine of the turbocharger, the overall structure becomes relatively voluminous and complicated. Further, the increased length of the flow path between the exhaust valves and the turbocharger decreases the responsiveness of the turbocharger to acceleration events. Further, this engine recovers only heat, whilst there are no provisions to convert the recovered energy to a more useful form of energy, such as rotary power or electricity.
DISCLOSURE OF THE INVENTION
On this background, it is an object of the present invention to provide a turbocharged diesel engine of the kind referred to initially, which is more compact and less complicated to construct. This object is achieved in accordance with claim 1 by providing a turbocharged diesel engine of said kind comprising a plurality of cylinders that are each connected to an exhaust gas receiver via respective manifold pipes, an upstream exhaust gas conduit for leading the exhaust gases from the exhaust gas receiver to the inlet of the turbine of the turbocharger, a downstream exhaust gas conduit for leading the exhaust gases from the outlet of the turbine of the turbocharger to the atmosphere, one or more exhaust gas heated boilers or heat exchangers for recovering heat energy from the exhaust gases, wherein at least one of said boilers or heat exchangers is disposed within said exhaust gas receiver.
By placing one of the boilers physically inside the exhaust gas receiver a component of the system is effectively no longer requiring any space in the cramped area at the top of a large turbocharged diesel engine. This measure therefore creates more space around the engine and it also reduces the amount of piping. Further, a housing for the boiler is saved since the housing of the exhaust gas receiver has now two functions: to provide a cavity for receiving and collecting the exhaust gases from the individual cylinders, and to provide a cavity for housing a boiler. Another advantage is that the pressure drop across the boiler can be allowed to be three times higher than in the convention constructions without a reduction in engine performance. An increased pressure drop allows in turn increased gas speed, which allows a significant reduction
of the heat exchange surface (all other parameters being equal), thus resulting in a much smaller boiler.
The large turbocharged diesel engine may further comprise a preheating/evaporating boiler at the low-pressure side of the turbocharger . In this case the boiler that is disposed within the exhaust gas receiver is used to overheat steam produced by the boiler at the low-pressure side of the turbocharger. Thereby, the quality of the steam is improved, in particular in view of using the superheated steam in a steam turbine.
The large turbocharged diesel engine may further comprise a steam turbine driven by steam produced by the boiler or boilers. Thereby, the energy to recover from the exhaust gases is converted to the more useful power form of energy. The power turbine may drive an electric generator for converting the rotary power into electricity.
The exhaust gas receiver may house a plurality of boilers, or several stages of a single boiler. Thus, the energy in the exhaust gas can be transferred more effectively to the steam.
The plurality of boilers may form a multistage steam superheated steam production system including preheating/evaporating and superheating/evaporating boilers .
It is another object of the present invention to provide a large turbocharged diesel engine with improved energy recovery from the exhaust gas. This object is achieved in accordance with claim 9 by a large turbocharged diesel engine comprising a turbocharger with an exhaust gas driven
turbine that is connected to a charging air compressor, a first exhaust gas heated boiler on the high pressure side of the turbocharger, a second exhaust gas heated boiler on the low-pressure side of the turbocharger, and a power turbine driven by a portion of the exhaust gases that is branched off from the high pressure side of the turbocharger.
By using a combination of a boiler on the high pressure side of the turbocharger turbine and branching off a part of the exhaust gases flow from the high pressure side of the turbocharger turbine, the overall amount of energy that can be recovered from the exhaust gases is improved in amount, in particular in a greater variation of operating conditions, because the system can adapt to a greater production of heat via the boilers as opposed to a greater production of rotary energy via the power turbine. Thus, the system can equally be of help to improve the overall fuel efficiency in stationary power plants as in propulsion systems for oceangoing vessels.
On one hand, the first boiler may be flowed through by all of the exhaust gasses and the exhaust gas portion for the power turbine is branched off downstreams of the first exhaust gas heated boiler. In this way the overall amount of energy that can be recovered is maximized.
On the other hand, the first boiler may be flowed through by the branched off portion of the exhaust gases only, leaving the heat balance of the turbocharger unaffected and in this way the responsiveness of the turbocharger turbine is guaranteed during acceleration events.
The exhaust gases leaving the power turbine may be reintroduced to the main exhaust gas stream at the low- pressure side of the turbocharger . In this way it can be ensured that all the exhaust gases receive the proper after-treatment in e.g. an SCR reactor and/or a silencer.
Preferably, the power turbine drives an electric generator. Thus, the recovered energy can be used to produce a very- attractive and flexible form of energy.
It is another object of the present invention to provide a large two-stroke diesel engine that is flexible to operate and has a good energy recovery from the exhaust gas .
This object is achieved in accordance with claim 23 by providing a large charged two-stroke diesel engine comprising an exhaust gas driven turbine driving an electric generator, a charging air compressor driven by an electric drive motor, and a heat exchanger on the high pressure side of the turbocharger for extracting heat from the exhaust gas.
Due to the absence of a shaft connecting the turbine to the compressor, the operating conditions of the engine can be controlled with a greater degree of freedom, whilst the use of a heat exchanger at the high pressure side of the turbine ensures a good recovery of the energy contained in the exhaust gas.
Preferably, the engine does not comprise a turbo charger.
The heat exchanger can be used to produce steam.
The engine may further comprise means for accumulating a portion of the electrical energy produced by the generator, and means for supplying the stored electrical energy to the electric drive motor.
Preferably, the engine comprises means to control the distribution of the electric energy produced by the electric generator and the stored electric energy.
The engine may further comprise a steam turbine driven by steam generated with the help of heat from the heat exchanger.
Preferably, the heat exchanger is configured to lower the temperature of the exhaust gasses leaving the heat exchanger to an extend that results in the exhaust gases leaving the turbine downstreams of the heat exchanger to have a temperature below ambient.
