DESCRIPTION

Field of application

The present invention relates to positive-displacement piston pumps. More specifically, the invention regards an axial piston pump including a housing, a cylinder block in that housing, a plurality of cylinders arranged around a longitudinal axis of the cylinder block, a corresponding plurality of pistons, each piston being slidable in the respective cylinder of the plurality of cylinders, and a fluid distributor, which is associated with the cylinder block.

Prior art

As it is well known in this specific technical field, a positive-displacement piston pump generally comprises a plurality of pistons housed in a corresponding plurality of rotating cylinders. The pistons, sliding inside the rotating cylinders, cause a suction with an increase of volume in a pumping chamber inside the cylinder when a suction port is open, creating a slight drop in pressure that allows a fluid to be drawn into the pumping chamber. Alternatively, those pistons cause a compression with a decrease of volume in the pumping chamber when a delivery port is open, i.e. when their motion is reversed.

In the axial piston pumps, the piston stroke occurs in the same direction of the axis of a cylinder block that rotates within a fixed housing, the motion of the pistons being caused by their sliding along a tilted plate (swash plate), whose tilt is adjustable in variable flow rate configurations. The main problem that the piston pump manufacturers must deal with is the noisiness of the pumps themselves. In fact, there are two fundamental causes of noise in piston pumps, namely the noise generated in the fluid (also called "fluidborne noise") and the structure noise (also called "structureborne noise"). The fluidborne noise is mainly due to irregularities, or "ripples" during fluid transfer. In particular, the fluidborne noise is due to the structure of the pumps themselves, which transfer the fluid irregularly and discontinuously, that irregularity being called ripple.

The ripple is caused both by the fluid transfer law (and therefore by the system geometry, in this case it is referred to as primary ripple) and by the irregularities of the switching between delivery and suction (in this case it is referred to as secondary ripple), this latter phenomenon also being related to the fluid compressibility. In general, the ripple induces an operating noise caused by fluctuations of the instantaneous flow rate over time. The fluctuations of the instantaneous flow rate over time generate a pulsating wave which is transmitted through the fluid to the surrounding environment and in particular to the pump walls, to the pipes and to the delivery ducts. The induced noise can reach unpredictable levels, particularly if the foregoing members resonate with the oscillation frequency.

The fluid compressibility exerts influence upon the sudden opening of the piston pumping chamber at the delivery port causing a sudden compression of the fluid, which leads to a consequential fluctuation of the instantaneous flow rate over time. Generally, a pressure of 150 bar causes a reduction in volume of approximately 1-2% of the fluid, generally oil, if devoid of entrapped air, but also a double reduction of volume even if there is a small amount of entrapped air.

The switching between delivery and suction occurring when the pump flow rate (i.e. the linear velocity of the piston) is other than zero, renders the compensation of the foregoing fluid compressibility phenomenon with the change in velocity impossible, particularly when said switching takes place in a phase in which the acceleration of the pistons is at its maximum, causing considerable irregularities in the flow rate itself.

The structureborne noise originates from the imbalance and from the fluctuation of the forces acting in the pump during the movement of the pistons, causing structural vibrations. In particular, that source of noise is inherent to the structure of the pump and has its origin both in the delivery irregularities, causing the development of oscillating forces, and in the pulsating frictional forces acting within the pump, such as for example the force of friction acting between the pistons and the swash plate.

In the past, some piston pump solutions were proposed, aimed to resolve the foregoing issues, in particular the issue of eliminating the causes of structureborne noise. In particular, figure 1A schematically shows a sectional view of an axial piston pump in accordance with the known techniques, indicated globally with 1. The pump 1 comprises a fixed housing 2 within which a rotating cylinder block 3, which houses a plurality of pistons 4. Each piston of the plurality of pistons 4 is coupled, generally via ball joints, to a corresponding plurality of hydrostatic slipper pads 5 that are arranged externally to the cylinders and slide on a swash plate 6. The rotation of the rotating cylinder block 3 is caused by a driving shaft 7. In this way, the pistons of the plurality of pistons 4 slide along the swash plate 6 through the plurality of hydrostatic slipper pads 5. The presence of the hydrostatic slipper pads 5 is advantageous for the reduction of the friction between the plurality of pistons 4 and the swash plate 6 and therefore it is advantageous for the reduction of the wear of pump 1 as a whole.

Furthermore, in the axial piston pump 1 , a fluid distributor 8, associated with the rotating cylinder block 3 and integral with the fixed housing 2, is provided, recesses corresponding to suction and delivery ports being located on such a distributor.

Though advantageous under various aspects, this solution has various drawbacks, in particular due to the presence of the hydrostatic slipper pads 5, which on one hand are necessary to reduce the friction and the wear and tear, but on the other hand lead to a substantially sinusoidal- type law of motion of the pistons, causing a non-constant instantaneous flow rate and therefore a ripple.

The hydrostatic slipper pads 5 slide in fact with circular motion over the swash plate 6, those hydrostatic slipper pads requiring a support whose contour does vary with rotation and requiring a minimum and constant opening (meatus) between themselves and the bearing surface. The motion of the slipper pads in any case leads to a disadvantageous substantially sinusoidal-type law of motion of the pistons, causing the foregoing ripple.

With reference to figure IB, a dimensionless graph of the instantaneous velocity of a single piston (i.e. the instantaneous flow rate of a single piston) is shown in the suction phase and in the delivery phase, as a function of the angle of rotation (expressed in radians) of a piston pump manufactured according to the prior art. In that figure, the suction phase corresponds to the negative portion of the graph and the delivery phase corresponds to the positive portion of the graph.

Considering, in general, a positive-displacement pump comprising a plurality of pistons, said velocity law does not allow a constant sum of the piston velocities (and therefore a constant flow rate) to be obtained for any angular phase shift of the pistons and therefore it causes ripple.

