WO2025158239A1

WO2025158239A1 - An internal combustion engine with variable intake valve actuation, having an improved efficiency at low engine loads, and related control method

Info

Publication number: WO2025158239A1
Application number: PCT/IB2025/050407
Authority: WO
Inventors: Raffaele Ricco
Original assignee: Centro Ricerche Fiat SCpA
Current assignee: Centro Ricerche Fiat SCpA
Priority date: 2024-01-25
Filing date: 2025-01-14
Publication date: 2025-07-31
Anticipated expiration: 2026-07-25
Also published as: IT202400001431A1

Abstract

An internal combustion engine has at least one intake valve (V1, V2) for each cylinder, actuated by a cam (14) adapted for providing a lift profile having two maximum points spaced apart from each other, for example a profile with two lobes, or a boot profile. Each intake valve is actuated by a respective cam via a hydraulic circuit that can be pressurized or discharged by an electrically actuated control valve (24), controlled by an electronic controller. When the electronic controller detects an engine operating condition with a low engine load, the controller actuates each intake valve (V1, V2) with a first opening period and a second opening period interspersed with a phase in which the intake valve is closed, while the respective piston is descending towards its BDC, so as to create a depression in the respective cylinder which generates a jet of air entering the cylinder during the subsequent second opening period. The second opening period begins when the piston has passed the BDC and is ascending towards the TDC. In this way, during a first part of the second opening period, while the piston is ascending towards the TDC, said jet of air is created entering the cylinder due to the depression created in the cylinder between the first opening period and the second opening period, while during a second part of the second opening period the depression in the cylinder is progressively reduced until it is eliminated and a portion of air is pushed back into the intake duct by the piston that continues to rise towards the TDC. The second opening period has a duration sufficient to generate an effective compression ratio which is at least 5 and sufficiently high to generate a temperature of the mixture within the cylinder, when the piston reaches the TDC, of at least 500 K.

Description

“An internal combustion engine with variable intake valve actuation, having an improved efficiency at low engine loads, and related control method”

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TEXT OF THE DESCRIPTION

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” trademark. 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 is possible provide 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 of the engine camshaft.

In this case, if it is desired to actuate the two intake valves of the same cylinder in a differentiated way, 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 both 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 (see figure 20 attached hereto, corresponding to figure 4 of the document cited above).

Alternatively, it is possible to provide the further solution known from document EP 3 832 078 A1 , also from the same Applicant, which provides for 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, one communicating with the hydraulic line between the two solenoid valves and the other with the pressure chamber (see figure 20 attached hereto, corresponding to figure 11 of EP 3 832 078 A1)

However, for the purposes of the present invention, it is also possible to 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 is possible to benefit from total flexibility in differentiating the openings of the two intake valves of each cylinder.

In the Italian patent applications IT 102022000025410, IT 102023000003450 and 102023000013266 the Applicant has presented progression of the known system described above.

Starting from these previous proposals, the Applicant has conducted further studies in order to further improve them.

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 that is characterized by high combustion efficiency in all engine operating conditions and specifically in conditions of low engine loads.

In particular, an object of the invention is to create an internal combustion engine in which the intake valves of each cylinder can be controlled according to an innovative strategy, which achieves maximum advantages in terms of combustion efficiency even in said conditions of low engine loads.

Finally, an important aim of the invention is to achieve the above objects with a solution that is simple in construction and easy to apply even on an already existing 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 12.

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 engine intake valves, according to a first embodiment of the present invention, in which the two intake valves of each cylinder are actuated by two separate cams, via respective tappets, respective pumping plungers and respective hydraulic circuits, and in which each cam is shaped with two lobes,

- figure 3 is a schematic view of the variable actuation system of the engine intake valves, in a second embodiment of the present invention, in which the two intake valves of each cylinder are controlled by two separate cams, via respective tappets, respective pumping plungers and respective hydraulic circuits, and in which each cam is shaped with a “boot” type profile,

- figure 4 is a schematic view of the variable actuation system of the intake valves of the engine, in a third embodiment of the present invention, in which the two intake valves of each cylinder are controlled by a single cam, via a single tappet, a single pumping plunger and a single pressure chamber communicating with the hydraulic actuators of the two intake valves, and

- figures 5-15 are diagrams illustrating the operating principle and advantages of the present invention.