It is a further object of the present invention to provide a combustion engine that can be used in a combined power and heating plant with very high fuel efficiency.
This object is achieved in accordance with claim 23 by providing a charged internal combustion engine for use in a combined heating and power plant, said engine comprising an intake system for taking in air at ambient pressure and temperature, the intake system including a compressor for delivering charging air with a pressure above ambient to the cylinders of the internal combustion engine, a turbine driven by exhaust gas, and a heat exchanger on the high pressure side of the turbine for extracting heat from the exhaust gas, the heat exchanger and the turbine being
configured to obtain an exhaust gas temperature at the low pressure side of the turbine below ambient.
By extracting a large amount of energy in the exhaust gas heated boiler at the high pressure side of the turbine and by using a turbine with a relatively small effective turbine area, the expansion of already relatively cool exhaust gases in the turbine will result in temperatures of the exhaust gases at the low pressure side of the turbine that are well below ambient. Thus, the combustion engine itself is turned into a heat pump that extracts low grade energy from the environment and turns it into high grade energy. An overall fuel efficiency well above 100% can be obtained, thereby rendering the very economical and environmental friendly. The temperatures of the exhaust gases can be as low as -40 0 C. The exhaust gases leaving the chimney on the power plant using such an engine may therefore contain snow or similar ice crystals.
An exhaust gas temperature below ambient is preferably- obtained by a large capacity of the heat exchanger for increasing the temperature fall of the exhaust gas passing though the heat exchanger and by a low effective turbine area for increasing the fall in the temperature of the exhaust gas during its expansion in the turbine.
Preferably, the temperature of the exhaust gases leaving the cylinders is between 400 and 500 °C, the temperature of the exhaust gases leaving the exhaust gas heated boiler is below 110 °C, and the pressure of the exhaust gas leaving the boiler is above 2 bar.
The turbine and the compressor can be connected by a shaft to form a turbocharger . In this case, the engine may
further comprise a power turbine driven by exhaust gas branched off from the exhaust gas flow to the turbocharger turbine downstream of the boiler.
The engine may comprise a steam turbine powered by steam produced with heat extracted from the exhaust gas by the heat exchanger.
Preferably, the engine may further comprise a charging air humidification unit at the high pressure side of the compressor.
The pressure of the exhaust gases leaving the turbine is preferably equal to or slightly above ambient pressure.
The temperature of the exhaust gases leaving the turbine is preferably between -5 and -40 °C.
The temperature of the exhaust gases leaving the turbine is below ambient at least when the engine is running at its maximum continuous rating.
The temperature of the exhaust gases leaving the turbine may range -5 and -40 °C, at least when the engine is running at its maximum continuous rating.
According to a further aspect of the invention there is provided a charged combustion engine comprising an intake system for taking in air at ambient pressure and temperature, the intake system including a compressor for delivering charging air with a pressure above ambient to the cylinders of the internal combustion engine, a first turbine with a given effective turbine area driven by- exhaust gas, a second turbine with a given effective
turbine area driven by exhaust gas and a heat exchanger on the high pressure side of the turbine for extracting heat from the exhaust gas, and means for selectively using either or both turbines in order to operate the engine with different exhaust gas temperatures at the low pressure side of the turbine.
According to another aspect of the invention there is provided a charged combustion engine comprising an intake system for taking in air at ambient pressure and temperature, the intake system including a compressor for delivering charging air with a pressure above ambient to the cylinders of the internal combustion engine, a turbine with a variable effective turbine area driven by exhaust gas and a heat exchanger on the high pressure side of the turbine for extracting heat from the exhaust gas.
According to yet another aspect of the invention there is provided a method of operating a charged combustion engine, said charged combustion engine comprising an intake system for taking in air at ambient pressure and temperature, the intake system including a compressor for delivering charging air with a pressure above ambient to the cylinders of the internal combustion engine, a first turbine with a given effective turbine area driven by exhaust gas, a second turbine with a given effective turbine area driven by exhaust gas and a heat exchanger on the high pressure side of the turbine for extracting heat from the exhaust gas, comprising the steps of selectively using turbines to obtain different exhaust gas temperatures at the low pressure side of the turbine or turbines.
Further objects, features, advantages and properties of the charged combustion engines according to the invention will become apparent from the detailed description.
BRIEF DESCRIPTION OF THE DRAWINGS
In the following detailed portion of the present description, the invention will be explained in more detail with reference to the exemplary embodiments shown in the drawings, in which:
Fig. 1 is a partial side view of a large turbocharged diesel engine according to a first embodiment of the invention,
Fig. 2 is a longitudinal sectional view through the engine of Fig. 1,
Fig. 3 shows schematically a large turbocharged diesel engine with heat energy recovery provisions according to a second embodiment of the invention,
Fig. 3a is a graph illustrating the operating parameters of the engine of Fig. 3,
Fig. 4 shows schematically a large turbocharged diesel engine with heat energy recovery provisions according to a third embodiment of the invention,
Fig. 4a is a graph illustrating the operating parameters of the engine of Fig. 4,
Fig. 5 shows schematically a large turbocharged diesel engine with heat energy recovery provisions according to a fourth embodiment of the invention,
Fig. 5 a is a graph illustrating the operating parameters of the engine of Fig. 5,
Fig. 6 shows another embodiment of the invention, in which the engine is operated as a heat pump,
Fig. 7 shows a further embodiment of the invention that does not use a turbocharger, but which is instead provided with a turbine and a blower that are electrically connected, and
Fig. 8 shows another embodiment of the invention that uses exhaust gas recirculation.
DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS
In the following detailed description, the large turbocharged diesel engine according to the invention in the form of a large two-stroke diesel engine will be described by the preferred embodiments.
The construction and operation of large turbocharged diesel engines, such as large two-stroke diesel engines of the cross-head type are as such well-known and should not require further explanation in the present context. Further details regarding the operation of the charging and exhaust gas systems are provided below.
Fig. 1 shows a first embodiment of an upper area of a large two-stroke diesel engine 1 according to the invention. This engine is provided with a plurality of cylinders arranged besides one another in line. Each cylinder is provided with an exhaust valve (not shown) associated with their cylinder cover. The exhaust channels can be opened and closed by the exhaust valve. Manifold pipes connect to the respective exhaust channels to an exhaust gas receiver 3. The exhaust gas receiver 3 is disposed in parallel to the row of cylinders. The manifold pipes 40 open into the exhaust gas receiver 3 and an exhaust conduit leads from the exhaust gas receiver to the turbine of a turbocharger . In engines with a very large number of cylinders (for example 10 or more cylinders) the exhaust gas receiver may be longitudinally divided into two or more parts (not shown) .
The exhaust gas receiver 3 has in this embodiment a cylindrical housing 42 that is, as shown in Fig. 2, provided at its ends with removable covers 44. The cylindrical housing 42 contains a heat exchanger 23 through which the exhaust gases can flow for producing superheated steam. The heat exchanger 23 therefore acts as a boiler. The cylindrical housing 42 also contains a collection duct 46 in which the manifold pipes 40 release the exhaust gases .
The cylindrical housing 42 of the exhaust gas receiver is, as shown in Fig. 2, divided in two heat exchanger portions 50a and 50b and collection duct portions 46a and 46b that are juxtaposed to a central outlet chamber 52 from which the exhaust gases depart via an exhaust conduit. Therefore, the construction of the exhaust gas receiver 3 is symmetrical with respect to its central radial plane.
Both sections 50a, 50b of the heat exchanger arrangement consist of several sequentially arranged and as such well- known heat exchanging elements that are separated by spacers 49. Each section 50a, 50b includes two heat exchanging elements 57a, 58a, 57b, 58b each comprising a large number of tubes that extend in the direction of the gas flow that is indicated by an arrow drawn as a continuous line, parallel with the longitudinal axis of the cylindrical housing. The flow direction in the respective heat exchanger portions 50a and 50b is opposite and towards one another.
The cross-sectional outline of the excentrically disposed heat exchanger elements 57a, 58a, 57b, 58b is in the shape of a ring segment that abuts with the inner circumference of the cylindrical housing 42. The ring segments may be
divided into sub segments to improve the ease of assembly (not shown) .
The cylindrical housing 42 of the exhaust gas receiver 3 is provided with a partition wall 63 separating the heat exchanger elements from the rest of the cross-section of the interior of the exhaust gas receiver 3, thereby dividing the cross-section of the interior of the exhaust gas receiver into a channel for receiving the heat exchanger elements and a channel for collecting and guiding the exhaust gases towards the channel with the heat exchanger elements 57a, 58a, 57b, 58b.
In the latter channel (into which the manifold pipes 40 open) guides the exhaust gases in the direction of the arrow that are indicated by interrupted lines.
The heating elements are retractable into the channel for receiving the heating elements. The longitudinally outer heating elements are separated from the inner heating elements by spacers 49. The complete assembly is kept in place by locking plates 66.
The collection channels 46a, 46b have a funnel shaped cross- sectional shape that opens up in a radially outward direction. The manifold pipes 40 are arranged such that they blow the exhaust gases into the respective collection channels 46a, 46b.
The collection channels 46a, 46b are separated from the central outlet chamber 52 by sidewalls 69 that are connected to frontal ends of the collection channels. The collection channels 46a, 46b are open at their opposite end at some distance from the removable covers 44. Thereby,
reversal chambers 71a, 71b are formed in the area of the ends of the exhaust gas receiver housing 42. The reversal chambers 71a, 71b connect the collection channels 46a, 46b to the channels in which the heat exchanger sections are received. Thus, flow paths are formed on both sides of the outlet chamber 52 that connect the receiving channels 46a, 46b via the channels containing the heat exchanger elements with the outlet chamber. The exhaust gases that leave the manifold pipes 40 in the respective collection channels 46a, 46b flow, as shown in Fig. 2 by the arrows indicated with interrupted lines, to the reversal chambers 71a, 71b and therefrom, as indicated by the arrows indicated by continuous lines, through the respective heat exchanger elements 57a, 58a, 57b, 58b towards the outlet chamber 52.
Thus, the housing 42 of the exhaust gas receiver 3 serves to contain both an exhaust gas receiving cavity and a boiler for recovering heat energy from the exhaust gases. By including the boiler inside the exhaust gas receiver the space required for an exhaust boiler and the housing for an exhaust gas boiler can be saved.
Fig. 3 shows a second embodiment of a large turbocharged two-stroke diesel engine of the crosshead type 1 with its intake and exhaust systems. The engine 1 has a charging air receiver 2 and an exhaust gas receiver 3. The exhaust gas receiver 3 can be of the type described in the first embodiment, but this need not be the case. The engine is provided with exhaust valves (one or more per cylinder) which are not shown. The engine 1 may e.g. be used as the main engine in an ocean going vessel or as a stationary engine for operating a generator in a power station. The total output of the engine may, for example, range from 5,000 to 110,000 kW, but the invention may also be used in
four-stroke diesel engines with an output of, for example, 1,000 JcW.