The ripple tends to zero as the number of pistons approaches infinity, however said increase of the number of the pistons obviously leads to an increase of the cost, to an increase of the opening and closing noises of the suction and delivery ports, and also to the increase of the volumetric losses.

In the known designs, the ripple also leads to harmful cusps and discontinuities that also reflect in an irregularity in the torque absorbed by the pump.

Furthermore, the substantially sinusoidal trend of the piston motion law entails a high noise due to the switching between the suction and delivery phase, the switching taking place in the phase of maximum acceleration of the pistons and with a non-zero velocity of the pistons (and therefore of the fluid), this velocity being equal to zero only in one point, as it can be seen in figure IB.

In an additional known solution proposed in the Parker pumps, a pre- compensation chamber is used, said chamber reducing the fluidborne noise generated by the compression of the fluid due to the sudden opening of the piston pumping chamber at the delivery port.

The technical problem of the present invention is to provide an axial piston pump, having structural and functional characteristics so as to allow overcoming the limits and the difficulties that still afflict the axial piston pumps according to the prior art, in particular eliminating the most important causes of fluidborne noise and of structureborne noise.

The purpose of the present invention is to devise an axial piston pump that can provide a constant flow rate and therefore a null ripple.

Another purpose of the present invention is to devise an axial piston pump able to reduce the effect of the switching irregularity between the delivery phase and the suction phase and to reduce the fluid-compressibility effect.

Another purpose of the present invention is to devise an axial piston pump featuring a mechanical design having balanced forces to eliminate vibrations and featuring a mechanical system apt to minimize friction and wear and tear.

Summary of the invention

The solution idea at the basis of the present invention is to provide an axial piston pump whose law of motion is determined by a cam profile on which the pistons slide without friction, providing a constant flow rate, reducing the effect of the switching irregularity and guaranteeing a proper mechanical operation.

Based on this solution idea, the foregoing technical problem is solved by an axial piston pump including a housing, a cylinder block in the housing, a plurality of cylinders arranged around an axis of the cylinder block, a corresponding plurality of pistons, each piston being slidable in the respective cylinder of the plurality of cylinders, a fluid distributor, which is associated with the cylinder block, and a frontal cam profile in turn including a plurality of cam lobes, each lobe comprising a delivery ramp and a suction ramp, characterized in that at least two pistons of the plurality of pistons have an end portion that protrudes from the respective cylinder and is a follower simultaneously engaging with a same type of ramp of a relative lobe of said plurality of cam lobes. It is noted that the follower, engaging with the frontal cam profile, may comprise a rolling member associated with the end portion of each piston of the plurality of pistons, such a rolling member being selected among a roller or a sphere.

Furthermore, a support bearing can be interposed between each piston of the plurality of pistons and the respective rolling member.

In particular, the support bearing can be a hydrostatic support bearing or a hydrodynamic support bearing.

Preferably, the support bearing is a hydrostatic support bearing having a hydrodynamic component being less than 50% of the total support.

Suitably, the fluid distributor may comprise delivery ports and a plurality of pre-injection holes that are interposed between the delivery ports, the plurality of pre-injection holes being apt to put in communication the cylinders of the plurality of cylinders in advance with the delivery ports of the fluid distributor.

Furthermore, the end portion of the pistons of the plurality of pistons can define a surface engaging with a cam counter-profile, the cam counter- profile being integral with the frontal cam profile and being conjugated to that frontal cam profile, realizing a desmodromic control.

According to an aspect of the present invention, the pump may comprise twelve pistons arranged around the axis of the cylinder block and separated from each other by a relative shift of 30°, the frontal cam profile comprising three cam lobes equidistant from each other on an inner wall of the housing, each of them occupying a portion of the inner wall corresponding to an angle of 120.

According to a further aspect of the present invention, a piston can be relatively shifted from another piston, with respect to the start of a same type of ramp of a relative lobe with which they are simultaneously engaged, by an angle that corresponds to half of an active angular extension of the delivery (or suction) portions of the frontal cam profile.

Moreover, according to the present invention, the frontal cam profile can be apt to impose to the at least two pistons, having the respective followers simultaneously engaging with a same type of ramp of a relative lobe, a law of motion such that the sum of the velocities of those at least two pistons is substantially constant for each angle of rotation of the axial piston pump.

Suitably, the law of motion can provide, in the function describing the velocity νι(φ) of the at least two pistons of the plurality of pistons, a horizontal inflection point between a suction phase and a delivery phase, with null velocity and null acceleration for those at least two pistons of the plurality of pistons during the switching between the suction phase and the delivery phase.

Advantageously according to the present invention, the law of motion may be substantially described by the integral of the function νι(φ) = C * (1 + οοβ(φ)), where φ is the angle of rotation of the pump or a linear function of the angle of rotation of the pump and C is a proportionality factor.

It is also noted that the law of motion can be an even function when the motion of the at least two pistons is considered only along the delivery ramp or only along the suction ramp of a lobe of the plurality of cam lobes, as well the law of motion can be an odd function when the motion of the at least two pistons is considered along an entire lobe of the plurality of cam lobes.

Finally, the law of motion can be changed based on the fluid compressibility. The features and advantages of the axial piston pump according to the invention will become apparent from the following description of an embodiment thereof, given by way of non-limiting example with reference to the accompanying drawings.

