WO2024127136A1

WO2024127136A1 - Internal combustion engine with variable intake valve actuation and engine control method

Info

Publication number: WO2024127136A1
Application number: PCT/IB2023/062008
Authority: WO
Inventors: Raffaele Ricco; Marcello Gargano
Original assignee: C.R.F. Società Consortile Per Azioni
Priority date: 2022-12-13
Filing date: 2023-11-29
Publication date: 2024-06-20

Abstract

In an internal combustion engine, the two intake valves (V1, V2) of each cylinder are controlled by a single cam (14), or by respective cams, of a camshaft (11) by means of a single hydraulic circuit, or respective hydraulic circuits, which can be pressurized or discharged by means of one or more electrically actuated control valves (24), controlled by an electronic controller. The, or each, cam has two lobes (14A, 14B) arranged and configured to tend to cause a first opening period and a second opening period of each of the two intake valves of each cylinder at each revolution of the cam (14). The electronic controller implements different opening strategies of the two intake valves (V1, V2) of each cylinder depending on the engine load and the engine rotation speed. The different strategies correspond to different combinations of presence or absence of the first opening period and presence or absence of the second opening period of the first intake valve and the second intake valve of each cylinder. The strategies may also differ from each other due to a different timing of the second opening period.

Description

“Internal combustion engine with variable intake valve actuation and engine control method”

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Field of the invention

The present invention relates to internal combustion engines of the type comprising:

- one or more cylinders and a piston movable in each cylinder and operatively associated with a crankshaft, wherein each engine cylinder has respective operating cycles comprising an intake stage, a compression stage, an expansion stage and an exhaust stage,

- a first intake valve and a second intake valve associated with each engine cylinder, for controlling a flow of intake air from a first intake duct and a second intake duct, respectively, during the intake stage into the cylinder in each cylinder operating cycle,

- a camshaft driven by the crankshaft, carrying a single cam, or two respective cams, to actuate the two intake valves of each engine cylinder, by means of a common tappet, or two respective tappets,

- wherein the two intake valves of each cylinder are actuated by said common tappet, or by said respective tappets, against the action of respective return springs, by interposition of a common hydraulic circuit, or respective hydraulic circuits,

- wherein said common hydraulic circuit or each of said respective hydraulic circuits includes:

- a pumping plunger (16) associated with each tappet, actuated by the respective tappet and configured to transfer pressure fluid, through a pressure chamber, to two hydraulic actuators respectively associated with the two intake valves of each engine cylinder, or to a single respective hydraulic actuator associated with one respective of the two intake valves of each engine cylinder,

- an electrically actuated control valve adapted to connect said pressure fluid chamber with a low pressure drain channel connected with a fluid pressure accumulator, such that when said control valve is opened, pressure fluid is discharged from the pressure chamber into said low pressure drain channel and the, or each, intake valve, controlled by said hydraulic circuit is closed due to the effect of the respective return spring, independently of the action of the respective cam, said engine further comprising an electronic controller for controlling the electrically actuated control valve associated with said common hydraulic circuit, or each electrically actuated control valve associated with each of the two hydraulic circuits of each cylinder, as a function of a plurality of operating parameters of the engine, including engine load and engine rotation speed,

Engines of this type are described for example in documents EP 0 803 642 B1 , EP 1 555 398, EP 1 508 676 B1 , EP 1 674 673 B1 and EP 2 261 471 A1 of the same Applicant.

Prior art

The Applicant has long developed internal combustion engines including a variable intake valve drive system of the type indicated above, marketed under the “MULTIAIR” brand. The same Applicant is the owner of various patents and patent applications relating to engines equipped with a system of the type specified above.

Figure 1 of the attached drawings shows a sectional view of an engine equipped with the “MULTIAIR” system, as described in the European patent EP 0 803642 B1.

With reference to this figure 1 , the engine illustrated therein is a multicylinder engine, for example a four-cylinder in-line engine, comprising a cylinder head 1. The head 1 includes, for each cylinder, a cavity 2 formed by the base surface 3 of the head 1 , defining the combustion chamber, into which two intake ducts 4, 5 and two exhaust ducts 6 lead. The connection of the two intake ducts 4, 5 with the combustion chamber 2 is controlled by two intake valves 7, of the traditional mushroom type, each comprising a stem 8 mounted for sliding in the body of the head 1 .

Each valve 7 is returned to the closed position by springs 9 placed between an internal surface of the head 1 and an end cup 10 of the valve. The connection of the two exhaust ducts 6 with the combustion chamber is controlled by two valves 70, also of the traditional type, which are associated with return springs 9 for returning to the closed position.

The opening of each intake valve 7 is controlled, in the way that will be described below, by a camshaft 11 mounted rotatable around an axis 12 within supports of the head 1 , and comprising a plurality of cams 14 for actuating the intake valves 7.

Each cam 14 which controls an intake valve 7 cooperates with the plate 15 of a tappet 16 mounted to slide along an axis 17 which, in the case of the example illustrated in the cited previous document, is directed substantially at 90° with respect to the valve axis 7. The plate 15 is returned against the cam 14 by a spring associated with it. The tappet 16 constitutes a pumping plunger slidably mounted within a bushing 18 carried by a body 19 of a pre-assembled group 20, incorporating all the electrical and hydraulic devices associated with the actuation of the intake valves, as described in detail below.

The pumping plunger 16 is able to transmit a thrust to the stem 8 of the valve 7, so as to cause the opening of the latter against the action of the elastic means 9, by means of pressure fluid (preferably oil coming from the lubrication circuit of the engine) present in a pressure chamber C which the pumping plunger 16 faces, and by means of a piston 21 mounted to slide in a cylindrical body consisting of a bushing 22 which is also carried by the body 19 of the pre-assembled group 20.

Still in the known solution illustrated in figure 1 , the pressure fluid chamber C associated with each intake valve 7 can be connected with an exhaust channel 23 by means of a solenoid valve 24. The solenoid valve 24, which can be of any known type, suitable for the function illustrated here, is controlled by electronic control means, indicated schematically with 25, as a function of signals S indicative of engine operating parameters, such as the accelerator position and the number of engine revolutions.

When the solenoid valve 24 is opened, the chamber C is connected with the channel 23, whereby the pressure fluid present in the chamber C flows into this channel and a decoupling of the cam 14 and the respective tappet 16 from the intake valve is obtained 7, which then quickly returns to its closed position by the action of the return springs 9. By controlling the connection between the chamber C and the discharge channel 23, it is therefore possible to vary, as desired, the opening time and stroke of each intake valve 7.

The exhaust channels 23 of the various solenoid valves 24 all flow into the same longitudinal channel 26 connected with pressure accumulators 27, only one of which is visible in figure 1 .

All the tappets 16 with the associated bushings 18, the pistons 21 with the associated bushings 22, the solenoid valves 24 and the relative channels 23, 26 are carried and obtained from the aforementioned body 19 of the pre-assembled group 20, to the advantage of speed and ease of assembly of the engine.

The exhaust valves 70 associated with each cylinder are controlled, in the embodiment illustrated in figure 1 , in a traditional way, by a respective camshaft 28, by means of respective tappets 29, although in principle it is not excluded, in the case of the document mentioned above, an application of the hydraulic drive system also to the control of the exhaust valves.

