DESCRIPTION

Fields of application

The invention relates to the technical field of reciprocating internal- combustion engines; in particular the invention relates to an engine having an expansion ratio greater than the compression ratio.

Prior art

The term "reciprocating engine" denotes a cylinder and piston engine, namely an engine comprising one or more plungers (pistons) slidable with a reciprocating movement inside respective cylinders.

The vast majority of reciprocating engines use an ordinary centred crank mechanism for connecting the piston to the driving shaft. Said crank mechanism is formed by the known connecting rod / crank system which constrains the piston to perform a predefined and constant stroke between a top dead centre (TDC) and a bottom dead centre (BDC). Said stroke determines the compression ratio of the engine and obviously the displacement.

The resultant thermodynamic cycle (for example Otto or Diesel) has a compression ratio equal to the expansion ratio, owing to the fact that the piston stroke between the BDC and the TDC is always the same. It is known that a drawback of these thermodynamic cycles is that the gases expelled from the cylinder still contain a considerable fraction of the energy provided by the fuel, which means, in other words, the expansion phase does not reach a full exploitation of the energy released by the combustion.

In order to overcome this drawback a thermodynamic cycle with an expansion ratio greater than the compression ratio has been proposed, known as Atkinson cycle. The Atkinson cycle can extract more energy from the combusted gases before expelling them, compared to a conventional Otto cycle, and consequently has a greater efficiency. It is therefore very attractive from a thermodynamic point of view; however, it requires a crank mechanism able to provide the piston with a different stroke depending on the cycle phase, in particular a longer stroke during the expansion phase.

The construction of such a crank mechanism introduces some complications and so far has prevented the widespread use of engines operating with this principle. For example a known type of crank mechanism able to perform the Atkinson cycle has a toggle system which, however, is practically unusable in modern-day internal-combustion engines owing to its weight and dimensions. Another possible solution consists in a crank mechanism comprising an epicyclic gearing. Epicyclic systems are more compact and efficient than a toggle system, but are however costly and are subject to greater internal losses (friction and viscous resistance) than the original crank mechanism.

These difficulties have discouraged so far the use of Atkinson cycle engines. The technology of the ordinary crank mechanism is nowadays a well- established one and is preferred despite its drawbacks.

In the automotive field, especially in hybrid vehicles, there are engines which, while adopting the ordinary crank mechanism, implement a variant of the Atkinson cycle by increasing the opening time of the intake valves. This measure reduces the compression ratio even though the mechanical piston stroke between the TDC and the BDC is always the same; it therefore does not manage to exploit fully the energy of the combustion gases, leaving a margin for improvement. Moreover it tends to reduce the power and an increase in the valve opening time is only possible within certain limits. Other ways of making use of the residual energy of exhaust gases comprise for example the supercharging or, in cogeneration systems, the heat recovery, but require additional components other than the engine and are therefore costly and not always applicable. Summary of the invention

The invention aims to solve the aforementioned problem of how to provide a reciprocating engine having a varied stroke, more specifically with an expansion stroke greater than the intake/compression stroke, able to perform the thermodynamic cycle known as the Atkinson cycle, in an efficient manner and with low costs, dimensions and weight.

The object is achieved with a reciprocating internal-combustion engine comprising: a driving shaft; at least two cylinders with axes angularly spaced by 120 degrees; a respective piston slidable in each of said cylinders; an epicyclic crank mechanism which connects the pistons to the driving shaft and which converts the reciprocating movement of the pistons into a rotational movement,

characterized in that:

said epicyclic crank mechanism comprises a fixed ring gear coaxial with the driving shaft and having inner teeth, a pinion which meshes with said ring gear, at least one eccentric and at least one crank with a respective crank pin;

said eccentric is rigidly connected to said pinion and has an eccentricity distance with respect to the axis of the said pinion;

the eccentric is pivoted rotatably about said crank pin so as to rotate about the axis of the pinion, and the pistons are connected to the eccentric by means of connecting rods.

Preferably the pinion has a radius Rp which is twice the crank radius Rm, while the ring gear has a radius Rc which is three times the crank radius Rm. Said radii Rp and Rc are referred to the pitch circle of the outer teeth of the pinion and of the inner teeth of the ring gear, respectively. The crank radius Rm is the distance between the axis of the driving shaft and the axis of the crank pin.

The crank mechanism also has, advantageously, a suitable angular phase difference between crank and eccentric. Said phase difference may be defined as the angle formed between the crank and a line joining together the centre of the pinion and the centre of the eccentric. Preferably said phase difference is equal to 90° or about 90°.