Brief description of the drawings In those drawings:

- figure 1A shows a schematic sectional view of an axial piston pump according to the prior art;

- figure IB shows a graph of the instantaneous velocity of a single piston as a function of the angle of rotation of a piston pump according to the prior art, in the delivery phase and in the suction phase; - figure 2 shows a schematic sectional view of an axial piston pump in accordance with the present invention;

- figure 3A, shows a schematic sectional view of a port ion corresponding to a frontal profile of the axial piston pump of figure 2; - figure 3B shows a two -dimensional plane development view of the frontal cam profile of the axial piston pump according to an embodiment of t he present invention;

- figure 4 shows a schematic sectional view of a portion of the axial piston pump of figure 2. - figure 5 shows a schematic sectional view of a fluid distributor of the axial piston pump of figure 2;

- figure 6A shows a graph of a function describing the velocity of a piston of the piston pump according to the present invention;

- figure 6B shows a graph of a superimposition of two functions equal to the function represented in figure 6A but phase-shifted by π; figure 7 shows a graph of a function describing the velocity of a piston of the piston pump in accordance with the present invention between -π and +π;

- figure 8 A shows a graph of the law of motion of a single piston as a function of the angle of rotation of a piston pump according to the present invention, in the delivery phase and in the suction phase;

- figure 8B shows a graph of the instantaneous velocity of a single piston as a function of the angle of rotation of a piston pump in accordance with the present invention, in the delivery phase and in the suction phase; and - figure 9 shows a detail of the graph of the instantaneous velocity of a single piston of figure 8B as a function of the angle of rotation of a piston pump in accordance with the present invention.

Detailed description

With reference to those figures, and in particular to figure 2, a positive- displacement axial piston pump manufactured according to the present invention is globally and schematically indicated with 10.

It is worth noting that the figures represent schematic views and are not drawn to scale, but instead they are drawn so as to emphasize the important features of the invention. Moreover, in the figures, the different elements are depicted in a schematic manner, their shape varying depending on the application desired. Finally, it is noted that in the figures the same reference numbers refer to elements that are identical in shape or function. In the embodiment illustrated in figure 2, provided as a non-limiting example of the present invention, the axial piston pump 10 includes a fixed housing 1 1 , a rotating cylinder block 12 in the fixed housing 1 1 , a plurality of cylinders 13 arranged parallel around an axis (the longitudinal axis or, in this embodiment, the rotation axis) of the rotating cylinder block 12, a corresponding plurality of pistons 14, each piston being slidable in the respective cylinder of the plurality of cylinders 13, and a fluid distributor 15, which is associated with the rotating cylinder block 12 and is integral with the fixed housing 1 1. The fluid distributor 15 houses suction ducts 16 and delivery ducts 17 communicating with the cylinders of the plurality of cylinders 13, the suction ducts 16 and delivery ducts 17 terminating with respective suction ports 18 and delivery ports

19. The rotating cylinder block 12 is placed into rotation by a driving shaft

20, similar to that described in the prior art section.

Advantageously according to the invention, each piston of the plurality of pistons 14 has an end portion 21 , which protrudes from the respective cylinder and is a follower 22 engaging with a frontal cam profile 23, which is fixed and integral with the fixed housing 1 1.

In particular, the engagement with the frontal cam profile 23 is a rolling engagement, the rolling engagement comprising a rolling member 22' that supports the end portion 21 of each piston of the plurality of pistons 14.

The follower 22 therefore comprises a rolling member 22', associated with the end portion 21 of each piston of the plurality of pistons 14, the rolling member 22' preferably being selected from a roller or a sphere. In particular, it's preferable to adopt a single roller as rolling member 22', since a roller allows simpler mechanical machining, is more inexpensive, more reliable, and causes less Hertzian pressure and therefore less stress on the frontal cam profile 23. Furthermore, a support bearing 24 is interposed between each piston of the plurality of pistons 14 and the respective rolling member 22'. In this way, in accordance with the invention, the rolling member 22' on one hand rolls without scraping on the frontal cam profile 23 and on the other hand is supported by the support bearing 24, preventing a scraping contact with its housing in the respective piston. Consequentially, the frictions between the follower 22, the plurality of pistons 14, and the frontal cam profile 23 are minimized, thus minimizing the noise and the wear of the axial piston pump 10.

If a roller is adopted as the rolling member 22', the support bearing 24 can be more easily built.

The support bearing 24 can be a hydrostatic support bearing or a hydrodynamic support bearing, for example. The use of a hydrodynamic support bearing 24 entails on one hand a lower loading capacity, but on the other hand less volumetric loss. It is also possible to use a support bearing 24 with a predominantly hydrostatic component and with a smaller hydrodynamic component, preferably less than 50% of the total support.

With reference to figure 3A, a schematic sectional view of a portion corresponding to the frontal cam profile 23 of the axial pist on pump 10 of figure 2, is shown. The frontal cam profile 23 is made on at least one inner portion of the fixed housing 1 1 and figure 3B shows a two-dimensional plane development view thereof. The frontal cam profile 23 includes a plurality of cam lobes 25, each lobe comprising a delivery ramp Ri and a suction ramp R2, as shown in figure 3B. Obviously, a piston is in the delivery phase when its follower 22 is engaged with the delivery ramp Ri, whereas such a piston is in the suction phase when the follower 22 is engaged with the suction ramp R2, in accordance with the travel direction of the pistons indicated in that figure by the arrow P, which corresponds to the direction of rotation of the pump. It is therefore obvious that the pistons 14 are driven by the delivery ramp Ri of the frontal cam profile 23 towards the delivery duct 17 during the delivery phase.

In figure 3B the frontal cam profile 23 comprises three cam lobes, but the number of lobes of the plurality of cam lobes 25 may vary according to specific needs and/ or circumstances, the figure 3B being provided only as an example that does not limit the scope of the present invention.

As can be seen in figure 4, which shows a schematic sectional view of a portion of the axial piston pump of figure 2, the end portion 21 of the pistons of t he plurality of pistons 14 is substantially T-shaped or mushroom-shaped, the end portion 21 therefore defining a surface 26 engaging with a cam counter- pro file 27, which is fixed and integral with the frontal cam profile 23, allowing a cont rolled stroke of the pistons of t he plurality of pistons 14 and therefore a positive control t hereof both in the delivery phase and in the suction phase, realizing a so-called desmodromic cont rol.

