DK180131B1

DK180131B1 - A large two-stroke uniflow scavenged gaseous fueled engine and method for reducing preignition/diesel-knock

Info

Publication number: DK180131B1
Application number: DKPA201870710A
Authority: DK
Inventors: Kjemtrup Niels; Hult Johan; Mørch Christian
Original assignee: MAN Energy Solutions
Priority date: 2018-10-31
Filing date: 2018-10-31
Publication date: 2020-06-08
Also published as: KR20200050371A; KR102279454B1; DK201870710A1; CN111120081A; CN111120081B; JP2020070804A; JP6807443B2

Abstract

A large two-stroke turbocharged uniflow scavenged internal combustion engine comprising a combustion chamber delimited by a cylinder liner (1), a piston (10) and a cylinder cover (22), scavenge air ports (18) arranged in the cylinder liner (1), an exhaust gas outlet arranged in the cylinder cover (22) and controlled by an exhaust valve (4), a plurality of gaseous fuel admission openings (34) arranged in the cylinder liner (1) for admitting a gaseous fuel received via the gaseous fuel valve (30) from a supply of pressurized gaseous fuel (40) into the combustion chamber, a combustion gas outlet connected to the combustion chamber and controlled by a combustion gas valve (26), a combustion gas flow path extending from an outlet of the combustion gas valve (26) to an inlet of the gaseous fuel valves (30), or to an inlet of a dedicated combustion gas injection valve (36), the engine being configured to inject the gaseous fuel with the combustion gas admission openings (34) and the combustion gas with the gas admission openings (34) or with the dedicated combustion gas injection valve (36) simultaneously, sequentially or with overlap into the combustion chamber in a fuel injection event to increase the momentum of the matter injected into the combustion chamber in the fuel injection event.

Description

A LARGE TWO-STROKE UNIFLOW SCAVENGED GASEOUS FUELED ENGINE AND METHOD FOR REDUCING PRE-IGNITION/DIESEL-KNOCK

TECHNICAL FIELD

The disclosure relates to large two-stroke gaseous fueled internal combustion engines, in particular large two-stroke uniflow scavenged internal combustion engines with crossheads running on gaseous fuel injected from fuel valves arranged in the cylinder liner.

BACKGROUND

Large two-stroke turbo charged uniflow scavenged internal combustion engines with crossheads are for example used for propulsion of large oceangoing vessels or as primary mover in a power plant. Not only due to sheer size, these twostroke diesel engines are constructed differently from any other internal combustion engines. Their exhaust valves may weigh up to 400 kg, pistons have a diameter up to 100 cm and the maximum operating pressure in the combustion chamber is typically several hundred bar. The forces involved at these high pressure levels and piston sizes are enormous.

Large two-stroke turbocharged internal combustion engines that are operated with gaseous fuel that is injected by fuel valves arranged medially along the length of the cylinder liner, i.e. engines that inject the gaseous fuel during the upward stroke of the piston starting approximately when the exhaust valve closes, compress a mixture of gaseous fuel and scavenging air in the combustion chamber and ignite the compressed mixture at or near top dead center (TDC) by timed ignition means, such as e.g. pilot oil injection.

This type of gas injection, using fuel valves arranged in the cylinder liner, has the advantage that a much lower fuel injection pressure can be used, since the gaseous fuel is injected when the compression pressure is relatively low, when compared to large two-stroke turbocharged internal combustion engines which inject gaseous fuel when the piston is close to its top dead center (TDC), i.e. when the compression pressure in the combustion chamber is at or close to its maximum. The latter type of engine needs fuel injection pressures that are significantly higher than the already high maximum combustion pressure. Fuel systems that can handle gaseous pressures at these extremely high pressures are expensive and complicated due to the volatile nature of the gaseous fuel and its behaviour at such high pressures, which include diffusion into and through the steel components of the fuel system.

Thus, the fuel supply and system for engines that inject gaseous fuel during the compression stroke are significantly less expensive when compared to engines that inject the gaseous fuel at or near TDC.

However, when injecting gaseous fuel during the compression stroke, the piston compresses a mixture of gaseous fuel and scavenging air and consequently there is a risk of preignition and/or diesel-knock.

Problems with pre-ignition/diesel-knock can be reduced by ensuring that the charge in the combustion chamber is as homogeneous as possible. However, obtaining homogeneous scavenging air and gaseous fuel charge is challenging since only a very short window of the engine cycle is available for obtaining the homogeneous charge due to the fact that the window in the engine cycle from the time where the exhaust valve closes to top dead centre is relatively small, typically 70-90° crankshaft angle, compared to the portion of the engine cycle available in e.g. four-stroke engines, where the gaseous fuel and the charging air can really be mixed in the intake system or at the least during most of the opening phase of the inlet valve, typically during 40-160° crankshaft angle.

This relatively short window of the engine cycle available for obtaining homogeneous charge increases the challenge of avoiding pre-ignition/diesel-knock in large two-stroke diesel engines.

A non-homogeneous charge of gaseous fuel and scavenging air inside the combustion chamber, increases the risk of preignition/diesel-knock, with potential serious damage to the engine as a result.

The prior art has attempted to solve the problem of preignition/diesel-knock in engines the following manner.

A large uniflow scavenged two-stroke engine having a piston moving inside a cylinder liner, a cylinder head comprising an exhaust valve, and scavenge air ports arranged circumferentially in the cylinder liner. Several fuel injection valves are circumferentially distributed around the cylinder liner above the air scavenge ports. The fuel is injected at crank angles of at least 90° before TDC.

A large uniflow scavenged two-stroke engine in which a gaseous fuel is injected at the scavenge ports, into the air flowing into the combustion chamber. Furthermore, water injection nozzles are provided at the cylinder head. Water is injected into the combustion chamber during compression in order to lower the temperature of the fuel/air mixture, and thereby to prevent pre-ignition/diesel-knock.

The above solutions, however, have shown that they do not satisfactorily serve to effectively prevent preignition/diesel-knock in a large two-stroke compressionignited internal combustion engine.