In a 4-stroke engine, the kinematic system of the invention produces a reciprocating movement of the piston with an expansion stroke greater than the compression stroke. Consequently the expansion ratio is greater than the compression ratio. Said movement of the piston is the result of the combination of rotation of the crank with rotation of the eccentric about the crank pin and the fact that the pistons are connected via connecting rods to the said eccentric. In greater detail, the difference between the compression and expansion strokes is related to the eccentricity.

The main advantage of the invention is that the kinematic mechanism is simple, compact and light; it also allows the connection to the same eccentric of 2 pistons in a 120° V arrangement or 3 pistons in a star (radial) arrangement.

The aforementioned ratios between crank radius Rm, pinion radius Rp (equal to 2 Rm) and ring gear radius Rc (equal to 3 Rm) provide the following advantages: the pinion has a relatively large diameter to reduce the stress on the teeth during the combustion phase; the viscous friction losses are limited and comparable with those of an ordinary crank mechanism owing to the fact that the relative speed between the rod and the eccentric is relatively low, being half the speed of the driving shaft, while the pin between the crank and the eccentric, which rotates at a higher speed, has a small diameter to limit the losses.

It should also be noted that the crank mechanism according to the invention comprises only two additional components compared to the ordinary crank mechanism, that is the fixed ring gear and the pinion with integral eccentric. Said two components may be realized with known methods which are not particularly expensive, and therefore the cost of the crank mechanism according to the invention is similar to the cost of an ordinary crank mechanism.

Even more advantageously, the engine comprises two sets of cylinders in a V or radial arrangement on opposite sides of said fixed ring gear of the crank mechanism. These embodiments may be referred to as twin V or twin radial respectively.

In a twin V or twin star engine, the crank mechanism comprises preferably two eccentrics which are fixed to the same pinion, that is one eccentric for each of the two sets of cylinders. The invention therefore makes it possible to provide multi-cylinder engines (up to 6 cylinders) making use of a single epicyclic crank mechanism and a single crank, with a very small axial dimension significantly smaller than conventional V or in-line engines. The shorter driving shaft is less subject to torsional stress and more generally the reduced axial dimension facilitates installation of the engine.

It should be noted that the engine according to the invention is modular and may be made with an even greater number of cylinders. An engine with up to 6 cylinders may be made with a single crank; thus, for example, with two cranks it is possible to operate 12 cylinders obtaining an extremely compact and light structure in relation to the number of cylinders. The multi-cylinder embodiments are particularly attractive because the crank mechanism according to the invention may have a cost slightly higher than a conventional crank mechanism; in a multi-cylinder engine, however, this additional cost is offset by the major advantage of compactness and superior efficiency of the Atkinson cycle. A multi-cylinder engine according to the invention has a high efficiency and a favourable weight/power ratio which makes it highly competitive compared to the conventional technology. The radial cylinder arrangement is particularly preferred, making it possible to have 3 cylinders (single radial arrangement) or 6 cylinders (twin radial arrangement) per single crank; however, the engine according to the invention may comprise one or more sets of two cylinders at 120 degrees (V arrangement).

The connecting rods between the piston and the eccentric may comprise a single connecting rod for each piston or a system comprising a master connecting rod and secondary connecting rods. These terms refer to a system comprising a main rod which is suitably shaped and has a base connected to one of the pistons and a head pivotably mounted directly on the eccentric and respective secondary rods for the other pistons, said secondary rods being pivotably mounted on the master rod instead of directly on the eccentric. This variant is also applicable to multi-V or multi-radial engines.

The advantages of a master and secondary rod connection system are a reduction in the axial dimensions and the constructional simplicity since only the master rod is connected to the eccentric. The connection to the eccentric must be precise and typically requires a hydrodynamic bearing, while the connection between the secondary rods and the master rod may generally be performed using cheaper means. A hydrodynamic bearing introduces losses due to viscous friction; the invention allows these losses to be reduced since there is a single hydrodynamic bearing for 3 cylinders.

As a result of the invention it is possible to provide internal-combustion engines which are superior to the engines at present widely used and which are characterized by a high efficiency, due to implementation of the Atkinson cycle, and a high power compared to the size and weight, due to the compact arrangement of the cylinders, all of which with costs similar to those of conventional engines.

Another significant advantage is that good efficiency is achieved also with relatively low compression ratios. This is an advantage in the case where low-quality fuels with a low number of octanes are used. Moreover, with low compression ratios, the maximum temperatures are reduced, improving the working life of the "hot" components of the engine which are subject to less thermal stress.