Starting from the known solution described above with reference to figure 1 , the invention allows efficient combustion to be achieved even at low engine loads and even in the case in which the engine operates with a fuel having a lower vapor pressure than that of gasoline, such as methanol.

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 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 lubrication oil.

In this embodiment, furthermore, each of the two cams 14 that 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 for each rotation 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, spaced apart, corresponding to a conventional opening phase of an intake valve.

In the second embodiment illustrated in figure 3, each of the two cams 14 that control the two intake valves V1 , V2 of each engine cylinder is configurated so that, when the pressure chamber is constantly maintained under pressure, a lift profile of said at least one intake valve is generated, as a function of the engine crank angle, having a boot shape with a first part defining a first lift level, having a first maximum point, followed by a second part defining a second maximum point, higher than the first maximum point.

In all of the above embodiments, the electronic controller 25 is configured for the detection, when the engine load is relatively low, i.e., below a predetermined threshold value and/or when the engine rotational speed is below a predetermined threshold value (e.g., 3.000 rpm), and in this condition, the electronic controller 25 is configured to control the electrically actuated control valve 24 so as to cause, at each engine cycle, a first opening period and a second opening period of each intake valve, with an intermediate phase in which the intake valve is fully closed while the respective piston is descending towards its BDC, so as to create a depression in the respective cylinder which generates a jet of air entering the cylinder during the subsequent second opening period of the intake valve.

According to the invention, the first opening period of each intake valve of each cylinder begins substantially when the respective piston is at its Top Dead Center (TDC) and ends when the piston is substantially midway between its TDC and its Bottom Dead Center (BDC).

The second opening period of each intake valve begins when the piston has passed the BDC and is ascending towards TDC.

In this description, and in the claims that follow, the convention has been adopted whereby the engine crank angle is considered equal to 360° and 540° when the piston is at TDC and BDC, respectively, at the beginning and end of a conventional intake stage.

In the invention, the values of the engine crank angle at which the second opening period begins and ends may vary from case to case, as they depend on various factors.

The values of the engine crank angle at which the second opening period begins and ends depend first of all on the desired engine load, that is, on the amount of air to be introduced into the cylinder. Then, depending on the desired air/fuel ratio, the necessary amount of fuel will be entered.

The values of the engine crank angle at which the second opening period begins and ends also depend on the rotational speed of the engine. Since during the second opening period there is initially an entry of air into the combustion chamber and subsequently an exit of air from the combustion chamber, it follows that the amount of air that remains in the engine is a function of the duration (expressed in seconds and not in degrees of engine crank angle) of the second opening period. Increasing the duration of the second opening reduces the engine load. For example, if the engine speed is 2.000 rpm, for an engine with a geometric compression ratio (CR) of 15, to achieve a load of 1 .5 bar of Indicated Mean Pressure (IMP), without the aid of internal EGR (Exhaust Gas Recirculation), the intake valves must open at approximately 40° of engine crank angle after BDC and close at approximately 140° of engine crank angle after BDC, with an overall duration of the second opening period of 100° of engine crank angle. To achieve the same load point at 1.000 rpm, the duration of the second opening should be reduced to 80° of engine crank angle.

The engine crank angle values at which the second opening period begins and ends also depend on the geometric compression ratio of the engine, which is the ratio between the volume of the combustion chamber when the piston is at its BDC and the volume of the combustion chamber when the piston is at its TDC.