The charging air is passed from the charging air receiver 2 to the scavenging air ports (not shown) of the individual cylinders. When the exhaust valve 4 is opened, the exhaust gas flows through manifold pipes into the exhaust receiver 3 and from there onwards through a first exhaust conduit 5 to a turbine 6 of a turbocharger, from which the exhaust gas flows away through a second exhaust conduit 7. Through a shaft 8, the turbine 6 drives a compressor 9 supplied via an air inlet 10. The compressor 9 delivers pressurized charging air to a charging air conduit 11 leading to the charging air receiver 2.
The intake air in the conduit 11 passes through an intercooler 12 for cooling the scavenging air - that leaves the compressor 9 at approximately 200 "C - to a temperature of approximately 36 "C.
The cooled scavenging air passes via an auxiliary blower 16 driven by an electric motor 17 that pressurizes the scavenging air flow (often only in low or partial load conditions) to the scavenging air receiver 2. At higher loads the amount of scavenging air delivered by the turbocharger compressor 9 is sufficient for operating the engine and the auxiliary blower 16 is stopped. In this state the auxiliary blower 16 is bypassed via conduit 15.
A first boiler 23, preferably in the form of a heat exchanger, e.g. of the pipe or fin type, is disposed in the first exhaust conduit 5, i.e. upstream of the turbine 6 using heat energy in the exhaust gases to produce steam. The exhaust gases have when entering the exhaust gas
receiver 3 a temperature of about 455°C and the temperature at the entry of the first boiler 23 is only insignificantly lower. The first boiler 23 can be an integral part of the exhaust gas receiver 3, as shown and explained with reference to the first embodiment above.
Downstreams of the boiler 23 the exhaust conduit is branched off, whereby the major portion of the exhaust gases continue via exhaust conduit 5 towards the turbine 6 and a minor portion of the exhaust gases flow via a conduit
30 towards a power turbine 31. The additional power turbine
31 drives an electric generator 32.
A surplus of energy in the exhaust gas stream is thus converted to electric power, i.e. energy with a high exergy. The amount of exhaust gas that is branched off to the power turbine 31 can be regulated by a variable flow regulator (not shown) in conduit 30. The exhaust gases that leave the power turbine 31 are led to the second exhaust conduit 7 and there reintroduced into the main exhaust gas stream.
The second exhaust conduit 7 directs the exhaust gases to the inlet of a second boiler 20 that comprises a heat exchanger, e.g. of the tube- or fin type. A third exhaust conduit 21 leads the charging air from the outlet of the second boiler 20 to the atmosphere. Before arriving at the atmosphere the exhaust gases may be purified in an SCR reactor (not shown) for reducing e.g. NOx levels and pass through a silencer (not shown) for reducing noise pollution.
The second boiler 20 uses the heat in the exhaust gas stream to produce steam under pressure. At this stage the
exhaust gas temperature is lower than when leaving the cylinders, typically the temperature at the outlet of the turbocharger turbine 6 is in the range between 250 and 300 0 C.
A conduit 22 leads the steam produced by the second boiler 20 to the inlet of the first boiler 23. The first boiler is heated with exhaust gases that have a temperature of about 450 0 C and thus is a very effective medium to evaporate/superheat the water/steam coming into the first boiler 23.
The superheated steam is directed via conduit 34 to a steam turbine 37 that converts the energy in the steam to rotary mechanical power. The steam turbine 37 drives an electric generator 35 for producing electric energy, that can be used aboard an ocean going vessel, e.g. for powering cooling equipment, or be added to the electricity produced in a stationary power station. Although not shown for this nor any of the other embodiments, it is understood that the boilers and the steam turbine are part of a steam circuit including a condenser, cooler and other components well known in the field of steam power.
An example of the operating parameters of the second embodiment with a MAN B&W® 12K98ME engine is provided in table 1 below. This is an engine with 12 cylinders with a cylinder bore of 98 cm. It is noted that the turbocharger compressor plus possible auxiliary blower requires a power input of approx. 25000 kW. This power is extracted from the exhaust gas and/or supplied by the auxiliary blowers.
Based on energy equations it is possible to determine an optimum with regard to power extraction from the complete
system. This will ultimately depend on circumstances, such as the type of boiler, the type of steam turbine and the conditions of use of the large two-stroke diesel engine. In oceangoing vessels the main focus will be on providing rotary power, whereas an application in a stationary power plant focuses equally on heat production (for district heating) as on electricity production.
The system can be operated at various operation points, with a variable amount of power that is taken out of the exhaust gas by the first boiler 23 and the power turbine 31.
The power extracted in the first boiler 23 upstreams of the turbocharger turbine 6 will reduce the power available for the turbocharger turbine 6 and the power turbine 31, whereas power extracted in the second boiler 20 will have no influence on the turbocharger and power turbine power.
In the example in table 1 an amount of energy of 10.000 kW is extracted in the first boiler 23 for delivery to the steam turbine 37 (this amount has been arbitrarily chosen for this example and other amounts could be selected as illustrated in Fig. 3A.
Fig. 3A is a graph showing the calculation results for different values for the amount of power extracted at the first boiler 23. The graph indicates the power of the various components as a percentage of the engine shaft power, to illustrate the fact that the invention can be applied to various sizes of engines. On the graph it can be seen that the power, which can be extracted from the power turbine is decreased when the power extracted in the first boiler 23 upstreams of the turbocharger turbine 6 is
increased. The optimum operation position can be determined in accordance with the type of power that is required (heat or rotary power/electricity) .
If both heat and rotary power is required, such as in stationary power plants that provide both electricity and heat, the optimum operation point is most likely more towards the maximum power extraction via the first boiler 23. This operation point will require that the auxiliary blower 16 is operated even at full load conditions.
On in oceangoing vessel the main energy that is required is propulsion power, i.e. rotary power to drive the propeller (not shown) . The amount of heat that is required around the vessel is typically relatively low, whereas the amount of electricity that is required varies with the type of vessel. In bulk carriers the amount of electricity required is relatively low.
