OF NITROGEN OXIDES AND PARTICULATE , AND RELATED METHOD

TECHNICAL SECTOR OF REFERENCE AND STATE OF THE ART

The present invention relates to a compres sion ignition endothermic engine , provided with a special combustion chamber able to contain the formation of nitrogen oxides (N0 _x ) and particulates and related method .

It may seem counterintuitive to continue developing heat engines , as there is a plan to cease production within a few years . Diesel engines are particularly adverse for their N0 _x and particulate emissions . The fuel used by these engines is diesel fuel which has low volatility; this fact limits the possibility of having clean combustion, however the high ef ficiency of these engines allows low consumption and, consequently, low carbon dioxide ( CO2 ) emissions .

Alternative fuels to diesel are being studied, such as hydrotreated vegetable oil (HVO) and dimethyl ether ( DME ) , which is a gas easily liquefiable under pressure . These fuels have a high cetane number, so they can be used in diesel engines . In particular, DME has the advantage of mixing rapidly with air . On the other hand it has a lower calori fic value which is about two thirds that of diesel fuel .

The total replacement of heat engines with electric motors is not free from rather serious problems to be solved . These problems are both of a technical nature and of an organi zational nature and concern the availability of materials for the production of batteries , the long recharging times , the construction of a very large number of recharging columns , said number being made even higher precisely due to due to the long reload times .

Supercapacitors are currently being studied which could replace batteries and which would have much shorter recharging times , but which, consequently, would require much higher recharging currents , with the consequent safety problems due to the risk of overheating of the conductors .

Furthermore , the problems of a geopolitical nature connected with the supply of batteries , nor those of an environmental nature for the extraction of lithium and the rare earths necessary for the construction of batteries and engines should not be underestimated .

However, these are problems that are not easy to solve and risk prolonging the time needed to replace the internal combustion engine with an electric one .

So-called hybrid cars have existed for some time , i . e . equipped with an endothermic engine , usually with controlled ignition, which in addition to providing traction, also recharges the batteries suitable for operating one or more electric motors . This type of vehicle , while maintaining an endothermic engine with the relative emissions , has considerable advantages , both in terms of emissions , because in urban centres it can only operate with electric traction, substantially solving the problems of pollution in the most critical areas . Furthermore , it has no recharging problems because the internal combustion engine takes care of this in a completely automatic way . Finally, thanks to the fact that there is an endothermic engine , it needs a much smaller battery pack, therefore it generates much less supply and disposal problems .

In other words , the future is probably electric, but until the numerous problems it entails are solved, the hybrid could still have many years of use .

The problem therefore arises of which heat engine to use .

A hybrid car will tend to travel on electric traction in urban centres , where gaseous emissions are more harmful , but the internal combustion engine will still be used to recharge the batteries . Nitrogen oxides (NO _X ) and particulate matter are particularly harmful in urban centres , so a heat engine should produce as few of them as possible . Carbon dioxide ( CO2 ) is not harmful to humans , but it is harmful to the environment . Hence the compression ignition engine would be preferable to the spark ignition one , compared to which it produces much less CO2 , i f it were possible to signi ficantly reduce the production of NO _X and particulate matter . PURPOSE OF THE INVENTION

The obj ect of the present invention is to provide a compression ignition heat engine which is capable of operating by producing a lower quantity of NO _X and particulate than that produced by spark ignition engines .

It is known that compression ignition endothermic engines adj ust the power delivered by dosing the fuel inj ected into the combustion chamber, unlike spark ignition endothermic engines , in which the power delivered is adj usted by choking the flow of aspirated air and the consequent dosage of the inj ected fuel .

When both said engines work at full power, the di f ferences between said two types are lessened, since, in both cases , the cylinders will be filled with the maximum possible quantity of air and the fuel will be dosed according to a substantially stoichiometric ratio .