For certain operating conditions (pressure and temperature of the intake gases, intake profiles of the valves, percentage of internal EGR) and for a certain value of the engine crank angle at which the intake valves close definitively, the higher the geometric ratio of the engine, the higher the temperature of the air/fuel mixture at the end of the compression stage (i.e. at TDC). A sufficiently high value of this temperature is important to ensure robust combustion.

Actually, in the case of the invention, what is more relevant is the “effective” compression ratio of the air/fuel mixture (and of the possible EGR), which is different and lower than the geometric compression ratio. The effective compression ratio is the ratio between the volume of the combustion chamber when the piston has risen to a position for which the pressure in the cylinder exceeds the value of 1 barA (after which a compression stage begins with a monotonic increase in pressure up to TDC) and the volume of the combustion chamber at TDC.

If the effective compression ratio is not sufficient, the mixture temperature at TDC will be low, for example below 500 °K, or below the limit value of the mixture ignitability. An engine with a higher geometric compression ratio, for a given value of the engine crank angle at which the intake valves close definitively, will determine a higher effective compression ratio of the mixture: this means that for example, for an engine with a geometric compression ratio (CR) equal to 15, at 2.000 rpm, the intake valves can close at 670° of engine crank angle while still ensuring a final temperature that allows ignition. Otherwise, for an engine with CR=10, the engine crank angle within which the intake valves must be closed will be, for example, 630°.

The values of the engine crank angle at which the second opening period begins and ends also depend on the possible presence of internal EGR (i.e. exhaust gases that return directly from the exhaust duct into the combustion chamber to participate in a new combustion) and on the temperature of the gases in the intake stage. For the reasons indicated above, the higher the temperature of the gas mixture in the intake, the higher the temperature of the mixture at TDC, for the same value of the engine crank angle at which the intake valves are closed definitively, or for the same value of the effective compression ratio of the gases.

In this regard, it should be considered that the internal EGR also has the advantage of increasing the mass of the mixture at TDC, with a further increase in temperature and pressure. It can be said that the EGR allows the engine crank angle at which the intake valves are closed to be advanced, so as to end the phase in which part of the air flows back into the intake duct. As a disadvantage, the EGR is an inert material and therefore requires sufficient turbulence to support combustion and miscibility with air and fuel.

Taking into account all the above requirements, according to the invention the engine crank angles at which the second intake period begins and ends (which in any case takes place during the piston rising phase from BDC to TDC), are chosen in such a way that the effective compression ratio (which as indicated above is the ratio between the volume of the combustion chamber when, during the piston rising phase and after the end of the second intake valve opening period, the pressure exceeds the value of 1 barA (after which a compression stage with monotonic pressure variation begins up to TDC), and the volume of the combustion chamber at TDC) is at least greater than 5.

This condition, in particular if at TDC the gas temperature has exceeded 550K, is sufficiently robust to ensure correct ignition and combustion. Of course, the stability of the combustion then also depends on the turbulence and the achieved degree of homogeneity of the mixture.

As a result of the above features, during a first part of said second opening period, while the piston is ascending towards the TDC, the above jet of air is created entering the cylinder due to the depression created in the cylinder between the first opening period and the second opening period, while during a second part of the second opening period the depression in the cylinder is progressively reduced until it is eliminated and a portion of air is pushed back into the intake duct by the piston continuing to ascend towards the TDC.

In this way the invention allows a plurality of advantages to be achieved simultaneously:

- the combustion, even at low engine loads, is very efficient, since the jet of air entering the cylinder during the second opening period, due to the depression created in the cylinder between the first opening period and the second opening period, gives rise to a high Turbulence Kinetic Energy or TKE;

- the availability of high TKE, not obtainable with known intake cycles, also allows to increase the amount of internal EGR that can be intaken, even at the lowest loads; as already indicated, the introduction of internal EGR allows to increase the mass of gases to be compressed as well as the temperature at TDC, with a reduction in pumping losses; the difficult combustion of internal EGR is compensated by the higher turbulence;

- the depression in the cylinder, compared to a Miller cycle implementation, characterized by an early closing (reference), is lower (i.e. at the end of the piston stroke, at BDC, a higher pressure is obtained): this is because - as can be seen from figure 15 - in the Miller case, to introduce little air one is forced to intake at 0.5 barA, while according to the invention the intake during the first opening period occurs at 1 barA

- the amount of air introduced into the cylinder at low engine loads is reduced to a measure suitable for this operating condition.