Again with reference to figure 1 , the variable volume chamber defined inside the bushing 22 and facing the piston 21 (which in figure 1 is illustrated in its minimum volume condition, the piston 21 being in its stroke end upper position) is connected with the pressure fluid chamber C by means of an opening 30 obtained in an end wall of the bushing 22. This opening 30 is engaged by an end nose 31 of the piston 21 in such a way as to achieve hydraulic braking of the movement of the valve 7 in the closing phase, when the valve is near the closed position, as the oil present in the variable volume chamber is forced to flow into the pressure fluid chamber C passing through the gap existing between the end nose 31 and the wall of the opening 30 engaged by it. In addition to the connection constituted by the opening 30, the pressure fluid chamber C and the variable volume chamber of the piston 21 are connected with each other by means of internal passages obtained in the body of the piston 21 and controlled by a non-return valve 32 which allows the passage of fluid only from the pressure chamber C to the variable volume chamber of the piston 21 .

During the normal operation of the known engine illustrated in figure 1 , when the solenoid valve 24 excludes the connection of the pressure fluid chamber C with the discharge channel 23, the oil present in this chamber transmits the movement of the pumping plunger 16, imparted by the cam 14, to the piston 21 which controls the opening of the valve 7. In the initial phase of the valve opening movement, the fluid coming from the chamber C reaches the variable volume chamber of the piston 21 passing through the non-return valve 32 and further passages which connect the internal cavity of the piston 21 , which has a tubular shape, with the variable volume chamber. After an initial movement of the piston 21 , the nose 31 comes out of the opening 30, so that the fluid coming from the chamber C can pass directly into the variable volume chamber through the opening 30, now free.

In the reverse closing movement of the valve, as already mentioned, during the final phase the nose 31 enters the opening 30 causing the hydraulic braking of the valve, so as to avoid impacts of the valve body against its seat, for example following an opening of the solenoid valve 24 which causes the immediate return of the valve 7 to the closed position.

In the system described, when the solenoid valve 24 is activated, the engine valve follows the movement of the cam (full lift). An early closing of the valve can be achieved by deactivating (opening) the solenoid valve 24, so as to empty the hydraulic chamber and obtain the closing of the engine valve by the action of the respective return springs. Similarly, a delayed opening of the valve can be achieved by delaying the actuation of the solenoid valve, while the combination of a delayed opening with an early closing of the valve can be achieved by activating and deactivating the solenoid valve while pushing the relevant cam. According to an alternative strategy, according to the teachings of the patent application EP 1 726 790 A1 of the same applicant, each intake valve can be controlled in “multi-lift” mode, i.e. according to two or more repeated “sub-cycles” of opening and closing. In each sub-cycle, the intake valve opens and then closes completely. The electronic control unit is therefore able to obtain a variation of the opening time and/or the closing time and/or the lift of the intake valve, depending on one or more engine operating parameters. This allows maximum engine efficiency and the lowest fuel consumption to be achieved in all operating conditions.

In the known system described above it can be provided that the two intake valves 7 associated with the same engine cylinder are controlled by a single pumping plunger 16 in turn controlled by a single cam on the engine camshaft.

In this case, if it is desired to actuate the two intake valves of the same cylinder in a differentiated manner, the solution known from document EP 2 693 007 A1 of the same Applicant can be provided, in which the electrically actuated control valve is a three-way and three-position solenoid valve, with an inlet connected to both the pressure chamber and the hydraulic actuator of one of the two intake valves, an outlet connected to the fluid accumulator and a further outlet connected to the hydraulic actuator of the other intake valve (see figure 20 attached here, corresponding to figure 4 of the document cited above).

Alternatively, the further solution known from document EP 3832078 A1 , also from the same Applicant, can be provided, which provides two solenoid valves arranged in series in the connection between the pressure chamber and the hydraulic accumulator and with the two hydraulic actuators of the two intake valves connecting, one with the hydraulic line between the two solenoid valves and the other with the pressure chamber (see figure 20 attached here, corresponding to figure 11 of EP 3 832 078 A1).

However, for the purposes of the present invention, it can also be provided that each intake valve of each engine cylinder is controlled by a respective cam of the camshaft and by a respective hydraulic circuit including a respective pumping plunger, in which case it can have total flexibility in differentiating the openings of the two intake valves of each cylinder.

Document EP 2 796 675 A1 of the same Applicant illustrates a solution according to the preamble of claim 1 , limited to the embodiment with a single pumping plunger for the two intake valves of each cylinder. Other solutions are known from EP 3 832 077 A1 , also by the same Applicant, and from EP 1 063 394 A2

Object of the invention

The main object of the invention is to provide an internal combustion engine of the type indicated at the beginning of this description which is characterized by high combustion efficiency in all engine operating conditions.

In particular, an object of the invention is to provide an internal combustion engine in which the two intake valves of each cylinder can be controlled independently of each other, but with strategies which combined together achieve the maximum advantages in terms of combustion efficiency in all engine operating conditions. Summary of the invention

In order to achieve the aforementioned purposes, the invention has as its object an internal combustion engine having the features indicated in claim 1.

In a preferred embodiment, said electronic controller is programmed to implement, depending on the load and engine revolutions, one or more of the following strategies:

- a first strategy in which both intake valves (V1 , V2) have both said first opening period and said second opening period of the first type (2A)

- a second strategy, in which a first intake valve (V1 ) of each cylinder has both said first opening period and said second opening period of the first type (2A), while the second intake valve (V2) has none of said opening periods,

- a third strategy, in which the first intake valve (V1 ) of each cylinder has both said first opening period and said second opening period of the first type (2A), while the second intake valve (V2) has only said second opening period of the first type (2A),

- a fourth strategy, in which the first intake valve (V1 ) of each cylinder has only said first opening period, while the second intake valve (V2) has only said second opening period of the first type (2A),

- a fifth strategy, in which the first intake valve (V1 ) of each cylinder has both said first opening period and said second opening period of the first type (2A), while the second intake valve (V2) has only said first opening period.

Preferably, said electronic controller is programmed to implement, depending on the engine load and the engine rotation speed, also one or more of the following further strategies:

- a sixth strategy, in which the first intake valve (V1 ) of each cylinder has only the second opening period of the first type (2A), and the second intake valve (V2) has only the second opening period of the first type (2A),

- a seventh strategy, in which the first intake valve (V1 ) of each cylinder has only said first opening period, while the second intake valve (V2) has none of said opening periods,

- an eighth strategy, in which the first intake valve (V1 ) of each cylinder has only said first opening period, and the second intake valve (V2) has only said first opening period,

- a ninth strategy, in which the first intake valve (V1 ) of each cylinder has only said second opening period of the first type (2A), while the second intake valve (V2) has none of the said opening periods,

- a tenth strategy, in which both intake valves (V1 , V2) have both the first opening period and the second opening period of the second type (2B)

- an eleventh strategy, in which a first intake valve (V1 ) of each cylinder has both said first opening period and said second opening period of the second type (2B), while the second intake valve (V2) has none of said opening periods,

- a twelfth strategy, in which the first intake valve (V1 ) of each cylinder has both said first opening period and said second opening period of the second type (2B), while the second intake valve (V2) has only said second opening period of the second type (2B).