When traveling at low speed, as usually occurs in city traf fic, the power required is a fraction of the maximum power that can be supplied, so the following operating conditions will apply : spark-ignition engines will have highly choked intake , therefore for each cycle , a quantity of air will be involved, and therefore of nitrogen, much lower than the volume of the cylinder ; compression ignition engines , by regulating the power with fuel dosing only, will instead suck in a quantity of air of a volume equal to the volume of the cylinder, or greater i f the engine is supercharged, as almost always happens in the more recent engines .

This fact has two important implications since , precisely in the conditions of use in city traf fic, in compression ignition engines combustion occurs in conditions of a strong excess of air, and therefore of nitrogen . This entails a high combustion temperature, not limited by the thermal capacity of the fuel which evaporates , and a high quantity of nitrogen which participates in the reaction . The consequence is that there is a signi ficantly higher formation of nitrogen oxides (NO _X ) than that which occurs in spark ignition engines . In order to reduce the maximum temperatures in the combustion chamber, the prior art provides for the recycling of a variable percentage between 5 and 15% of the exhaust gases which, not participating in the combustion, reduce the maximum temperature of the cycle , which is favourable for the reduction of NO _X , but causes a signi ficant increase in speci fic consumption . Since a greater quantity of recycled gases would cause alterations in the regularity of operation, the formation of nitrogen oxides remains conspicuous , in operation at low power, compared to engines with spark ignition .

The fact that compression-ignition engines produce a higher quantity of nitrogen oxides than spark ignition engines and that this phenomenon occurs to a greater extent when running at low power , therefore above all in city traf fic, has caused a strong aversion towards compression ignition engines , commonly called diesel engines .

It is known that nitrogen oxides are formed when the combustion temperature is high enough ( above 2000 °K) and, since the phenomenon is exponential , a small di f ference in temperature is enough for a reduced or abundant formation of NO _X .

Also the charge formation, i . e . of the air- fuel mixture , has an influence on the formation of NO _X and particulates and it has been found that the homogeneity of the charge , combined with a low equivalence ratio , favours a low formation of NO _X and particulates .

For this purpose , the so-called LTC ( Low Temperature Combustion) and HCCI (Homogeneous Charge Compression Ignition) charges were studied which, in principle , solve the problem posed . However, there are many problems that arise , both intrinsic and of practical application .

One method, already mentioned, to reduce the maximum combustion temperature consists in recycling a certain percentage of exhaust gas (EGR - Exaust Gas Recycling) in the combustion chamber . The recycled gases do not participate in the combustion and this makes it possible to reduce the combustion temperature . However, this method has a strong limitation, since in the case o f recirculation percentages higher than 15% there are combustion di f ficulties , moreover it becomes more complex to control the start of the combustion itsel f .

In cars compliant with Euro VI standards , recourse is made to the recycling of burnt gases , but the consequence is that , in parallel with an appreciable reduction in NO _X and particulate matter, there is an increase in speci fic consumption which translates into a higher cost of use and, above all , in a greater production of CO ₂ .

Even the formation of the homogeneous charge is very problematic as diesel fuel has a low volatility and the fact of having to contain the temperatures does not help . It would be necessary to have a suitable period of time between the inj ection in the combustion chamber and the ignition of the air- fuel mixture to allow the air- fuel mixture to homogeni ze . This last aspect is particularly critical since the air- fuel mixture , which is formed upon inj ection, ignites instantly, without allowing the mixing to be completed . Said incomplete mixing results in a strati fication of the charge , i . e . with areas with a high concentration of fuel which give rise to the formation of particulate matter. Finally, the homogeneous mixture, if it were possible to form it, would burn explosively, which, if on the one hand more closely approximates an isochoric transformation, on the other hand it is particularly noisy and stresses the mechanical organs to such an extent that it could only be used at very reduced power .

In principle, engines with HCCI charge would have particularly low particulate and NO _X emissions, but present such problems that in practice prevent their use.