Container ships with cargo that needs to be cooled or liquid natural gas carriers require a substantial amount of electrical power. In those situations it is from an overall energy efficiency viewpoint of advantage to operate with 5.000 to 10.000 kW that is extracted by the first boiler.
Fig. 4 shows a third embodiment of the invention. This embodiment corresponds substantially to the second embodiment, except that the scavenge air cooler 12a is of a different type. The scavenge air cooler is a scrubber in which large amounts of water are injected and evaporated. The injected water is preferably relatively warm, e.g. by heating up ocean water (when the engine is installed in an ocean going vessel) or river water (when the engine is installed in a stationary power plant near a river) with
waste heat from the (water) cooling system (not shown) of the engine 1. The scrubber 12a is operated to result in air leaving the scrubber outlet that has a temperature of approximately 70 "C and a substantially 100% relative humidity. The absolute humidity of this scavenging air is about five times higher than the scavenging air leaving the intercooler 12 of the second embodiment. Therefore, the amount of energy contained in the scavenging air and also in the exhaust gases is significantly increased. Thus, there is more energy available for extraction from the exhaust gases by the boilers 20,23 and the power turbine 31.
An example of the operating parameters of the third embodiment with a MAN B&W® 12K98ME engine is shown in table
1.
In order to produce this scavenging air condition the turbocharger compressor plus possible auxiliary blower require a power input of approx. 25.000 kW and further injection of approx 7,5 kg/s water evaporating in the compressor outlet air must be realized.
This (25000 kW) power has to be extracted from the exhaust gas and/or supplied by aux. blowers.
In this example 10.000 kW is extracted in the first boiler 23 for delivery to the steam turbine 37 (this amount has been arbitrarily chosen for this example and other amounts could be selected as illustrated in Fig. 4A.
Fig. 4A is a graph showing the calculation results for various values of the amount of energy that is withdrawn inside the first boiler. The graph indicates the power of
the various components as a percentage of the engine shaft power, to illustrate the fact that the invention can be applied to various sizes of engines. On the graph it can be seen that the power that can be extracted from the power turbine 31 is decreased when the power extracted in the first boiler 23 upstreams of the turbocharger turbine 6 is increased. In the present example more than 25.000 kW can be extracted in the first boiler 23 without the need for putting power into the auxiliary blower 16. In the engine according to the second embodiment only about 14.000 kW can be extracted in the first boiler without the need for putting power into the auxiliary blower 16. Since the fuel efficiency of the engine itself is only very slightly deteriorated by the humid and warm scavenging air, the overall fuel efficiency of the engine 1 in combination with the exhaust gas energy recovery system according to the present invention is significantly more efficient than a conventional engine with an exhaust gas energy recovery system (e.g. the second embodiment). The ideal operating points for the engine according to the third embodiment are similar to the operating points for the engine according to the second embodiment.
In a variation of the third embodiment, the engine is operated with very low exhaust gas temperature at the outlet. These temperatures can be as low as -40 °C which means that the water in the exhaust gas will go through two phase changes: from steam to liquid and from liquid to solid, e.g. the exhaust gas leaving the engine will contain snow or similar form of ice. Thus, the engine acts as a heat pump, which is particularly interesting for applications in which both mechanical energy and heat are required, such as in a combined heat and electricity power plant that is used for providing electricity and district
heating. This operating state is obtained by extracting a very large amount of energy at the first boiler 23, in the example of table 1 72.000 kW is extracted. Further, the effective area of the turbine 6 is reduced by about one third when compared with the above described examples/embodiment, resulting in an exhaust gas temperature of -25°C. As a consequence of the reduced effective turbine area, the amount of energy available for the compressor 9 is significantly reduced (the temperature drop of the exhaust gas over the turbine (due to gas expansion) increases when the effective turbine area decreases) . Thus, the capacity and the power consumption of the auxiliary blower is increased. In this embodiment, the auxiliary blower 16 is operative at all load conditions, e.g. also at full load, since the power produced by the turbine 6 is even at full engine load not enough for the compressor 9 to produce all the scavenging air required.
When the engine is running on heavy fuel oil or diesel oil the exhaust part components downstream of the dew point are constructed with corrosion resistant materials so that they can cope with the acidic deposits that are the result of the sulfur content in these fuels (the condensate contains sulphuric acid) .
When the engine is operated with natural gas or another fuel that is substantially sulphur-free such measures are not required.
An example of the operating parameters of this variation of the third embodiment with a MAN B&W® 12K98ME engine is given in table 1 in the column "3 cold".
In this variation of the third embodiment there is no second boiler at the low pressure side due to the low temperatures of the exhaust gas after the turbocharger turbine. Thus, the system comprises only the first boiler 23 at the high pressure side of the turbine.
In another variation of this embodiment (not shown) , the engine is provided with a second turbine, for operation with higher exhaust gas temperatures at both the high and low pressure side of the turbine (e.g. between 50 and 200 °C at low pressure side and between 150 and 350 °C at high pressure side) when the demand for heat is less and the focus on rotary power is higher, e.g. summer operation of a combined power and heating plant. The system may either switch to a second turbine with a larger effective turbine area that the turbine used to obtain exhaust gas temperatures below ambient, or the second turbine may also have a relatively small effective turbine area, and the two turbines with each a small effective area are used in parallel, each receiving a portion of the exhaust gas stream. In the operation with higher exhaust gas temperatures the turbine with a larger effective turbine area, or the two turbines with a small effective turbine area operated in parallel will deliver enough power to the compressor so that the auxiliary blower only needs to be active at low load conditions. The power extracted at the boiler 23, is lowered accordingly, to obtain a temperature of the exhaust gases leaving the boiler 23 that suits the desired temperature of the exhaust gas at the low pressure side of the turbine 6. Alternatively, a single turbine with a variable effective turbine (not shown) can be used as opposed to two turbines to obtain the required flexibility in effective turbine area. Thus, this second variation is able to operate in a mode focused on heat production and a
very high overall energy efficiency, whilst the other mode focuses on rotary power production, and the system is in this mode optimized to have maximum efficiency of the amount of rotary power that can be extracted from the fuel.