JP2012154188 discloses a two-stroke engine comprising a cylinder, a piston sliding in the cylinder, a discharge port formed at one end in the stroke direction of the cylinder and opened/closed for discharging the exhaust generated in the cylinder; a scavenging port formed in the inner circumferential surface on the other end side in the stroke direction of the cylinder for sucking active gas into the cylinder corresponding to the sliding operation of the piston; a fuel injection valve for injecting fuel gas in the inner circumferential surface of the cylinder; and an inert substance injection valve for injecting the inert substance so as to collide with the fuel gas thus injected. Thus, nitrogen oxide emissions are reduced by low pressure injecting a small amount of inert substance, in operation of an engine. More inert substance is supplied to the region where the concentration of the fuel gas is locally high in the combustion chamber and unnecessary supply of the inactive substance is suppressed in a region where the concentration of the fuel gas is low. This reduces nitrogen oxides with a small amount of inert material and reduced pre-ignition/diesel-knock caused by deviation of the ignition timing of the fuel gas when the inside of the cylinder liner becomes hot. However, pre-ignition/dieselknock problems are not satisfactory solved by injecting a small amount of inert substance to collide with the fuel gas, and the engine proposed in JP2012154188 requires a supply of pressurized inert gas without providing a workable solution for a source of inert gas.

Thus, there is a need for an improvement in the injection of fuel in such large two-stroke turbocharged internal combustion engines in order to overcome or at least reduce the problems relating to pre-ignition/diesel-knock.

SUMMARY

It is thus an object to provide a large uniflow scavenged two-stroke turbocharged engine operated with gaseous fuel injected from fuel valves in the cylinder liner that inject the gaseous fuel during the compression stroke, in which pre-ignition/diesel-knock can be prevented, or at least reduced.

The foregoing and other objects are achieved by the features of the independent claims. Further implementation forms are apparent from the dependent claims, the description and the figures.

According to a first aspect, there is provided a large twostroke turbocharged uniflow scavenged internal combustion engine comprising:

a combustion chamber delimited by a cylinder liner, a piston and a cylinder cover, scavenge air ports arranged in the cylinder liner, an exhaust gas outlet arranged in the cylinder cover and controlled by an exhaust valve, one or more of gaseous fuel admission openings arranged in the cylinder liner for admitting a gaseous fuel received from a supply of pressurized gaseous fuel via a gaseous fuel valve into the combustion chamber, a combustion gas outlet connected to the combustion chamber and controlled by a combustion gas valve, a combustion gas flow path extending from an outlet of the combustion gas valve to:

an inlet of the gaseous fuel valve, or to an inlet of a dedicated combustion gas injection valve, the engine being configured to:

admit the gaseous fuel and to inject the combustion gas simultaneously, sequentially or with overlap into the combustion chamber in a fuel injection event to increase the momentum of the matter injected into the combustion chamber in the fuel injection event.

The purpose of the injected combustion gas is to increase the momentum of the matter injected into the combustion chamber by injecting a reactive substance that does not change the calorific value of the matter injected into the combustion chamber. Increasing the momentum improves mixing of the gaseous fuel with the scavenging air, which in turn results in a more homogenous charge and a reduced risk of pre-ignition/diesel-knock .

The combustion gas is a reactive substance, but this does not change the calorific value of the matter injected into the combustion chamber. However, the additional momentum created by injecting the combustion gas increases the overall momentum of the matter injected, thereby reducing the risk of knock or premature combustion.

Momentum (p) is the product of mass m (kg) and velocity v (m/s) of an object: p = mv, (kg-m/s)

Thus, the overall momentum of the matter injected into the combustion chamber in a fuel injection event is the product of the mass of the fuel that is injected multiplied by the speed of the fuel injected combined with the product of the mass of the combustion gas injected multiplied by the speed of the injected combustion gas.

The speed of the injected gaseous fuel is limited by the speed of sound. Thus, the speed of the injected gaseous fuel cannot be increased infinitely. The mass of the gaseous fuel injected during an injection event/per engine cycle is determined by the engine load. Thus, the mass of the gaseous fuel injected cannot be changed without changing the engine load, and in most operating conditions the engine load will determine the amount of fuel injected and not vice versa. Consequently, increasing the momentum of the injected gaseous fuel any further after the gaseous fuel has reached the speed of sound is normally not possible.

However, the inventor arrived at the insight that momentum can be increased by injecting combustion gas in addition to the injected gaseous fuel, to thereby increase the mass injected, and thus increase the momentum of the matter injected during an injection event. Thus, the momentum is increased by high velocity injection of additional gas, in particular combustion gas.

The inventor also arrived at the insight that the additional combustion gas lowers the temperature of the charge in the combustion chamber during compression, thereby further reducing risk of pre-ignition/dieselknock.

The combustion gas that is injected with the gaseous fuel to increase the momentum of the injected matter, is combustion gas that is obtained from the combustion chamber. The combustion gas obtained from the combustion chamber at a high pressure, thereby reducing the pumping effort required to inject the combustion gas back into the combustion chamber.

Further, injecting combustion gas into the combustion chamber in a fuel injection event is a form of exhaust gas/combustion gas recirculation that reduces the oxygen concentration of the mixture in the combustion chamber and thereby lowers combustion temperature and consequently lowers formation of NOx.

Moreover, the resulting reduction oxygen concentration the combustion chamber in itself also reduces preignition/diesel-knock

Thus, there are three effects that reduce preignition/diesel-knock, one being the reduction in oxygen concentration, the temperature reduction, and the third being the increased momentum of the matter injected during the fuel injection event, which leads to in improved mixing between the scavenging air and the gaseous fuel resulting in a more even mixture in the combustion chamber which in itself reduces pre-ignition/diesel-knock since an uneven mixture has a higher tendency to pre-ignition/diesel-knock.

Reducing the tendency or risk of pre-ignition/diesel-knock in turn allows increasing compression pressure which is advantageous in connection with power output and fuel efficiency.