The cam counter-profile 27 has a profile that is conjugated to the frontal cam profile 23, thus allowing the drive of t he pistons of the plurality of pistons 14 towards said frontal cam profile 23 during the suction phase, so as t o allow the suct ion of fluid t hrough t he suction line 16. In a preferred embodiment , represented in the figures 2 and 4, the cam counter-profile 27 (and t herefore the desmodromic control) is made internally, i.e. acting on the portion of the end portion 21 located closer to the axis of rotation of t he rotating cylinder block 12 (or axis of rotat ion of the pump). Alternatively, it is certainly possible to use, instead of the foregoing desmodromic control, other control methods of the suction phase, for example based on springs.

As previously mentioned, in the axial piston pump 10 according to the present invention, suction ports and delivery ports are created and suitably arranged in the fluid distributor 15, as shown with greater detail in figure 5, which shows a schematic sectional view of the fluid distributor 15 of the piston pump of figure 2. As is evident in figure 5, a plurality of small pre-injection holes 28, properly calibrated, is interposed between the delivery ports 19 of the fluid distributor 15, the plurality of small pre- injection holes 28 being apt to put in communication the cylinders of the plurality of cylinders 13 in advance with respect to the delivery ports 19 of the fluid distributor 15. Advantageously, the foregoing configuration allows to reduce the fluidborne noise, which noise is due to the sudden opening of the piston pumping chamber, situated inside cylinder, at the delivery ports 19, the plurality of small pre-injection holes 28 anticipating in a gradual and controlled manner a miniscule passage of fluid that is used to compensate the compressibility thereof in the piston pumping chamber.

Furthermore, in the embodiment of the axial piston pump 10 of figure 2, the law of motion of the pistons of the plurality of pistons 14 is determined by the frontal cam profile 23.

Conveniently, at least two pistons of the plurality of pistons 14 have the respective followers 22 engaging simultaneously with a same type of ramp of a relative lobe of the plurality of cam lobes 25. Clearly, the at least two pistons may have the respective followers 22 engaging simultaneously with a same lobe of the plurality of cam lobes 25. But it can be provided also a configuration in which at least two pistons are respectively engaging with different lobes, but in any case engaging with the same type of ramp and having an appropriate relative shift with respect to the start of the respective ramp, said ramp thus possibly belonging to different lobes and said shift being such as to guarantee the constancy in the sum of the flow rates of the pistons, as it will be specified below. In others words, the invention provides that there are at least two pistons 14 with followers 22 simultaneously engaging with a frontal cam profile 23 in a same active phase (i.e. engaging with the same type of ramp) and with a relative shift (or relative distance) so as to guarantee a substantially constant sum of the overall pump flow rate. In this way, as it will be clearly evident from the following description, it is possible to produce the axial piston pump 10 in which the at least two pistons of the plurality of pistons 14 follow a law of motion such that the sum of their velocities is substantially constant, and therefore such that the instantaneous flow rate is substantially constant, the change of instantaneous flow rate of a piston in the delivery phase (or suction phase) being substantially compensated by the change of the instantaneous flow rate of the other piston in the delivery phase (or suction phase), the followers 22 of these at least two pistons engaging simultaneously with the same type of ramp of a relative lobe of the plurality of cam lobes 25.

The at least two pistons of the plurality of pistons 14, having the respective followers 22 engaging simultaneously with the same type of ramp of the relative lobe of the plurality of cam lobes 25, have a relative (angular) shift with respect to the start of the delivery ramp Ri or suction ramp I¾ of the relative lobe with which the respective follower 22 is engaged, so as to guarantee the constancy of the sum of the velocities of the pistons 14.

In particular, if two pistons of those pistons having the respective followers 22 engaging simultaneously with the same type of ramp of a relative lobe of the plurality of cam lobes 25 are considered, the relative angular shift necessary to guarantee the constancy of the sum of their velocities corresponds to half of the active angular extension of the delivery (or suction) portions of the frontal cam profile 23, as measured along the rotation direction of the pump. For the case in which at least two pistons are engaging with a same type of ramp of a same lobe of the plurality of cam lobes, the relative shift between two adjacent pistons corresponds to the ratio between the active angular extension of the delivery (or suction) portions of the frontal cam profile 23 and the total number of pistons of the plurality of pistons 14 having the respective followers 22 engaging simultaneously with the same type of ramp (i.e. always a delivery or suction ramp) of the same lobe of the plurality of cam lobes 25.

Here and hereinafter, the term "active angular extension" of the delivery (or suction) portions of the frontal cam profile 23 means the angle of rotation of the pump necessary for a single piston of the plurality of pistons 14 to carry out a full stroke (i.e. the complete delivery phase or suction phase) and hereafter it will be indicated with e _c . The number of lobes of the plurality of cam lobes 25 determines the number of complete strokes of each piston of the plurality of pistons 14 for every complete revolution of the cylinder block 12 and therefore determines the value of the active angular extension e _c .

As indicated in figure 3B, in an embodiment of the present invention, the frontal cam profile 23 comprises three cam lobes equidistant from each other on the inner wall of the fixed housing 1 1 , and therefore e _c = 60°.

In a preferred embodiment, the axial piston pump 10 comprises twelve pistons equally spaced, separated from each other by a relative shift of 30°, as measured in the rotation direction of the pump. In this way, there are always two pistons having the respective followers 22 engaging simultaneously with the same type of ramp of a relative lobe of the plurality of cam lobes 25, the pistons having a relative shift equal to 30° with respect to the start of the delivery ramp Ri or suction ramp I¾ of the relative lobe with which they are engaged. Consequentially, for this pair of pistons, the relative shift of 30° corresponds to e _c /2, and therefore is capable of guaranteeing constancy in the sum of the velocities.

Note that the number of pistons of the plurality of pistons 14, the value of the active angular extension e _c of the delivery (or suction) portions of the frontal cam profile 23, as well as the number of the lobes of the plurality of cam lobes 25, may vary according to specific needs and/ or circumstances, the above mentioned embodiment being provided only as an example and not limiting the scope of the present invention.

Advantageously according to the present invention, the axial piston pump 10 of the present invention allows an axial balancing of the forces since there is always a same number of pistons in the suction phase and in the delivery phase.

