Field of the invention

The present invention relates to internal combustion engines supplied with gaseous fuel.

Specifically, the invention relates to an internal combustion engine supplied with gaseous fuel, of the type comprising one or more cylinders and respective pistons sliding in the cylinders between a top dead centre (TDC) and a bottom dead centre (BDC) and operatively connected to a crankshaft, said engine being configured to actuate subsequent intake, compression, expansion and exhaust stages during each operating cycle in each cylinder, wherein the engine includes, for each cylinder:

- a first intake duct and a second intake duct opening into the cylinder and both communicating with the same intake manifold, so as to receive air at the same pressure,

- a first intake valve and a second intake valve associated with the cylinder, to control entering into the cylinder of a flow of intake air from the first intake duct and from the second intake duct, respectively, during each operating cycle of the cylinder,

- an actuation device of said first intake valve and of said second intake valve, to control, during each operating cycle of the cylinder, an opening movement and a subsequent closing movement of said first intake valve and of said second intake valve, and

- a gaseous fuel supply system comprising at least one injector associated with each cylinder, for injecting gaseous fuel directly into the respective cylinder or into an intake duct associated with the respective cylinder, during each operating cycle of the cylinder.

Prior art

In engines of the type indicated above there is the problem of avoiding backfiring phenomena in the intake ducts and knocking and/or preignition phenomena in the combustion chamber.

A known solution to this problem consists in cooling the exhaust gases remaining in the cylinder and the hottest surfaces of the combustion chamber, for example those close to the exhaust valves, by using noncombustible gas (fresh air) at the beginning of the intake stroke. This prevents the combustible mixture from being in a high temperature environment, which could generate preignition, and/or that during the combustion triggered by the spark plug electrodes, other ignition points form in correspondence with the so-called “hot spots” of the combustion chamber, giving rise to irregular combustion.

Furthermore, if the fuel is hydrogen, characterized by the ability to bum in the presence of extremely lean mixtures and in very confined spaces (such as the volume defined between the cylinder, the piston rings and the piston), it could happen that the combustion has not finished and could trigger the combustion of the incoming mixture during the intake stroke and the subsequent backfiring towards the intake duct: also in this case, a high motion, in particular a swirl motion, of fresh air could quench these residual combustions before introducing fuel in the intake ducts.

Another method to eliminate the backfiring consists in the direct injection (or DI) of the gaseous fuel into the combustion chamber instead of the port fuel injection (or PFI) when the intake valves are closed: this strategy could avoid backfiring towards the intake manifold, but it may not be sufficient to ensure the quenching of the residual combustions in the gaps and the cooling of the exhaust valves.

Even in the case of ultra-lean combustion, which can be useful for increasing thermal efficiency, both in the case of PFI and in the case of DI, a misfiring phenomenon can occur, resulting in unwanted backfiring or ignition during the subsequent air and fuel intake stroke.

If the gaseous fuel is hydrogen, it should also be considered that the minimum ignition energy is about ten times lower than that of petrol. This characteristic, together with the tendency of hydrogen to bum within wide flammability limits and with high diffusivity, determines a drastic increase in the risk of a backfiring in the intake ducts. Countermeasures normally effective for petrol or natural gas engines may therefore be ineffective in the case of an engine supplied with hydrogen.

Furthermore, the engines supplied with hydrogen need to operate with ultra-lean dosage to avoid NOx emissions and knocking phenomena (due to the low ignition energy of hydrogen). Burning of a lean mixture is generally slower, which means that there may still be some air and unburned fuel in the cylinder when the intake valves open, which leads to backfiring. This condition is more likely to occur if the engine is operating at a low load or with a throttled intake. In this case, there is a lower pressure in the intake ducts than in the combustion chamber and part of the exhaust gases and unburned hydrogen can flow back into the intake ducts.