Further, the reduced risk of pre-ignition/diesel-knock can be exploited by reducing the injection pressure of the gaseous fuel, thereby reducing the costs for the gaseous fuel supply system, both in construction and operation.

According to a first possible implementation of the first aspect the combustion gas flow path comprises a combustion gas receiver and a combustion gas feed conduit connecting the combustion gas valve with an inlet of the combustion gas receiver.

According to a second possible implementation of the first aspect the combustion gas flow path comprises a combustion gas control valve for ensuring stable control of the combustion gas flow by creating a back pressure.

According to a third possible implementation of the first aspect the combustion gas flow path comprises a wet scrubber for cleaning the combustion gas, the wet scrubber preferably being arranged downstream of the combustion gas receiver.

According to a fourth possible implementation of the first aspect the combustion gas flow path comprises a combustion gas cooler for cooling the combustion gas, the combustion gas cooler preferably being arranged downstream of the wet scrubber.

According to a fifth possible implementation of the first aspect the combustion gas flow path comprises a shutdown valve.

According to a sixth possible implementation of the first aspect the combustion gas flow path comprises a combustion gas supply conduit that connects the combustion gas flow path to the fuel valves or to the combustion gas valves.

According to a seventh possible implementation of the first aspect the combustion gas outlet is arranged in the cylinder cover or in the upper region of the cylinder liner.

According to an eighth possible implementation of the first aspect the engine is configured to open the combustion gas valve before or simultaneously with, the exhaust valve.

According to a ninth possible implementation of the first aspect the gaseous fuel and the combustion gas are simultaneously injected into the combustion chamber from the at least one fuel valve as a mixture.

According to a tenth possible implementation of the first aspect the gaseous fuel and the combustion gas are mixed inside the at least one fuel valve.

According to an eleventh possible implementation of the first aspect the gaseous fuel and combustion gas are mixed upstream of the at least one fuel valve.

According to a twelfth possible implementation of the first aspect the gaseous fuel and the combustion gas are simultaneously injected from nozzle holes in a nozzle of the at least one fuel valve.

According to a thirteenth possible implementation of the first aspect of the engine comprises combustion gas supply conduits for supplying the combustion gas to the at fuel valves or to the combustion gas injection valves, and separate supply conduits for supplying the gaseous fuel from the source of pressurized gaseous fuel to the fuel valves.

According to a fourteenth possible implementation of the first aspect the engine comprises a control unit configured for controlling the amount of combustion gas injected into the combustion chamber in a fuel injection event.

According to a fifteenth possible implementation of the first aspect the fuel valves are evenly distributed over the circumference of the cylinder liner.

According to a sixteenth possible implementation of the first aspect the fuel valves are arranged at a medial position along the length of the cylinder liner.

According to a seventeenth possible implementation of the first aspect the simultaneous, sequential or overlapping injection of the gaseous fuel and the combustion gas is initiated during the stroke of the piston towards the cylinder cover, preferably after the piston has passed the scavenge air ports, and even more preferably at or just before the exhaust valve is closed.

According to an eighteenth possible implementation of the first aspect the engine comprises an ignition system for initiating ignition, preferably at or near TDC.

According to a nineteenth possible implementation of the first aspect the engine comprises knock sensors and the engine is configured to control the amount of combustion gas added in response to a signal from the knock sensors.

According to a twentieth possible implementation of the first aspect the engine is configured to increase the mass of combustion gas injected in a fuel injection event when pre-ignition/diesel-knock is detected by the knock sensors.

According to a twenty-first possible implementation of the first aspect the engine is configured to decrease the mass of combustion gas injected during a fuel injection event when no knock has been detected by the knock sensors for a predetermined period of time or for a predetermined number of engine revolutions.

According to a twenty-third possible implementation of the first aspect each dedicated combustion gas valve is arranged in the cylinder liner in close proximity of one of the gaseous fuel valves.

According to a twenty-fifth possible implementation of the first aspect each dedicated combustion gas injection valve is provided with a nozzle with one or more nozzle holes.

According to a twenty-sixth possible implementation of the first aspect scavenging air is added to the combustion gas and the mixture of scavenging air and combustion gas is injected in a fuel injection event.

According to a twenty-seventh possible implementation of the first aspect the combustion gas injection valves are arranged at a medial position along the length of the cylinder liner.

According to a twenty-eighth possible implementation of the first aspect, the engine is configured admit said gaseous fuel with said fuel admission openings, and to inject said combustion gas with said fuel admission openings or with said dedicated combustion gas injection valve, with the admission of the gaseous fuel and the injection of the combustion gas taking place simultaneously, sequentially or with overlap into said combustion chamber in a fuel injection event to increase the momentum of the matter injected into the combustion chamber in said fuel injection event.

According to a twenty-ninth possible implementation of the first aspect, the cylinder liner is provided with one or more combustion gas injection openings, that are connected to the combustion gas injection valve.

According to a second aspect there is provided a method of avoiding or reducing pre-ignition/diesel-knock by improving the mixing of gaseous fuel with scavenging air in a combustion chamber of a large two-stroke turbocharged uniflow scavenged internal combustion engine, the engine comprising:

a combustion chamber delimited by a cylinder liner, a piston and a cylinder cover, scavenge air ports arranged in the cylinder liner, an exhaust gas outlet arranged in the cylinder cover and controlled by an exhaust valve, at least one of gaseous fuel admission opening arranged in the cylinder liner (1), for admitting a gaseous fuel received via a fuel valve from a supply of pressurized gaseous fuel into the combustion chamber, a combustion gas outlet connected to the combustion chamber and controlled by a combustion gas valve, a combustion gas flow path extending from an outlet of the combustion gas valve to:

the inlet of the gaseous fuel valve, or to the inlet of one or more dedicated combustion gas injection valves, the method comprising increasing the momentum of the matter injected into the combustion chamber during a fuel injection event by admitting the gaseous fuel with the admission openings and by simultaneously, sequentially or with overlap injecting the combustion gas with the admission openings or with the one or more dedicated combustion gas injection valves into the combustion chamber.