In this way, the constraints on the equilibrium of the forces are respected at every moment, the resultant of the forces being substantially null for every angle of rotation of the pump, and frictions and wear of the axial piston pump 10 are minimized. As a consequence, the most important causes of structureborne noise are eliminated in the axial piston pump 10.

As previously mentioned, advantageously according to the present invention, at least two pistons of the plurality of pistons 14 have the respective followers 22 engaging simultaneously with the same type of ramp (delivery ramp Ri or suction ramp I¾) of a relative lobe of the plurality of cam lobes 25, those at least two pistons following a law of motion and having a relative shift such that the sum of their velocities is substantially constant for each angle of rotation of the pump.

As a consequence, it is possible to realize an axial piston pump in which the sum of the velocities of all pistons whose followers are engaging with the delivery ramp Ri, as well as the sum of the velocities of all the pistons whose followers are engaging with the suction ramp R2, is substantially constant for each angle of rotation of the pump, thus resulting in a substantially constant instantaneous flow rate and eliminating the ripple.

As previously mentioned, in the known designs of the axial piston pumps, the presence of hydrostatic slipper pads leads to a disadvantageous, substantially sinusoidal-type, law of motion of the pistons. Advantageously, the axial piston pump 10 in accordance with the present invention allows a free selection of the law of motion of the pistons of the plurality of pistons 14 through the selection of the geometry of the frontal cam profile 23. It is therefore possible to select an appropriate shape for the frontal cam profile 23 so that the instantaneous flow rate of the axial piston pump 10 is substantially constant.

It is in fact known that the instantaneous flow rate Q(t) at time t of a positive-displacement piston pump is defined as:

Q(t) =∑iQi(t) =∑iAiVi(t) where the index i moves from 1 to n, n being the total number of pistons simultaneously involved in the delivery (or suction), A being the cross- sectional area of the i ^th piston, and Vi(t) being the instantaneous velocity of the i ^th piston.

As a consequence, in order to obtain a constant instantaneous flow rate and therefore in order to eliminate the ripple, ∑iVi(t) must be constant, i.e. invariant with t, and therefore∑iVi(t) = k must hold, where k is a constant.

In order to relate the above formula with the axial piston pump 10 that is physically manufactured, the instantaneous velocity Vi(t) of the i ^th piston will be linked to the angle of rotation φ of the axial piston pump 10, by calling βί(φ) the spatial coordinate of the i ^th piston as a function of the angle φ and by setting:

Vi(t) = 5si(t)/5t = δβί(φ)/δφ * δφ/δΐ = δβί(φ)/δφ * _ω = νΐ(φ) * ω where ω is the rotational velocity of the pump (ω = δφ/δΐ), which is considered to be constant, any changes thereof being irrelevant. The analysis thus switched from the time t domain to the angle of rotation of the pump φ domain, and therefore the condition for achieving a null ripple is written as: Σΐνΐ(φ) = Σίδ φΚδφ = k

As previously illustrated, in the known designs of the axial piston pumps, the condition expressed by the above equation (∑ίνί(φ) = ∑ίδβί(φ)/δφ = k) is never met. This happens because the known designs involve a sinusoidal or sinusoidal-like law of motion of the piston, such a law being imposed by the fact of having the pistons supported by a pad in circular motion on a swash plate. Such a law of motion, for a single i ^th piston, may be expressed by the following equation: leading to: νΐ(φ) = δβί(φ)/δφ = k * οοβ(φ)

As already observed with reference to figure I B, considering a pump comprising a plurality of pistons, the trend of the pistons velocity as a function of the angle of rotation of the pump does not allow to achieve a constant sum of the velocities for any relative shift of the pistons and therefore it causes ripple.

Furthermore, again referring to figure I B, it is observed that the switching between delivery and suction takes place in the moment of maximum acceleration of the pistons and that the point at which their velocity is equal to zero is only a single time instant having a null angular width. Alternatively, advantageously according to the present invention, the movement of the pistons of the axial piston pump 10 inside the cylinders is determined by the frontal cam profile 23, which is shaped so as to impose upon every i ^th piston of the plurality of pistons 14 a law of velocity as a function of the angle of rotation of the pump of the following kind: such a function νι(φ) being defined between -π and +π, where the angle of rotation of the pump φ = -π corresponds to the start of the piston stroke and the angle φ = +π corresponds to the end of the piston stroke. In this way, suitably, νι(φ) = 0 at the start of the piston stroke (φ = -π), there is a maximum at φ = 0, and finally νί(φ) = 0 at the end of the piston stroke (φ = +π); in fact, it is impossible to start a motion with velocity other than zero and, upon termination of the motion, the velocity must return to zero. The angular interval [-π, +π] therefore corresponds to the individual delivery or suction phase. Figure 6A illustrates the function νι(φ) = 1 + οοβ(φ) for a single piston, whereas figure 6B illustrates the superposition of two functions νι(φ) = 1 + οοβ(φ) that are phase shifted by an angle φ = π, the sum S of these two functions being constant for every angle φ, therefore demonstrating the possibility of cancelling the ripple in the case of two pistons engaging simultaneously with the frontal cam profile 23 with an appropriate phase shift of their respective laws of motion, said phase shift corresponding to the relative shift between said pistons with respect to the start of the relative delivery ramp Ri or suction ramp I¾.

The angle of rotation φ of the pump is therefore the argument of the function that describes the velocity. Alternatively, the argument of the function that describes the velocity may be a linear function of the angle of rotation of the pump, in such a way that the angular interval [-π, +π] corresponds to a rotation of the pump corresponding to the delivery or suction phase.

Conveniently, the function νί(φ) is an even function for the individual delivery phase (or suction phase) when the reference is in the middle of said phase. As shown in figure 7, which shows the graph of the function νι(φ) = 1 + οοβ(φ) between -π and +π, if a horizontal line passing through the average value of said function is drawn, the motion of the piston during the delivery phase (or suction phase) is divided into four sections Xi , X2, X3 and X4. The first section Xi is substantially a mirror image of section X3, as well as the section X2 is substantially a mirror image of the first section X ₄ .