An attempt to solve this problem has been proposed in the document US 7,980,220 B2. This known solution provides an engine of the type indicated above, with two intake ducts and two respective intake valves for each cylinder. The injection of hydrogen is carried out only in one of the two intake ducts. At least at higher engine loads and speeds, the intake valve, which controls the intake duct into which hydrogen is injected, is opened, during each operating cycle of the cylinder, with a delay compared to the other intake valve. In this way, in the stage in which only the intake valve of the intake duct with which the injector is not associated is open, only fresh air enters the combustion chamber. Theoretically, this solution should eliminate the risk of backfiring, thanks to the cooling obtained with the fresh air introduced in the first part of the intake stroke. However, a solution of this type is not fully satisfactory, due to the fact that the aforementioned cooling is obtained only during a short first part of the intake stroke: in fact, as soon as the second valve is opened, the gas flow introduced by it cancels swirl motion and hot spot cooling.

Documents US 2012/023935 A1 and JP S60 212619 A disclose conventional internal combustion engines, supplied with liquid fuel, in which each cylinder has two intake valves which have different opening times during each operating cycle of the engine.

Object of the invention

The object of this invention is to provide an internal combustion engine supplied with gaseous fuel in which the risk of backfiring is drastically reduced.

A further object is to provide an internal combustion engine supplied with gaseous fuel, in particular hydrogen, in which the risk of knocking and preignition (so-called “mega-knock” phenomenon) is also drastically reduced. Specifically, the invention aims to provide an internal combustion engine supplied with gaseous fuel which is capable of:

• obtain effective cooling of hot spots in the combustion chamber (so as to avoid knocking and mega-knock);

• prevent a flow of fuel to the intake manifold (so as to avoid backfiring).

Furthermore, a further object of the invention is to provide an internal combustion engine supplied with gaseous fuel in which NOx emissions are significantly reduced, also making it possible to operate with ultra-lean dosage, exploiting an increase in turbulent kinetic energy in the cylinder, resulting in better homogeneity of the fuel-air mixture.

Summary of the invention

In view of achieving one or more of the aforementioned objects, the invention relates to an internal combustion engine supplied with gaseous fuel, of the type indicated at the beginning of the present description and further characterized in that:

- the actuation device of the first intake valve and of the second intake valve is configured to control, in each operating cycle of the cylinder, at least under certain operating conditions of the engine, the first opening movement and the subsequent closing movement of said first intake valve only, while the second intake valve remains closed, and then an opening movement and a subsequent closing movement of the second intake valve only, while the first intake valve remains closed, and

- the gaseous fuel supply system is configured in such a way that:

- in each operating cycle of each cylinder, no gaseous fuel is supplied when the first intake valve is open,

- in each operating cycle of the cylinder, gaseous fuel is supplied directly into the cylinder or into said second intake duct associated with the cylinder, only when the second intake valve is open.

Thanks to the characteristics indicated above, for the entire opening stage of the first intake valve, the cylinder receives only fresh air. The swirl motion of the air flow introduced into the cylinder during the opening of the first intake valve has a strong positive effect on cooling the hot spots in the combustion chamber, reducing the tendency for knocking. The reduction of the risk of knocking allows to operate with a higher compression ratio, which gives a positive benefit on the indicated thermal efficiency. The same intense swirl motion also has the value of quenching any residual combustion of hydrogen trapped in the gaps between the piston and the cylinder, avoiding backfiring and/or preignition phenomena.

In the case of port fuel injection (PFI), the high pressure gradient generated by the opening of the second intake valve after the closing of the first intake valve avoids the back-flow of gaseous fuel into the intake manifold and reduces the risk of backfiring.

Instead, in the case of direct injection of the gaseous fuel into the combustion chamber, the opening in subsequent stages of the first intake valve and of the second intake valve causes a combination of swirl and tumble motions of the air flow which reduces the risk of a back-flow of the gaseous fuel in the supply ducts, resulting in a reduction of the risk of backfiring.

