Volumetric gear machine with helical teeth

A volumetric gear machine interacting with a working fluid comprising:      At a portion of the meshing area (2), the first and the second flank (311, 312) being in simultaneous contact with the second wheel (4).

TECHNICAL FIELD

The present invention relates to a volumetric gear machine, typically a pump or an engine.

PRIOR ART

Pumps are well-known comprising a first and a second toothed wheel with helical teeth that mesh each other so as to make the mechanical contact of the gears more gradual. They are interposed between a suction and a delivery conveying a working fluid from the former to the latter.

A drawback of this type of pumps is connected with the fact that special attention must be paid to the hydraulic seal between the helical teeth. In fact, to prevent the delivery and suction becoming directly connected for a certain angular operating range, the helical extension of the tooth must be carefully studied and constraints must be complied with which drastically reduce the designer's freedom. Known solutions prevent hydraulic seal problems by using teeth that extend according to not very accentuated helices. It would be useful to be able to have high helices so as to have more gradual contact, lower contact pressure between the teeth and a more gradual variation of the transferred fluid volumes.

Pumps with gears that have straight teeth (therefore not helical) with double contact (the teeth that mesh come into contact in two distinct zones on opposite sides) are also known. In pumps with straight teeth, double contact cannot be used to improve the hydraulic seal for the reasons set out below. In pumps with straight teeth and single contact, in order to guarantee the hydraulic seal, the condition εTR≥1 must be satisfied (εTR indicating the transverse contact ratio defined as the ratio between the rotation of the wheel so that a tooth thereof can travel along the entire action line and the angular step; action line means the segment in which the toothed wheels come into contact during operation).

Double contact envisages two lines of action and would theoretically allow the hydraulic seal if the relationship εTR≥0.5 is satisfied (and not εTR≥1 as in the case of single contact teeth) hence leaving greater freedom in the shape of the tooth with respect to a pump with straight teeth and single contact. But, in fact, such freedom cannot be used as another essential condition for these types of pumps would be compromised and that is the continuous transmission of the motion of the driving wheel to the contact wheel; in the case of pumps with straight tooth gears such a condition translates into respect for the following mathematical condition: εTR≥1 Respect for such relationship therefore thwarts the advantages that double contact could offer for the hydraulic seal.

Pumps, as disclosed in US2011/223051 and WO96/01950, are known.

OBJECT OF THE INVENTION

The object of the present invention is to propose a gear machine that overcomes the above-illustrated drawbacks connected with the mechanical and hydraulic optimisation of the gears, in particular of the tooth helix.

The stated technical task and specified objects are substantially achieved by a gear machine comprising the technical features disclosed in one or more of the appended claims.

In the accompanying figures reference number1denotes a volumetric gear machine. Such machine1is a pump or an engine. The machine1is intended to convey a working fluid (typically incompressible, preferably oil).

The machine1comprises a working fluid inlet and a working fluid outlet. In the case of a pump the inlet is usually called suction whereas the outlet is called delivery. In the case of an engine the inlet is called induction and the outlet is called exhaust.

The machine1comprises a first toothed wheel3with helical teeth. Appropriately all the teeth of the first wheel3are the same as each other.

The helical teeth of the first wheel3comprise a first tooth31which in turn comprises a first and a second flank311,312opposite each other. The first and second flank311,312contribute to defining two compartments intended to convey the working fluid. Appropriately at least one section of the first and second flank311,312are involutes of a circle.

The portion of the first flank311that extends as an involute of a circle affects advantageously more than ⅓, preferably at least ½, of the height of the first tooth31. The height of the tooth means the difference between the tip radius and the root radius.

The description with reference to the first tooth31may also be repeated for the other teeth of the first wheel3.

The machine1comprises a second toothed wheel4with helical teeth. The helical teeth of the second toothed wheel4appropriately comprise an involute profile. Also in that case the teeth of the second wheel4have two opposite flanks at least one portion of which has an involute shape (the involute portion advantageously affects at least ⅓, preferably at least ½ of the height of the tooth). Appropriately, the teeth of the first and second wheel3,4are the same as each other. As exemplified in the figures, the machine1advantageously has external gears (the first wheel3and the second wheel4are therefore flanked externally to each other). In an alternative solution one of the two gears could be at least partially internal to the other.

The use of an involute profile allows friction, vibrations, noise and wear to be minimised. In line with common practice in the technical sector, involute profile also means profiles that have a correction of a few tenths of a millimetre with respect to the theoretical involute line (in the case in question the displacement is less than 5% of the normal module of the tooth). It is underlined that in the technical sector the normal module of a tooth is defined as: d/Z·cos β wherein:

Z: number of teeth;

β: angle of the helix at the primitive diameter.

The first tooth31periodically comes into contact with the second wheel4only at the first and second flank311,312.