Note that also other laws of motion that satisfy the equation Σίνί(φ) = ∑ίδβί(φ)/δφ = k may be used, as well as small deviations from the above law cause small ripple effects.

Furthermore, advantageously in accordance with the present invention, the function νί(φ) = 1 + οοβ(φ) is continuous and is differentiable infinite times and therefore has derivatives with small jerk, snap, crackle and pop values, guaranteeing the continuity and a particularly smooth variation of the forces acting within the axial piston pump 10.

The total active angular extension of the frontal cam profile 23 is now defined as 2ec, therefore taking into account both the suction phase and the delivery phase of an individual piston, and said total active angular extension will now be linked to the round angle 2π. In the case of the frontal cam profile 23 comprising the plurality of lobes 25, by calling L the total number of lobes, the total active angular extension is parameterized to the round angle by setting 2ec = 2 fL. This indicates that the law of motion for a single piston will be repeated L- times on the round angle 2π. Consequentially, ec = π/L represents the active angular extension of the delivery (or suction) portions of the cam profile.

Furthermore, since the angle of rotation φ, or a linear function thereof such that the rotational angular interval [-π, +π] corresponds to the delivery phase (or suction phase), is the argument of the function that describes the velocity, the law of the velocity νι(φ) is rewritten as: so that when φ = ± βο/2, the argument of the cosine is equal to ± π. The parameter C is a proportionality factor necessary to generalize said law of velocity. The integration of the previous equation with respect to the angle φ yields the lift law βί(φ) of the frontal cam profile 23 for each piston of the plurality of pistons 14 of the axial piston pump 10: ¾(φ) = ί νί(φ)άφ = K + C * φ + C * (ε ₀ /2π) * where Κ is the integration constant.

The term "lift law βί(φ)" of the frontal cam profile 23 here and hereafter indicates the law of motion of a single piston of the plurality of pistons 14, in the delivery phase and in the suction phase.

As a consequence, the law of motion of the pistons of the plurality of pistons 14 is substantially described by the integral of the function νί(φ) =

Remembering that 2ec = 2 lh, the law of motion of a single piston may also be rewritten as a function of L as: φ) = J νί(φ)άφ = Κ + 0 * φ + 0 * ( 1 /2Ι * sin(2(|>L)

It is preferable to keep e _c as variable instead of L when βί(φ) describes the lift law for angular extensions of the frontal cam profile 23 that do not correspond to the exact division of the round angle by the total number of cam lobes L.

The constants C and K are derived by imposing the boundary conditions on the lift law βί(φ) of the frontal cam profile 23.

In particular, in a first phase of the motion of the pistons, from a minimum lift to a maximum lift, it is imposed that: - if φ = -ec/2 (start of the piston stroke), β(φ) = 0 (i.e. minimum lift); if φ = +ec/2 (end of the piston stroke), β(φ) = A (i.e. maximum lift), A being the total travel distance (stroke) of each piston of the plurality of pistons 14. Alternatively, in a second phase of the piston motion from a maximum lift to a minim lift, i.e. in the switching between delivery and suction (or between suction and delivery), the sign of the piston velocity is inverted and the law of motion is rewritten as:

¾(φ) = ί νί(φ)άφ = Ki + C^ + Ci * ( l /2L) * sin^L) under the following conditions: if φ = +ec/2, β(φ) = A, i.e. maximum lift; if φ = +3ec/2, β(φ) = 0, i.e. minimum lift. Such conditions lead to:

• during the phase of the motion from minimum lift to maximum lift:

C = A/ e _c ⁼ A*L/ π;

K = A/2;

• in phase of the motion from maximum lift to minimum lift (i.e. after the inversion of the sign of the piston velocity) :

Hereinafter, the phase from minimum lift to maximum lift indicates the delivery phase, whereas the phase from maximum lift to minimum lift indicates the suction phase.

Such law can be further slightly modified to take into account the fluid compressibility phenomenon or also dynamic delays in the fluid behaviour. In particular, with the purpose of eliminating or at least reducing the fluidborne noise induced by the sudden opening of the piston cylinder at the delivery port, said law may be modified by the amount by which the fluid is compressed during the first communication of the cylinders of the plurality of cylinders 13, full of uncompressed fluid, with the delivery, so as to gradually regulate its compression. Given that the fluid is generally compressed on the order of 1-2% every 150 bar, the modification to said law shall be substantially modest and in line with this small compressibility.

The lift law βί(φ) of the frontal cam profile 23 as defined according to the above equations, apart from the small corrections related to the fluid compressibility, is represented in figure 8A as a function of the angle of rotation (expressed in degrees) of the axial piston pump 10 of figure 2. In the figure 7A, provided as a non-limiting example of the present invention, each piston of the plurality of pistons 14 has a total stroke A and the frontal cam profile 23 comprises three equally spaced cam lobes, with a total active angular extension 2e _c equal to 120°. The graph is thus defined over the entire total active angular extension 2e _c = 120°. In particular, the graph is defined between -30° and 90° and therefore the switching between delivery and suction takes place at the angle φ = 30°. In this example, the delivery phase is characterized by an increase of βί(φ) and is represented in the left portion of the graph in figure 8A, whereas the suction phase is characterized by a decrease of βί(φ) and is represented in the right portion of the graph in figure 8A.

Again with reference to figure 8A, advantageously according to the present invention, the sections characterizing the start and the end of the lift are extremely smooth and can be assimilated to sections with null flow rate, which is very advantageous in the design and in the proportioning of the fluid distributor 15.