Furthermore, the swirl motion, which can be modulated according to the engine operating point, at the end of the intake stroke, thanks to the opening in separate and subsequent stages of the two intake valves, has significant effects on the homogeneity of the air/fuel mixture, resulting in a reduction of NOx emissions, and on the reduction of thermal losses during combustion (on the contrary, a too high swirl, during combustion, increases thermal losses due to the contact of the flow with the wall of the combustion chamber).

Studies and experiences of the Applicant have shown that the opening, at subsequent times, of the two intake valves of each cylinder of the engine generally allows to modulate the intensity of the macro-motions organized inside the combustion chamber, i.e. the so-called swirl, (flow rotating around the cylinder axis), tumble (flow rotating around an axis orthogonal to the cylinder axis), cross-tumble (flow rotating around an axis orthogonal to the cylinder axis and orthogonal to the of tumble) motions, so as to favor the homogeneity of the charge in the case of fuel injection directly into the combustion chamber with benefits on NOx emissions. The aforesaid advantages derive from the fact that the two intake ducts flow into the cylinder in tangential directions with respect to the cylinder axis. In the initial stage of the intake stroke, which begins when the piston is approximately at TDC, only the first intake valve is opened. The air flow introduced into the cylinder through the first intake duct generates a fluid-dynamic field including a high swirl component, as well as tumble and cross-tumble components. After the first intake valve closes, which occurs before the piston has completed 2/3 of its intake stroke, the second intake valve is opened, combined with fuel injection into the intake duct; one gets a double effect:

1 ) a flow field is generated, which contrasts the component of the flow field of the swirl generated by the opening of the first intake valve, so that the pre-existing swirl is attenuated or even reversed in its direction of rotation. This phenomenon is particularly desired at higher engine loads, in fact at higher loads the temperature of the walls of the combustion chamber and of the exhaust valves is higher; the high swirl is also desired in the presence of mixtures for which a low ignition energy is sufficient to trigger combustion, for example in the presence of hydrogen. The high intensity swirl generated by opening the first valve is therefore desired as it effectively cools these hot surfaces, with the aim of avoiding knocking and megaknock. The delayed opening of the second valve has the purpose of attenuating the intensity of the previously generated swirl to avoid excessive heat exchanges during combustion. In fact, one of the problems associated with a flow field with an intense swirl component consists in the fact that the swirl is not very sensitive to the position of the piston: while the tumble motions are converted into turbulent kinetic energy as the piston moves towards TDC, the swirl and the kinetic energy associated with it continue to exist even after TDC and are the cause of an higher heat dissipation during combustion, through the walls of the combustion chamber. Unlike the turbulent kinetic energy, which is desired to increase the combustion propagation rate, the kinetic energy associated with the swirl is therefore undesired and the attenuation of the swirl, obtained with the opening of the second intake valve, is instead desired. Furthermore, the transformation of the swirl into a macro-motion organized with a tumble component (or crosstumble, but in any case around an axis orthogonal to the axis of the cylinder) allows the transfer of kinetic energy from the swirl to the tumble with the possibility of converting the turbulent kinetic energy, useful for increasing the combustion rate, when the piston is close to TDC and supporting the combustion of ultra-lean mixtures, particularly desired if the fuel is hydrogen. 2) due to the depression generated in the cylinder during the angular interval in which both intake valves were closed, when the second valve opens, a high instantaneous flow is generated from the corresponding intake duct towards the cylinder: injecting the fuel into the intake manifold during this interval, or even a few moments before opening, all the fuel, mixed with the air, is sucked into the cylinder, without the risk of any reflux towards the intake manifold. In this way, thanks to the high relative rate of the incoming air/fuel mixture with respect to the gases present in the cylinder, the risk of fuel reflux towards the intake manifold and subsequent backfiring is avoided. Furthermore, thanks to the high traversing rate of the air-fuel mixture of the valve, an extremely homogeneous mixture between air and fuel is also obtained, with enormous benefits on NOx emissions.