In practice, the following conditions must be fulfilled: homogeneous charge, to have the same equivalence ratio in each point of the mixture: low combustion temperature (below 2000°K) , to limit the formation of NO _X as much as possible; low equivalence ratio (less than 1) because an excess of oxidant allows for more complete combustion .

At present, HCCI engines have the following main disadvantages : difficulty in controlling the start of combustion; low values of MEP (Mean Effective Pressure) and therefore limited maximum power; narrow limit of the operating field; explosive combustion.

As already mentioned, in a conventional Diesel engine the formation of the homogeneous charge is greatly hindered by the fact that the air-fuel mixture, which is formed upon injection, instantaneously burns, without allowing the completion of the mixing with the air itsel f .

Furthermore , since there is no external start , the ignition of the charge is excessively conditioned by the chemical kinetics of the fuel-air mixture .

It is therefore very clear that the main problem of HCCI type combustions , since there is no external start that influences the start of the combustion process , is that of controlling the way in which the combustion itsel f evolves .

The start of HCCI combustion mainly depends on the thermodynamic conditions and on the characteristics of the air- fuel mixture during the intake and compress ion phases . Said start is therefore influenced by pressure , temperature , local air- fuel ratios , etc . inside the cylinder and from the interactions between the mixture itsel f and the system in which it evolves . Among these, the temperature of the cylinder walls and the heat exchange are relevant .

In HCCI type combustion the fuel is premixed with the oxidant ; the combustion starts at a low temperature in a well diluted mixture , therefore the danger of NO _X formation is avoided .

Furthermore , thanks to the low equivalence ratios , combustion is more complete and therefore both the production of particulate matter and the emis sion of unburnt hydrocarbons (HC ) in the exhaust are greatly reduced .

The fundamental technical problem to be solved in the case of an engine with HCCI operation is that o f governing the start of combustion, freeing onesel f from the complex chemical kinetics which controls the instant in which sel f-ignition of the charge takes place in the combustion chamber .

A first obj ect of the present invention is to provide a method for producing an air- fuel mixture which is as homogeneous as possible .

A second purpose of the present invention is to govern the start of combustion by freeing onesel f from the chemical kinetics of the mixture .

A third obj ect of the present invention is to use the homogeneous mixture ignited as a strati fied charge for completing the combustion .

A fourth obj ect of the present invention is to greatly limit the negative ef fects of explosive combustion, i . e . noise and excessive stress on the mechanical members .

Ultimately, the following conditions must be met : homogeneous charge , to have the same equivalence ratio in each point of the mixture : low temperature , to limit the formation of NO _X as much as possible ; low equivalence ratio ( less than 1 ) because an excess of oxidant allows for more complete combustion . DEFINITION OF THE INVENTION

The above and other obj ects , as will appear clear from the following description, are achieved by means of a method and a device for implementing said method, respectively in accordance with claims 1 and 2 .

The method for achieving low nitrogen oxide and particulate emissions in a compression ignition internal combustion engine , of the type using HCCI (Homogeneous Charge Compression Ignition) combustion, in which generation is provided within a combustion chamber of a substantially homogeneous charge , or air- fuel mixture , said substantially homogeneous charge being obtained by prolonging the mixing time at a temperature lower than the ignition temperature of said charge , is characteri zed in that : said extension of the mixing time to a temperature lower than the ignition temperature of said charge is obtained in a first zone of the combustion chamber whose volumetric compres sion ratio is such that the adiabatic heating of the mixture is not suf ficient to cause it to ignite ; the ignition o f said homogeneous charge occurs through further mixing with a superheated gaseous fluid at a higher temperature than the ignition temperature of said charge , said superheated gaseous fluid coming from a second zone of the combustion chamber, the volumetric compress ion ratio of which is such that the adiabatic heating of a gaseous fluid contained therein makes it possible to reach a temperature higher than the ignition temperature of said charge and, correspondingly, a higher pressure which determines said mixing between the gaseous fluids existing in the two parts of the combustion chamber .