The invention is applicable to small or large-size internal-combustion engines intended for any use, including for example stationary applications, the driving of operating machines, cogeneration units, and vehicle traction. In connection with vehicles, the invention is also applicable with success in the automotive sector, but also in the railway and naval sectors; in this two latter sectors the reduction in the amount of fuel required (owing to the greater efficiency) has a significant impact. Moreover, an engine according to the invention is fully compatible with the various technical measures already known in the sector of internal-combustion engines (e.g. supercharging, etc.) for increasing the power or the efficiency or for reducing fuel consumption, which may be applied as required.

According to a further aspect of the invention a system for varying the compression ratio is provided, said system acting by rotation of the ring gear by a predetermined angle relative to the axis of the driving shaft. By modifying the angular position of the ring gear, in fact, the top dead centre and/or bottom dead centre of the pistons and is/are displaced and, consequently, the compression ratio is increased or decreased.

Said system preferably comprises a control unit which determines an angular position of the ring gear depending on one or more operating parameters of the engine.

Said system preferably acts by continuously varying the compression ratio depending on the conditions of use and achieves the advantage of optimizing the torque curve of the engine and further improving the efficiency of the said engine. Preferably the system modifies the compression ratio depending at least on the number of revolutions and the position of the accelerator. In some embodiments, the system may modify the compression ratio depending on further parameters.

The advantage of a constructional solution with variable compression ratio is essentially an increase in the efficiency, especially under partial load conditions. Moreover, it will be noted that, since the compression ratio may be modified by rotating the ring gear, an engine according to the present invention may be provided with a variable compression ratio using a system which is relatively simple from a mechanical point of view.

The advantages will become even clearer from the following description which illustrates a number of preferred embodiments with the aid of the figures.

Description of the figures

Fig. 1 is a front view of the main components of a piston-type reciprocating engine according to an embodiment of the invention.

Fig. 2 shows an exploded view of the components of Fig. 1.

Fig. 3 is a diagram which illustrates a number of operating parameters of the crank mechanism according to the invention.

Fig. 4 is a diagram which illustrates another parameter of the crank mechanism, in particular the phase difference of the eccentric.

Fig. 5 is a detail, on a larger scale, of Fig. 3.

Fig. 6 is a diagram which illustrates the movement of the piston in an engine according to the invention.

Fig. 7 shows the components of an engine according to another embodiment, comprising two sets of cylinders on a single crank.

Fig. 8 is an exemplary cross-section of an engine of the type shown in Fig. 7. Fig. 9 shows a further embodiment of the invention.

Fig. 10 is a diagram illustrating the movement of the piston in the embodiment of Fig. 9.

Detailed description

Fig. 1 shows the components of an engine with three 120° radial cylinders and indicates the cylinder axes 1.1 , 1.2 and 1.3. Respective pistons 2.1 , 2.2 and 2.3 slide inside the cylinders. The driving shaft is perpendicular to the plane of Fig. 1 and has an axis of rotation passing through the point 3 of intersection of the cylinder axes.

The head of the engine (valves, ducts, etc.) is not essential for the purposes of the invention, it can be constructed using known techniques and therefore will not be described in detail.

For the understanding of the invention, instead, it should be noted that the engine comprises an epicyclic crank mechanism which converts the reciprocating movement of the pistons 2.1 , 2.2 and 2.3 into a rotational movement of the driving shaft about the axis 3. Said crank mechanism essentially includes a ring gear 4, a pinion 5, an eccentric 6 and a crank 7.

The ring gear 4 has inner teeth 8 with a pitch circle radius Rc; the pinion 5 meshes with said ring gear 4 via outer teeth 9 with a pitch circle radius Rp which mesh with the teeth 8. The ring gear 4 in this embodiment is fixed, i.e. is for example fastened to the chassis or to the engine base.

The eccentric 6 is rigidly connected to said pinion 5 and has an eccentricity distance ε with respect to the axis of the said pinion.

The crank 7 supports a pin 10 and has a crank radius Rm which corresponds to the distance between the axis 3 and the axis of said pin 10. The eccentric 6 is mounted rotatably on said pin 10. In particular the pin 10 is received in the seat 22 which, as can be noted in Fig. 2, is axially offset from the substantially cylindrical body of the eccentric 6 and coincides with the axis of the pinion 5.