The lift law βί(φ) defined in accordance with the above equations leads to a law for the velocity νί(φ) of a single piston of the plurality of pistons 14 that is illustrated in figure 8B as a function of the angle of rotation (expressed in degrees) of the axial piston pump 10, said figure representing again, as a nonlimiting example, the law of velocity of a piston in the suction and delivery phases in the case of a frontal cam profile comprising three cam lobes with total active angular extension 2e _c = 120°. In the example considered, the negative portion of the graph corresponds to the suction phase whereas the positive portion corresponds to the delivery phase.

Advantageously in accordance with the present invention, as it is clearly evident in figure 8B, the above discussed lift law of the frontal cam profile 23 provides, in the function that describes the velocity νι(φ) of each piston of the plurality of pistons 14, a horizontal inflection point (i.e. with δβί(φ)/δφ = 0 e δ¾(φ)/δφ ² = 0) between the suction phase and the delivery phase, i.e. the pistons of the plurality of pistons 14 have null velocity and null acceleration during the switching between suction and delivery, i.e. when their motion is inverted.

Furthermore, the function νι(φ) is an even function for the individual delivery and suction phases, whereas it is an odd function when the reference is taken in the middle of the total active angular extension (corresponding to the point of connection between the delivery ramp Ri and the suction ramp R2) . In others words, the law of motion is an even function whenever the motion of the piston is considered only along the delivery ramp Ri or only along the suction ramp R2 of a lobe of the plurality of cam lobes 25, whereas it is an odd function whenever the motion of the piston is considered along an entire lobe of the plurality of cam lobes 25, i.e. along the entire total active angular extension 2e _c . Such features of the function νί(φ) are indispensable to ensure that when at least two pistons of the plurality of pistons 14 have the respective followers 22 engaging simultaneously with the same type of ramp of a relative lobe of the plurality of cam lobes 25 (i.e. in the same active phase of the cam), they have a constant sum of their velocities and therefore the ripple is null.

To understand the trend of the sum of the velocities of two pistons, it is convenient to imagine the advancement of the pistons on the frontal cam profile 23 (or vice versa) by superimposing two curves as the one shown in figure 8B, such curves being appropriately phase shifted, where the phase shift between one curve and the other corresponds to the relative shift between the two pistons with respect to the start of the relative delivery ramp Ri or suction ramp R2 with which the pistons are engaged. Obviously, for appropriate phase shifts between the two curves, the sum of the instantaneous flow rates is null, eliminating the ripple. As previously indicated, in the case of two pistons with respective followers 22 engaging simultaneously with the same type of ramp of a relative lobe of the plurality of cam lobes 25, in order to eliminate the ripple they must have a relative shift, with respect to the start of the respective ramp, equal to ec/2, as it can be easily seen in figure 9. A portion of the law of velocity of fig. 8B (the delivery phase, in particular) is shown in that figure. In particular, the number 29 indicates a portion of the area of the graph that corresponds to the flow rate of a first piston and the number 30 indicates a portion of the area of the graph that corresponds to the flow rate of a second piston, the follower of which is engaging with a same type of ramp (and therefore in a same active phase of the frontal cam profile 23), the second piston being angularly shifted by ec/2 (30° in the example of figure 9) with respect to the first piston. It is evident that the curve in delivery phase tends to flatten in the center, as well as it flattens at the ends. Consequentially, if the two pistons are relatively shifted by ec/2, the sum of their flow rates is constant for every angle φ, since the decrease in flow rate of the second piston is perfectly compensated by the increase in flow rate of the first piston.

The velocity of the first piston being defined as νι(φ) and the velocity of the second piston being defined as ν2(φ + e _c /2), the following equation is therefore obtained thanks to the law of the velocity as represented in figure 8B: for each angle φ, k being a constant, in such a way that as the velocity (i.e. the instantaneous flow rate) of the first piston increases, the velocity of the second piston decreases by the same amount, the first piston starting to slide on a ramp when the second piston begins to surpass the mid-point of said ramp, thus guaranteeing the constancy of the sum of the velocities.

Based on the foregoing submissions, it is therefore clear that whenever the foregoing function νι(φ), or any other trigonometric function similar thereto, is used, the argument of the trigonometric functions describing the velocity of two pistons simultaneously engaging with a same type of ramp of a relative lobe will be phase shifted by π, said phase shift by π therefore corresponding to the relative shift by e _c /2=30° of figure 9.

Suitably, in the axial piston pump 10, there are always two pistons having the respective followers 22 engaging simultaneously with the same type of ramp (delivery or suction) of a relative lobe of the plurality of cam lobes 25 and having a relative shift equal to 30° with respect to the start of the relative ramp on which they are engaged, which corresponds to a relative shift of e _c /2, therefore guaranteeing the constancy of the sum of the velocities.

Furthermore, the inflection point between the suction phase and the delivery phase implies the fact that the switching between the suction phase and the delivery phase takes place advantageously with null velocity and with null acceleration, the velocity also remaining substantially null in a significant neighborhood of the inflection point, causing said switching to occur in a longer time interval than that occurring in the known solutions and thus reducing the pump noise. This is a considerable advantage with respect to the known solutions, where, as shown in figure IB, the velocity of the piston is null only at one instant, the moment of the switching between suction and delivery being the moment of maximum acceleration for the piston, the fast switching therefore thwarting any attempt of dampen the fluid compressibility phenomenon. In the known designs, the switching therefore takes place at a velocity other than zero, the velocity being null only at one point, while the switching is a phenomenon that should occur within a time interval (or angular interval) having appreciable width (few degrees), and therefore greater than zero. As previously shown, in a particularly preferred embodiment of the present invention, the frontal cam profile 23, which is fixed and integral with the fixed housing 1 1 , comprises three cam lobes equidistant from each other on the inner wall of the fixed housing 1 1 and with a total active angular extension e _c equal to 120°, the axial piston pump comprising twelve pistons having a relative shift of 30°.

In a further alternative embodiment not shown in the figures, the pistons may be tilted by an average value of the pressure angle of the frontal cam profile, so as to reduce the stresses between the piston and cylinder body induced by lateral forces acting on the follower of the piston, such a solution being advantageous for high pressures though implying the non- symmetry and non-reversibility of the pump itself.