In one embodiment, the engine includes a variable actuation device of the first intake valve and of the second intake valve which is configured to achieve, under different operating conditions of the engine, either a first operating mode with the opening, at subsequent times, firstly of the first intake valve only and then of the second intake valve only, or a second operating mode with the opening of the first intake valve and the second intake valve at crank angles identical or substantially identical to each other and the closing of the first valve intake and the second intake valve at crank angles identical or substantially identical to each other, or a third operating mode, with the opening and the closing of the second intake valve only.

The variable actuation device of the intake valves can be an electronically controlled hydraulic actuation device or a device with electromagnetic or electro-pneumatic actuators or a device including actuation cams having different profiles which can be activated selectively.

The invention also relates to the method for controlling the engine described above.

Further advantageous characteristics of the invention are indicated in the attached claims.

Detailed description of the invention

Further characteristics and advantages of the invention will emerge from the following description with reference to the attached drawings, provided purely by way of non-limiting example, in which: figures 1 , 2 are a perspective view and a plan view of the combustion chamber of a cylinder of an internal combustion engine according to the invention, figure 3 is a perspective view of a conventional type actuation system for the intake valves associated with a cylinder of a conventional internal combustion engine, figure 3A illustrates a variant of figure 3, corresponding to an embodiment of the motor according to the invention, figures 4-6 are diagrams illustrating various examples of opening and closing cycles of the two intake valves associated with a cylinder of the engine according to the invention, compared with a conventional cycle, figures 7-11 are diagrams illustrating the advantages of this invention, and figure 12 is a diagram of a variable actuation system of the intake valves of the known type developed by the Applicant and marketed under the MultiAir trademark, which can be used in an embodiment of the invention.

With reference to figures 1 , 2, the reference number 2 indicates as a whole the combustion chamber associated with a cylinder 1 of an internal combustion engine supplied with hydrogen, according to the invention.

Figures 1 , 2 show two intake ducts 3A, 3B, opening into the combustion chamber 2 and shaped and directed according to any known technique, so as to introduce respective air flows into the cylinder, in directions spaced apart from the cylinder axis C1 . The two ducts 3A, 3B are both in communication with the same intake manifold 30 (illustrated only partially), and therefore receive air at the same pressure from the air supply line to the engine. Figures 1 , 2 also show two exhaust ducts 4A, 4B associated with the cylinder 1 and converging in an exhaust manifold 5 (partially visible in figure 2) of the internal combustion engine.

The engine according to the invention is provided with a hydrogen supply system of any known type. The hydrogen supply system comprises a hydrogen injector H2 for each cylinder, shown in Fig. 2, which is arranged and configured to supply hydrogen into the intake duct 3B. The injector H2 is shown only schematically in figure 2, as it can be made according to any known technique. In the drawings only the parts which are relevant for the purposes of this invention are illustrated, it being understood that the structure and the general configuration of the engine can also be made in any known way.

Associated with the two intake ducts 3A, 3B are two intake valves of the conventional mushroom type, with stem and circular head: a first intake valve VA, and a second intake valve VB.

As will be further illustrated hereinafter, for the purposes of this invention, the actuation system of the intake valves VA, VB can be of any known type. Purely by way of example, figure 3 shows an actuation device for the intake valves VA, VB of the conventional type, comprising a camshaft 6 rotatably supported in the cylinder head structure of the engine and controlled to rotate, in the conventional manner, by a drive device (such as a toothed belt drive device) from the crankshaft (not shown) of the internal combustion engine. The camshaft 6, of which only a portion is shown in figure 3, comprises two cams 6A, 6B for operating the first intake valve VA and the second intake valve VB, respectively. In the conventional example illustrated in figure 3, the two cams 6A, 6B actuate the two valves VA, VB by means of respective rocker arms 7A, 7B, each of which has one end mounted in an oscillating manner on a support 8 carried by the cylinder head structure and the opposite end acting on the respective intake valve.