Since the two zones of the combustion chamber are communicating, means are obviously provided which are able to make ef fective the di f ferent compression implemented in said two zones .

As will be better speci fied hereinafter, the fuel i s inj ected, towards the end of the compression phase , into the first zone of the combustion chamber, whereby the second zone is virtually fuel- free . However, especially when high power is required from the engine , small quantities of fuel can escape from the first zone and end up in the second zone of the combustion chamber . For this reason we have spoken more properly of gaseous fluid which is found in the second zone of the combustion chamber . Even i f said fluid were to consist of air alone , at least in the case of operation at reduced power, hereinafter we will always speak of a gaseous fluid .

After ignition of the fuel-air mixture , the pressure in the first combustion chamber zone rises sharply and the burnt gases flow from the first to the second combustion chamber zone . I f the engine is delivering a power suf ficiently higher than the minimum, among the burnt gases there is a quantity of unburned fuel which, in contact with the gaseous fluid present in the second zone of the combustion chamber, continues to burn forming a strati fied charge .

The method according to the invention is implemented by constructing the combustion chamber in such a way as to obtain two di f ferent compression ratios : a lower first compression ratio in the first zone of the combustion chamber, so as to contain the end compression temperature below the fuel flash point ; a second higher compression ratio in the second zone of the combustion chamber, such that the end-of- compression temperature is high enough to ignite combustion of the fuel-air mixture .

In this way the ignition of the mixture is released from the chemical kinetics and is obtained when the superheated gaseous fluid seeps from the second zone to the first one, said seepage occurring when the piston is near top dead centre (TDC) . By adequately sizing the passage clearance, it is possible to obtain the passage of the gaseous fluid heated to the desired temperature for a given crank angle. Furthermore, since compression always takes place under the same conditions, i.e. always on air and always in the same quantities, as it is not partialised, once the temperature has reached full speed, leakage will always occur at the same crank angle. In this way it was possible to control with sufficient precision the instant of the start of combustion.

Therefore, combustion in the first zone occurs on a homogeneous mixture and is triggered at a relatively low temperature, so as to minimize the formation of NO _X . By starting combustion as close to TDC as possible, combustion will continue simultaneously with expansion, so as to avoid a large temperature rise. The continuation of combustion in the second zone will take place when the expansion phase has already started and therefore, also in this case, the danger of NO _X formation will be averted.

The equivalence ratio of the mixture, which is voluntarily kept low (preferably less than 1) , together with the homogeneity of the charge, guarantees a very low formation of particulate matter.

The application of said method will allow to realize a device, compliant with claim 2.

Said device is a compression ignition endothermic engine, of the type comprising a piston, sliding in a cylinder, which drives a connecting rod-crank mechanism, said endothermic engine being of the HCCI (Homogeneous Charge Compression Ignition) combustion type , in which : a recess is made in the piston crown; a protuberance is obtained in the head of said endothermic engine which, due to the ef fect of the stroke of said piston towards top dead centre ( TDC ) , fits into said recess , thus forming : a first zone o f a combustion chamber compri sed between the internal walls of said recess and said protuberance ; a second zone of said combustion chamber, comprised between said pi ston, cylinder, head and sides of said protuberance of the head; said first and second zones of the combustion chamber being placed in communication with each other by an annular duct , included between the sides of said protuberance and the internal walls of said recess and is characterized in that : said first zone of the combustion chamber, in which there is an inj ector, is si zed so as to obtain, by ef fect of the adiabatic heating of an air- fuel mixture , a temperature lower than the ignition temperature of said air- fuel mixture , in so as to prolong the mixing time to obtain HCCI (Homogeneous Charge Compression Ignition) combustion; said second zone of the combustion chamber is si zed so as to have , due to the adiabatic heating of a gaseous fluid present inside said second zone , a temperature higher than the ignition temperature of said air- fuel mixture , at said temperature corresponding to a higher pressure than the pres sure simultaneously existing in the first zone of the combustion chamber, so that said gaseous fluid at a temperature higher than the ignition temperature of the air- fuel mixture can leak, through said annular duct , from the second annular zone to the first central zone of the combustion chamber, igniting the air- fuel mixture present in the first zone of the combustion chamber .