Fig. 5 illustrates a fourth embodiment of the invention. This embodiment corresponds substantially to the second embodiment, except that the first boiler 23 is placed in the exhaust gas stream that is branched off from the exhaust gas conduit 5. Therefore, only the branched off portion of the exhaust gases passes through the first boiler 23. A conduit 30 leads the exhaust gases from the outlet of the first boiler 23 to the power turbine 31. The advantage of this embodiment is the fact that the exhaust gases can flow from the exhaust gas receiver 3 directly to the turbocharger turbine 6, which means that the engine will have better responsiveness to acceleration events. The outlet of the power turbine 31 is either connected to the inlet of the second boiler 20 or the final part of the exhaust conduit 21 as indicated by the interrupted line. The choice of connection depends on the outlet temperature of the power turbine 31. If the outlet temperature of the power turbine 31 is significantly lower than that of the turbocharger turbine 6, the outlet of the power turbine is connected to the final part of the exhaust conduit 21.
An example of the operating parameters of the fourth embodiment with a MAN B&W® 12K98ME engine, is shown in table 1 column "4".
In this example 20% of the exhaust gases are branched off towards the power turbine the possible power turbine power output (POp T ) or auxiliary blower input power.
It is possible to determine an optimum with regard to power extraction from the complete system. This will ultimately depend on circumstances, such as the type of boiler, the type of steam turbine and the conditions of use of the large two-stroke diesel engine. In oceangoing vessels the main focus will be on providing rotary power, whereas an application in a stationary power plant may focus equally on heat production (district heating) as on electricity production.
The available power in the exhaust gas stream (160 kg/s at 455 "C and 3.35 bar (abs.) can be utilized in 4 devices.
1) the first boiler 23 upstreams of the turbocharger turbine 6;
2) the power turbine 31;
3) the second boiler 20 downstreams of the turbocharger turbine 6; and
4) the turbocharger turbine 6.
The system can be operated at various operation points, with a variable amount of power that is taken out of the exhaust gas by the first boiler 23 and the power turbine 31.
The power extracted in the first boiler 23 upstreams of the turbocharger turbine 6 will reduce the power available for the turbocharger turbine 6 and the power turbine 31, whereas power extracted in the second boiler 20 will have no influence on the turbocharger and power turbine power.
The results for other amounts of energy extracted from the first boiler 23 are shown in the graph of Fig. 5A.
In a variation of the fourth embodiment (not shown) the cooling unit 12 is replaced by a cooling and humidification unit 12a, which adds a substantial amount of water (vapor) to the charging air. The charging air is in this embodiment not cooled to as low a temperature as in the embodiments without humidification of the charging air. The operating parameters of this embodiment are shown in table 1 in the column "4 humid".
Fig. 6 illustrates a fifth embodiment of the invention. This embodiment corresponds substantially to the second embodiment, except that the second boiler 20 is not present. Further, the engine is operated with very low exhaust gas temperature at the outlet. These temperatures can be as low as -40 'C which means that the water in the exhaust gas will go through two phase changes: from steam to liquid and from liquid to solid, e.g. the exhaust gas leaving the engine will contain snow or similar form of ice. Thus, the engine acts as a heat pump, which is particularly interesting for applications in which both mechanical energy and heat are required, such as in a combined heat and electricity power plant that is used for providing electricity and district heating.
The low temperature of the exhaust gas is obtained by extracting a large amount of energy at the boiler 23, so that the temperature of the exhaust gases leaving the boiler 23 is relatively low. The subsequent expansion of the exhaust gas in the turbocharger leads to a further fall in the exhaust gas temperature. This fall in temperature is not limited to the ambient temperature, but may fall significantly below ambient. Thus the combustion engine is turned into a so called heat pump, in which low grade heat
energy is extracted from the environment to produce high grade heat .
When the engine is running on heavy fuel oil or diesel oil the exhaust part components downstream of the dew point are constructed with corrosion resistant materials so that they can cope with the acidic deposits that are the result of the sulfur content in these fuels (the condensate contains SO 3 - sulphuric acid) .
When the engine is operated with natural gas (LNG) , LPG, DME, alcohol or other fuel that is substantially sulphur- free such measures are not required.
In the fifth embodiment there is no boiler at the low pressure side due to the low temperatures of the exhaust gas after the turbocharger turbine. Thus, the system comprises only the first boiler 23 at the high pressure side of the turbine.
An example of the operating parameters for the fifth embodiment when using a MAN B&W® 12K98ME engine is shown in table 1 in column "5&6".
The available power in the exhaust gas stream (160 kg/s at 455 °C and 3.30 bar (abs.) is utilized in 3 devices.
1) the first boiler 23 upstreams of the turbocharger turbine 6; 2) the power turbine 31; and 3) the turbocharger turbine 6.
The system can be operated at various operation points, with a variable amount of power that is taken out of the
exhaust gas by the first boiler 23 and the power turbine 31.
The power extracted in the first boiler 23 upstreams of the turbocharger turbine 6 will reduce the power available for the turbocharger turbine 6 and the power turbine 31.
