EP4673642A1

EP4673642A1 - Internal combustion engine with variable intake valve actuation and engine control method

Info

Publication number: EP4673642A1
Application number: EP24701523.3A
Authority: EP
Inventors: Raffaele Ricco; Marcello Gargano
Original assignee: Centro Ricerche Fiat SCpA
Current assignee: Centro Ricerche Fiat SCpA
Priority date: 2023-02-27
Filing date: 2024-01-23
Publication date: 2026-01-07
Also published as: WO2024180392A1; IT202300003450A1

Abstract

In an internal combustion engine, having a single intake valve (V) for each cylinder, this intake valve is controlled by a cam (14) of a camshaft (11) by means of a hydraulic circuit which can be pressurized or discharged by means of an electrically actuated control valve (24), governed by an electronic controller. The cam has two lobes (14A, 14B) arranged and configured to tend to cause a first opening period and a second opening period of said single intake valve of each cylinder at each revolution of the cam (14). The electronic controller implements different opening strategies of the intake valve (V) of each cylinder as a function of the engine load and the engine rotational speed. The different strategies are characterized by an actuation of both the first and second opening periods, or by an actuation of only the first opening period, or by an actuation of only the second opening period. 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 refers to internal combustion engines of the type indicated in the preamble of claim 1 .

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, and two exhaust ducts 6 lead. The connection of the two intake ducts 4 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 .

The known solution illustrated in figure 1 also includes the case wherein each cylinder is associated with a single intake valve.

Returning to the case wherein two intake valves are provided, 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.

An engine of the type indicated above using a cam with two lobes, where one of the two lobes has the purpose of causing an early opening of the intake valve during the exhaust stage, in order to obtain an internal EGR, is known from the document EP 0 803 642 B1 already mentioned above. Further solutions are known from documents EP 4 043 700 A1 , US 2006/102157 A1 , DE 102 012 188 A1 and EP 1 063 394 A2.

In the international patent application PCT/IB2023/062008, the Applicant disclosed a progression of the known system described above, specifically dedicated to an engine with two intake valves per cylinder. The present invention is based on the need to apply a progression of the known system described above specifically to an engine with a single intake valve for each cylinder.

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 wherein the single intake valve provided for each cylinder can be controlled with strategies that achieve the maximum advantages in terms of combustion efficiency in all operating conditions of the engine.

Summary of the invention

In order to achieve the aforementioned purposes, the invention has as its object an internal combustion engine having the features of claim 1 and a control method according to claim 11 .

Preferred and advantageous features of the invention are indicated in the dependent claims.

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 intake valves of the engine according to the invention, which has a single intake valve for each cylinder, and wherein the intake valve of each cylinder is controlled by a respective two-lobe cam, by means of a tappet, a pumping plunger and a respective hydraulic circuit,

- figure 3 illustrates different opening strategies of the single intake valve of each cylinder of the engine,

- figure 4 is a diagram illustrating different opening strategies of the single intake valve of each cylinder of the engine, that are implemented in different areas of the engine load/engine speed diagram, and

- figures 5, 6 are diagrams that illustrate opening strategies of the single intake valve of each cylinder of the engine,

- figures 7-12 are diagrams showing the advantages of the invention, with reference to different operating parameters of the engine,

- figure 13 shows a diagram of a cylinder and the relative intake duct, in an exemplary embodiment.

Starting from the known solution described above with reference to figure 1 , the invention allows the provision of a plurality of new strategies for opening the engine intake valves, with specific reference to an engine having a single intake valve for each 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 single intake valve of each cylinder of the engine.

According to the invention, each cylinder of the engine has a single intake valve V.

The single intake valve V of each cylinder is controlled by a respective cam 14 of the camshaft by means of a respective hydraulic circuit, comprising a pumping plunger 16, a pressure chamber C, a hydraulic actuator 21 of the intake valve, a respective two-position solenoid valve 24 capable of controlling the communication between the pressure chamber C and a pressure accumulator 270 which is also in communication with a low pressure circuit of the engine lubricating oil.

