TECHNICAL FIELD

The present invention relates to a -dry screw compressor for a gas, in particular air, for use in pressure applications (e.g. in the conveyance of granulates or powders, or in water treatment, where large amounts of air must be conveyed to start and help aerobic reactions) and in vacuum applications (e.g. in gas, fume or steam exhaust systems) . In particular, the present dry screw compressor is used in applications with low differential pressures comprised between 1 bar and 3 bars and under vacuum to a threshold absolute pressure of 150 mbars .

BACKGROUND ART

As already known, low differential pressure applications (lower than 1 bar) use lobe compressors.

These are compressors wherein two lobe rotors

(usually two or three lobes) with parallel axes mesh together and synchronically rotate in opposite directions .

However, these lobe compressors, although being structurally simple, economical and able to guarantee a good flow, have a scarce thermodynamic efficiency.

Therefore, it has been projected a screw compressor which could work under low pressure, with a high flow and a thermodynamic efficiency characteristic of an internal compression machine, but whose structural characteristics were as similar as possible to a lobe compressor.

As known, a normal screw compressor under high pressures comprises at least a male rotor and at least a female rotor meshing together during the rotation around respective axes and housed within a casing body. Each of the two rotors has screw-shaped ribs that mesh with corresponding screw-shaped grooves of the other rotor. Both the male and female rotor show, in cross section, a predetermined number of teeth corresponding to their ribs and a predetermined number of valleys corresponding to their grooves. The casing body has an inlet for the gas to be taken in and an outlet (also called "delivery outlet") for the compressed gas. The intake gas is compressed between the two moving rotors and arrives to the outlet under the requested pressure.

Furthermore, it is known that dry screw compressors, generally indicated as ^w oil-free", as opposed to oil injection compressors, are largely used in applications wherein the level of contaminants must be kept below a determined percentage threshold (usually very low) .

In recent years some manufacturers have proposed dry screw compressors for differential pressures comprised between 3 and 10 bars, thus re-adjusting the technology of oil injection screw compressors for applications under high pressures (higher than 10 bars) .

However, the manufacture of such dry screw compressors is quite sophisticated and expensive, since it must take into account the remarkable mechanical and thermal stresses to which the rotors are subjected. In particular, in order to avoid an excessive bending under load, the ratio between the length and the external diameter of the male rotor is usually comprised between 1.5 and 1.8; this requirement strongly limits the compressor capacity and requires the insertion in the compressor structure of a gear multiplier in order to start the rotors at very high peripheral speeds, usually > 150 m/s.

By altering the delivery outlets the aforesaid compressors can be also used under differential pressures comprised between 1 bar and 3 bars. However, the drawback of these low pressure compressors is represented by their having the same structural complexity of high-pressure compressors.

OBJECT OF THE INVENTION

Therefore, the technical task object of the - A - present invention is proposing a dry screw compressor which can work under low pressure, with a high flow and a thermodynamic efficiency typical of this kind of machines .

In particular, the main aim of the present invention is providing for a dry screw compressor under low differential pressures (comprised between 1 and 3 bars) and with a high flow, which is structurally simple, economical and easy to maintain.

Moreover, a further aim of the present invention is proposing a dry screw compressor which is also suitable for under vacuum applications, up to a threshold of 150 mbar of absolute pressure.

The defined technical task and the specified objects are substantially reached by means of a dry screw compressor comprising the technical characteristics claimed in the independent claim 1 and in further claims dependent on said claim 1.

BRIEF DESCRIPTION OF DRAWINGS

Further characteristics and advantages of the present invention will become more apparent from the following approximate, and hence non-restrictive, description of a dry screw compressor, as illustrated in the appended drawings, wherein:

- figure 1 illustrates a longitudinal cross section of a dry screw compressor according to the present invention;

- figure 2 illustrates a three-dimensional view, in longitudinal section, of some details belonging to the dry screw compressor belonging in turn to the dry screw compressor of figure 1;

- figure 3 illustrates a cross section (not in scale) of the rotors used in the compressor according to figures 1, 2; and

- figure 4 illustrates a three-dimensional schematic side view (not in scale) of a male rotor used in the dry screw compressor according to the invention.

PREFERRED EMBODIMENT OF THE INVENTION

With reference to the alleged figures, 1 indicates a dry screw compressor for gas, in particular air, according to the invention.

The compressor 1 can be used both under pressure and under vacuum.

The compressor 1 comprises at least a male rotor 2 and at least a female rotor 3, meshed together (figures 1, 2, 3) .

The embodiment here described and illustrated provides for a single male rotor 2 and a single female rotor 3 housed within a single casing body 4.

In particular, this casing body 4 is obtained by coupling two communicating cylinders (not shown) so that they define a single cavity 5 housing the rotors

2, 3.

An alternative embodiment (not shown) provides for a plurality of conjugated pairs of male rotors 2 and female rotors 3.

As shown in figure 1, the female rotor 3 is keyed on a shaft 17 (having an axis of rotation (01)), whereas the male rotor 2 is keyed on a shaft 18 (having an axis of rotation (02)). In particular, the first axis of rotation (01) is arranged at a certain distance (I) (generally known as "centre distance") from the second axis of rotation (02). The first axis of rotation (01) and the second axis of rotation (02) are parallel to each other.