Since the two zones of the combustion chamber communicate with each other, means are obviously provided for making the di f ference in compression ef fective between said two zones of the combustion chamber .

ADVANTAGES OF THE INVENTION

Combining the positive characteristics of the homogeneous charge and the strati fied charge , the engine according to the invention emits a reduced quantity of nitrogen oxides , particulates , unburned hydrocarbons and is not penali zed in consumption, so as to emit even smaller quantities of CO2 .

In particular, when the engine will operate by delivering a small fraction of its maximum power, as usually happens in city use , the quantity of fuel inj ected per cycle is very limited so that the fact of limiting the maximum temperature , with consequent lowering of ef ficiency thermodynamics of the cycle , is compensated by the fact that the trans formation more closely approximates the theoretical isochore , while when the engine operates at speeds suf ficiently higher than idle , the unburned fuel in the first zone of the combustion chamber is proj ected into the second zone where it forms a strati fied charge burning completely . In other words , the fact that there are no hydrocarbons in the flame extinguishing layer on the walls produces the double advantage of using the fuel to the maximum and greatly reducing pollution from unburnt hydrocarbons (HC ) in the exhaust .

With regard to the reduction of the negative ef fects of explosive combustion, it should be noted that this explosive combustion occurs exclusively in the first zone of the combustion chamber . It follows that the interested area in the piston crown is only a small fraction o f the entire area of the piston itsel f . Furthermore , explosive combustion af fects only a minor part of the fuel inj ected when the engine is running at full power .

Ultimately, the engine according to the invention has the following advantages over the high dilution EGR system : lower particulate , NO _X and HC emissions ; minor decrease in thermodynamic ef ficiency and, therefore , lower speci fic consumption, or reduced CO2 production DESCRIPTION OF AN EXAMPLE OF IMPLEMENTATION

The invention will now be described, for illustrative and non-limiting purposes , according to a preferred embodiment and with reference to the attached figures , in which : figure 1 shows a scheme of the engine according to the invention; figures 2 to 11 show the operation of the engine according to the invention .

With reference to the attached figures , ( 1 ) indicates the scheme of a compression ignition engine according to the invention. In Fig. 1 said engine is represented in any position of the cycle, during the compression phase.

Said engine (1) comprises a piston (2) , sliding in a cylinder (3) , which drives a connecting rod (4) and crank

(5) mechanism. A recess (6) is made in the piston crown (2) into which a protuberance (7) of the cylinder head (8) is inserted. In the head (8) there are then obtained one or more intake ducts (9) , closed by as many intake valves (9a) , and one or more exhaust ducts (10) , closed by as many exhaust valves (10a) .

The part included between the piston crown (2) , the internal walls of the cylinder (3) and the lower surface of the head (8) constitutes the combustion chamber (60) .

Following the stroke of the piston (2) towards TDC, said protuberance (7) penetrates said recess (6) , dividing the combustion chamber (60) into two zones ( Figs . 4 to 10) : a first central zone (60a) , included between the internal walls of the recess (6) and the protuberance (7) , which constitutes the central part of the combustion chamber (60) (Fig. 4) , into which the fuel is injected , by means of an injector (11) inserted in the protuberance (7) of the head (8) ; a second zone (60b) , of annular shape, comprised between the piston (2) , the cylinder (3) , the head (8) and the sides of the protuberance (7) , which constitutes the annular zone of the combustion chamber (60 ) .