In a concrete example for a minimum engine load equal to IMEP=1 .5 bar (therefore PME«1 bar), for an engine with geometric CR equal to 15, at 2000 rpm, intaking only air at a temperature of 310 K, with internal EGR (“iEGR”) equal to zero, and assuming a stoichiometric mixture, the second opening period begins at an engine crank angle equal to 580° ± 15° and ends at an engine crank angle equal to 670°± 10°.

In practice, the piston behaves like an air spring that first creates depression in the cylinder and then, during the ascent, the piston is pushed upwards by the depression generated, so the compression work begins only when the pressure in the cylinder returns to being greater than 1 barA: this means that the later the second suction phase begins, the greater the pumping work will be recovered.

This is made clear in figure 14 of the attached drawings, where 01 and 02 are the lift profiles of the intake valves in the first opening period and in the second opening period, E is the lift profile of the exhaust valve and the area in which the gas spring effect is indicated is the one corresponding to the pumping work recovered.

In the case of the embodiment of figure 2, where the cams have double-lobe profiles, with zero lift between one lobe and the other, said result is obtained by always keeping the control valve 24 closed.

In the case of the solution of figure 3, where the cams have boot-type profiles, where the lift does not return to zero between the first maximum point and the second maximum point of the cam profile, the result is obtained by keeping said control valve 24 closed at the first opening period and the second opening period and open at an intermediate phase between the first opening period and the second opening period, so as to keep the respective intake valve closed.

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 is between 1/4 and 1/5 of the diameter of the circular head of the respective intake valve

Figure 5 shows the lift of each intake valve as a function of the engine crank angle, in a specific embodiment of the invention. The first opening period is indicated by 01 and the second opening period is indicated by 02. The diagram refers to an operating condition at low engine rotational speed (2.000 rpm). Figure 5 also shows the lift of the intake valve in a conventional Miller cycle engine (M curve) under the same operating conditions.

In the case of the invention, during the first opening period 01 , an amount of air, and possibly - by means of a boot profile of the intake valve in overlap with the exhaust stage of the previous cycle - a certain amount of recirculated exhaust gas (EGR), sufficient to prevent the pressure in the cylinder from falling to values that are too low (i.e. for example to values lower than 0.45 bar A) are introduced into the cylinder. In this way, it is avoided that there may be oil leakages through the oil-scraper rings associated with the piston, with consequent passage of oil from the crankcase to the combustion chamber. Advantageously, the closing angle of this first event is between 75° and 90° after the TDC: in particular, advancing this value will generate greater turbulence, i.e. the intensity of the air jet will be greater.

As already indicated, due to the depression created in the cylinder between the first opening period 01 and the second opening period 02, the second opening period 02 generates a jet of air entering the cylinder, which produces a TKE sufficient to support both an efficient mixing of the air with the fuel and with the eventual amount of EGR intended, and the propagation of the flame during combustion.

The very delayed start of the second opening period allows to reduce the pumping losses, making the piston work like an air spring: the duration of the second opening period allows the excess air to return to the intake duct, leaving only the necessary amount of air (and eventual EGR) trapped in the cylinder, which again reduces the pumping losses.

Especially for engines with a high geometric compression ratio (e.g. above 12), the very late end of the second opening period still ensures an effective compression ratio sufficient to ensure that at ignition at TDC the temperature and pressure are sufficiently high to ensure robust combustion, also thanks to the high TKE generated by the second opening period; this value being at least 5 as indicated.