The cam 14 which controls the single intake valve V of each cylinder of the engine is provided with two lobes 14A, 14B, configured to tend to cause two opening periods of the valve V 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 the intake valve of each cylinder of the engine, and thanks to the provision of a hydraulic circuit interposed between the cam and the intake valve, the intake valve of each cylinder 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 illustrates five 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 the single 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 the single 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 the single 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 the single 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 5 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 6, 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 valve.

In particular, in addition to making the actuation angles of the pumping cam coincide with those of the intake valve (opening and closing), it is also possible to significantly reduce the maximum lift in the first opening period (as visible in figure 5), 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.

With reference to figure 3, in the case of strategy I, the single intake valve of each cylinder of the engine has both the first opening period and the second opening period, said second opening period being of type 2A.

Thanks to the possibility of discharging the hydraulic circuit associated with each intake valve, for the intake valve of each cylinder it is possible to completely cancel the first opening period or the second opening period.

Again with reference to figures 3, strategy II provides that the single intake valve V of each cylinder of the engine has only the first opening period.

Strategy III provided that the single intake valve of each cylinder of the engine has only the second opening period of type 2A.

Strategy IV provides that the single intake valve of each cylinder of the engine has both the first opening period and the second opening period, said second opening period being of the 2B type. This can be achieved for example with the use of a cam having portions with different bilobed profiles, axially contiguous, which can be activated selectively by means of an axial movement of the cam.

Strategy V provides that the single intake valve of each cylinder of the engine has only the second opening period of type 2B.

In the diagrams of figure 3, for each strategy, 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 figure 3 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 single intake valve of each cylinder of the engine with a reduced angular extent (less than 120 degrees of crank angle),

- in the possibility of providing opening and closing profiles of the single intake valve of each cylinder of the engine characterized by steeper rising and falling sides compared to a traditional system with valve actuated directly by cam, whose steepness of rising and falling depends by the ratio between the diameter of the pumping element 16 and the diameter of the hydraulic brake 21

-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 degrees of crank angle,

- in the possibility of modifying and/or modulating, with respect to the opening and closing values indicated in figure 3, at each engine cycle, the opening profile of the single intake valve of each cylinder of the engine: for example, for strategy I, it is however possible to reduce the closing angle from 570 to a smaller value (consequently the valve lift will also be reduced); ditto for the closing angle of the first opening. This advantage is further 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. However, it is understood that the design of the lobes 14A and 14B of the actuating cam of the pumping plunger must be designed to minimize hydraulic losses: in other words, if it is provided that the engine must never operate according to an over-expanded cycle (closing of the intake valve much beyond the BDC), then it is useless to draw the profile 14B with closure beyond the same angle.

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); anyway, in the case of ultra-lean combustion, the increase in the end compression temperature is also obtained due to the different chemical composition compared to the stoichiometric case, since the compression polytropic exponent increases in an inversely proportional manner to the equivalence ratio. 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 ultralean mode is permitted: according to the prior art, modem 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 for increased turbulence, while strategy I also allows for increased mass of intake air and is therefore optimal for ultra-lean mode operation.

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 strategy V allows 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 exists with the strategies illustrated in figure 3, if combined with an asymmetric position of the duct and intake valve with respect to the symmetry plane of the cylinder. This solution is schematized in figure 13, which is a view of a cylinder C in a plane orthogonal to the cylinder axis. The single intake duct 4 associated with cylinder C has a main axis 5 lying in a plane spaced by a distance “d” from the cylinder axis (and from the symmetry plane X-X of the cylinder parallel to the axis 5): in this way the incoming air would always have a swirl component which however would then be coupled to the high TKE generated by the opening strategies. The swirl, in addition to making stratified combustion possible, also has anti-knocking value as it can cool the hot spots near the exhaust valves and the dead space between the cylinder and piston.

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. Higher TKE values, especially if generated during the compression stage, promote the homogeneity: furthermore, being able to have higher tumble values during opening 2A and 2B (delayed compared to conventional actuations) allows fuel injection to be delayed, with beneficial effects on the cooling of the mixture itself with an anti-knocking perspective.