According to a first possible implementation of the second aspect the combustion gas is injected only when the high engine load is high, preferably only when the engine load is more than 60 % of the maximum continuous rating of the engine, and even more preferably only when the engine load is more than 70% of the maximum continuous rating of the engine.

According to a second possible implementation of the second aspect each dedicated combustion gas valve is are arranged in the cylinder liner in close proximity of one of the gaseous fuel valves.

These and other aspects will be apparent from and the embodiment(s) described below.

BRIEF DESCRIPTION OF THE DRAWINGS

In the following detailed portion of the present disclosure, the aspects, embodiments and implementations will be explained in more detail with reference to the example embodiments shown in the drawings, in which:

Fig. 1 is a front view of a large two-stroke diesel engine according to an example embodiment,

Fig. 2 is a side view of the large two-stroke engine of Fig. 1,

Fig. 3 is a diagrammatic representation the large twostroke engine according to Fig. 1,

Fig. 4 is a sectional view of the cylinder frame and a cylinder liner according to an example embodiment with a cylinder cover and an exhaust valve fitted thereto and a piston shown both in TDC and BDC,

Fig. 5 is a partial sectional view of the cylinder liner of Fig. 4,

Fig. 6 is a cross-sectional view the cylinder liner of Fig. 5 along the line VI - VI' with a fuel valve arrangement according to an embodiment in which gaseous fuel and injected combustion gas are delivered to the combustion chamber via one and the same fuel valve,

Fig. 7 is a side view of a fuel supply and fuel valve arrangement according to an embodiment in which gaseous fuel and injected combustion gas are mixed before delivery to the fuel valve,

Fig. 8 is a side view of a fuel supply and fuel valve arrangement according to an embodiment in which gaseous fuel and injected combustion gas are mixed inside the fuel valve,

Fig. 9 is a side view a fuel supply and fuel valve arrangement according to an embodiment in which gaseous fuel and combustion gas are not mixed before delivery to the nozzle of the fuel valve,

Fig. 10 is a side view of a combustion gas injection valve, Fig. 11 is a cross-sectional view the cylinder liner of

Fig. 5 along the line VI - VI' with a fuel valve arrangement according to another embodiment in which gaseous fuel and combustion gas are delivered to the combustion chamber via separate valves, and

Fig. 12 is a graph illustrating a gas exchange and fuel injection cycle.

DETAILED DESCRIPTION

In the following detailed description, an internal combustion engine will be described with reference to a large two-stroke low-speed turbocharged internal combustion crosshead engine in the example embodiments. Figs. 1, 2 and 3 show an embodiment of a large low-speed turbocharged two-stroke diesel engine with a crankshaft 8 and crossheads 9. Figs. 1 and 2 are front and side views, respectively. Fig. 3 is a diagrammatic representation of the large low-speed turbocharged two-stroke diesel engine of Figs. 1 and 2 with its intake and exhaust systems. In this example embodiment, the engine has four cylinders in line. Large low-speed turbocharged two-stroke diesel engines have typically between four and fourteen cylinders in line, carried by an engine frame 11. The engine may e.g. be used as the main engine in a marine 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 1,000 to 110,000 kW.

The engine is in this example embodiment an engine of the two-stroke uniflow scavenged type with scavenge ports 18 in the lower region of the cylinder liners 1 and a central exhaust valve 4 at the top of the cylinder liners 1. The scavenge air is passed from the scavenge air receiver 2 through the scavenge ports 18 of the individual cylinders 1 when the piston is below the scavenge ports 18. Gaseous fuel and combustion gas are injected from fuel injection valves 30 (in an embodiment (Fig. 11) the combustion gas is injected by dedicated combustion gas injection valves 36, with each dedicated combustion gas injection valve 36 preferably in close proximity of a fuel valve 30 in the cylinder liner 1) when the piston is in its upward movement and before the piston passes the fuel valves 30/combustion gas injection valves 36. Both the fuel valves 30 and the combustion gas injection valves 36 are preferably evenly distributed around the circumference of the cylinder liner and placed somewhere in the central area of the length of the cylinder liner 1. Thus, the injection of the gaseous fuel (and the injection of the combustion gas) takes place when the compression pressure is relatively low, i.e. much lower than the compression pressure when the piston reaches TDC.

A piston 10 in the cylinder liner 1 compresses the charge of gaseous fuel, injected combustion gas and scavenge air, compression takes place and at or near TDC ignition and is triggered by e.g. injection of pilot oil (or any other suitable ignition liquid) from pilot oil fuel valves 50 that are preferably arranged in the cylinder cover 22, combustion follows and exhaust gas is generated. Alternative forms of ignition systems, instead of pilot oil fuel valves 50 or in addition to file fuel valves 50, such as e.g. pre-chambers (not shown), laser ignition (not shown) or glow plugs (not shown) can also be used to initiate ignition.

When the exhaust valve 4 is opened, the exhaust gas flows through an exhaust duct associated with the cylinder 1 into the exhaust gas receiver 3 and onwards through a first exhaust conduit 19 to a turbine 6 of the turbocharger 5, from which the exhaust gas flows away through a second exhaust conduit via an economizer 20 to an outlet 21 and into the atmosphere. Through a shaft, the turbine 6 drives a compressor 7 supplied with fresh air via an air inlet 12. The compressor 7 delivers pressurized scavenge air to a scavenge air conduit 13 leading to the scavenge air receiver 2. The scavenge air in conduit 13 passes an intercooler 14 for cooling the scavenge air.

The cooled scavenge air passes via an auxiliary blower 16 driven by an electric motor 17 that pressurizes the scavenge air flow when the compressor 7 of the turbocharger 5 does not deliver sufficient pressure for the scavenge air receiver 2, i.e. in low- or partial load conditions of the engine. At higher engine loads the turbocharger compressor 7 delivers sufficient compressed scavenge air and then the auxiliary blower 16 is bypassed via a non-return valve 15.