The 01 , 02 lift profiles shown in figure 5 are purely qualitative, but the values of the engine crank angles corresponding to the beginning and end of each opening period are real. In figure 5, the example refers to an operating condition in which the engine rotational speed is 2.000 RPM and the geometric compression ratio of the engine is 12.5. The maximum lift in the second opening period can be higher than that in the first opening period.

In the case where two intake valves are provided for each cylinder, and where the two valves are actuated by means of hydraulic cylinders separated by two separate chambers, it is still possible to actuate the two intake valves in an identical manner, corresponding to what is illustrated in figure 5, even if in the case of this embodiment the two intake valves of the cylinder can also be controlled differently, for example by opening both valves in the first opening period, and by opening only one valve in the second opening period, as will be illustrated in more detail below.

Figure 6 of the attached drawings shows the mass flow rate of air in the cylinder as a function of the engine crank angle both in the case of the invention (curves 01 , 02) and in the case of a conventional Miller cycle engine (curve M). Figure 6 also shows the variation of the piston speed V as a function of the engine crank angle.

As can be seen from figure 6, the second opening period initially generates a positive mass flow rate (from the intake ducts to the cylinder), resulting in TKE generation. After this initial phase, there is a negative return flow from the cylinder to the intake ducts. The engine crank angle at which the second opening period ends predefines the amount of air that remains trapped in the cylinder together with the possible amount of EGR.

It should be noted that, in case of fuel introduction by means of a direct injection system, it is appropriate and advantageous to introduce the necessary fuel, or at least most of it, during or even better after the second opening period: among the advantages is the possibility of avoiding that the fuel spray hits the piston (which is close to the BDC or in any case far from the TDC), furthermore, unlike the well-known Atkinson cycle characterized by late closing of the intake valves (LIVC), delaying the fuel injection will reduce the amount of fuel that could flow back to the intake manifold. The large amount of turbulence generated (but also tumble) allows for excellent mixing despite the fuel being injected so late.

Figure 7 shows, both in the case of the conventional Miller cycle engine (M curve) and in the case of the invention (I curve), the amount of air mass in the cylinder as a function of the engine crank angle. As is evident from this figure, thanks to the return flow of air from the cylinder to the intake ducts in the second phase of the second opening period, only the amount of air (plus any EGR) that is necessary remains trapped in the cylinder.

Figures 8 and 9 show the tumble index and the TKE in the cylinder as a function of the engine crank angle, both in the case of the conventional Miller cycle engine (M curve) and in the case of the invention (I curve).

As can be seen from figures 8, 9, the invention achieves a higher TKE during the second opening period, which can allow a delayed fuel injection into the cylinder (for example, fuel injection can start at 610° of engine crank angle): the TKE and the high tumble ensure a good air-fuel mixing. Furthermore, the delayed fuel injection avoids a fuel flow towards the intake ducts, as can instead occur in conventional Atkinson cycle engines.

The presence of a higher TKE at the engine crank angle at which ignition occurs (typically 680°) allows to support combustion and the presence of internal EGR.

Figure 10 illustrates the variation of pressure in the cylinder as a function of engine crank angle, both in the case of a conventional Miller cycle engine (M curve) and in the case of an engine according to the invention (I curve). With reference to figure 10, the minimum value of the pressure in the cylinder in the case of the engine according to the invention is equal to 0.48 bar A, while in the case of the conventional Miller cycle engine (where the throttle valve at intake is kept only slightly open at low engine loads) it is equal to 0.2 bar A. Consequently, in the case of the invention, the risk of oil leakages (cylinder breathing) from the crankcase to the combustion chamber, through the oil-scraper rings due to the presence of a depression in the cylinder, is greatly reduced.

Still referring to figure 10, at the end of the pumping phase, corresponding to the return of the pressure to the value of 1 bar A, (“lx” in the case of the invention and “Mx” in the case of the conventional Miller cycle engine) the engine crank angle is respectively 614° (in the invention) and 668° (in the Miller cycle engine). The engine according to the invention therefore produces lower pumping losses. The differences in pumping losses are even more evident in a classic representation of the P x V cycle illustrated in figure 15.