When the protuberance (7) is inserted in the recess

(6) , an annular duct (6a) is identified, included between the sides of said protuberance (7) and the internal walls of said recess (6) , apt to put in communication between them said central zone (60a) and annular zone (60b) of the combustion chamber (60) .

Starting from the instant in which the protuberance (7) begins to fit into the recess (6) , the first central zone (60a) and second annular zone (60b) of the combustion chamber (60) are in fact formed, therefore it can be define a volumetric compression ratio of said two zones (60a) and (60b) , as a function of the minimum and maximum volumes of said zones (60a) and (60b) . Due to the presence of the annular duct (6a) which connects said two zones (60a) and (60b) , the pressure and temperature inside the two zones (60a) and (60b) will not vary according to a strictly adiabatic law, however, the strong load losses in the passage in the narrow annular duct (6a) will cause significantly different temperatures and pressures to settle in the two zones (60a) and (60b) in any case.

By adequately dimensioning the volumes of said zones (60a) and (60b) it is possible to easily obtain, in the first central zone (60a) , a temperature lower than the ignition temperature of the air-fuel mixture and, in the annular zone (60b) , a temperature higher than the ignition temperature of the fuel-air mixture. This can be achieved with a suitably higher volumetric compression ratio in the annular zone (60b) of the combustion chamber (60) . Obviously also the pressure in the annular zone (60b) will be correspondingly higher than the pressure in the central zone (60a) . The difference in pressures will cause flows between the two zones (60a) and (60b) , as will be better specified below.

Figures 2 to 11 show the operation of the engine (1) according to the invention, for a rotation of the crank (5) of about 100° around the top dead centre (TDC) , during the compression and combustion phases.

In Fig. 2 the piston (2) is moving up in the cylinder (3) towards TDC during the compression stroke, therefore the valves (9a) and (10a) are closed.

In Fig. 3 the piston (2) continues to rise in the cylinder (3) towards the TDC, during the compression phase. Since the piston (2) is close to the TDC, the centripetal squish motions have started so that the injection of fuel can begin since said centripetal motions favour the thickening of the fuel itself in the central zone (60a) of the combustion chamber (60) .

In Fig. 4 the piston (2) continues to rise in the cylinder (3) towards TDC. In this position, identified by a crank angle (a) , the protuberance (7) of the head (8) begins to fit into said recess (6) obtained in the crown of the piston (2) , identifying said annular duct (6a) , between the sides of the protuberance (7) and the internal walls of the recess (6) of the piston (2) . In this way also said central zone (60a) and annular zone (60b) of the combustion chamber (60) are identified. The injection of fuel continues but does not ignite because in the central zone (60a) of the combustion chamber (60) , due to the low compression ratio, the temperature is lower than the ignition temperature of the fuel. In this situation continues the mixing of the fuel with the air so as to obtain a homogeneous or almost homogeneous mixture, said mixing being favoured by the cusp shape (61) of the bottom of the recess (6) .

In Fig. 5, with the piston (2) now approaching TDC, the gaseous fluid present in the annular zone (60b) of the combustion chamber (60) , due to the high compression ratio, has reached a temperature higher than the ignition temperature of the fuel-air mixture. The gaseous fluid superheated at high pressure then propagates towards the central part (60a) of the combustion chamber (60) , through said annular duct (6a) , said propagation however being limited by the load losses in the passage of the gases through the annular duct (6a) .

In Fig. 6 the overheated air continues to seep from the annular zone (60b) to the central zone (60a) of the combustion chamber (60) . Said seepage is increasingly turbulent as the piston (2) approaches TDC, whereby there is rapid mixing with the air-fuel mixture present in the central zone (60a) of the combustion chamber (60) and an explosive combustion of said mixture, said mixing being favoured by the cusp (61) obtained on the bottom of the recess (6) . The sudden increase in pressure almost exclusively concerns the central part (60a) of the combustion chamber (60) , by virtue of the fact that the passage from the central zone (60a) to the annular zone (60b) is slowed down by the passage through the narrow passage (6a) , allowing the system to run out the effects of explosive combustion.