The helical teeth of the first wheel3and of the second wheel4are truncated at the tip. The tip of the teeth is therefore substantially flat.

As exemplified inFIG. 2, the first and/or the second wheel3,4are cylindrical toothed wheels. The first and second wheel3,4have parallel rotation axes. Preferably, the first and second wheel3,4are counter-rotating.

The machine1comprises a casing7that houses the first and the second wheel3,4. Appropriately the inlet5and the outlet6are afforded in said casing7.

The first and the second wheel3,4are interposed between the inlet5and the outlet6.

The first and the second wheels3,4are operatively coupled at a meshing area2. The meshing area2is interposed between the outlet6and the inlet5of the working fluid. In particular the meshing area2is located along an imaginary band that connects the inlet5and the outlet6of the working fluid.

At a portion of the meshing area2, the first and the second flank311,312are in simultaneous contact with the second wheel4. This allows an inherent hydraulic property of the double contact to be exploited which is not possible on straight teeth. In fact, an important intuition of the Applicant derives from the following theoretical analysis. For double contact helical teeth the hydraulic seal can be guaranteed by the condition εTR−εEL≥0.5; the case of symmetrical teeth has been considered for simplicity purposes but similar considerations can be repeated in the case of non-symmetrical teeth. In fact, in that case, both lines of contact (lines of action) cooperate for the seal. εTR means the transverse contact ratio i.e. the minimum value between εTRsxand εTRdx(that coincide in the case of symmetrical teeth i.e. wherein the first and the second flank311,312are identical along each contact section orthogonally to the rotation axis of the first wheel3).

εTRsxmeans the ratio between:the rotation of the first wheel3necessary so that the point of contact between the first tooth31and the second wheel4travels the entire line of action C of the first flank311andthe angular pitch.

εTRdxmeans the ratio between:the rotation of the first wheel3necessary so that the point of contact between the first tooth31and the second wheel travels the entire line of action (D) of the second flank312andthe angular pitch.

The line of action of the first flank311is the line drawn by the points of contact of the first flank311with the second wheel4; the line of action of the second flank312is the line drawn by the points of contact of the second flank312with the second wheel4. Appropriately, the first and/or the second line of action are rectilinear segments.

εEL indicates the helical contact ratio defined as the ratio between the phase shift of the helix and the angular pitch. The phase shift of the helix corresponds to the angular displacement between the first and the last section of the toothed wheel (evaluated orthogonally to the rotation axis) and is in turn defined as:
S=360·L/(2π·rb/tan(βb)
where:
L: longitudinal length of the tooth;
rb: base radius (at the base of the involute);
βb: helix angle at the base diameter (at the base of the involute).
Angular pitch means the ratio between 360° and the number of teeth.

In the case of single contact helical teeth, to guarantee the hydraulic seal the relationship would be much more disadvantageous: εTR−εEL≥1.

Therefore, a εEL value equal to about 0 should be adopted in order to have a εTR value equal to 1. There would therefore be a good hydraulic seal, but the helix would not be thrust much and the performance would be low.

With double contact for obtaining similar results in terms of hydraulic seal, a εTR value equal to 1 can be adopted and εEL values equal to about 0.5 can be used, which allow a high helix angle and tooth sizing without too many restrictions in order to maintain the hydraulic seal. In the case of double contact helical teeth, to have a high helix it is therefore advisable to comply with the following condition: εTR−εEL≤1.

In fact, with a higher helix angle it is possible to obtain more gradual contact, lower contact pressure between the teeth and a more gradual variation of the transferred fluid volumes.FIGS. 3a, 3band 3cindicate with references30and40the points of contact between the first tooth31and the second wheel4. The threeFIGS. 3a, 3b, 3crefer to a same angular position of the first and the second toothed wheel3,4but refer to different cross sections of the first helical tooth31.FIG. 3arelates to a cross section placed half way along the longitudinal length of the first tooth31,FIG. 3bat 25% or 75% of the longitudinal length of the first tooth31(according to whether the helix of the tooth31is right- or left-handed),FIG. 3cis taken at one of the two longitudinal ends of the first tooth31(according to whether the helix is right- or left-handed). Longitudinal extension of the first tooth31means the extension line of the tooth that connects the two opposite shims of the pump1. In fact, the first and the second wheel3,4are axially interposed between the two shims.

InFIG. 4, the references30and40indicate again the points of contact of the first tooth31with the second wheel4. Furthermore, a first and a second line of action are shown in broken lines and indicated by references300and400. They highlight the movement of the points of contact between the first tooth31and the second wheel4during the rotation of the wheels.

As mentioned previously, preferably but not necessarily, the first and the second flank311,312are symmetrical.

The teeth of the first toothed wheel3mesh in double contact with the teeth of the second wheel4.