Even though the invention has the end of the pumping phase at a lower engine crank angle, the starting point of the compression is substantially the same. For the conventional Miller cycle engine, the start of the compression stage coincides with the end of the pumping phase, while in the case of the invention, the start of the compression stage is delayed with respect to the end of the pumping phase. At the end of the second opening period (at an engine crank angle of 670°) both curves I, M, have the same pressure value. This means that the final values of temperature and pressure, at the engine crank angle at which the ignition of the air-fuel mixture occurs, are substantially identical.

Figure 11 shows the variation of the temperature in the cylinder as a function of the engine crank angle for the invention (curve I) and for the conventional Miller cycle engine (curve M). Figure 11 makes it clear that, in the case of the invention, due to the lower pumping work, the temperature is lower than in the Miller case: however, thanks to the higher TKE that occurs in the invention, it is possible to introduce an amount of internal (hot) EGR to further reduce the pumping losses and increase the ignition temperature.

Figure 12 illustrates the variation of the effective pumping work as a function of the engine crank angle in the case of the invention (curve I) and in the case of the conventional Miller cycle engine (curve M). Figure 12 shows that the effective pumping work in the case of the invention (about 3J) is drastically reduced compared to the case of the conventional Miller cycle engine (about 7J).

As is evident from the above, the engine according to the invention has the following important advantages:

- the air jet at low loads allows to increase the minimum pressure in the cylinder and avoid an oil breathing phenomenon from the crankcase to the combustion chamber; - the air jet at low loads supports the dilution with internal EGR thanks to a higher TKE (which is about double compared to the conventional Miller cycle engine).

- the air jet at low loads allows to drastically reduce (by 50%) the pumping losses.

Figure 13 is a variant of figure 5, which refers to an embodiment in which the cam has a double boot profile that allows to implement a first opening period 01 with an initial part having the function of allowing to introduce internal EGR into the cylinder.

Furthermore, as already mentioned above, in the embodiments that provide two valves per cylinder actuated by means of separate cams and hydraulic circuits, it is possible to differentiate the actuation of the two intake valves of each cylinder, opening both valves in the first opening period and opening only one of the two valves in the second opening period.

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

For example, the two intake valves could be opened with profiles of different shapes, as for example illustrated in figure 5: a valve could be opened according to profile 01 +02 (which would be the profile that actuates the air jet) and the second valve could be opened with profile M, provided that the closure of this valve satisfies the specified criterion of closing no more than 80° after TDC.

Asymmetric intake valve openings, while still meeting the defined criteria, can bring a further benefit, that is, generating swirl motions in the combustion chamber that support stratified combustion well: that is, especially at low loads, introducing the fuel in a swirling motion, better by means of an injector placed in a central position, just before the piston reaches the TDC. The swirl acts as a sort of gas layer that isolates the propagation of the flame to the cylinder walls, with benefits on the heat released to the walls: this is particularly advantageous for low loads, as long as there is sufficient turbulence (in practice it is as if the combustion chamber is virtually reduced).

Claims

1. An internal combustion engine, 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 including an intake stage, a compression stage, an expansion stage and an exhaust stage,

- at least one intake valve (V1 , V2) associated with each engine cylinder to control a flow of intake air from a respective intake duct (5) during the intake stage into the cylinder in each cylinder operating cycle,

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

- wherein said at least one intake valve (V) of each cylinder is actuated by said tappet (15), against the action of a return spring (9), with the interposition of a hydraulic circuit including:

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

- an electrically actuated control valve (24) adapted to connect said pressurized fluid chamber (C) to a low-pressure drain channel (23) communicating with a pressurized fluid accumulator (270), such that when said control valve (24) is open, pressurized fluid drains from the pressure chamber (C) into said low-pressure drain channel and said at least one intake valve (V) closes due to the action of the respective return spring (9), regardless 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, based on a plurality of engine operating parameters, including engine load and engine rotational speed, said engine being characterized in that:

