Field of the invention The present invention relates of a dual fuel internal combustion engine operative to burn in a combustion chamber a main charge comprising a mixture of a gaseous hydrocarbon fuel, preferably natural gas, and air and to initiate combustion of the main charge by means of a liquid fuel that is injected directly into the combustion chamber.

Background of the invention

It is desirable to be able to run an engine on gaseous fuels, such as natural gas (which is predominantly methane), or propane which are readily available in large quantities and have the potential to burn more cleanly than gasoline or diesel with very low hydrocarbon an carbon monoxide

emissions. However, using normal combustion, engines running on such gases will produce NOx and it is desirable to use leaner mixtures (below stoichiometry) in order to reduce NOx by reducing combustion temperature.

The problem with lean mixtures is that they are difficult to ignite reliably by spark ignition alone. For this reason, dual fuel engines have previously been proposed that use diesel injection to initiate combustion of the lean gaseous fuel and air mixture. Such engines must of necessity employ a high compression ratio for the injected diesel fuel to ignite by compression ignition.

If the compression ignition of the diesel fuel acts in the same way as a powerful spark, the homogeneous lean charge is burnt by flame propagation and this results in localised areas of high temperature that create NOx. Also if the compression temperature of the homogeneous lean charge is high, there is a risk of auto-ignition occurring

prematurely and causing engine damage.

Object of the invention

The present invention seeks therefore to provide an engine that can reliably ignite a lean mixture of a gaseous hydrocarbon fuel and air without having to resort to a high compression ratio.

Summary of the invention

According to a first aspect of the present invention, there is provided a dual fuel internal combustion engine operative to burn in a combustion chamber a main charge comprising a mixture of a gaseous hydrocarbon fuel and air and to initiate combustion of the main charge by igniting a liquid fuel injected directly into the combustion chamber, characterised in that the liquid fuel is gasoline and a spark plug is provided within the combustion chamber to ignite the gasoline, and in that the gaseous hydrocarbon fuel and air mixture is a lean mixture that is pressurised by means of a turbocharger or supercharger. In the invention, a lean homogeneous mixture of gaseous hydrocarbon fuel and air is used in order to reduce NOx emissions and to avoid the need for a high compression ratio, combustion is initiated by injecting gasoline and igniting the gasoline using a spark. In order to compensate for the reduced output power that results from burning a weak mixture in a cylinder having a low compression ratio, the intake charge is pressurised by means of a turbocharger or supercharger.

Conveniently, the gasoline is injected directly into the combustion chamber in such a manner as to create a stratified charge with the gasoline concentrated in the vicinity of the spark plug.

The gaseous hydrocarbon fuel may suitably be introduced into the combustion chamber by port injection so that a homogeneously premixed mixture of gaseous fuel and air is admitted into the combustion chamber.

It is desirable to promote auto-ignition of the homogeneous lean charge, which is a known low temperature combustion process resulting in negligible amounts of NOx . Such combustion can be achieved by a combination of a high concentration of hot EGR and low compression ratio but the timing of the auto-ignition is unpredictable and may occur prematurely resulting in engine damage.

In some embodiments of the invention, exhaust gases from a first combustion cycle are retained within, or reintroduced into, the combustion chamber during a

subsequent cycle to raise the temperature of the charge in the subsequent cycle so as to promote auto-ignition of the homogeneous gaseous fuel and air mixture but only subsequent to spark ignition of the gasoline mixture. In this case, the selected compression ratio and the regulated amount of hot EGR set the temperature and pressure of the charge to a point slightly below the conditions necessary for auto- ignition. The spark ignition of the injected gasoline then raises the temperature and pressure sufficiently to create the necessary conditions for auto-ignition of the lean homogenous mixture, thereby enabling the timing of the auto- ignition to be controlled.

The gaseous fuel is preferably natural gas but other hydrocarbon gases such as propane may be used.