Said explosive combustion takes place in a very short time, but not zero, so the fact that the combustion is triggered before the piston (2) has reached TDC, is equivalent to the ignition advance of a spark ignition engine . In Fig. 7 the piston has reached TDC; following the increase in pressure in the central zone (60a) of the combustion chamber (60) , the burnt gases seep through the annular duct (6a) from the central zone (60a) to the annular zone (60b) of the combustion chamber (60) , where they find overheated air so that the combustion of any fuel present in the burnt gases can continue. This combustion takes place with a stratified charge.

In Fig. 8 the piston (2) starts the stroke towards the bottom dead centre (BDC) ; the flow of burnt gases continues from the central zone (60a) to the annular zone (60b) of the combustion chamber (60) .

In Fig. 9 the piston (2) continues the stroke towards the BDC; the flow of burnt gases continues from the central zone (60a) to the annular zone (60b) of the combustion chamber (60) .

In Fig. 10 the piston (2) continues its stroke towards the BDC, the seep of burnt gases continues from the central zone (60a) to the annular zone (60b) of the combustion chamber (60) . The amount of oxygen present in the central zone (60a) of the combustion chamber (60) is limited, therefore any fuel in excess, present in the flow that crosses the annular duct (6a) , will start to burn as it enters the annular zone (60b) of the combustion chamber (60) in which there is a lot of oxygen. In this situation the pressure existing in the central zone (60a) of the combustion chamber (60) , also due to the load losses in the passage through the annular duct (6a) , is much higher than the pressure existing in the annular zone (60b) of the combustion chamber (60) . The fuel that has not yet been burned will therefore undergo a sudden reduction in pressure which will facilitate its complete evaporation, to the advantage of complete combustion.

In Fig. 11 the piston (2) continues the stroke towards the BDC; in this phase the oxygen present in the central zone (60a) of the combustion chamber (60) is completely exhausted, so that the combustion continues in the annular zone (60b) of the combustion chamber (60) , where there is a significant quantity of oxygen. The fuel projected into the annular zone (60b) of the combustion chamber (60) forms a stratified charge and burns simultaneously as it enters the annular zone (60b) of the combustion chamber (60) . This combustion mode has several relevant advantages.

Firstly, the stratification of the charge ensures that there is no fuel in the flame extinguishing layer on the walls, therefore the presence of unburned hydrocarbons in the exhaust will be greatly reduced.

Secondly, the combustion proceeds gradually, as the fuel enters the annular zone (60b) of the combustion chamber (60) . This fact has two important advantages: the first advantage consists in the fact that combustion occurs during the expansion phase, and this limits the temperature increase; the second advantage consists in the fact that this graduality ensures less noise and less stress on the mechanical parts.

As can be seen from the preceding description, when fuel injection begins (Fig. 3) there is not yet a clear separation between the two zones (60a, 60b) of the combustion chamber (60) , but the fuel thickens in the 1 central zone (60a) of the combustion chamber (60) , because the squish motions have already started.

In the event that the engine operates at very limited power, as normally occurs in the city, due to the reduced quantity of injected fuel, the quantity of oxygen present in the central zone (60a) of the combustion chamber (60) , is capable of oxidizing practically all of the injected fuel.

In the case of progressively higher power delivery, the quantity of fuel injected is correspondingly higher, so that the injection lasts longer, even at the beginning of the expansion phase.

A small amount of air is contained in the small volume of the central zone (60a) of the combustion chamber (60) . Consequently, the evaporation of the fuel removes a high amount of heat, the temperature is limited and so is the formation of nitrogen oxides, also in consideration of the limited amount of nitrogen present in the small volume of the central zone (60a) of the combustion chamber (60) . Proceeding with the injection, the piston (2) begins the expansion stroke. This fact causes the adiabatic expansion of the gases contained in the cylinder, with a consequent reduction in temperature, so that the continuation of combustion also takes place at a reduced temperature, thus limiting the formation of nitrogen oxides.