- said cam (14) associated with said at least one intake valve (V) of each engine cylinder is configurated so that, when the pressure chamber is constantly maintained under pressure, a lift profile of said at least one intake valve is generated, as a function of the engine crank angle, which lift profile has two maximum points spaced apart from each other,

- said electronic controller (25) is configured to detect when the engine load is below a predetermined threshold value and/or when the engine rotational speed is below a predetermined threshold value, and, in this condition, is configured to control the electrically actuated control valve (24) associated with said hydraulic circuit of each cylinder, so as to cause, at each engine cycle, a first opening period and a second opening period of said at least one intake valve, with an intermediate phase in which said at least one intake valve is fully closed while the respective piston is descending towards its BDC, so as to create a depression in the respective cylinder which generates a jet of air entering the cylinder during the subsequent second opening period of said at least one intake valve, wherein said first opening period of said at least one 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 at least one intake valve of each cylinder begins when the piston has passed the BDC and is ascending towards the TDC, so that during a first part of said second opening period, while the piston is ascending towards the TDC, said jet of air is created entering the cylinder due to the depression created in the cylinder between the first opening period and the second opening period, while during a second part of the second opening period the depression in the cylinder is progressively reduced until it is eliminated and a portion of air is pushed back into the intake duct by the piston continuing to ascend towards the TDC, said second opening period having a duration sufficient to generate an effective compression ratio which is at least 5 and sufficiently high to generate a temperature of the mixture within the cylinder, when the piston reaches the TDC, of at least 500 K, where the effective compression ratio is defined as the ratio between the volume of the combustion chamber when the piston has risen to a position for which the pressure in the cylinder exceeds the value of 1 bar A, after which a compression stage begins with a monotonic increase in pressure up to when the piston reaches the TDC, and the volume of the combustion chamber when the piston is at TDC.

2. The engine according to claim 1 , characterized in that it includes a fuel injection system configured to perform a fuel injection during the rise of the piston to the TDC, during or after the end of the second opening period of said at least one intake valve.

3. The engine according to claim 1 , wherein with each cylinder there are associated two intake valves (V1 , V2), characterized in that the electronic controller (25) is configured to cause an asymmetric opening of the intake valves (V1 , V2), or in that the two intake ducts of each cylinder have different configurations, such that an air flow field within the cylinder is generated at TDC, having a swirl component with a swirl index greater than 1.

4. The engine according to any one of claims 1 -3, characterized in that with each cylinder there is associated a fuel injector located at the center of the cylinder and configured to inject fuel just before the piston reaches the TDC, at the center of a swirl of air in the cylinder, so as to achieve a stratified combustion.

5. The engine according to claim 1 , characterized in that it comprises a single intake valve for each cylinder.

6. The engine according to claim 1 , characterized in that it comprises a first intake valve (V1 ) and a second intake valve (V2) for each cylinder,

7. The engine according to claim 1 , characterized in that:

- said cam (14) associated with said at least one intake valve (V) of each engine cylinder is shaped with two lobes (14A, 14B) arranged and configured to tend to cause said first opening period and said second opening period of said at least one intake valve (V) at each revolution of the cam (14), and

- in said condition wherein the engine load is below a predetermined threshold and/or the engine speed is below a predetermined threshold, the electronic controller (25) maintains said electrically actuated control valve (24) constantly closed,

8. The engine according to claim 1 , characterized in that said cam (14) associated with said at least one intake valve (V) of each engine cylinder has a boot shape, with a first part defining a first maximum point (M1 ), followed by a second bell-shaped part, defining a second maximum point (M2), higher than the first maximum point, and

- in said condition wherein the engine load is below a predetermined threshold and/or the engine speed is below a predetermined threshold, the electronic controller (25) maintains said electrically actuated control valve (24) closed at the first opening period and the second opening period and open at an intermediate phase between the first opening period and the second opening period.