The invention is equally applicable to two stroke and four stroke engines . According to a second aspect of the invention, there is provided a method of operating an internal combustion engine having a combustion chamber which comprises filling the combustion chamber homogeneously with a lean mixture of a gaseous hydrocarbon fuel and ambient air, the pressure of the mixture being greater than the ambient atmospheric pressure, compressing the mixture, injecting gasoline directly into the combustion chamber towards the end of the compression step and igniting the gasoline and air charge by means of a spark plug to initiate combustion of the gaseous hydrocarbon fuel and air charge.

Brief description of the drawings The invention will now be described further, by way of example, with reference to the accompanying drawings, in which :

Figure 1 is a schematic diagram of one cylinder of an internal combustion engine having a gasoline direct injector and at least one natural gas injector located in the intake port, and

Figures 2a to 2d show various valve timing diagrams to achieve high temperature internal EGR (internal exhaust gas recirculation) in a four stroke engine.

Detailed description of the preferred embodiment ( s )

Figure 1 shows a schematic view of a cylinder 10 of a reciprocating internal combustion engine operating with a 2- stroke cycle or 4-stroke cycle. A piston 12, connected to a crankshaft by a connecting rod 14, reciprocates within the cylinder 10 to define a variable volume working chamber 16. The working chamber 16 has at least one intake valve 22 operated by a cam 28 and one exhaust valve 24 operated by a cam 30. A direct injection liquid fuel injector 20 in the cylinder head injects liquid fuel directly into the working chamber 16 and a spark plug 18 is provided for the spark ignition of the injected liquid fuel in the working chamber.

A port gaseous fuel injector 26 supplies natural gas into the intake port 32 to form a premixed homogeneous lean mixture of natural gas and air. The engine has at least one turbocharger and/or supercharger which is not shown in the drawing. Turbochargers and superchargers are well known equipment and may be assumed to operate in a conventional manner. The injector 20 is shown as being positioned immediately adjacent the spark plug 18 but it may

alternatively be possible to position an injector in the side of the combustion chamber and to rely on air motion to transport the injected liquid fuel to the vicinity of the spark plug. The engine may also have an exhaust back pressure valve which is not shown in the drawing. The back pressure valve serves to increase the exhaust back pressure in the engine in order to promote sufficient internal EGR into the pressurised intake charge when required.

The natural gas injected by the injector 26 mixes with pressurised air in the intake port 32 and fills the working chamber 16 with a premixed homogeneous lean mixture of gaseous fuel and air at the commencement of the compression process. Direct injection of gasoline from the injector 20 is timed to occur in the later part of the compression stroke in order to form a stratified charge with a rich or near stoichiometric gasoline and air region near the spark plug 18, which is optimised both by appropriate positioning and orientation of the injector 20 and the geometry of the piston crown and cylinder head. Combustion starts with spark ignition of the stratified gasoline and air mixture. The resulting combustion kernel formed around the spark plug then spreads to the homogeneous lean gaseous fuel and air mixture. The homogeneous lean gaseous fuel and air charge may be preheated by mixing with hot EGR gases, in the manner to be explained below. Its temperature will rise during the compression stroke to a value near TDC that is just short of the auto-ignition temperature of the gaseous fuel and air mixture. The increased temperature and pressure created by the ignition of the gasoline fuel will then raise the temperature to above the auto-ignition point to cause the entire contents of the combustion chamber to auto-ignite and undergo low temperature combustion.

Figure 2 shows various known valve timing diagrams for obtaining hot internal EGR. In all these valve timing diagrams, exhaust event are shown in solid lines and intake events in dashed lines.

In Figure 2a, hot internal EGR is achieved by a secondary opening of the exhaust valve during the intake stroke .

In Figure 2b, hot internal EGR is achieved by a secondary opening of the intake valve during the exhaust stroke . In Figure 2c, early exhaust valve closing results in a high proportion of hot exhaust gases being retained within the combustion chamber. The intake valve does not open until after TDC to prevent the compressed residual gases from being expelled through the intake valve into the intake port.

In Figure 2d, hot internal EGR is achieved by a large overlap between the opening periods of the intake and exhaust valves .

Internal EGR can also be implemented in a known manner in two-stroke engines .