In the preferred solution the first and/or the second toothed wheel3,4have/has a number of teeth comprised between 8 and 14, preferably between 9 and 12 teeth. Advantageously, the helix angle at the primitive diameter of the teeth of the first and/or of the second toothed wheel3,4is comprised between 8° and 20°, preferably between 12° and 16°. It indicates the angle between the extension direction of the helix and the direction identified by the rotation axis of the first and of the second wheel3,4. Appropriately, the angular phase shift of the helix (previously identified by the letter S) between the cross sections of opposite ends of the teeth of the first and/or of the second wheel3,4is comprised between 10° and 45°, preferably between 20° and 35°.

The involute portion of the first flank311extends between a first and a second edge313,314. The first edge313is radially closer to a rotation axis315of the first toothed wheel3with respect to the second edge314; the helical teeth of the first wheel3comprise a second tooth32consecutive to the first one and facing the first flank311; a first compartment33being afforded as the space interposed between the first tooth31and the second tooth32.

In a theoretically optimal solution the meshing of the first and of the second wheel3,4has a constant hydraulic seal between the inlet5and the outlet6. This means that there is always (i.e. for every angular position of the teeth) at least one pair of teeth of the first and of the second wheel3,4that are in contact along their entire length. This prevents a direct connection between the inlet5and the outlet6, minimising working fluid leakage and therefore optimising the volumetric performance.

However, this condition limits the designer's choice of the size of the first and the second toothed wheel3,4(in particular in the generation of the cross section of the tooth and in the angular definition β of the helix). In actual fact, through experimental tests the Applicant has verified that excellent results can still be obtained in the absence of a perfect constant hydraulic seal.

In that case, a profile (typically involute) of a tooth of the first wheel3and the profile (typically involute) of a tooth of the second wheel4, for at least one portion of the longitudinal length of the tooth, are no longer in contact and allow a hydraulic connection between the inlet5and the outlet6.

However, it is important to contain the extension of such hydraulic connection in order to prevent excessive leakages.

When the relationship 0.5≤εTR−εEL 1 is satisfied there is a constant hydraulic seal and therefore the optimal solution is obtained. However, the user could be pushed to size the teeth without satisfying the relationship 0.5≤εTR−εEL, but keeping leakages contained.

In order for the leakages not to be excessive the following condition must be respected in any case: in a configuration in which the volume of the first compartment33occupied by the second wheel4is maximum, no point of the first edge313is located at a radial distance from a rotation axis316of the second wheel4which is greater with respect to a tip radius of the second wheel4.

Should a hydraulic connection be accepted between delivery and suction the involute profile of a tooth of the first wheel3and the involute profile of a tooth of the second wheel4advantageously satisfy the following characteristics (in the configuration in which the volume of the first compartment33occupied by the second wheel4is maximum):they are opposite each other;they have a minimum distance which is less than 1 tenth of a millimetre.

Furthermore, with sizing of εTR−εEL≤0.5 a similar effect to the one exercised by noise control exhausts placed on the shims is obtained. Noise control exhausts normally place in communication a volume of fluid that is located in a compartment in the meshing area with the high pressure environment and/or the low pressure environment. In this way, it is possible to compensate for violent pressure variations that could be generated in an isolated compartment in the meshing area (and that could determine significant strain, cavitation, noise, localised erosion). If εTR−εEL≤0.5 there will not be a perfect seal and this will facilitate the work of noise control exhausts. In this way, noise control exhausts can be realised on the shims with less narrow dimensional tolerances.

Appropriately the relationship eTOT=εTR+εEL≥1 must be satisfied (to guarantee the continuous transmission of motion).

Hypothesizing operation as a pump, the working fluid at the inlet that is sucked by the first and by the second wheel3,4is positioned in the spaces between two consecutive teeth and is substantially conveyed along two alternative paths until the outlet (which is at a higher pressure than the suction-inlet). The fluid in the passage from the inlet5to the outlet6therefore follows the rotation sense of the first and of the second wheel3,4.

Exemplified, but non-limiting, solutions of a pump according to the present invention developed by the Applicant are summarised by the parameters indicated in the following table (the definition of such parameters has already been indicated previously or is well known to a person skilled in the art who is familiar with the main nomenclature of toothed wheels):

The present invention achieves important advantages.

The introduction of a helix on involute profiles on one hand improves the transmission of motion and on the other worsens the hydraulic seal along the toothed band. The analysis performed by the Applicant highlighted that the combination of the helical geometry with double contact operation leads to interesting potential. In fact, the Applicant theoretically demonstrated (and the experimental data confirm this) that combining helical teeth with double contact operation allows an intrinsic hydraulic property of double contact to be exploited that is not possible on straight teeth.

The invention as it is conceived is susceptible to numerous modifications and variations, all falling within the scope of the inventive concept characterising it. Furthermore, all the details can be replaced with other technically-equivalent elements. In practice, all the materials used, as well as the dimensions, can be any according to requirements.