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 III and V 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 I and II, with fuel injection during the second opening period is desired (the swirl, if generated as previously described, 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 strategy I, or strategy IV 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. 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 I and conventional implementation.

The comparison focuses in particular on the quantities that define the flow field and the turbulence, so in figures 8, 9 the following are shown, respectively: the instantaneous flow rate entering the cylinder and the mass of air (and any fuel or EGR) in the cylinder. 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 I, 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 extent. 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 valve 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 (figure 8). Thanks to the inertia of the air in the duct 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 shown in figure 9). Such a high and sudden instantaneous flow rate generates instantaneous, considerable increase in the tumble value (see figures 10 and 11 ): such a high value is due to the high instantaneous flow rate rather than to the shape of the intake duct. Therefore it is advisable to design the intake duct in such a way as to maximize the flow rate rather than the speed of the incoming air.

The high instantaneous flow rate shown in figure 8 also has the effect of instantly generating a high quantity of TKE which is optimal for charge homogeneity, especially if the injection event is timed with the second opening 2A and 2B: observing the graph showing the trend of the TKE (in figures 8-11 , the dotted line refers to the conventional actuation of the intake valve), it is noted that the maximum value is equal to more than 10 times the maximum value of the reference. It should be noted that, while the maximum TKE in the reference case occurs when the piston is halfway 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.

Figure 12 shows the increase in TKE compared to the reference, during an ignition at a crank angle of 700 degrees - 710 degrees. This high TKE value is optimal for fast combustion. At TDC (ignition angle), it is noted that strategy I achieves an increase of more than double in the TKE value compared to the reference at 700 degrees of crank angle. This translates into a potential higher combustion speed, particularly necessary in the presence of a diluted mixture with excess air or EGR. Greater TKE at startup also allows for a reduction in knocking at high loads for stoichiometric mixtures.

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 4 shows an engine load-RPM diagram divided into different areas where different strategies are implemented. In particular:

- strategy II of figure 3 is implemented if the engine rotational speed is lower than a threshold value RPM1 of the rotational speed and if the engine load is lower than a first minimum threshold value L2 of the engine load: this strategy has the advantage of reducing particularly high pumping losses in that region of the load/RPM diagram.

- strategy I of figure 3 is implemented if the engine rotational speed is lower than the RPM1 threshold value of the rotational speed and the engine load is between the minimum threshold value L2 of the engine load and a second threshold value L1 which is greater than the first minimum threshold value L2 of the engine load: this strategy allows maximizing both the quantity of intake air and the TKE. It is therefore suitable for working in ultra-lean mixture conditions when the quantity of air required is double that in stoichiometric conditions and therefore filling must be maximized. Furthermore, in ultra-lean mixture conditions it is also necessary to have a greater quantity of TKE to support combustion.

- strategy I can also be used when the engine load is higher than the L1 value, the mixture is stoichiometric (possibly with additions of EGR) and the compression ratio is less than 11 : in this way the specific power of the engine is maximized in addition to the combustion centroid being optimized thanks to the high TKE associated with strategy I.

- the second strategy of figure 3 is implemented even if the engine rotational speed is higher than the RPM1 threshold value of the rotational speed, and the engine load is between the first minimum threshold value L2 of the engine load and the second value of threshold L2 which is greater than L1 : also in this region of the load/RPM diagram, strategy II allows to reduce pumping losses, however the higher engine rotational speed is sufficient to increase the TKE and support combustion.