In the case of a first embodiment of the invention, the cams 6A, 6B are modified in the manner illustrated in figure 3A, in order to obtain lift profiles of the type illustrated in any of the figures 4-6.

Figure 4 shows the lift profiles of the two intake valves VA, VB according to a first embodiment of this invention. The diagram in figure 4 shows the displacement of each intake valve as a function of the engine rotation angle. In the convention used here, an engine rotation angle of 360° corresponds to the condition in which the piston inside the cylinder is at TDC. The piston position at BDC corresponds to a crank angle of 540°.

In figure 4, the line LC illustrates the lift diagram of the intake valves in the case of a conventional engine, equipped with conventional cams. In the case of the conventional solution, the two intake valves VA, VB are controlled simultaneously and in synchronism according to the profile LC. As can be seen, in the conventional solution the two valves begin to open just before TDC, reach the condition of maximum opening around a crank angle close to 470° and are closed again near a crank angle equal to 600°. To obtain this result, the two cams 6A, 6B have an identical profile, such as to generate the lift profile LC and also have an identical angular position on the cam shaft 6, as visible in figure 3.

In the embodiment of the invention which is illustrated in figure 4, the cams 6A, 6B have different configurations (as in figure 3A) and are angularly oriented in different ways on the camshaft 6. The conformation and orientation of the two cams 6A, 6B is such as to produce, for the intake valves VA, VB, the lift profiles indicated respectively by LA and LB in figure 4 (it should be noted that the lift profiles LA and LB shown in figure 4 and following have qualitative value only)

The first important characteristic to observe is that during each operating cycle of the cylinder, firstly an opening and closing movement of only the first intake valve VA is activated, while the second intake valve VB is kept closed, and then an opening and closing movement of only the second intake valve VB is activated, while the first intake valve VA is kept closed.

The second important feature to note is that the hydrogen supply system is controlled such that, in each operating cycle of each cylinder, no hydrogen is supplied when the first intake valve VA is open, such that the air flow introduced into the cylinder with the opening of the first intake valve VA is used, exploiting a swirl motion of the air flow, to cause a cooling of the cylinder such as to avoid subsequent preignition or backfiring phenomena, while, in each operating cycle of the cylinder, hydrogen is supplied by the injector H2 into the intake duct 3B associated with the cylinder, only when the second intake valve VB is open. In figure 4, the activation stage of the hydrogen injector H2 is represented by the dotted line H2.

In this way, the flow of only fresh air which enters the cylinder during the opening of the first intake valve VA produces an effective cooling of the combustion chamber, that allows to obtain a regular combustion, without preignition phenomena or backfiring phenomena, thanks to the cooling obtained in the opening stage of the first intake valve.

Figure 4 refers to a particularly preferred embodiment in which the first intake valve VA starts to open when the piston in the cylinder is near TDC (or is in the vicinity thereof, just before or just after TDC) and it is then closed before a crank angle of 540°, i.e. when the piston in the cylinder is still moving in the direction of the BDC, and has not yet reached the BDC. The second intake valve VB, on the other hand, is opened when the piston in the cylinder is near BDC (or is in the vicinity thereof, just before or just after the BDC) and is closed after a further rotation of the crankshaft, e.g. equal to about 90°.

Again with reference to figure 4, according to the preferential solution, the closing of the valve VA occurs when the piston is in the middle of its stroke, while the opening of the valve VB occurs at the BDC. Figures 5, 6 are diagrams similar to that of figure 4 which illustrate further embodiments of the invention, which differ from the example of figure 4 in the shape of the lift profiles.

All the above embodiments have in common the fact that during the intake stroke in the cylinder there is a first time in which substantially only the first intake valve VA is open, while the second intake valve VB is kept closed, while in a second time only the second intake valve VB is open, while the first intake valve VA remains closed.

When air coming only from the intake duct 3A is introduced into the cylinder, the introduced air flow gives rise to a fluid dynamic field with a swirl component (flow rotating around the cylinder axis C 1 ).