Combustion gas is extracted from the combustion chamber through an outlet that is controlled by a combustion gas valve 26. The outlet and the combustion gas valve 26 are preferably arranged in the cylinder cover 20. The combustion gas valves 26 are controlled by an electronic control unit (not shown) and the timing and the length of the opening of the combustion gas valve 26 is determined in accordance with operating conditions, i.e. the amount of combustion gas that is required to be injected back into the combustion chamber. Preferably, the combustion gas valve 26 is opened simultaneously with the exhaust gas valve or before the exhaust gas valve in order to extract combustion gas at a pressure higher than the cylinder pressure at the time of combustion gas injection.

The engine is provided with a combustion gas flow path that extends from the outlet of the combustion gas valve 26 to an inlet of a fuel valve 30 or to an inlet of a dedicated combustion gas injection valve 36.

The combustion gas flow path serves to supply the extracted combustion gas to the fuel valve 30 to a dedicated combustion gas injection valve 36 for injecting the combustion gas into the combustion chamber, simultaneously, sequentially or with overlap with the injection of the gaseous fuel into the combustion chamber.

The combustion gas flow path comprises a combustion gas receiver 27 for collecting the combustion gas and to minimize pressure pulsations from the extraction process, a control valve 28 for ensuring stable control of the recirculated combustion gas amount by creation of necessary back pressure, a wet scrubber 29 for cleaning the recirculated combustion gas so that SO2 and soot particles are removed to the necessary extent, a combustion gas cooler 39 for cooling the recirculated combustion gas down the same temperature as scavenge air, a shutdown valve 38 for closing down the system and tightening against the wet scrubber 29 during standstill of the combustion gas recirculation system 44 and a water treatment system 45 for cleaning of the scrubber water and separation of soot particles and clean water. The combustion gas flow path also comprises a combustion gas feed conduit 48 and a combustion gas supply conduit 37. The combustion gas feed conduit 48 connects in an embodiment the outlet of the combustion gas valve 26 to the inlet of the combustion gas receiver 27 and the outlet of the combustion gas receiver 27 to the inlet of the wet scrubber 29, with the control valve 28 placed in the feed conduit 48, preferably between the combustion gas receiver 27 and the wet scrubber 29. The combustion gas flow path also comprises a combustion gas supply conduit 37, and in an embodiment the combustion gas supply conduit 37 connects the outlet of the wet scrubber 29 to the inlet of the combustion gas cooler 39 and the outlet of the combustion gas cooler 39 to the inlet of the fuel valves 30 or to the inlet of the dedicated combustion gas injection valves 36. In an embodiment, the shutdown valve 38 is placed in the combustion feed conduit 37 between the combustion gas cooler 39 and the inlet of the fuel valves 30 or the inlet of the dedicated combustion gas injection valves 36.

There may be only one or the plurality of combustion gas valves 26 is provided for each cylinder of the engine.

In an embodiment, the combustion gas receiver 27 is connected to the combustion gas valves 26 of all of the cylinders of the engine, so that the combustion gas flow path has only a single string from the combustion gas receiver 27 to the shutdown valve 38, i.e. combustion gas flow path will thus only comprise a single combustion gas receiver 27, a single control valve 28, a single wet scrubber 29 a single exhaust gas cooler 39 and the single shutdown valve 38.

Hot combustion gas of approx. 700°C is extracted from the combustion chamber by the combustion gas valve(s) 26 and supplied into the combustion gas receiver 27, where the pressure pulsations resulting from the opening and closing of the combustion gas valves 26 are reduced, in order to achieve stable pressure before the gas enters the scrubbing system. For the cause of an efficient scrubbing process it is preferred to have stable pressure condition through the scrubber.

In the wet scrubber 29 the recirculated combustion gas is cleaned by re-circulated fresh water, preferably with addition of NaOH for neutralization of the Sulphuric acid generated during reaction between the water and the SO2 and SO3. By evaporation of scrubber water, the temperatures finds an equilibrium temperature around 90°C in the wet scrubber 29. Only 10-20% of the scrubber water evaporates and the surplus of dirty scrubber water is drained from the scrubber and collected in a buffer tank which is part of the water treatment system 45. The water treatment system 45 cleans the dirty scrubber water and supplies the scrubber with clean re-circulated scrubber water. The water treatment system 45 is provided with a sludge exhaust 46 and with a clean water exhaust 47, the latter may provide the scrubber with clean water.

The cleaned recirculated combustion gas is subsequently cooled in the combustion gas cooler 39. During the cooling a considerable amount of water will condensate in the combustion gas cooler 39.

In an embodiment the recirculated combustion gas is cooled down to scavenge air temperature approximately 35 to 40°C.

In an embodiment the recirculated combustion gas is mixed with cooled scavenging air from the compressor 7 and the mixed combustion gas and scavenging air are supplied to the fuel valves 30 or to the dedicated combustion gas injection valves 36 for simultaneous, sequential or overlapping injection with gaseous fuel in a fuel injection. The scavenging air is supplied though a scavenging air supply duct 52 and the amount of scavenging air is controlled by a scavenging air valve 53. The scavenging air valve 53 is used to shut off the supply of scavenging air to the combustion gas supply conduit 37, or a separate shut off valve (not shown) is provided in the scavenging air duct 52.

Control of the amount of recirculated combustion gas is achieved by timing of the opening and closing time of the combustion gas valves 26, so that it is possible to run in different modes i.e. International Marine Organization (IMO) Tier II and Tier III modes with combustion gas recirculation. Besides variations of the amount of recirculated combustion gas can be obtained over the load range.

In an embodiment, the combustion gas valve 26 starts opening 35-50 degrees before the exhaust valve 4 starts opening. In a typical large two-stroke in flow scavenging crosshead internal combustion engine this pressure will be approximately in the range of 15 to 30 bar.

In an embodiment where the engine is operated with gaseous fuel, a smaller amount of combustion gas is necessary for avoiding pre—combustion/diesel-knock, i.e. the combustion gas injection valve will be opened later at a lower cylinder pressure.