The need to provide, albeit in different engine operating conditions, the two different types described above of the second opening period of the intake valves derives from the fact that it is necessary to take into account the different types of engine.

In particular, an engine characterized by a high geometric compression ratio, for example a compression ratio greater than 13, will need to operate, at the highest loads, in operating conditions with an over- expanded cycle (Miller-Atkinson), i.e. with effective compression stroke smaller than the expansion stroke and in this case it will be necessary to implement a second opening period of the aforementioned second type.

In the present description, and in the claims that follow, where it is indicated that a cam has two lobes, it is intended to include both the case in which the cam has a single body configured with two lobes, and the case in which the two aforementioned lobes belong to two different bodies offset from each other in the direction of the camshaft axis and connected in rotation to each other.

According to a further feature of the invention, the electronic controller is programmed to implement:

- said first strategy or said third strategy when the engine compression ratio is less than 11 and the engine load is greater than a first threshold value, - said first strategy or said third strategy when the engine compression ratio is greater than 11 , the engine load is greater than said first threshold value, and the engine is equipped with a water injection system,

- the second strategy, or the third strategy, or the eleventh strategy or the twelfth strategy, if the engine compression ratio is between 11 and 13, the engine load is greater than said first threshold value, and an exhaust gas recirculation (EGR) is provided in the cylinder and/or a fuel injector device is associated with the cylinder in a central position with respect to the cylinder,

- said tenth strategy or said twelfth strategy if the engine compression ratio is greater than 13,

- said fourth strategy or said fifth strategy when the engine load is below said first threshold value and the engine rotation speed is above a threshold value,

- said first strategy or said third strategy when the engine rotation speed is lower than said threshold value and the engine load is between the first threshold value and a further threshold value lower than the first threshold value.

Preferably, the electronic controller is programmed to implement the seventh strategy when the engine rotation speed is lower than the rotation speed threshold value and the engine load is lower than said further threshold value lower than said first threshold value.

The invention also has as its object the engine control method described above.

Detailed description of the invention

Further features 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:

- figure 1 is a sectional view of the cylinder head of an internal combustion engine equipped with an electronically controlled hydraulic system for operating the engine intake valves, according to the prior art illustrated in document EP 0 803 642 B1 and discussed above,

- figure 2 is a schematic view of the variable actuation system of the engine intake valves, according to a first embodiment of the present invention, in which the two intake valves of each cylinder are controlled by two distinct cams, by means of respective tappets, respective pumping plungers and respective hydraulic circuits,

- figures 3, 4 illustrate different opening strategies of the two intake valves of each engine cylinder, and

- figure 5 is a diagram illustrating different engine intake valve opening strategies that are implemented in different areas of the engine load/engine speed diagram, and

- figures 6-8 are diagrams illustrating further strategies for opening the engine intake valves,

- figures 9-19 are diagrams showing the advantages of the invention, with reference to different operating parameters of the engine, and

- figures 20, 21 are diagrams of two known solutions, already mentioned above, for controlling the connection of the hydraulic actuators of the two intake valves of each cylinder with the pressure chamber of a common hydraulic circuit, which are adopted in the case of further embodiments of the invention in which the two intake valves of each cylinder are operated by a single cam, by means of a single tappet, a single pumping plunger and a single hydraulic circuit including a single pressure chamber.

Starting from the known solution described above with reference to figure 1 , the invention allows the provision of a plurality of new opening strategies of the two intake valves of each engine cylinder for the purpose of increasing combustion efficiency, with consequent advantages in terms of fuel consumption, and therefore a reduction in CO2 emissions, and the reduction of harmful exhaust gases, in all engine operating conditions.

Figure 2 shows a diagram of the variable actuation system of the intake valves of each engine cylinder in a first embodiment of the invention.

According to the invention, each engine cylinder has two intake valves V1 , V2.

In this first embodiment, each of the two intake valves V1 , V2 is controlled by a respective cam 14 with a respective hydraulic circuit, including a respective pumping plunger 16, a respective pressure chamber C, a respective hydraulic actuator 21 of the valve intake, a respective two- position solenoid valve 24 capable of controlling the connection between the pressure chamber C and a pressure accumulator 270 which is also in connection with a low pressure circuit of the engine lubricating oil.

Furthermore, in this embodiment, each of the two cams 14 which control the two intake valves V1 , V2 of each engine cylinder is provided with two lobes 14A, 14B, configured to tend to cause two opening periods of the respective valve V1 or V2 at each revolution of the cam 14.

The two lobes 14A, 14B can be an integral part of a single cam body or be part of two separate bodies, coupled onto the camshaft in such a way as to be integral in rotation with each other and with the camshaft. Preferably, the two lobes 14A, 14B are offset from each other in the direction of the camshaft axis and the tappet plate 15 (figure 1 ) is sufficiently extended to be able to cooperate with both lobes.

Thanks to the provision of a cam 14 with two lobes 14A, 14B for each intake valve of each engine cylinder, and thanks to the provision of a respective hydraulic circuit interposed between each cam and each intake valve, both intake valves of the engine can have a first opening period and a second opening period, at a distance from each other, at a conventional opening phase of an intake valve. Figures 3, 4 illustrate twelve different strategies that can be implemented in the engine according to the invention, characterized by different combinations of presence or absence of the first opening period and presence or absence of the second opening period.

Where it is provided, the first opening period of an intake valve of a given cylinder is always of the same type: said first opening period substantially begins when the respective piston is in its TDC and ends when the piston is substantially midway between the PMS and the BDC.

Where it is provided, the second opening period of an intake valve of a given cylinder can be of two types: a first type, referred to below for brevity as “type 2A” and a second type, referred to below for brevity as “type 2B”.

In the case of type 2A, the second opening period of an intake valve of a given cylinder begins when the piston has passed 4/5 of its path from the TDC to the BDC and has not yet reached its BDC, preferably not more than 30° of crank angle before the BDC, and at not less than 20° of crank angle before the BDC. Again in the case of type 2A, the second opening period ends when the piston has passed the BDC and is rising towards the TDC, preferably at least 20° of crank angle after the BDC, and at not more than 30° of crank angle after the BDC.

In the case of type 2B, the second opening period of an intake valve of a given cylinder begins when the piston has passed BDC and is rising towards the TDC, preferably at least 20° of crank angle after the BDC, and at not more than 30° of crank angle after the BDC, said second opening period ending at least 80° of crank angle after the BDC, and at not more than 100° of crank angle after the BDC.

In the same engine, the two different types of the second opening period described above can be obtained in various ways.

For example, it is possible to modulate the actuation of the electrically actuated control valve, to obtain a different lift profile, at the beginning and/or at the end of the second opening period of the intake valve, than that which would be determined by the second lobe of the respective actuating cam. In this case the second lobe is configured with sufficient extent to allow both types of lift in the second opening period. Figure 6 of the attached drawings shows the lift profiles C1 , C2 which would be determined by the two lobes of the cam, corresponding to the first opening period and the second opening period. Thanks to the intervention of the electrically actuated control valve, which can unload the respective hydraulic circuits, in the first opening period the lift profile can become the profile 1 B, while in the second opening period the lift profile can become the profile 2A of the first type or profile 2B of the second type.