In a variation of the fifth embodiment (not shown) the engine is as described above for the third embodiment provided with two turbines, to allow the engine also to be operated with higher exhaust gas temperatures and a focus on the efficiency of the amount of rotary power extracted from the fuel as opposed to the overall fuel energy (calculated relative to the combined heat and power produced by the engine) .
Fig. 7 illustrates a sixth embodiment of the invention. This embodiment is similar to the embodiment of Fig. 6 except that the turbocharger 8 is omitted. An electrically driven blower 16' (that no longer deserves the name "auxiliary blower") pressurizes the scavenging air. On the exhaust gas side an enlarged power turbine 31' takes over the role of the turbocharger turbine and supplies electricity via electric generator 32' to the electric drive motor 17' that powers the blower 16' . Any surplus electric power generated by the enlarged generator 32' can be deployed for other purposes. The management of the electric power generated by the generator 32' can be handled by a controller unit (not shown) that operates in accordance with a power management program or under direct instructions from a human operator. The lack of the fixed connection between the turbine and the compressor allows for a more flexible operation of this engine since the power generated with the power turbine can be distributed
more flexibly than with a fixed shaft connection between the turbine and the compressor. An accumulator system (not shown) such as an electric battery can be used to bridge fluctuations in the amount of energy required for the blower 16' , thereby improving the engine response upon acceleration since the blower output can be increased simultaneously with an increase of the amount of fuel injected, without having to wait for the response of the turbine on the increased exhaust gas flow.
The engine according to the sixth embodiment can be flexibly operated over a range of powers that can be withdrawn from within boiler 23. Thus, in a "winter" setting or operating state, where large amounts of heat are required for district heating the engine is operated as a heat pump, with exhaust gas temperatures at the outlet well below 0 °C, and a "summer" setting or operating state in which the engine is not operated as a heat pump and exhaust gas temperatures in the range of 50 to 200 °C. For the summer setting a second turbine (not shown) is used in combination with the turbine 31' or instead of turbine 31' so that the overall effective turbine area is increased. Alternatively, a single turbine with a variable effective turbine can be used. The change in operating state is also determined by the amount of energy extracted at boiler 23. The larger the amount of energy withdrawn at the boiler 23, the lower the temperature of the exhaust gas leaving the turbine becomes.
In the "winter" setting the various temperatures and pressures correspond to the example provided for the fifth embodiment, cf. table 1.
In a variation of the sixth embodiment (not shown) the turbine 31' drives a hydraulic pump and the blower 16 is driven by a hydraulic motor (instead of an electric generator and motor respectively) . The hydraulic pump and motor can be positive ' displacement devices eventually with variable stroke for flexibility. The hydraulic pump and motor are connected via conduits and valves that are operated by the controller 27 so that the hydraulic energy delivered by the pump is used to supply the hydraulic motor.
Another variation of the sixth embodiment (not shown) is operated with exhaust gas of 180 °C and a second boiler at the low pressure side of the power turbine 31' to maximize the efficiency for the "summer" setting. In this case the engine parameters will correspond to those of the third embodiment (cf. table 1) in column "3 cold".
The engine can not only be operated at the two extremes mentioned above, as a matter of fact the engine can be operated with exhaust gas temperatures leaving the turbine at any desired temperature therebetween, by adjusting the amount of energy extracted at boiler 23, and selecting the appropriate effective turbine area accordingly. Hereto, the engine may also include two turbines with different effective turbine areas, one turbine with a small effective turbine area, and one turbine with a larger effective turbine area. In this variation, the engine can be operated with the turbine with the small effective turbine area only for very low exhaust gas temperatures at the low pressure side thereof (winter setting in a combined heating and power plant) , with only the turbine with the larger effective turbine area for medium temperatures of the exhaust gas at the low pressure side thereof (spring/autumn
setting in a combined heating and power plant) , and with both turbines in parallel for high exhaust gas temperatures at the low pressure side if the turbines (summer setting in a combined heating and power plant) .
Fig. 8 illustrates a seventh embodiment of the invention. This embodiment is similar to the fourth embodiment. However, in the seventh embodiment the air flow to the turbocharger 8 and the exhaust gas flow from the turbocharger/power turbine is reduced with 20%, as 20% of the exhaust gas is recirculated via the first boiler 23, recirculation conduit 19, a blower 18 and a scrubber 18a, back to the scavenging system at conduit 11 upstreams of the intercooler 12. The outlet of the power turbine 31 is either connected to the inlet of the second boiler 20 or the final part of the exhaust conduit 21 as indicated by the interrupted line. The choice of connection depends on the outlet temperature of the power turbine 31. If the outlet temperature of the power turbine 31 is significantly lower than that of the turbocharger turbine 6, the outlet of the power turbine is connected to the final part of the exhaust conduit 21.
An example of the operating parameters for this embodiment using the same engine as in the previous embodiments is shown in table 1, column "7" .
In order to produce this air amount 128 kg/s with a scavenging air pressure of 3.6 bar the turbocharger compressor requires a power input of approx. 20.000 kW.
This power must be extracted from the exhaust gas by the turbocharger turbine. The exhaust gas contains 22.400 kW. The turbocharger turbine needs only to have
20000/22400*100% = 89% of the exhaust gas flow in order to produce the 20.000 kW required. The remaining 11% flow can be used in the power turbine 31. Further, the exhaust gas re-circulating flow will be 20% of the total exhaust gas flow, and all the energy in the flow line can be utilized in the first boiler 23.
The inlet temperature of the second boiler 20 is variable depending on power extracted in boiler 1 and should not be lower than approx. 300 0 C as lower temperature than the 300 0 C will result in outlet temperatures lower than 180 0 C (if natural gas or another sulphur free fuel is used the temperatures can be chosen lower with condensation and possible freezing of the exhaust gas to maximize overall energy efficiency)
The power turbine 31 power only depends on the power turbine inlet temperature or actually how much power is extracted in the first boiler 23 power turbine inlet string.