As previously mentioned, a flow field mainly characterized by swirl motion allows the hot spots of the combustion chamber to be significantly cooled and therefore has an anti-knocking value.

Thanks to the depression generated in the cylinder due to the early closing of the valve VA, the opening of the second intake valve VB determines an energetic entry of air into the cylinder from the second intake duct 3B: the gaseous fuel injected into the corresponding intake duct 3B, during the opening of the valve VB, is instantly transported inside the cylinder by the air flow. By virtue of the high instantaneous flow entering the cylinder, the permanence of gas in the same intake duct 3B after the closing of the valve VB is avoided, and therefore the risk of backfiring is avoided.

As already indicated above, the invention can be implemented both with an internal combustion engine having an actuation device of the conventional type intake valves, in which the lift profiles of the two intake valves are fixed and predetermined, and with internal combustion engines equipped with variable actuation systems of the intake valves.

Thus, for example, with reference to figure 6, the cam 6A which determines the opening and closing of the first intake valve VA can be controlled by a device, for example an electro-hydraulic one, in such a way as to have a profile such as to determine the lift profile indicated with LA1. In fact, thanks to the decoupling of the cam motion from the intake valve motion through a hydraulic means, it is possible to have a not monotonous motion law of the intake valve. However, the same variable actuation system can be used to vary the lift profile of the valve VA, for example according to the profile LA2. Similarly, the actuation device, combined with the cam 6B, can generate a motion law according to the lift profile LB1 shown in Figure 6. However, the engine can be equipped with a variable actuation system which allows to obtain an effective lift of the valve VB according to the profile LB2.

In one example, the invention is applied to an internal combustion engine equipped with a variable actuation system of the engines intake valves of the type developed by the same Applicant and marketed under the MultiAir trademark.

Figure 12 schematically shows an example of the variable actuation system MultiAir. In this case, each of the intake valves VA, VB (Figure 12 shows the device associated with the valve VA) is actuated by the respective cam 6A or 6B by means of an electronically controlled hydraulic device 8. The cam 6 actuates a tappet 9 maintained in contact with the cam 6 by a return spring 10. The tappet 9 is associated with the pumping piston 11 of a master cylinder which transfers pressurized fluid from a chamber 12 to the chamber of a slave cylinder 13 whose piston 14 acts as an actuator of the intake valve VA. The intake valve VA is recalled by a spring 15 towards a closed position of the intake duct 3A. All of the above components are carried by the engine cylinder head structure 16. An electronically controlled valve 17 is controlled by an electronic control unit E. When the electronically controlled valve is in a closed condition, it interrupts the communication between the pressurized fluid chamber 12 and a low pressure environment 18, communicating with an fluid accumulator 19 and with an inlet 20 intended to be in communication with the engine lubrication circuit. If the electronically controlled valve 17 is in the closed condition, the pressurized fluid chamber 12 is isolated, whereby the movements of the tappet 9 imparted by the cam 6 can be transferred, via the fluid in the chamber 12 and the slave cylinder 13 to the intake valve VA. In a condition in which the cam 6 is keeping the intake valve VA open, an opening of the electronically controlled valve 17 actuated by the electronic control unit E causes the discharge of the pressurized fluid chamber 12 and the consequent closing of the intake valve VA due to the effect of the return spring 15. In this condition, the intake valve VA is insensitive to the movements of the tappet 9 imparted by the cam 6.

This description is provided here purely as an indication of the basic operating principle of the MultiAir system. The Applicant has developed various embodiments of this system which have been the subject of various patent publications including those already mentioned above.

It is understood that the invention would also be usable in combination with variable actuation systems of the intake valves of any known type, such as electromagnetic actuation systems, or variable actuation systems of the type comprising multi-profile cams, for example.