Alternatively, to obtain one or the other type of second opening period in the same engine, a timing shifting device can be provided associated with the camshaft which controls the intake valves, capable of delaying the entire lift profile of approximately 40° of crank angle (obviously in this case the first opening would also be translated). This solution is schematized in figure 7, where C is the lift profile that would be determined by a conventional cam (with a single lobe), 1A, 1 B are the lift profiles in the first opening period before and after an intervention of the timing shifting device, and 2A, 2B are the two lift profiles of the second opening period before and after the intervention of the timing shifting device.

Still alternatively, to obtain one or the other type of the second opening period in the same engine, it is possible to use actuation devices of the type known per se in which different cam profiles can be activated selectively. In this way it is possible to provide both types 2A, 2B of the second opening period and possibly also the conventional profile C with a single opening (see figure 8)

In the event that the engine does not require the use of both types of the second opening period, it is obviously possible to optimize the design of the cam in order to limit the consumption absorbed for pressurizing the oil and therefore the cam profile can follow the desired opening profile of the intake valves.

In particular, in addition to making the actuation angles of the pumping cam coincide with those of the intake valves (opening and closing), it is also possible to significantly reduce the maximum lift in the first opening period (as visible in figure 6), even up to a value equal to 1/9-1/10 of the diameter of the circular head of the intake valve. This measure makes it possible to facilitate the production of the double lobe cam.

In the case of strategy I, both intake valves V1 and V2 have both the first opening period and the second opening period, said second opening period being 2A.

Thanks to the possibility of unloading the hydraulic circuit associated with each intake valve, for each of the two intake valves it is possible to completely cancel the first opening period or the second opening period, or both opening periods.

With reference to figures 3, 4, as already indicated, strategy I provides that both intake valves V1 , V2 of each engine cylinder have both the first opening period and the second opening period of type 2A.

Strategy II provides that the first intake valve V1 has both opening periods, with the second period 2A, while the second intake valve V2 always remains closed.

Strategy III provides that the first intake valve V1 has both opening periods, with a second period 2A, while the second intake valve V2 has only the second opening period 2A.

Strategy IV provides that the first intake valve V1 has only the first opening period, while the second intake valve V2 has only the second opening period 2A.

Strategy V provides that the first intake valve V1 has both opening periods, with a second period 2A, while the second intake valve V2 has only the first opening period.

Strategies X, XI and XII are similar to strategies I, II and III respectively, from which they differ in the second opening period which is 2B.

Strategies l-V, X and XII illustrated above are those considered most important for achieving the purposes of the invention.

However, preferably, the engine of the invention is programmed to also be able to select any of the further strategies VI -IX and XI illustrated in Figure 4.

Strategy VI provides that the first intake valve V1 has only the second opening period and that the second intake valve V2 has only the second opening period of type 2A.

Strategy VII provides that the first intake valve V1 has only the first opening period and that the second intake valve V2 always remains closed.

Strategy VIII provides that the first intake valve V1 has only the first opening period and that the second intake valve V2 has only the first opening period.

Strategy IX provides that the first intake valve has only the second opening period 2A and that the second intake valve V2 always remains closed.

In the diagrams of figures 3, 4, for each strategy and for each intake valve, exemplary values of the crank angles are shown, at which the start of the first opening period of the intake valve (Intake Valve Opening 1 = IVO1 ), the end of the first opening period (Intake Valve Closing 1 = IVC1 ), the beginning of the second opening period (Intake Valve Opening 2 = IVO2) and the end of the second opening period (Intake Valve Closing 2 = IVC2) occur. Consider that a crank angle of 360° corresponds to the TDC and a crank angle of 540° corresponds to the BDC.

As shown, the specific crank angle values indicated in the diagrams of figures 3, 4 are exemplary. Each of these values can fluctuate within a range of +/- 20° compared to the indicated values.

The advantages of the proposed invention consist of:

- in the possibility of providing opening periods of the intake valves with reduced angular extent (less than 120° of crank angle),

- in the possibility of providing opening and closing profiles of the intake valves characterized by steeper rising and falling sides compared to a traditional system with valve actuated directly by cam,

-in the possibility of providing two opening cycles of the same intake valve, substantially distinct from each other, in an angular duration period less than 320° of crank angle,

- in the possibility of modifying and/or modulating the opening profile of the intake valves at each engine cycle: this advantage is optimized if the combustion chamber is equipped with a pressure transducer so as to be able to control the filling of the cylinder and the tendency to closed loop knocking; the control system is able to correct the crank angle at which the second opening period ends, so as to allow air to flow back towards the intake manifold and keep the maximum temperature, at the time of ignition, within a limit value

Some of the objectives that can be achieved with the strategies described above are set out below

Objective 1 : promoting ultra-lean combustion both by increasing the amount of intake air and by increasing turbulence. As it is known, the combustion of a highly lean mixture (lambda>1.7) has the advantage of significantly reducing the losses of heat transferred to the walls and thus increasing efficiency: independently of the type of fuel, this mode is always desired, but, due to the amounts of air required (almost double compared to stoichiometric operation) it can only be achieved up to medium loads. Necessary conditions to be able to achieve this mode consist in the need for greater turbulence during combustion (this condition favors the combustion rate which cannot last long so as not to compromise the efficiency and stability of the engine’s operation) and in a higher temperature of the air-fuel mixture at the end of compression (this condition favors ignition).

To increase turbulence, the air motions and the turbulence generated during the loading of the air (and any EGR) into the combustion chamber are crucial, while the increase in temperature can be induced by the increase in the compression ratio (for example, for an aspirated engine, with values greater than 13). The ignition system must also be adequate to facilitate the ignition of the ultra-lean mixture, in particular due to the greater turbulence, the electric arc generated must be more powerful, but this is not the subject of the patent. The maximum amount of air that can be sucked in with the same engine layout is also relevant: assuming, for example, that the engine is aspirated, the higher the amount of air that the combination of the geometry of the intake ducts with the opening law of the valves allows sucking, the higher the engine load at which operation in ultra-lean mode is permitted: according to the prior art, modern ignition engines use high turbulence ducts, often combined with valve masking, to promote both rate than the stability of combustion, but these ducts worsen the filling of the cylinder. As will be shown, strategies I and III allow both increasing turbulence and increasing the mass of intake air and are therefore optimal for operation in ultra-lean mode.

Objective 2: increasing efficiency at maximum load by reducing the risk of knocking and supporting the use of high CR, possibly in combination with EGR. As it is known, the efficiency of a spark ignition engine is proportional to the value of the compression ratio: however, high CR values lead to the onset of knocking at higher loads: in order to avoid knocking, maintaining high expansion ratio values, it is necessary to differentiate the compression ratio from the expansion ratio. The cycle described is defined as an over-expanded cycle (Miller-Atkinson) and is obtained either by anticipating the closing of the intake valves, or by delaying them: the final effect is that the pressure, and therefore the temperature of the air-fuel mixture and any EGR (which being inert has an anti-knocking effect) is identical to the case of the engine with a reduced compression ratio ((the tendency to knocking is proportional to the end compression temperature). The main disadvantage of over-expanded cycles consists in a drastic reduction in turbulence which precludes combustion stability or in any case determines a slowing-down to the point of worsening the overall operating efficiency. As will be shown, implementation strategies X and XII allow supporting the implementation of over-expanded cycles without have a worsening of the rate/duration/robustness of combustion.