Further the boiler inlet temperature is now a mixture of turbocharger outlet temperature and power turbine outlet temperature.
This embodiment is particularly advantageous in that it obtains low NO x values for the exhaust gas.
Table 1
Embodiment 2 3 4 4 humid 7 3 cold 5&6
Pambient 1000 1000 1000 1000 1000 1000 1000 mbar
Tambient 25 25 25 25 25 25 25 dg.c
Tcom pressor out 184 184 184 184 184 184 184 dg.c
Pcom pressor out 3,62 3,62 3,62 3,62 3,62 3,62 3,62 bar
Mass flow compressor 156 156 156 156 125 156 156 kg/s
Water added 0 7,5 0 7,5 0 0 7,5 kg/s
Tscavenging 37 71 37 71 37 37 71 dg.c
Pscavenging 3,6 3,6 3,6 3,6 3,6 3,6 3,6 bar
Mass flow 156 163,6 156 163,5 156 156 163,6 kg/s
Tcyl out 464 463 464 463 464 464 463 dg.c
Pcyl out 3,35 3,35 3,35 3,35 3,55 3,35 3,35 bar
Mass flow 160 167 160 167 160 160 167 k/s
Tboiler 1 in 464 463 464 43 464 464 463 dg.c
Pbolier 1 in 3,35 3,35 3,35 3,35 3,35 3,35 3,35 bar
Mass flow 160 167 18 29 45 160 167 k/s
Power boiler 1 10000 10000 5000 5000 5000 72.000 77.500 kW
Condensation heat X X X X X 12.000 31.000 kW
Tboiler out 409 412 200 304 358 50 50 dg.c
Pboiler out 3,34 3,34 3,34 3,34 3,34 3,34 3,34 bar
Mass flow 160 167 18 29 45 160 167 k/s
T turbine in 409 412 464 463 464 50 50 dg.c
P turbine in 3,33 3,33 3,35 3,35 3,35 3,33 3,33 bar
Mass flow 154 149 142 138 115 160 167 k/s
T power turbine in 409 412 200 304 358 50 50 dg.c
P power turbine in 3,33 3,33 3,33 3,33 3,33 3,33 3,33 bar
Mass flow 6 18 18 29 14 0 0 kg/s
Tturbine out 259 263 303 304 304 -25 -25 dg.c
P turbine out 1 ,03 1 ,03 1 ,03 1 ,03 1 ,03 1 ,03 1 ,03 bar
Mass flow 160 149 142 138 115 160 167 k/s
Mass flow with condensation 155 155
T power turbine out 259 263 97 180 221 -25 -25 dg.c
P power turbine out 1 ,03 1 ,03 1 ,03 1 ,03 1 ,03 1 ,03 1 ,03 bar
Mass flow 6 18 18 29 14 0 0 k/s
Tboiler2in 259 263 280 282 294 X X dg.c
P-boiler2in 1 ,02 1 ,02 1 ,02 1 ,02 1 ,02 X X bar
Mass flow 160 167 160 167 129 160 167 k/s
Power turbine - - power 930 3100 1900 3900 2000 12.900 11.500 kW
If condensation 13.400 13.400 kW
P boiler 2 13000 14600 16500 17900 21200 X X
P boiler 2 alternative 13000 14600 18100 17900 20.800 X X
Tboiler out 180 180 180 180 180 X X dg.c
Pboiler out 1 1 1 1 1 X X bar
Mass flow 160 167 160 167 129 X X k/s
Caloric input 147000 147000 147000 147000 147000 147000 147000 kW
Shaft output 68640 68640 68640 68640 68640 68640 68640 kW
Power turbine power 930 3100 1900 3900 2000 -13400 -13400 kW
Cooling effect 28600 7000 28600 7000 28600 28600 7000 kW
Boiler 1 10000 10000 5000 5000 5000 84000 108500 kW
Boiler 2 to 50 dg.C 34400 37400 37900 40700 30000 0 0 kW
Mechanical eff. 47,3 48,8 48,0 49,3 48,1 37,6 37,6 %
Thermal eff. 49,7 37,0 48,6 35,9 43,3 76,6 78,6 %
Total 97,0 85,8 96,6 85,2 91 ,3 114,2 116,1 %
The embodiments described above have been illustrated with two stage steam system. However, the steam system can also be carried out as a single stage system or as a system with more than two stages.
The embodiment with the boiler disposed inside the exhaust gas receiver as illustrated with reference to Figs. 1 and 2 can be combined with the other embodiments illustrated in Figs. 3, 3a, 4, 4a, 5-8.
The examples above are all for an engine running at its maximum continuous rating (MCR) . It is noted that these engines can run under different loads, which will result in other values for the temperatures and pressures in the intake and exhaust systems.
Although the above embodiments and examples are based on one specific model large two-stoke diesel engine, other sizes and types of combustion engines can advantageously be used in connection with the inventions described herein.
Typically, the temperatures of the exhaust gases leaving the cylinders of a large two-stoke diesel engine are between 400 and 500 °C. The pressure of the exhaust gases leaving the cylinders of such an engine is normally above 2 bar, typically between 3 and 4 bar.
Specifically, the concept of expanding the exhaust gas over the turbine to temperatures below ambient can be used for 2-stroke and 4-stoke combustion engines.
The term "comprising" as used in the claims does not exclude other elements or steps. The term "a" or "an" as used in the claims does not exclude a plurality.
The reference signs used in the claims shall not be construed as limiting the scope.
Although the present invention has been described in detail for purpose of illustration, it is understood that such detail is solely for that purpose, and variations can be made therein by those skilled in the art without departing from the scope of the invention.