Finally, according to a further alternative embodiment not shown in the figures, it is provided an axial piston pump wherein the frontal cam profile and the desmodromic control, which is integral with the frontal cam profile, rotate inside the fixed housing, whereas the cylinder block is fixed and integral with the fixed housing. In this embodiment, the cylinder block is fixed and therefore a fluid distributor housing check valves is provided, in particular two check valves for each piston, one for the delivery phase and the other for the suction phase. Note that in this embodiment, in order to withstand very high axial thrusts, an additional hydrostatic (or hydrodynamic) support bearing is provided for the axial support of the rotating frontal cam profile.

Now a method for manufacturing a positive-displacement axial piston pump 10 is described below, wherein each piston of a plurality of pistons 14 is structured with an end portion 21 protruding from a respective cylinder until it engages with a frontal cam profile 23 as a follower 22. As previously mentioned, the frontal cam profile 23 includes a plurality of cam lobes 25, each lobe comprising a delivery ramp Ri and a suction ramp I¾, at least two pistons of the plurality of pistons 14 having the respective followers 22 simultaneously engaging with a same type of ramp of a relative lobe of the plurality of cam lobes 25.

Such a method allows eliminating the ripple in an axial piston pump and simultaneously to reduce the noise occurring during the switching between the delivery and suction phases.

In particular, the method provides for the realization of a frontal cam profile (i.e. the frontal cam profile 23), the shape of which allows to apply a law of motion to the pistons such that:

- the start and the end of the delivery phase, as well as the start and the end of the suction phase, have null velocity and null acceleration;

- with respect to the vertical axis passing through its maximum, the function that describes the velocity in the delivery phase is an even function, as well as, with respect to the vertical axis passing through its minimum, the function that describes the velocity in the suction phase is even;

- the overall function that describes the velocity of a single piston both in suction and in delivery phase exhibits a horizontal inflection point that connects the two phases; this inflection point, where velocity and acceleration are null, ensures that even in a large neighborhood thereof the velocity and the acceleration are so low such that they can be considered null;

- it is substantially described by the integral of the function νι(φ) = C * ( 1 +

- the pistons of the plurality of pistons 14 having the respective followers 22 engaging simultaneously with a same type of ramp of a relative lobe of the plurality of cam lobes 25 have a relative shift such that, with reference to figure 7, whereas one piston starts to travel along a portion of a ramp corresponding to the section Xi, the second piston starts to travel along a portion of a same type of ramp corresponding to the section X3; in this way, whenever the foregoing function νί(φ) or any other trigonometric function similar thereto is used, the argument of the trigonometric functions of two pistons simultaneously engaging with the same type of ramp of a relative lobe is phase shifted by π.

The foregoing method therefore allows manufacturing an axial piston pump in which the sum of the velocities of the pistons involved in the delivery phase (or suction phase) is constant. In conclusion, the axial piston pump according to the present invention comprises a frontal cam profile including a plurality of cam lobes, each lobe comprising a delivery ramp and a suction ramp, at least two pistons of the plurality of pistons having an end portion that protrudes from the respective cylinder and is a follower simultaneously engaging with a same type of ramp of a relative lobe of the plurality of cam lobes. Therefore, the hydrostatic slipper pads, sliding on a swash plate, that characterize the positive-displacement piston pumps in accordance with the prior art are no longer used, and the pump configuration in accordance with the invention is such that the sum of the velocities of all the pistons is always constant for every angle of rotation of the pump.

The frontal cam profile is suitably shaped in such a way that the lift law of the cam profile is substantially described by:

¾(φ) = J νί(φ)άφ = Κ + 0 * φ + 0 * (1 /2L) * sin(2(|>L) during the phase of the piston motion from minimum lift to maximum lift and, ¾(φ) = J νί(φ)άφ = Κι + Οι * φ + Οι * (1 /2Ι * sin(2(|>L) during the phase of the piston motion from maximum lift to minimum lift, the constants C, K, Ci and Ki being defined as above. Advantageously, the foregoing frontal cam profile allows to obtain a law of velocity for each piston of the plurality of pistons of the axial piston pump capable of providing, on one hand, a constant pump flow rate and of allowing, on the other, a proper mechanical operation of the pump itself, that is meeting the following conditions from a mathematical point of view:

- providing a constant instantaneous flow rate (∑iVi(t) = k);

- the function βί(φ) being continuous;

- the function νι(φ) being continuous;

- the function cceleration) being continuous; - the function erk) being continuous;

- the function (snap) being continuous;

- the function (crackle) being continuous; and

- the function pop) being continuous.

Particularly advantageously according to the present invention, in addition to the shape of the frontal cam profile imposing the above law of motion, the pistons are suitably shifted relatively to each other so that the sum of their velocities is substantially constant.

Furthermore, advantageously the axial piston pump features a mechanical design with balanced forces, since there is always a corresponding number of pistons in the suction phase and in the delivery phase, such pistons being spaced from each other by a same relative shift. In this way, the axial resultant of the forces is substantially constant and barycentric for each angle of rotation, eliminating the overturning torques on the rotating cylinder, so that all that remains is a resistant torque that, the instantaneous flow rate being constant, is also constant. In this way, the pump supports are subject to much less stress, vibrations are practically absent or much smaller, and there are no damaging fluctuations of forces and torque; furthermore, the smoothness of the switching due to inflection point eliminates most of the vibrations due thereto. Advantageously, the axial piston pump in accordance with the present invention is symmetric and in this case there is no obligatory direction of rotation of the rotating cylinder block and a "four-quadrants" configuration may be achieved, where the axial piston pump is reversible and may also act as a motor. Obviously, a person skilled in the art, in order to meet particular needs and specifications, can carry out several changes and modifications to the axial piston pump described above, all included in the protection scope of the invention as defined by the following claims.