Objective 3: supporting stratified combustion. Stratified combustion, often used to support the combustion of ultra-lean mixtures (at medium loads) or with high EGR rate (at high loads), can be achieved with a fuel injector, located in a central position in the combustion chamber, which perform at least one injection when the piston is near the TDC, just before ignition. A flow field with a high swirl index is desired and favors this type of combustion; however, a high swirl index, outside of this operating mode, could increase heat transfer to the walls and is undesirable. The possibility of implementing this stratified combustion mode only when required is made possible by implementing strategies II, III, XI and XII, in which operation is also optimized thanks to the possibility of modulating the intensity of the swirl.

Objective 4: promoting air-fuel mixing at higher engine speeds. Due to the more stringent emission regulations and due to the need to use injection systems that introduce the fuel directly into the combustion chamber (to exploit its anti-knocking properties and at the same time guarantee the correct cylinder/cylinder dosage), in particular, at higher engine revolutions, when the time to generate the mixture is reduced, it is necessary to adopt strategies that favor charge homogeneity. The spark ignition engines in production today mainly base the flow field only on tumble motion. Inversions of the main flow field, during the same combustion cycle, for example from tumble to swirl, and higher values of the instantaneous flow rate of air entering the cylinder (this also instantly induces greater turbulence), greatly favor the homogeneity of the charge: it will be shown that strategies IV and V allow these flow field reversals as well as high instantaneous flow rate values.

Objective 5: promoting cold engine starting and promoting warm-up of the exhaust gas treatment system. As it is known, the greatest amount of emissions is produced during engine starting, especially in the presence of particularly cold atmospheric conditions: this is due both to the inefficiency of the injector spray (especially when the injection system is directed into the chamber) with accentuated wall impingement phenomenon and is due to the inefficiency of the exhaust gas treatment system which, to be operational, requires working above a temperature threshold. These considerations are further aggravated when the type of fuel is characterized by lower vapor pressures (for example methanol). One of the methods adopted today to raise the temperature of the exhaust gases and accelerate the warm up of the catalyst consists in delaying the center of gravity of combustion, but this leads to high instability in the operation of the engine, also due to the reduced turbulence. It will be demonstrated that strategies VI and IX allow to maximize the evaporation phase of the spray as well as to increase the temperature of the exhaust gases in order to reduce the warm up times of the catalyst.

Objective 6: avoiding backfiring phenomena if the fuel is characterized by reduced ignition energy (for example hydrogen or a mixture of hydrogen and natural gas). Some fuels, in particular hydrogen, are characterized by a reduced ignition energy: this entails the risk of unwanted ignition due, for example, to the presence of hot spots in the combustion chamber, even with open intake valves and consequent backfiring in the intake manifold, especially if the injector type is port fuel. To avoid this phenomenon and also to maximize charge homogeneity, strategies III, IV and XII, with fuel injection during the second opening period is desired (the initial swirl cools the hot spots).

Objective?: increasing power for high-performance engines. In the case of an high performance engine, which therefore needs to maximize the mass of air and fuel, in stoichiometric conditions, when it operates at maximum load, it is possible to combine strategies I and III, or strategies X and XII in the case of an engine characterized by a high compression ratio (>11 ) in combination with water injection with anti-knocking value, as well as possibly the intake charge cooling by means of a heat exchanger. The water injection system can be either of the direct injection type in the combustion chamber or with a water injector placed on the intake manifold (in the case of strategies III and XII the injector is placed on the intake manifold corresponding to the valve which is opened). Furthermore, it is possible to activate this closed loop water injection system by introducing a pressure transducer into the combustion chamber and programming the electronic controller so that the temperature of the gases at the time of ignition does not exceed a predefined threshold (the temperature being proportional to the pressure value in the chamber).

To understand how the strategies presented allow the achievement of the objectives listed above, consider, by way of example (similar considerations are possible for all other strategies), the comparison between strategy III and conventional implementation.

The comparison focuses in particular on the quantities that define the flow field and the turbulence, so in figures 9-14 the following are shown, respectively: the instantaneous flow rate entering the cylinder, the mass of air (and any fuel or EGR) in the cylinder, the Swirl, Tumble and CrossTumble indices, the turbulent kinetic energy (TKE). The curves reported derive from CFD analyzes relating to a reference engine with a displacement of 500 cc and a compression ratio of 10: the following considerations are intended to have a qualitative value, the numerical values of which strongly depend on the type of intake ducts of the engine analyzed, but are well suited to highlight the advantages of the proposed solution.

In summary:

- Higher intake mass compared to the conventional reference case and high propensity for charge homogeneity: observing the graph relating to the instantaneous flow rate, during the first opening period of strategy III, the instantaneous flow rate is identical to that of the reference (dashed), this because during the first half of the piston stroke the intake mass depends on the opening and closing angle rather than on the valve lift. Furthermore, although according to this strategy only one valve opens, thanks to the reduced speed values, the air is comparable to an incompressible fluid and therefore for the same piston stroke the same intake mass will correspond (except for losses due to outflow coefficients, however reduced): this is also confirmed by observing the graph which shows the intake mass into the cylinder. Due to the depression caused by the closing of the intake valves during the second part of the piston stroke from TDC to BDC, when the second intake period opens, a high instantaneous flow rate occurs, with a maximum value triple compared to the maximum value of the reference. Thanks to the inertia of the air in the ducts due to such a high instantaneous flow rate, a greater amount of air is sucked in than the reference (see the graph of the intake mass).

- Flow field with multiple components, mixture homogeneity and antiknocking properties: observing the graphs of the intensity indicators of the vortex, swirl, tumble and cross tumble macro-motions, it is evident how the opening of the valve alone during the initial period generates an intense swirl flow field (which could be useful for cooling hot spots), while the high instantaneous flow rate that occurs when the piston is at BDC involves the generation of a high tumble motion, also with the maximum intensity tripled compared to the reference case. In addition to the swirl and tumble components, now, compared to the reference case, the presence of the cross tumble component is also notable: the fact that the flow field is characterized by all three components favors, among other benefits, greatly the homogeneity of air-fuel mixing. Other advantages derive from the greater kinetic energy associated with the vortices which transforms into TKE during the final compression stage.

- High TKE peak optimal for charge homogeneity: observing the graph showing the trend of the TKE, it is noted that the maximum value is equal to 10 times the maximum value of the reference: obviously timing the fuel injection with the angle in which you have the maximum TKE (i.e. at BDC) you can further optimize the air-fuel homogeneity. It should be noted that, while the maximum TKE in the reference case occurs when the piston is midway through the stroke, the fact that thanks to strategy III the maximum TKE corresponds to the BDC reduces the risk that the fuel spray, which must appropriately be injected in conjunction with the maximum TKE, impacts the piston which is located at the furthest distance.

- Greater TKE, compared to the reference, during optimal ignition for fast combustion: looking again at the TKE diagram, in particular the box showing the zoom at the TDC (ignition angle), we note that strategy III allows more than doubling the TKE value compared to the reference to 700CA. This translates into a potential for greater combustion rate, which is particularly necessary in the presence of a diluted mixture with excess air or EGR. Greater TKE at start-up also allows for a reduction in knocking at high loads for stoichiometric mixtures.