The length of the opening of the combustion gas valve 26 will depend on the dimensions of the combustion gas valves 26 and the number of combustion gas valves 26. The opening time and period varies depending on the amount of recirculated combustion gas needed.

Fig. 12 is a graph illustrating the open and closed periods of the scavenge ports 18 (inlet port), the exhaust valve 4 and the gas valves (gaseous fuel valves 30 and combustion gas injection valves 36), respectively, as a function of the crank angle. The graph shows that the window for injecting gaseous fuel and combustion gas is very short, allowing very short time for the gaseous fuel to mix with the scavenging air in the combustion chamber. The gaseous fuel, and the injected combustion gas must be injected in this very short window.

The amount of combustion gas injected is substantial and the pressure at which the combustion gas that is injected is high, in order to obtain a relatively large mass that is injected in a relatively high speed, in order to obtain a significant momentum of the combustion gas injected.

The momentum of the combustion gas injected combines with the momentum of the gaseous fuel injected to create a total momentum that is significantly higher than the momentum of the gaseous fuel alone.

The combustion gas injected is a reactive substance but it does not have any calorific value and thus, the calorific value of the total matter injected into the combustion chamber is not substantially different from the caloric value of the gaseous fuel injected into the combustion chamber alone.

The amount of gaseous fuel injected per engine cycle is dictated by the engine load. The amount of combustion gas to be injected per engine cycle will depend on the velocity of the injected gaseous fuel (which relates to the pressure with which the gaseous fuel is injected and relates to the nozzle of the fuel valves and the geometry of the nozzle holes) and on the need to prevent pre-ignition/diesel-knock of a particular engine running on a particular type of

4 gaseous fuel and can be determined by simple trial and error.

Preferably, combustion gas is injected for each engine cycle. For low engine loads there is normally a much lesser risk of pre-ignition/diesel-knock. Thus, in an embodiment, the combustion gas is only injected into the combustion chamber when the engine load is high, e.g. above 60 - 70 % of the maximum continuous rating of the engine.

In an embodiment, the engine is provided with knock sensors (not shown) and the amount of combustion gas added is controlled in response to the signal form the knock sensors, i.e. the amount (mass) of combustion gas injected is increased when pre-ignition/diesel-knock is detected (and lowered after a while when no pre-ignition/dieselknock is detected).

The combustion gas is in an embodiment injected simultaneously with the gaseous fuel, as a mixture with the gaseous fuel or separately injected from the gaseous fuel valve 30 through separate nozzle holes in a single nozzle or through separate nozzles of the single fuel valve.

The combustion gases are in an embodiment injected simultaneously, sequential or overlapping with the gaseous fuel separately from a dedicated combustion gas injection valve 36.

Figs. 4 and 5 and 6 show a cylinder liner 1 generally designated for a large two-stroke crosshead engine. Depending on the engine size, the cylinder liner 1 may be manufactured in different sizes with cylinder bores typically ranging from 250 mm to 1000 mm, and corresponding typical lengths ranging from 1000 mm to 4500 mm.

In Fig. 4 the cylinder liner 1 is shown mounted in a cylinder frame 23 with the cylinder cover 22 placed on the top of the cylinder liner 1 with the gas tight interface therebetween. In Fig. 4, the piston 10 is shown diagrammatically by interrupted lines in both bottom dead center (BDC) and top dead center (TDC) although it is of course clear that these two positions do not occur simultaneously and are separated by a 180 degrees revolution of the crankshaft 8. The cylinder liner 1 is provided with cylinder lubrication holes 25 and cylinder lubrication line 24 that provides supply of cylinder lubrication oil when the piston 10 passes the lubrication line 24, next piston rings (not shown) distribute the cylinder lubrication oil over the running surface of the cylinder liner 1.

In the shown embodiment, the thinnest portion of the wall 49 is at the bottom of the cylinder liner 1, i.e. the portion below the scavenge ports 18. The thickest portion of the wall 49 of the cylinder liner 1 is in the upper portion of the axial extent of the cylinder liner 1. A sharp transition in the thickness of the cylinder liner 1 around the middle of the axial extent of the cylinder liner 1 serves as a shoulder that allows the cylinder to rest on the cylinder frame 23. The cylinder cover 22 is pressed with great force applied by tensioning bolts onto the upper surface of the cylinder liner 1.

The pilot oil valves 50 (typically more than one per cylinder), or pre-chambers with pilot oil valves 50, are mounted in the cylinder cover 22 and connected to a source of pilot oil (not shown). The timing of the pilot oil injection is in an embodiment controlled by an electronic control unit (not shown).

The fuel vales 30 are installed in the cylinder liner 1, with their nozzle substantially flush with the inner surface of the cylinder liner 1 and with the rear end of the fuel valve 30 protruding from the outer wall of the cylinder liner 1. Typically, one or two, but possibly as much as three or four fuel valves 30 are provided in each cylinder liner 1, circumferentially distributed around the cylinder liner 1. The fuel valves 30 are in an embodiment arranged substantially medial along the length of the cylinder liner 1.

Figs. 5 and 6 show the cylinder liner 1 and the fuel valves 30 in greater detail. In this embodiment, a cylinder liner 1 is provided with four fuel valves 30. The fuel valves 30 are shown radially directed in Fig, 6, but it is understood that the fuel valves 30 can be arranged in another angle relative to the cylinder liner 1. In this embodiment the engine is not provided with dedicated combustion gas injection valves 36 and the combustion gas is injected by the gaseous fuel valves 30.

Fig. 6 is a cross-sectional view of the cylinder liner 1 at the level of the fuel injection valves 30. Figure 6 schematically shows the gaseous fuel supply system including a source of pressurized gaseous fuel 40 connected via a gaseous fuel supply conduit 41 to an inlet of each of the gaseous fuel valves 30. An inlet of the gaseous fuel valves 30 is also connected to a combustion gas supply conduit 37.