According to a preferred embodiment of the invention, said annular duct (6a) , connecting the first and second combustion chambers, is particularly thin, so as to introduce high load losses as the burnt gases pass from the first to the second zone (60a, 60b) of the combustion chamber (60) . This means that in the instant in which, due to the effect of the piston stroke, there is no longer any separation between the two zone of the combustion chambers, there is a sudden drop in pressure which facilitates the completion of the evaporation of the injected fuel.

According to a further preferred embodiment, the transversal area of said duct (6a) is equal to or less than the critical section necessary for the velocity of the gases passing through it to be equal to the local speed of sound, so as to create a seal between the two zones of the combustion chamber (60) . In this way, the passage of gases from the first to the second zone (60a, 60b) of the combustion chamber (60) is further limited, so as to make the pressure drop more abrupt and, therefore, the vaporization of the fuel and the completion of combustion. SIZING

A correct sizing of the engine according to the invention provides for its CFD (Computational Fluid Dynamics) modelling; however, some general indications can be given. It is evident that the annular duct (6a) must be as thin as possible, to minimize leakage in the compression phase, therefore a passage clearance of 0.4e0.6 mm can be indicated as a minimum value to avoid interference problems between the edge of the recess (6) and the wall of the protuberance (7) .

As regards the compression ratio of the two parts of the combustion chamber (60) , it should be borne in mind that the following events occur due to leakage: the pressure and temperature in the central zone (60a) of the combustion chamber (60) , will be higher than those calculated taking into account only the volumetric compression ratio, due to the contribution from the annular zone (60b) of the combustion chamber (60) ; the pressure and temperature in the annular zone (60b) of the combustion chamber (60) , will be lower than those calculated taking into account only the volumetric compression ratio, due to the leakage of air from the annular zone (60b) towards the central zone (60a) of the combustion chamber (60) .

From the above it follows that the volumetric compression ratio should be increased by le5 points in the annular zone (60b) and reduced by the same amount in the central zone (60a) of the combustion chamber (60) , with respect to those deemed suitable without said leaks.

The variation of the volumetric compression ratio must be on the minimum values, if the thickness of the annular duct (6a) is on the minimum values, as only a small leakage is allowed. Said variation is obviously on the maximum values in the opposite case.

In practice, the following guidelines can be indicated : volumetric compression ratio of the annular zone (60b) of the combustion chamber (60) equal to 17e21, increased by le5 points; volumetric compression ratio of the central zone (60a) of the combustion chamber (60) equal to 5el0, decreased by le5 points; passage clearance of the annular duct (6a) equal to 0.4e0.8 mm. As far as the sizing of the recess (6) is concerned, it should be noted that this must affect a small area of the piston crown (2) , to minimize the effects of explosive combustion; for this purpose, said area is 8el5% of the total area of the piston crown (2) .

Finally, as regards the crank angle (a) in which the protuberance (7) of the head (8) begins to fit into the recess (6) obtained in the crown of the piston (2) , identifying the annular duct (6a) , can advantageously be 20e25° . CONCLUSIONS

Preferred embodiments of the invention have been described, but naturally they are susceptible to further modifications and variations within the same inventive idea. In particular, numerous variants and modifications, functionally equivalent to the previous ones, which fall within the scope of protection of the invention, will be immediately apparent to those skilled in the art, as evidenced in the attached claims in which the reference signs placed in brackets cannot be interpreted in the sense to limit the claims themselves. Furthermore, the word "comprising" does not exclude the presence of elements and/or phases other than those listed in the claims. The article "a", "an" preceding an element does not exclude the presence of a plurality of such elements. The mere fact that some features are mentioned in different dependent claims does not indicate that a combination of these features cannot be used to advantage .