Figures 15-19 show graphs comparing the most significant values relating to TKE@700CA, Flow Rate, max Swirl Index, max Tumble index, Temperature@700CA as a function of the air mass in the cylinder: the graphs show the values of some strategies according to the invention compared with those of the reference engine at 2000rpm, Full load and 2000x7bar. In summary:

• The TKE@700CA diagram shows that, with the exception of strategy VII, for all the others reported the value is significantly greater than the reference; in particular, strategy X, with the same mass of the full load reference, has an almost triple TKE. The same diagram shows that strategies I and III allow introducing 20% more mass than the full load reference.

• The diagram of the maximum instantaneous flow rate, proportional to the maximum TKE, a determining parameter for obtaining good air/fuel homogeneity, strategies I, III and X have values almost triple compared to the reference. In any case, all the other strategies, except VII, are characterized by higher values than the reference.

• The diagram of the maximum value of the swirl, relevant both because the swirl allows cooling of the combustion chamber with antiknocking value, and because the swirl enables stratified combustion (in the presence of an injector located in a central position), shows the maximum value relating to strategy IX, and significant values for strategies II, III, IV, V

• The diagram of the maximum value of the tumble, relevant both because during compression the tumble supports the TKE (therefore an indicator of fast combustion and anti-knock), and because it is an index of mixing and therefore of homogeneity of the charge: all the strategies, except VII, are characterized by a Tumble greater than the reference. In particular, strategy IX is characterized by the highest tumble value: the same strategy also has the maximum value for swirl and is therefore suitable for supporting fuel sprays with low vapor pressure

• The diagram of the temperature value at 700CA shows that strategies I and III involve higher values compared to the reference case, therefore they could lead to the onset of knocking in the case of a stoichiometric mixture, but on the contrary they favor combustion in the case of an ultra-lean mixture. Strategies II, V and X, with similar mass values, have only an increase of 20° compared to the full load reference: this is due to the pumping work, generated during the interval between the closing of the first period and the beginning of the second opening period, transformed into an increase in turbulence: this is due to the pumping work, generated during the interval between the closing of the first period and the beginning of the second opening period, transformed into an increase in turbulence. Such a small increase does not lead to an increase in the tendency to knocking.

If a pressure transducer is introduced in the combustion chamber, the controller is able to modulate the closing angle of the second intake period in such a way as to limit the gas temperature value within a predefined threshold, such as to avoid the onset of knocking.

Figure 5 shows an engine load-RPM diagram divided into different areas where different strategies are implemented. In the illustrated example, when the engine load is above a first threshold value L1 , the following is implemented:

• strategy I or III if the engine has a compression ratio <11 and needs to increase power, maximizing the amount of intake air

• strategy I or III if the engine has a compression ratio >11 , is equipped with a water injection system (direct or indirect) and needs to increase power, maximizing the amount of intake air

• strategy II, III, XI o XII if the engine has a compression ratio between 11 and 13 and needs to maximize efficiency, with combustion diluted with EGR and stratified charge achieved by means of an injector in a central position

• strategies X or XII if the engine has a compression ratio greater than 13, and needs to implement the Miller-Atkinson cycle with an effective compression stroke smaller than the expansion stroke. In particular, strategy XII is optimal if you want to achieve stratified combustion, with the fuel injector in a central position. The same strategies are particularly desired in the case of dilution of the mixture with excess air or EGR

When the engine load is below the L1 threshold value, if the engine rotation speed is above a threshold value RPM1 , strategy IV or strategy V is implemented, in order to optimize the homogeneity of the charge and at the same time obtain faster and more robust combustion.

When the engine rotation speed is below the threshold value RPM1 and the engine load is between the first threshold value L1 and a further threshold value L2 lower than the first threshold value L1 , strategy I or strategy III, in the presence of a diluted mixture, with air (lambda >1 .7), with EGR (EGR%>20%) or with both, is implemented.

Again in the case of the example illustrated in figure 5, when the engine rotation speed is lower than RPM1 and the engine load is lower than L2, strategy VII is implemented to minimize the pumping work. Alternatively, if the engine has passed the warm up phase, the strategy IV can be implemented with a lean mixture, with lambda >1.7 or with uncooled EGR, recovered from the exhaust gases by means of the EGR valve.

In order to optimize the benefits of the actuation profiles described in this patent, it is also possible to adopt other measures, for example the connecting rod inter-axis length/crank axis length ratio: in fact, once the length of the crank is fixed, the shorter the connecting rod, the higher the speed will be and the acceleration of the piston at TDC, while the opposite happens at BDC. Greater acceleration at TDC allows both to suck in more air with greater turbulence during the first opening period and to reduce the time the piston remains at TDC during combustion, reducing the risk of knocking. At the same time, a longer residence time at BDC allows the depression caused by the closing of the first opening period to be exploited and maximizes the amount of intake air during the second. Optimal values of the connecting rod/crank ratio, unless there are contraindications due to other problems, are those lower than 1 .6.

Of considerable importance, for the optimal functioning of the engine with the claimed strategies, is the use of intake ducts with a reduced tumble index, suitable, namely, to maximize filling rather than turbulence: in fact the generation of high turbulence is responsible for implementation strategies.

To this end, in a preferred example, the intake valves of each cylinder have axes inclined with respect to the axis of the exhaust valve (70 in figure 1 ), or the exhaust valves, by an angle thetal less than 46°, and each intake duct (4 in figure 1 ), for the main portion of its length that precedes the inlet curve into the cylinder, has an axis inclined with respect to the axis of the respective intake valve by an angle theta2 such that the sum of the angles thetal and theta2 is less than 80°.

In one example, each cylinder is associated with a fuel injector device located centrally with respect to the cylinder.

Preferably, the maximum lift of the first opening period is between 1/8 and 1 Z10 of the diameter of the circular head of the respective intake valve, and the maximum lift of the second opening period, of the first type (2A) and second type (2B), is between 1 /4 and 1 /5 of the diameter of the circular head of the respective intake valve.

In the case of an engine configured to be fed with a low vapor pressure fuel, the controller is programmed to implement said ninth strategy during the start-up and warm-up phase, in particular in the case of cold weather conditions.

In the case of an engine configured to be fed with hydrogen or a mixture of hydrogen and another gas, the controller is programmed to implement said third strategy or said fourth strategy to avoid backfiring phenomena, or unwanted pre-ignitions.

It is possible to inject water into each cylinder, and at the same time to implement said first strategy or said third strategy.

In an example, the pressure in each cylinder is detected, water is injected into each cylinder, and at the same time one of the said first, third, tenth and twelfth strategies is implemented for modulating the amount of water injected at maximum load in order to limit the temperature of the gases at the time of ignition within a predetermined value, so as to avoid knocking, the injection of water preferably being carried out during said second period of opening of an intake valve of each cylinder.

It is also possible to detect the pressure in each cylinder and modify the crank angle at which said second opening period ends so as to allow air to flow back towards the intake manifold and keep the maximum temperature at the time of ignition within a limit value.