9. The engine according to claim 6, 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.

10. The engine according to claim 6, characterized in that it comprises 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.

11. The engine according to claim 6, characterized in that said with camshaft (11 ) there is associated a phase variator device configured to vary the angular position of said camshaft with respect to said crankshaft.

12. The engine according to claim 6, characterized in that the cam (14) associated with each intake valve (V1 ; V2) of each engine cylinder is of a type with multiple profiles which can be rendered selectively operative.

13. The engine according to claim 9, characterized in that the electronic controller (25) is configured such that in said condition in which the engine load is below a predetermined threshold and/or the engine speed is below a predetermined threshold both intake valves are opened during the first opening period and only one intake valve is opened during the second opening period.

14. A method for controlling an internal combustion engine, wherein the engine includes:

- 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 including an intake stage, a compression stage, an expansion stage and an exhaust stage, - at least one intake valve (V1 , V2) associated with each engine cylinder to control a flow of intake air from a respective intake duct (5) during the intake stage into the cylinder in each cylinder operating cycle,

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

- an electrically actuated control valve (24) adapted to connect said pressurized fluid chamber (C) to a low-pressure drain channel (23) communicating with a pressurized fluid accumulator (270), such that when said control valve (24) is open, pressurized fluid drains from the pressure chamber (C) into said low-pressure drain channel and said at least one intake valve (V) closes due to the action of the respective return spring (9), regardless 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, based on a plurality of engine operating parameters, including engine load and engine rotational speed, said method being characterized in that:

- said cam (14) associated with said at least one intake valve (V) of each engine cylinder is configurated so that, when the pressure chamber is constantly maintained under pressure, a lift profile of said at least one intake valve is generated, as a function of the engine crank angle, which lift profile has two maximum points spaced apart from each other, said method comprising controlling, by means of said electronic controller, the electrically actuated control valve (24) associated with said hydraulic circuit of each cylinder, so as to cause, at each engine cycle, when the engine load is below a predetermined threshold value and/or when the engine rotational speed is below a predetermined threshold value, a first opening period and a second opening period of said at least one intake valve, with an intermediate phase in which said at least one intake valve is fully closed while the respective piston is descending towards its BDC, so as to create a depression in the respective cylinder which generates a jet of air entering the cylinder during the subsequent second opening period of said at least one intake valve, wherein said first opening period of said at least one 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 at least one intake valve of each cylinder begins when the piston has passed the BDC and is ascending towards the TDC, so that during a first part of said second opening period, while the piston is ascending towards the TDC, said jet of air is created entering the cylinder due to the depression created in the cylinder between the first opening period and the second opening period, while during a second part of the second opening period the depression in the cylinder is progressively reduced until it is eliminated and a portion of air is pushed back into the intake duct by the piston continuing to ascend towards the TDC, said second opening period having a duration sufficient to generate an effective compression ratio which is at least 5 and sufficiently high to generate a temperature of the mixture within the cylinder, when the piston reaches the TDC, of at least 500 K, where the effective compression ratio is defined as the ratio between the volume of the combustion chamber when the piston has risen to a position for which the pressure in the cylinder exceeds the value of 1 bar A, after which a compression stage begins with a monotonic increase in pressure up to when the piston reaches the TDC, and the volume of the combustion chamber when the piston is at TDC.

15. The method according to claim 14, wherein with each cylinder there are associated two intake valves (V1 , V2), characterized in that the electronic controller (25) is configured to cause at least one of the intake valves (V1 , V2) to open during the final part of the exhaust stage of the combustion gases, so as to cause the intake of combustion gases inside the intake ducts, thus providing an internal EGR.

16. The method according to claim 14, characterized in that during the first opening period and/or the second opening period of the intake valves, a predetermined amount of internal EGR is drawn into the combustion chamber, which contributes to reaching the minimum temperature of 500K at the end of the compression stage, said amount of internal EGR being preferably 35% of the mass of air drawn into the cylinder.