The pistons are connected to the eccentric 6 by means of connecting rods which in the example of Fig. 1 are formed by a master (or main) rod 11 for the piston 2.1 and two small (secondary) rods 12 for the remaining pistons 2.1 and 2.3. The master rod 11 is directly pivoted on the eccentric 6, while the secondary rods 12 are connected to the master rod 11 via pins 13. The connection of the rod base to the pistons may be performed in a conventional manner, for example by means of gudgeon pins 18.

The components of the crank mechanism are even more clearly visible in the exploded view of Fig. 2 which shows the journals 14, corresponding to the engine axis 3, and the pin 10 on which the eccentric 6 is rotatable. It is possible to note also the particular form of the head 15 of the master rod 11 which for example has a pair of flanges 16 which seat the pins 13 of the secondary rods 12, and the bearing 17 which forms the rotary coupling between the head of the master rod 11 and the eccentric 6.

The main parameters which define the geometry of the crank mechanism are the said radii Rc, Rp and Rm and the eccentricity ε. Said parameters are shown in the diagram of Fig. 3 and are even more clearly visible in Fig. 5. The following relations exist between the radii Rp, Rm and Rc:

Rp = 2 Rm;

Rc = 3 Rm;

Fig. 4 shows another geometric parameter of the crank, i.e. the phase difference δ between eccentric and crank. Preferably said phase difference is equal to 90° as shown in the figure.

Figs. 3 and 4 for the sake of simplicity show only one piston and do not show the secondary rods 12. With reference in particular to Figs. 3-5, the mode of operation of the crank mechanism according to the invention can be understood.

It is assumed for example that the crank has a speed of rotation (angular speed) ω about the engine axis 3. The pin 10 rotates with said speed ω following a circular path 19 (Fig. 5) of radius Rm, imposed by the constraint of the crank. The eccentric 6, being mounted on the pin 10, follows the same path and also, owing to meshing with the fixed ring gear 4, rotates about the pin 10 in the opposite direction at said speed ω. Owing to the dimensional relationship between the radii, the speed of rotation of the eccentric 6 with respect to the fixed reference is half the speed of rotation of the crank 7. Therefore the following is obtained: angular speed of the crank = ω angular speed of the eccentric = - ω / 2.

Since the rod 11 is pivotably mounted on the eccentric 6, the head of the rod follows the movement of the eccentric which is composed of the circular movement of the pin 10 and the rotation of the eccentric 6 about the pin at a speed which is halved and opposite.

Owing to the existing constraints as well as the eccentricity ε, the resultant movement of the piston 2.1 is derived approximately from the combination of: a harmonic motion with amplitude Rm and angular frequency ω, generated by the movement of the crank pin 10 relative to the fixed chassis; a harmonic motion with amplitude ε and angular frequency ω / 2 generated by the rotation of the eccentric on the pin 10.

This relation is approximate since it ignores the inclination of the rod 11 with respect to the cylinder axis, but in any case allows the operating principle to be understood. With suitable synchronism between these two movements, determined by the geometry of the system and in particular by the phase difference between crank and eccentric, the resultant movement of the piston reaches alternately, every two revolutions, a first bottom dead centre and a second bottom dead centre, at a different distance from the top dead centre. Fig. 6 is a graph showing on the horizontal axis the angle of rotation of the driving shaft from 0 to 720° (two revolutions) and on the vertical axis the linear displacement of the piston. The broken lines I and II represent the two aforementioned harmonic motions, and the continuous line III represents the motion resulting from their combination (movement of the piston). Introducing the time t and the angle of rotation a defined as α = ω t as well as the angular phase difference δ between crank and eccentric, the position of the piston along the axis 1 may be described by the equation: ε sin (a/2) + Rm sin (a + δ) which corresponds to the line III. The difference between the compression stroke and the expansion stroke is equal to 2 ε.

The engine operates with the known intake, compression, expansion and exhaust phases. The first bottom dead centre is reached at the end of the intake phase, while the second bottom dead centre is reached at the end of the expansion phase. Consequently, the expansion stroke is greater than the compression stroke and the engine performs the desired Atkinson cycle.

In the figures, and in particular in Fig. 2, the compactness of the engine according to the invention can be appreciated.

Fig. 7 shows an embodiment with two sets of pistons 20, 21 on opposite sides of a single fixed ring gear 4. In the example, the two sets 20, 21 have three pistons each, i.e. the first set 20 comprises the pistons 2.1 - 2.3 and the second set comprises the pistons 2.4 - 2.6. The pistons 2.1 - 2.3 and 2.4 - 2.6, respectively, have a 120-degree radial arrangement and consequently the engine is a 6-cylinder twin-radial engine.