Further embodiments of the invention provide that the two intake valves V1 , V2 of each cylinder are controlled by a single cam having the two lobes 14A, 14B described above, by means of a single pumping plunger 16, a common hydraulic circuit, with a single electrically actuated control valve (24).

Said single electrically actuated control valve, in a first example (figure 20), is a three-way, three-position solenoid valve, with an inlet connected to the pressure chamber and to the hydraulic actuator of one of the two intake valves, an outlet connected to the fluid accumulator and a further outlet connected to the hydraulic actuator of the other intake valve, according to a solution known per se from document EP 2 693 007 A1 of the same Applicant.

In a second example (figure 21 ) said single electrically actuated control valve is a system of two two-way, two-position solenoid valves arranged in series with each other, in the connection between the pressure chamber and the hydraulic accumulator and with the two hydraulic actuators of the two intake valves connecting, one with the hydraulic line between the two solenoid valves and the other with the pressure chamber.

Of course, notwithstanding the principle of the invention, the construction details and the embodiments may vary widely with respect to what has been described and illustrated purely by way of example, without thereby departing from the scope of the present invention, as defined by the attached claims.

Claims

1. Internal combustion engine, comprising:

- a first intake valve (V1 ) and a second intake valve (V2) associated with each engine cylinder, for controlling a flow of intake air from a first intake duct (5) and a second intake duct, respectively, (6) during the intake stage into the cylinder in each operating cycle of the cylinder,

- a camshaft (11 ) driven by the crankshaft, carrying a single cam, or two respective cams, to actuate the two intake valves of each engine cylinder, by means of a common tappet (15), or two respective tappets (15),

- wherein the two intake valves (V1 , V2) of each cylinder are actuated by said common tappet (15), or by said respective tappets (15), against the action of respective return springs (9), by interposition of a common hydraulic circuit, or respective hydraulic circuits,

- a pumping plunger (16) associated with each tappet (15), actuated by the respective tappet (15) and configured to transfer pressure fluid, through a pressure chamber (C), to two hydraulic actuators (21 ) respectively associated with the two intake valves (V1 , V2) of each engine cylinder, or to a single respective hydraulic actuator (21 ) associated with one respective of the two intake valves (V1 ; V2) of each engine cylinder,

- an electrically actuated control valve (24) adapted to connect said pressure fluid chamber (C) with a low pressure drain channel (23) connected with a fluid pressure accumulator (270), so that when said control valve (24) is opened, pressure fluid is discharged from the pressure chamber (C) into said low pressure drain channel and the, or each, intake valve (V1 ; V2), controlled by said hydraulic circuit is closed by the respective return spring (9), independently of the action of the respective cam, said engine further comprising an electronic controller (25) for controlling the electrically actuated control valve (24) associated with said common hydraulic circuit, or each electrically actuated control valve (24) associated with each of the two hydraulic circuits of each cylinder, based on a plurality of engine operating parameters, including engine load and engine rotation speed, said engine being further characterized in that:

- said single cam (14) associated with the two intake valves (V1 , V2) of each engine cylinder, or each of the two cams (14) associated with the two intake valves (V1 , V2) of each engine cylinder, has two lobes (14A,14B) arranged and configured to tend to cause a first opening period and a second opening period of each intake valve (V1 , V2) at each revolution of the cam (14), said electronic controller is programmed to control the electrically actuated control valve (24) associated with said common hydraulic circuit of each cylinder, or each electrically actuated control valve (24) associated with each of the two hydraulic circuits of each cylinder, so as to actuate the two intake valves (V1 , V2) of each engine cylinder, independently of each other, according to a number of strategies, depending on the engine load and the engine rotation speed, said strategies being characterized by different combinations of presence or absence of the first opening period and presence or absence of the second opening period of the first intake valve (V1 ) and the second intake valve (V2) of each cylinder, wherein said first opening period of each intake valve of each cylinder begins substantially when the respective piston is at its TDC (Top Dead Center) and ends when the piston is substantially midway between its TDC and its BDC (Botton Dead Center), wherein said second opening period of each intake valve of each cylinder is of a first type (2A) or of a second type (2B), wherein said second opening period of the first type (2A) begins when the piston has passed 4/5 of its path from the TDC to the BDC and has not yet reached its BDC, preferably at not more than 30° of crank angle before the BDC, and at not less than 20° of crank angle before the BDC, wherein said second opening period of the first type (2A) ends when the piston has passed the BDC and is rising towards the TDC, preferably at least 20° of crank angle after the BDC, and at not more than 30° of crank angle after the BDC, wherein the second opening period of the second type (2B) begins when the piston has passed the BDC and is rising towards the TDC, preferably at least 20° of crank angle after the BDC, and at not more than 30° of crank angle after the BDC, and wherein said second opening period of the second type (2B) ends at least 80° of crank angle after the BDC, and at not more than 100° of crank angle after the BDC.

2. Engine according to claim 1 , characterized in that said electronic controller is programmed to implement, depending on the load and engine revolutions, one or more of the following strategies:

3. Engine according to claim 2, characterized in that said electronic controller is programmed to implement, depending on the engine load and the engine rotation speed, one or more of the following further strategies:

- a sixth strategy, in which the first intake valve (V1 ) of each cylinder has only said second opening period of the first type (2A), and the second intake valve (V2) has only said second opening period of the first type (2A),

- a tenth strategy, in which both intake valves (V1 , V2) have both said first opening period and said second opening period of the second type (2B)

4. Engine according to claim 3, characterized in that said electronic controller is programmed to implement:

- said first strategy or said third strategy when the engine compression ratio is less than 11 and the engine load is greater than a first threshold value (L1 ),

- said first strategy or said third strategy when the engine compression ratio is greater than 11 , the engine load is greater than said first threshold value (L1 ), and the engine is equipped with a water injection system,

- the second strategy, or the third strategy, or the eleventh strategy or the twelfth strategy, if the engine compression ratio is between 11 and 13, the engine load is greater than said first threshold value (L1 ), and exhaust gas recirculation (EGR) is provided in the cylinder and/or a fuel injector device is associated with the cylinder in a central position with respect to the cylinder,

- said tenth strategy or said twelfth strategy if the engine compression ratio is greater than 13, - said fourth strategy or said fifth strategy when the engine load is below said first threshold value (L1 ) and the engine rotation speed is above a threshold value (RPM1 ),

- said first strategy or said third strategy when the engine rotation speed is lower than said threshold value (RPM1 ) and the engine load is between the first threshold value (L1 ) and a further threshold value (L2) lower than the first threshold value (L1 ).

5. Engine according to claim 4, characterized in that said electronic controller is programmed to implement said seventh strategy when the engine rotation speed is lower than said threshold value (RPM1 ) and the engine load is lower than said further value threshold (L2) lower than said first threshold value (L1 ).