- strategy IV of figure 3 is implemented if the engine load is higher than the L1 threshold value of the engine load, and the compression ratio is higher than 11 , i.e. strategy IV is implemented to generate an over- expanded cycle wherein the actual compression stroke is shorter than expansion stroke. In this way, especially in the case of a stoichiometric mixture, the reduced compression stroke avoids the onset of knocking (the temperatures at the end of compression are reduced), however the TKE necessary to support combustion is not affected. The main advantage of the strategy IV is shown in figure 13: according to the state of the art, one of the ways to achieve an over-expanded cycle is to delay the closing of the intake valve beyond the BDC (540 degrees of crank angle), but this involves a significant backflow of air and fuel (fuel injection necessarily occurs before closing) to the intake manifold. This leads to an increase in emissions due to the loss of control of the amount of fuel in each cylinder. For example, with reference to the following figure, it can see how, in the ATKINSON case, almost the same amount of air is flowed as remains in the cylinder (almost 0.6 g of air are intaken in but then 0.3 g remain). The strategy IV allows to aspirate only and exactly the amount of air that is necessary. This strategy is also suitable for hydrogen-powered engines, where it is necessary to avoid any form of fuel backflow to the intake.

- strategy III of figure 3 is implemented during a warm-up phase after a cold start of the engine. This strategy, which generates the maximum quantity of TKE, is suitable for promoting the evaporation of fuels with reduced vapor pressure, such as, for example, methanol.

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 single intake valve of each cylinder has an axis inclined with respect to the axis of the exhaust valve (70 in figure 1 ) by an angle thetal less than 46 degrees, and the intake duct (4 in figure 1 ), for the prevailing portion of its length preceding the inlet curve into the cylinder, has an axis inclined with respect to the axis of the intake valve by an angle theta2 such that the sum of the angles thetal and theta2 is less at 80 degrees.

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 powered by hydrogen or a mixture of hydrogen and another gas, the controller is programmed to implement said strategies I and IV (strategy IV for higher compression ratio values) to avoid backfiring phenomena, or undesired ignition before ignition. In the case of hydrogen supply by means of an injector placed on the intake duct, the injection will take place after the first opening 1 A or 1 B.

Especially in the presence of low density gaseous fuel, such as hydrogen, it is possible to provide for working in stoichiometric conditions and injecting water into each cylinder, and at the same time implementing said first strategy or said third strategy: in this way it will be possible maximize specific power.

In an example, the pressure in each cylinder is detected, water is injected into each cylinder, and at the same time said strategy I 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.

The single intake duct associated with each cylinder can be shaped and arranged to direct an airflow along a main axis that intersects the cylinder axis or that is more or less spaced from the cylinder axis, in order to generate the desired degree of swirl and tumble of the airflow within the cylinder.

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. An internal combustion engine, comprising:

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

- a single intake valve (V) associated with each cylinder of the engine, for controlling a flow of intake air from a respective intake duct (5) during the intake stage into the cylinder in each operating cycle of the cylinder,

- a camshaft (11 ) driven by the crankshaft, carrying a cam, for actuating the intake valve of each cylinder of the engine, by means of a tappet (15),

- wherein the intake valve (V) of each cylinder is actuated by said tappet (15), against the action of a return spring (9), through the interposition of a hydraulic circuit which comprises:

- a pumping plunger (16) actuated by the tappet (15) and configured to transfer fluid under pressure, through a pressure chamber (C), to a hydraulic actuator (21 ) associated with the intake valve (V) of each cylinder of the engine,

- an electrically actuated control valve (24) for connecting said pressure fluid chamber (C) with a low pressure drain channel (23) communicating with a pressure fluid accumulator (270), so that when this control valve (24) is opened, pressure fluid is discharged from the pressure chamber (C) into said low pressure channel and the intake valve (V) closes due to 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 hydraulic circuit, as a function of a plurality of engine operating parameters, including engine load and engine rotational speed, wherein furthermore:

- said cam (14) associated with the single intake valve (V) of each cylinder of the engine has two lobes (14A,14B) arranged and configured to tend to cause a first opening period and a second opening period of said single intake valve (V) at each revolution of the cam (14), said electronic controller (25) is programmed to govern the electrically actuated control valve (24) associated with said hydraulic circuit of each cylinder, so as to actuate the single intake valve (V) of each cylinder of the engine according to a plurality of strategies, depending on the engine load and the engine rotational speed, each of said strategies being characterized by an actuation of both the first opening period and the second opening period of said single intake valve (V) of each cylinder, or by an actuation of only the first opening period of said single intake valve of each cylinder, or by an actuation of only the second opening period of said single intake valve of each cylinder, wherein said first opening period of said single intake valve of each cylinder begins substantially when the respective piston is at its TDC and ends when the piston is substantially midway between its TDC and its BDC, and wherein said second opening period of said single 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 TDC to BDC and has not yet reached its BDC, preferably at not more than 30 degrees of crank angle before BDC, and at not less than 20 degrees of crank angle before BDC, wherein said second opening period of the first type (2A) ends when the piston has passed the BDC and is in the process of rising towards the TDC, preferably at least 20 degrees of crank angle after the BDC, and at not more than 30 degrees of crank angle after BDC, wherein the second opening period of the second type (2B) begins when the piston has passed the BDC and is in the process of rising towards the TDC, preferably at least 20 degrees of crank angle after the BDC, and at not more than 30 degrees of crank angle after BDC, and wherein said second opening period of the second type (2B) ends at not less than 80 degrees of crank angle after BDC, and at not more than 100 degrees of crank angle after BDC.

2. An engine according to claim 1 , characterized in that said electronic controller (25) is programmed to implement, depending on the load and the engine rotational speed, one or more of the following strategies:

- a first strategy, wherein said single intake valve (V) of each cylinder has both said first opening period and said second opening period of the first type (2A)

- a second strategy, wherein said single intake valve (V) of each cylinder has only said first opening period,

- a third strategy, wherein said single intake valve (V) of each cylinder has only said second opening period of the first type (2A),

- a fourth strategy, wherein said single intake valve (V) of each cylinder has both said first opening period and said second opening period of the second type (2B), and

- a fifth strategy, wherein the first intake valve (V1 ) of each cylinder has only the second opening period of the second type (2B).

3. An engine according to claim 2, characterized in that said electronic controller (25) is programmed to implement:

- said second strategy if the engine rotational speed is lower than a rotational speed threshold value (RPM1 ) and the engine load is lower than a first minimum engine load threshold value (L2),

- said first strategy if:

- the engine rotational speed is lower than said rotational speed threshold value (RPM1 ), and

- the engine load is between said first minimum engine load threshold value (L2) and a second threshold value (L1 ) which is greater than said first minimum engine load threshold value (L2), and

- the lambda ratio between the air/fuel metering and the stoichiometric metering is greater than 1 .8;

- said fourth strategy if:

- the lambda ratio between the air/fuel metering and the stoichiometric metering is less than 1 .8;

- said first strategy or said second strategy if the engine rotational speed is greater than said rotational speed threshold value (RPM1 ), and the engine load is lower than said second engine load threshold value (L1 ),

- said first strategy if the engine load is greater than said second engine load threshold value (L1 ) and the compression ratio (CR) is lower than 11 ,

- said fourth strategy if the engine load is greater than said second engine load threshold value (L1 ) and the compression ratio (CR) is greater than 11 ,

- said third strategy during a warm-up phase after a cold start of the engine.

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

5. An engine according to claim 1 , characterized in that the cam (14) associated with said single intake valve (V) of each cylinder of the engine is of the type with multiple selectively activatable profiles.

6. An engine according to claim 1 , wherein each piston is operatively connected to a crankshaft by means of a connecting rod-and-crank assembly, characterized in that the ratio between the operating lengths of the connecting rod and the crank is lower than 1 .6.

7. An engine according to claim 1 , characterized in that with each cylinder there is associated a fuel injector device located in a central position with respect to the cylinder.

8. An engine according to claim 1 , characterized in that said single intake valve (V) associated with each cylinder has an axis 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 degrees, and the intake duct, for the prevailing portion of its length preceding the inlet curve into the cylinder, has an axis inclined with respect to the intake valve axis by an angle theta2 such that the sum of the angles thetal and theta2 is lower than 80 degrees.

9. An 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 intake valve, and the maximum lift of the second opening period, either of the first type (2A) or of the second type (2 B ) , is between 1 /4 and 1 /5 of the diameter of the circular head of the intake valve.