The set 20 is similar to that shown in Figs. 1 and 2 and is mounted on the eccentric 6; the set 21 constructionally is identical to the set 20 and is mounted on a second eccentric 6a (Fig. 8), also integral with the pinion 5. This embodiment is extremely compact and provides the engine with a considerable power compared to weight and size, allowing the six pistons 2.1 - 2.6 to be operated with a single crank. Basically, with the embodiment shown in Fig. 7 the number of cylinders may be doubled by adding only one eccentric.

It should be noted that, in order to obtain the correct kinematics, the angular phase difference between the two eccentrics 6 and 6s must have a value defined on the basis of the angular phase difference of the two sets of cylinders. For example, given a generic angular phase difference β between said two eccentrics, the pistons of the second set connected to the second eccentric 6a (and respective cylinders) have axes offset by a suitable angle φ with respect to the pistons and cylinders of the first set connected to the eccentric 6, and said angle (in degrees) may be calculated with the formula: φ = 120° - 2/3 β.

From the point of view of balancing, it may be pointed out that each set of 3 radial cylinders has the inertia forces of the 1st order balanced (by means of counterweights on the driving shaft), while those of the 2nd order may be balanced for example with a counter-rotating shaft.

In other versions (not shown) the engine according to the invention may comprise one or more sets of two 120-degree cylinders (V arrangement). Embodiments with two or more radial or V arrangements of cylinders are therefore possible.

The invention achieves the objects mentioned above, providing an attractive solution for implementing the Atkinson thermodynamic cycle. Figs. 9 and 10 refer to a further embodiment of the invention which envisages a variable compression ratio. The variation in the compression ratio is obtained by controlling the angular position of the ring gear 4 which, in this embodiment, may rotate at least through a certain angle with respect to the engine axis.

The engine comprises a system 20 able to perform a rotation of the ring gear 4 about the engine axis and relative to the base 30.

The ring gear 4 has a reference position which determines the nominal (rated) compression ratio of the engine. Said system 20 is able to move the ring gear 4 into an angular position different from the reference position, which corresponds to a compression ratio greater or lesser than the nominal compression ratio. The angle of rotation defined for the ring gear 4 preferably is comprised between +3° and -3° with respect to said reference position ("zero" position).

Said system 20 in greater detail comprises: a processing and control unit 21 ; a controller (driver) 22, an electric servomotor 23. The servomotor 23 has a shaft 24 provided with teeth 25 of the involute or "endless screw" type which mesh with teeth 26 formed on the outer edge of the ring gear 4.

The system in Fig. 9 is shown purely by way of example: other embodiments, such as a recirculating ball screw operating system, are possible.

As a result of meshing between the teeth 25 and 26, a rotation of the shaft 24 causes a rotation of the ring gear 4 about the engine axis, as indicated by the arrows in Fig. 9. In this way the system 20 may move the ring gear into an angular position calculated by the unit 21.

Advantageously, the unit 21 determines the position of the ring gear 4 on the basis of data relating to operation of the engine, such as one or more of the following parameters: number of revolutions, position of the accelerator; pressure and temperature of the feed air; signal of a knock sensor, able to detect the risk of detonation, etc. This input data is indicated by the arrow 27. Moreover, as indicated by the line 28, the unit 21 may receive a feedback signal relating to the current position of the ring gear 4. On the basis of these input signals, the unit 21 produces a signal 29 which corresponds to the calculated position (target position) of the ring gear 4 and which is transmitted to the controller 22. The controller 22, on the basis of the signal 29, controls operation of the servomotor 23, causing a rotation of the shaft 24 so as to reach the target position of the ring gear 4.

The adjustment described above is controlled entirely by the unit 21 and does not require any action on the part of the driver.

Fig. 10 is a graph similar to that of Fig, 6 and shows the effect of the angular displacement of the ring gear 4. The curve III indicated by a continuous line represents the movement of a piston with ring gear 4 in the reference position (zero), while the curve lll-a indicated by a broken line represents the movement of the same piston with ring gear 4 rotated through an angle Θ (theta). The figure shows how the top dead centre of the piston is raised by a quantity s. Consequently, the compression ratio is modified.

Advantageously, the control system 20 adapts continuously the position of the ring gear 4, and therefore the compression ratio, depending on the operating conditions such as the number of revolutions, load, etc. In this way a further benefit in terms of energy efficiency of the engine is obtained and the consumption is therefore reduced.