The fuel valves 30 are in an embodiment connected to a common (mixed) supply of gaseous fuel and combustion gas.

Fig. 7 shows the fuel valve 30 connected to both a source of pressurized fuel 40 and to the combustion gas extraction and recirculation system 44 via a single supply line 42 that is connected to a mixing point to which both the combustion gas supply conduit 37 and the gaseous fuel supply conduit 41 connected. In an embodiment, valves (not shown) are provided to ensure the desired ratio between gaseous fuel and combustion gas delivered to the fuel valves 30. A common conduit 32 transports the mixture to an admission opening 34 of the gaseous fuel valve 30. The mixture is admitted into the combustion chamber from the admission opening 34. The fuel valve 30 is provided with means for timed injection of the mixture to the combustion chamber, e.g. under control from an electronic control unit

In a variation of the embodiment of Fig. 7, the gaseous fuel and the combustion gas are not mixed and instead supplied to the fuel valve 30 sequentially and injected sequentially, either the gaseous fuel first or the injected combustion gas first, but without any substantial pause therebetween.

In another embodiment, illustrated by Fig. 8, the source of gaseous fuel 40 is connected by a dedicated supply line 41 to a dedicated port in the fuel valve 30. A dedicated conduit 31 leads the gaseous fuel to a mixing point 33 inside the fuel valve 30. The combustion gas extraction and recirculation system 44 is connected by the combustion gas supply conduit 37 to a dedicated port of the fuel valve 30. A dedicated conduit 35 leads the combustion gas to the mixing point 33 inside the fuel valve 30. In the mixing point 33 the gaseous fuel and the combustion gas are mixed and from the mixing point 33 the mixture is transported to the admission opening 34 by a common conduit 32. From the admission opening 34 the gaseous fuel that is admitted to the combustion chamber. The fuel valve 30 is provided with means for timed injection of the mixture to the combustion chamber, e.g. under control from an electronic control unit.

In another embodiment, illustrated by Fig. 9, the source of gaseous fuel 40 is connected by a dedicated supply line 41 to a dedicated port in the fuel valve 30. A dedicated conduit 31 leads the gaseous fuel to a first nozzle 39. The combustion gas extraction and recirculation system 44 is connected by the combustion gas supply conduit 37 to a dedicated port in the fuel valve 30. A dedicated conduit 35 leads the combustion gas to a combustion gas admission opening 51. The admission opening 34 admits the gaseous fuel to the combustion chamber and the combustion gas admission opening 51 injects the combustion gas into the combustion chamber. The fuel valve 30 is provided with means for timed injection of the gaseous fuel and of the combustion guns into the combustion chamber, e.g. under control from an electronic control unit.

In another embodiment, shown in Fig. 11, the cylinder liner 1 is provided with dedicated fuel valves 30 for injection of the gaseous fuel and with dedicated combustion gas injection valves 36 for the injection of combustion gas into the combustion chamber.

The dedicated combustion gas injection valves 36 are installed in the cylinder liner 1, with their combustion gas injection opening 51 substantially flush with the inner surface of the cylinder liner 1 and with the rear end of the dedicated combustion gas injection valves 36 protruding from the outer wall of the cylinder liner 1. Typically, one or two, but possibly as much as three or four dedicated combustion gas injection valves 36 are provided in each cylinder liner 1, circumferentially equally distributed around the cylinder liner 1. The dedicated combustion gas injection valves 36 are in an embodiment arranged substantially medial along the length of the cylinder liner 1, preferably in close proximity to the dedicated gaseous fuel valves 30.

The source of pressurized gaseous fuel 40 is connected to the (in this embodiment four) gaseous fuel valves 30 and the combustion gas extraction and recirculation system 44 is connected to the (in this embodiment four) dedicated combustion gas injection valves 36. The gaseous fuel valves and the combustion gas injection valves 36 are provided with means for timed injection of the gaseous fuel and of the combustion gas into the combustion chamber, e.g. under control from an electronic control unit. The fuel valves and the combustion gas injection valves 36 are shown as closely spaced pairs in Fig. 11, but it is understood that this arrangement is merely an example and that the fuel valves 31 and the combustion gas injection valve 36 do not need to be arranged in pairs and can be more widely spaced.

Fig. 10 is a side view of a combustion gas injection valve 36 with its at least one combustion gas injection opening 51 for injecting the combustion gas into the combustion chamber. The combustion gas injection valve 36 is shown connected with its proximal end to the combustion gas supply conduit 37.

In an embodiment the pressure of the combustion gas boosted by a compressor (not shown) to a suitable injection pressure. Since the combustion gas is already pressurized the energy required to bring the air or gas to the injection pressure is less compared to bringing the pressure up to injection pressure when starting from atmospheric pressure.

In an embodiment, the gaseous fuel is admitted before the exhaust valve closes, to provide more time for mixing. The timing of the injection of the combustion gas is adjusted accordingly.

In an embodiment (not shown) the cylinder is provided with a pre-chamber for ignition, which is fueled by a pilot injection from a separate pilot liquid (pilot oil) injection system.

In an embodiment (not shown) is one gaseous fuel valve supplies gaseous fuel to a plurality of gaseous fuel admission openings that admit the gaseous fuel to the combustion chamber.

The various aspects and implementations have been described in conjunction with various embodiments herein. However, other variations to the disclosed embodiments can be understood and effected by those skilled in the art in practicing the claimed subject-matter, from a study of the drawings, the disclosure, and the appended claims. In the claims, the word comprising does not exclude other elements or steps, and the indefinite article a or an does not exclude a plurality. A single processor or other unit may fulfill the functions of several items recited in the claims. The mere fact that certain measures are recited in mutually different dependent claims does not indicate that a combination of these measured cannot be used to advantage.

The reference signs used in the claims shall not be construed as limiting the scope.