10. An engine according to claim 3, configured to be fuelled with hydrogen or a mixture of hydrogen and another gas, characterized in that the controller (25) is programmed to implement said first strategy or said third strategy to avoid backfiring phenomena, or undesired ignition before ignition.

11. A method for controlling an internal combustion engine, wherein the engine comprises:

- a single intake valve (V) associated with each cylinder of the engine, for controlling a flow of intake air from an intake duct (5) during the intake stage into the cylinder in each operating cycle of the cylinder,

- a camshaft (11 ) driven by the crankshaft, carrying a cam for actuating said single intake valve of each cylinder of the engine, by means of a tappet (15),

- wherein said single intake valve (V) of each cylinder is actuated by said tappet (15), against the action of a respective return spring (9), through the interposition of a hydraulic circuit which comprises:

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

- an electrically actuated control valve (24) for connecting said pressure fluid chamber (C) with a low pressure drain channel (23) communicating with a pressure fluid accumulator (270), so that when said control valve (24) is opened, pressure fluid is discharged from the pressure chamber (C) into said low pressure channel and said single intake valve (V) 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 hydraulic circuit of each cylinder, as a function of a plurality of engine operating parameters, including engine load and engine rotational speed, wherein furthermore:

- said cam (14) associated with said single intake valve (V) of each cylinder of the engine has two lobes (14A,14B) arranged and configured to tend to cause a first opening period and a second opening period of said single intake valve (V1 , V2) at each revolution of the cam (14), said process comprising governing, by means of said electronic controller, the electrically actuated control valve (24) associated with said hydraulic circuit of each cylinder, so as to actuate said single intake valve (V) of each cylinder of the engine according to different strategies, as a function of the engine load and the engine rotational speed, each of said strategies being characterized by an actuation of both the first opening period and the second opening period of said single intake valve (V) of each cylinder, or by an actuation of only the first opening period of said single intake valve of each cylinder, or by an actuation of only the second opening period of said single intake valve of each cylinder, wherein said first opening period of said single intake valve of each cylinder begins substantially when the respective piston is at its TDC and ends when the piston is substantially midway between its TDC and its BDC, and wherein said second opening period of said single 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 TDC to BDC and has not yet reached its BDC, preferably at not more than 30 degrees of crank angle before BDC, and at not less than 20 degrees of crank angle before BDC, wherein said second opening period of the first type (2A) ends when the piston has passed the BDC and is in the process of rising towards the TDC, preferably at least 20 degrees of crank angle after the BDC, and at not more than 30 degrees of crank angle after BDC, wherein the second opening period of the second type (2B) begins when the piston has passed the BDC and is in the process of rising towards the TDC, preferably at least 20 degrees of crank angle after the BDC, and at not more than 30 degrees of crank angle after BDC, and wherein said second opening period of the second type (2B) ends at not less than 80 degrees of crank angle after BDC, and at not more than 100 degrees of crank angle after BDC.

12. The method according to claim 11 , characterized in that, by means of said electronic controller, one or more of the following strategies are implemented, as a function of the load and the engine rotational speed:

13. The method according to claim 12, characterized in that it comprises implementing, by means of said electronic controller:

- said first strategy if:

- said fourth strategy if:

- said third strategy during a warm-up phase after a cold start of the engine.

14. The method according to claim 13, comprising the operation of injecting water into each cylinder, and at the same time implementing said first strategy or said fourth strategy.

15. The method according to claim 14, comprising detecting the pressure in each cylinder, injecting water into each cylinder, and at the same time implementing said first strategy to modulate the amount of water injected at maximum load so as to limit the temperature of the gases at the time of the ignition within a predetermined value, so as to avoid knocking, the injection of water being preferably carried out during said second opening period of the intake valve of each cylinder.

16. The method according to claim 14, 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 the ignition within a limit value.

17. An engine according to any one of claims 1 -10, characterized in that the single intake duct (4) associated with each cylinder (C) has a main axis (5) lying in a plane which is spaced from the cylinder axis.