In the case of adoption of a variable actuation system it is possible that the operating mode described above, with an actuation, at subsequent times, firstly of the first intake valve only and then of the second intake valve only, is actuated only in correspondence with certain conditions of engine operating conditions, while under other engine operating conditions the two intake valves of each cylinder are controlled in the conventional way, causing them to open and close simultaneously.

Figures 7 - 11 show the main benefits obtainable with the invention compared to a conventional standard actuation where both intake valves open at TDC and close at BDC).

Figure 7 is a diagram showing the variation of the mean value of the turbulent kinetic energy (TKE) in the combustion chamber in the case of the standard embodiment and in the case of the invention. The diagram in figure 7 shows how the delayed opening of the second intake valve generates a new increase in the turbulent kinetic energy near the BDC, so that, even if it dissipates, at the next TDC the TKE value is considerably higher than the standard case.

Figures 8 a, b and c schematically illustrate the different organized macro-motions that take place in the combustion chamber. These figures show a reference system in which the axis X lies in the symmetry plane of the intake valve ducts and is in accordance with the introduction of air into the combustion chamber. Therefore, the motion that lies in the planes normal to the unit vector Y is defined as tumble motion; the one lying in the planes normal to the unit vector X is defined as cross-tumble motion. The one lying in the planes normal to the unit vector Z is defined as swirl. The so-called tumble, cross-tumble and swirl indices are defined as follows:

Tumble Index — UJTumble I CUEngine

Cross Tumble Index ^— CUcrossTumble I CUEngine

Swirl Index — CUswirl I CUEngine where coEngine is the engine rotation speed, WTumbie, wcrossTumbie and coswiri are the average angular speeds of the respective motions [rad/sec].

Figure 9 shows the effect of the opening of the second intake valve on the swirl generated by the previous opening cycle of the first intake valve: the dashed line refers to the trend that the swirl index would have if, after the opening movement of the first intake valve, any opening of the second intake valve did not occur. It can be seen that the swirl intensity would also be considerable during combustion, with an increase in heat exchanges and a worsening of engine efficiency. The solid line shows the advantages of the invention for the swirl: as soon as the opening of the second intake valve begins, the swirl is reduced proportionally to the air introduced during the opening of the second intake valve. It can be deduced that, having an actuation device for the intake valves capable of varying the profile of the opening movement of the second intake valve, it is possible to modulate the intensity of the swirl present in the chamber during combustion. The dash/dot line refers to the standard implementation which obviously does not provide for the formation of swirl.

Figure 10 shows the effect of the opening of the second intake valve on the tumble generated by the previous opening cycle of the first intake valve: note that, at the start of the opening of the second intake valve (which in this example occurs at a crank angle of 500°), there is a significant increase in the tumble index, against a sudden reduction in the swirl index (visible in figure 9). In a similar way to the standard case, starting from a crank angle of 660°, due to the motion of the piston which compresses the tumble, the tumble index decreases to zero, with associated conversion of the kinetic energy into turbulent kinetic energy: see also figure 7: from 660° to 700° of crank angle there is no further dissipation of TKE as the one that dissipates is replaced by that generated by the cancellation of the Tumble.

Figure 11 shows the trend of the Cross Tumble index: similarly to the swirl (figure 9), this motion is not present in the case of standard implementation. And similarly to the case of the Tumble motion (figure 10), the cross-tumble motion is enlivened in correspondence with the opening of the second intake valve following the opening movement of the first intake valve, due to the transformation of the swirl motion. Similarly to the case of the Tumble motion, the cross tumble motion also contributes to supporting the TKE in the range between 660° and 700° of crank angle.

In one embodiment, the first and second intake ducts are sized in such a way that at high engine loads the closing of the first intake valve generates a pressure wave which travels up the first intake duct and passes through the common intake manifold in the second intake duct, so as to maximize the filling of the cylinder.