6. Engine according to claim 1 , characterized in that it comprises:

- two respective cams (14), two respective pumping plungers (16) and two respective hydraulic circuits, with respective electrically actuated control valves (24) for operating the two intake valves (V1 , V2) of each engine cylinder, or: a single cam (14), a single pumping plunger (16), a common hydraulic circuit, with a single electrically actuated control valve (24) for actuating the two intake valves (V1 , V2) of each engine cylinder, said single electrically actuated control valve being i) a three-way, three-position solenoid valve, with an inlet connected to both the pressure chamber and the hydraulic actuator of one of the two intake valves, an outlet connected to the fluid accumulator and a further outlet connected to the hydraulic actuator of the other intake valve, or ii) said single electrically actuated control valve being a system of two two-way, two-position solenoid valves arranged in series with each other, in the connection between the pressure chamber and the hydraulic accumulator and with the two hydraulic actuators of the two intake valves connecting, one with the hydraulic line between the two solenoid valves and the other with the pressure chamber.

7. Engine according to claim 1 , characterized in that with said camshaft (11 ) there is associated a timing shifting device configured to vary the angular position of said camshaft with respect to said crankshaft.

8. Engine according to claim 1 , characterized in that the cam (14) associated with each intake valve (V1 ; V2) of each engine cylinder is of the type with multiple selectively actuatable profiles.

9. Engine according to claim 1 , wherein each piston is operationally connected to a crankshaft by means of a connecting rod-and-crank assembly, characterized in that the ratio of inter-axes lengths of the connecting rod and the crank is less than 1 .6.

10. Engine according to claim 1 , characterized in that each cylinder is associated with a fuel injector device located in a central position with respect to the cylinder.

11. Engine according to claim 1 , characterized in that the two intake valves (V1 , V2) associated with each cylinder have axes inclined with respect to the axis of the exhaust valve, or exhaust valves, of the same cylinder, by an angle thetal of less than 46°, and each intake duct, for the main portion of its length that precedes the inlet curve into the cylinder, has an axis inclined with respect to the axis of the respective intake valve by an angle theta2 such that the sum of the angles thetal and theta2 is less than 80°.

12. Engine according to claim 1 , wherein each intake valve has a circular head having a predetermined diameter, characterized in that the maximum lift of the first opening period is between 1/8 and 1/10 of the diameter of the circular head of the respective intake valve, and the maximum lift of the second opening period, either of the first type (2A) or of the second type (2B), is between 1/4 and 1/5 of the diameter of the circular head of the respective intake valve.

13. Engine according to claim 3, configured to be fed with a low vapor pressure fuel, characterized in that the controller is programmed to implement said ninth strategy during the start-up and warm up phases, in particular in cold weather conditions.

14. Engine according to claim 3, configured to be fed with hydrogen or a mixture of hydrogen and another gas, characterized in that the controller is programmed to implement said third strategy or said fourth strategy to avoid backfiring phenomena, or undesired pre-ignitions.

15. Method for controlling an internal combustion engine, wherein the engine comprises:

- a pumping plunger (16) actuated by the respective tappet (15) and configured to transfer pressure fluid, through a pressure chamber (C), to two hydraulic actuators (21 ) respectively associated with the two intake valves (V1 , V2) of each engine cylinder, or to a respective hydraulic actuator (21 ) associated with a respective intake valve (V1 ; V2) of each engine cylinder,

- an electrically actuated control valve (24) adapted to connect said pressure fluid chamber (C) with a low pressure drain channel (23) connected with a fluid pressure accumulator (270), so that when said control valve (24) is opened, pressure fluid is discharged from the pressure chamber (C) into said low pressure drain channel and the, or each, intake valve (V1 ; V2), controlled by said hydraulic circuit is closed by the respective return spring (9), independently of the action of the respective cam, said engine further comprising an electronic controller (25) for controlling the electrically actuated control valve (24) associated with said common hydraulic circuit, or each electrically actuated control valve (24) associated with each of the two hydraulic circuits of each cylinder, based on a plurality of engine operating parameters, including engine load and engine rotation speed, said method being characterized in that:

- said single cam (14) associated with the two intake valves (V1 , V2) of each engine cylinder, or each of the two cams (14) associated with the two intake valves (V1 , V2) of each engine cylinder, has two lobes (14A,14B) arranged and configured to tend to cause a first opening period and a second opening period of each intake valve (V1 , V2) at each revolution of the cam (14), said method comprising controlling, by means of said electronic controller, the electrically actuated control valve (24) associated with said common hydraulic circuit of each cylinder, or each electrically actuated control valve (24) associated with each of the two hydraulic circuits of each cylinder, so as to actuate the two intake valves (V1 , V2) of each engine cylinder, independently of each other, according to different strategies, depending on the engine load and the engine rotation speed, said different strategies being characterized by different combinations of presence or absence of the first opening period and presence or absence of the second opening period of the first intake valve and the second intake valve of each cylinder, wherein said first opening period of each intake valve of each cylinder begins substantially when the respective piston is at its TDC (Top Dead Center) and ends when the piston is substantially midway between its TDC and its BDC (Bottom Dead Center), wherein said second opening period of each intake valve of each cylinder is of a first type (2A) or of a second type (2B), wherein said second opening period of the first type (2A) begins when the piston has passed 4/5 of its path from the TDC to the BDC and has not yet reached its BDC, preferably at not more than 30° of crank angle before the BDC, and at not less than 20° of crank angle before the BDC, wherein said second opening period of the first type (2A) ends when the piston has passed the BDC and is rising towards the TDC, preferably at least 20° of crank angle after the BDC, and at not more than 30° of crank angle after the BDC, wherein the second opening period of the second type (2B) begins when the piston has passed the BDC and is rising towards the TDC, preferably at least 20° of crank angle after the BDC, and at not more than 30° of crank angle after the BDC, and wherein said second opening period of the second type (2B) ends at least 80° of crank angle after the BDC, and at not more than 100° of crank angle after the BDC.

16. Method according to claim 15, characterized in that it comprises implementing, by means of said electronic controller, one or more of the following strategies, depending on the engine load and engine revolutions:

17. Method according to claim 16, characterized in that it comprises implementing, by means of said electronic controller, one or more of the following further strategies, depending on the engine load and engine revolutions:

18. Method according to claim 17, characterized in that it comprises implementing, by means of said electronic controller:

- said fourth strategy or said fifth strategy when the engine load is below said first threshold value (L1 ) and the engine rotation speed is above a threshold value (RPM1 ), - said first strategy or said third strategy when the engine rotation speed is lower than said threshold value (RPM1 ) and the engine load is between the first threshold value (L1 ) and a further threshold value (L2) lower than the first threshold value (L1 ).

19. Method according to claim 18, characterized in that it comprises implementing, by means of said electronic controller, said seventh strategy when the engine rotation speed is lower than said threshold value (RPM1 ) and the engine load is lower than said further threshold value (L2) lower than said first threshold value (L1 ).

20. Method according to claim 16, comprising the operation of injecting water into each cylinder, and at the same time implementing said first strategy or said third strategy.

21. Method according to claim 17, comprising detecting the pressure in each cylinder, injecting water into each cylinder, and at the same time implementing one of said first, third, tenth and twelfth strategies for modulating the amount of water injected at maximum load so as to limit the temperature of the gases at the time of ignition within a predetermined value, so as to avoid knocking, the injection of water preferably being carried out during said second period of opening of an intake valve of each cylinder.

22. Method according to claim 15, comprising detecting the pressure in each cylinder and modifying the crank angle at which said second opening period ends so as to allow air to flow back towards the intake manifold and limit the maximum temperature, at the time of ignition within a limit value.