Claims

Large turbocharged two-stroke internal combustion engine with longitudinal flushing comprising:

a combustion chamber defined by a cylinder liner (1), a piston (10) and a cylinder cover (22), purge air ports (18) arranged in the cylinder liner (1), an exhaust gas outlet arranged in the cylinder cover (22) and controlled by an exhaust valve (4), one or more gas fuel inlet openings (34) arranged in the cylinder liner (1) for supplying a gas fuel received from a supply of gas fuel (40) under pressure via a gas fuel valve (30) in the combustion chamber, characterized by a flue gas outlet connected to the combustion chamber and controlled by a flue gas valve (26), a flue gas flow path extending from an outlet of the flue gas valve (26) to:

an inlet to the gas fuel valve (30) or to an inlet to an injection valve (36) intended for combustion gas, the engine being designed to: supply the gas fuel and to inject the combustion gas simultaneously, sequentially or with overlap into the combustion chamber of a fuel injector; to increase the momentum of the material injected into the combustion chamber in the fuel injection event.

The engine of claim 1, wherein the combustion gas flow path comprises a combustion gas receiver (27) and a combustion gas supply duct (48) connecting the combustion gas valve (26) to an inlet of the combustion gas receiver (27).

An engine according to claim 1 or 2, wherein the combustion gas flow path comprises a combustion gas control valve (28) to ensure stable control of the combustion gas flow by generating a back pressure.

An engine according to any one of the preceding claims, wherein the combustion gas flow path comprises a wet scrubber (29) for purifying the combustion gas, which wet scrubber (29) is preferably arranged on the downstream side of the combustion gas receiver (27).

An engine according to any one of the preceding claims, wherein the combustion gas flow path comprises a combustion gas cooler (39) for cooling the combustion gas, which combustion gas cooler (39) is preferably located on the downstream side of the wet scrubber (29).

An engine according to any one of the preceding claims, wherein the flow path of the combustion gas comprises a shut-off valve (38).

An engine according to any one of the preceding claims, wherein the combustion gas flow path comprises a combustion gas supply channel (37) connecting the combustion gas flow path to the inlet of the fuel valves (30) or to the inlet of the combustion gas injection valve (36).

An engine according to any one of the preceding claims, wherein the combustion gas outlet is located in the cylinder cover (22) or in the upper region of the cylinder liner (1).

An engine according to any one of the preceding claims, wherein the engine is designed to open the combustion gas valve (26) before, at the same time as, or after the exhaust valve (4).

An engine according to any one of the preceding claims, wherein the gas fuel and the combustion gas are simultaneously injected into the combustion chamber from the at least one fuel valve (30) as a mixture.

An engine according to claim 10, wherein the gas fuel and the combustion gas are mixed inside the at least one fuel valve (30).

An engine according to claim 10, wherein the gas fuel and combustion gas are mixed on the upstream side of the at least one fuel valve (30).

The engine of claim 1, wherein the gas fuel and the combustion gas are simultaneously injected from nozzle holes in a nozzle of the at least one fuel valve (30).

An engine according to any one of claims 10 to 13, comprising combustion gas supply ducts (37) for supplying the combustion gas to the fuel valves (30) or to the combustion gas injection valves (36), and separate supply ducts (41) for supplying the gas fuel from the source of gas fuel (40) under pressure to the fuel valves (30).

An engine according to any one of the preceding claims, comprising a control unit configured to control the amount of combustion gas injected into the combustion chamber in a fuel injection event.

An engine according to any one of the preceding claims, wherein the fuel valves (30) are evenly distributed over the periphery of the cylinder liner (1).

An engine according to any one of the preceding claims, wherein the fuel valves (30) are arranged at a central position along the length of the cylinder liner (1).

An engine according to any one of the preceding claims, wherein both the injection of the gas fuel and the injection of the combustion gas are initiated during the stroke of the piston (10) against the cylinder cover (22), preferably after the piston (20) has passed the purge air ports, and even more preferably at or just before closing the exhaust valve (4).

An engine according to any one of the preceding claims, provided with an ignition system for initiating ignition, preferably at or near the apex dead center.

An engine according to any one of the preceding claims, which is provided with knock sensors, and wherein the engine is designed to control the amount of added combustion gas in response to a signal from the knock sensors.

An engine according to claim 20, which is designed to increase the mass of injected combustion gas in a fuel injection event when the ignition / diesel bank is detected by the knock sensors.

An engine according to claim 20 or 21, designed to reduce the mass of injected combustion gas during a fuel injection event when no ignition / diesel bank is detected by the knock sensors for a predetermined period of time or for a predetermined number of engine revolutions.

An engine according to any one of the preceding claims, wherein each injection valve (36) intended for combustion gas is located in the cylinder liner in the immediate vicinity of one of the gas fuel valves (30).

A method of avoiding or reducing the ignition / diesel bank by improving the mixture of gas fuel with purge air in a combustion chamber of a large turbocharged two-stroke internal combustion engine, the engine comprising:

a combustion chamber defined by a cylinder liner (1), a piston (10) and a cylinder cover (22), purge air ports (18) arranged in the cylinder liner (1), an exhaust gas outlet arranged in the cylinder cover (22) and controlled by an exhaust valve (4), at least one gas fuel inlet opening (30) arranged in the cylinder liner (1), for supplying a gas fuel received via a fuel valve from a supply of gas fuel (40) under pressure into the combustion chamber, characterized by a flue gas outlet connected to the combustion chamber and controlled by a flue gas valve (26), a flue gas flow path extending from an outlet of the flue gas valve (26) to:

the inlet to the gas fuel valve (30) or to the inlet of one or more injection valves (36) intended for combustion gas, the method comprising increasing the momentum of the material injected into the combustion chamber during a fuel injection event by accessing the gas fuel with the inlet openings (30) and simultaneously , sequentially or by overlapping, injecting the combustion gas with the inlet openings (34) or with the one or more injection valves (36) intended for combustion gas into the combustion chamber.

The method of claim 24, wherein the combustion gas is injected only when the high engine load is high, preferably only when the engine load is more than 60% of the maximum continuous engine power, and even more preferably only when the engine load is more than 70% of the maximum continuous engine power. performance.