In a further example, the first and second intake ducts have different diameters and different lengths, chosen in such a way that, under conditions of maximum engine filling and full opening of the second intake valve, the flow of air entering the combustion chamber with the opening of the second intake valve does not cancel the swirl motion of the air flow previously introduced into the combustion chamber with the opening of the first intake valve.

In a further example, the actuation device of said first intake valve and of said second intake valve is configured to control a lift of the first intake valve significantly lower than the lift of the second intake valve, so that the filling of the cylinder is mainly obtained thanks to the opening of the second intake valve: in this motion the intensity of the initial swirl is reduced, but the mass of air and fuel introduced by depression during the second opening is increased. This condition may be desired for engines with high compression ratios which have extremely negative effects on the flow field (therefore on the combustion rate) and at low loads when the heat transferred to the walls is not relevant, but rather the residual swirl is necessary to increase the flow field and the TKE during combustion In the engine according to the invention, the aforementioned advantages add up to the advantage relating to the reduction of the backfiring risk, the reduction of NOx emissions and the reduction of the preignition and knocking risk, which would otherwise occur due to supply with hydrogen.

It should be noted that, as already indicated above, the closing of the first intake valve occurs when the piston has not yet reached the BDC, in particular the preferred embodiment provides for the closing of the first intake valve when the piston is approximately halfway through its intake stroke, whereby the subsequent movement of the piston creates a depression in the cylinder which favors the entry of flow from the second intake duct when the second intake valve is opened. The greater the opening delay of the second intake valve VB with respect to the closing of the first intake valve VA and the greater the depression created in the cylinder following the movement of the piston after the closing of the first intake valve VA and the greater the flow generated with the opening of the second intake valve VB. A greater flow corresponds to a higher momentum, a higher turbulent kinetic energy and a greater rate of the inlet flow: this high rate counteracts the motions already present in the combustion chamber which could favor the reflux of gaseous fuel into the intake ducts.

In the case of a variable actuation device of the intake valves, it is possible to adjust the degree of opening of the second intake valve VB in order to modulate the intensity of the swirl motion in the combustion chamber following the opening of the second intake valve. The modulation of the swirl motion has significant repercussions on the other components of the flow motion in the combustion chamber (tumble and cross-tumble) and also on the instantaneous generation of turbulent kinetic energy, which, in turn, produces an increase in the combustion rate.

Therefore, the optimal swirl intensity value, which can be modulated by adjusting the lift profile of the second intake valve, depends on various factors, in particular: on the gaseous fuel injection system (PFI or DI) and, in the case of direct injection (DI), the injector position in the combustion chamber (lateral or central), the crank angle at which the gaseous fuel injection begins and ends, the engine load and the desired type of combustion, stratified or homogeneous charge. The strong swirl that occurs in the cylinder following the opening of the first intake valve has a drastic cooling effect on the hot spots in the combustion chamber.

In the case of the port fuel injection (PF I), the advantage is obtained that, thanks to the depression in the cylinder, the output flow from the second intake duct is high, which avoids the risk of a backflow towards the intake and the backfiring risk.

In the case of using a variable actuation device of the intake valves, the advantage is obtained that the lift profile of each valve can be “almost rectangular”, with a particularly high valve closing speed, which is particularly important in order to obtain the advantages of this invention.

Both solutions described above, i.e. with the port fuel injection (PFI) or direct injection (DI) are advantageous for several reasons: in the case of PFI, the injection system is simple and low cost, while in the case of direct injection (DI), the gaseous fuel injection system is relatively more complex but is certainly more advisable in the case of a higher specific power of the engine.

Naturally, without prejudice to the principle of the invention, the embodiments and construction details may vary widely with respect to what is described and illustrated purely by way of example, without thereby departing from the scope of this invention, as defined in the appended claims.

Specifically, the invention is valid even if the intake duct is common for the two intake valves; moreover, the injection of fuel into the intake duct can start before the second intake valve opens, even during the final part of the opening of valve 1 .

Again, the fuel injector may be located at a portion of the intake manifold common to both intake valves.