Each of said rotors 2, 3 has screw-shaped ribs meshing with screw-shaped grooves formed between corresponding screw-shaped ribs of the other rotor 2,

3. In this way, in cross section (figure 3), the male rotor 2 shows lobes 6 (or teeth) and valleys 7 meshing with corresponding valleys 8 and lobes 9 (or teeth) of the female rotor 3.

Moreover, figure 3 shows some main dimensional parameters characterizing the profiles of the rotors 2, 3. In particular, it can be seen an external circumference (Cef) of the female rotor 3 and an external circumference (Cem) of the male rotor 2.

Moreover, as shown in figure 1, the length (Lm) of the male rotor 2 corresponds to the length (Lf) of the female rotor 3.

Conjugated profiles identical to those shown in the present figure 3 have been described and claimed in the international patent application PCT/IB2010/051416 in the name of the same Applicant, whose content must be considered as an integral part of the present detailed description since, in combination with the geometrical elements described hereinafter, it allows to maximise the compressor capacity and to minimise the gas leaks normally- occurring in the coupling areas between the rotors, and between the rotors and their casing.

In fact, in the present scope and with a particular reference to figure 4, the "winding angle"

(φ) is formed by the angle of a generic helix 40

(described by the head of a generic tooth) comprised between a segment OA, connecting the axis (02) of the male rotor 2 to the helix 40 on a first end plane

(πl) of the rotor 2, and a segment O'B', also connecting the axis (02) to the helix 40 on a second end plane (π2) of the rotor 2 opposed to the first end plane (πl) .

As also shown in figure 4, the rotor 2 comprises three helixes 30, 40, 50, parallel -to each other, described by the heads of the relative teeth.

Furthermore, the term "length (Lm)" of the male rotor 2 defines the distance between the two end planes (πl) , (π2), the term "pitch (Pz)" between two helixes 30, 40 defines the distance between point B and point Bl, and the term "angle of the helix" (ψ) defines the angle comprised between the tangent (r) to the helix 40 in any point (P) and the axis (02) of the male rotor 2.

In an inventive way, it has been found that the ratio between the length (Lm) and the external diameter (Dm) of the male rotor 2 (see also figure 4) must be higher than or equal to 2, to maximise the compressor capacity and, therefore, together with the conjugated profiles of the lobes of the rotors, to guarantee high gas flows. Preferably, said (Lm) / (Dm) ratio is comprised between 2 and 3. In this context, external diameter (Dm) means the diameter of the external circumference (Cem) of the male rotor 2 (figure 3) .

Moreover, it has been found that, in order to maximise the compressor capacity, if the other geometric dimensions are equal, the maximum value of the winding angle (φ) must be 300°; in fact, by increasing the value of the winding angle (φ) , and with an equal length (Lm) , an equal diameter (Dm) and an equal profile of the tooth of the male rotor 2, the overlap between the teeth of the two rotors 2, 3 consequently increases, with a following reduction of the total capacity of the compressor 1.

Moreover, the values (Lm) , (Pz) and the angles (φ) , (ψ) are geometrically related to each other.

Therefore, it is possible to projectually determine the optimal values of parameters (Lm) , (Dm) , (Pz) , (ψ) in order to define an optimal value of the "winding angle" (φ) giving the maximum gas flow at reduced peripheral speed of the male rotor 2 and under reduced pressure.

Preferably, the number of lobes 6 of the male rotor 2 is different from trie number of lobes 9 of the female rotor 3. In particular, the number of lobes 6 of the male rotor 2 is lower than the number of lobes 9 of the female rotor 3 by at least one unity. For example, in the embodiment here described and illustrated, the number of lobes 6 of the male rotor 2 corresponds to three , whereas the number of lobes 9 of the female rotor 3 corresponds to five. In another embodiment (not shown) , the number of lobes 6 of the male rotor 2 corresponds to four, whereas the number of lobes 9 of the female rotor 3 corresponds to six. The two rotors 2 , 3 are kept in the reciprocal position by means of the synchronization gear formed by two toothed wheels 20a and 20b of the known kind (figure 1) .

Obviously, in order to allow a correct working of the compressor 1, the transmission ratio between the synchronization gears 20a, 20b must be equal to the ratio existing between the number of teeth of the two rotors 2 , 3.

Advantageously, the driving shaft is the shaft 17 on which the female rotor 3 is keyed because it is the one with more teeth, so that each rotation of this shaft 17 corresponds to the filling of a larger number of gaps and, in short, to a larger volume conveyed by the compressor 1.

As shown in more detail in figure 2, the casing body 4 has an inlet 10 for a gaseous fluid to be taken in, flowing according to an arrow (Fl) , and at least an outlet 11 (or delivery outlet) for the compressed fluid flowing according to an arrow (F2) . Said outlet 11 defines an opening 12 formed in the casing body 4.

The compressor 1 uses bearings of a known kind. In particular, the radial loads are sustained by a first group 19a of radial ball bearings arranged close to the inlet 10 and by a second group 19b of cylindrical ball bearings arranged close to the outlet 11. The axial loads, on the other hand, are sustained by a third group ~19c of oblique contact ball bearings arranged . beside the bearings of the second group 19b.

In the particular embodiment shown in figure 1, the compressor 1 is provided with an electric motor

16 whose rotor is advantageously keyed on the shaft

17 of the female rotor 3 for starting its rotation around the first axis of rotation (01) . Preferably, the motor 16 is of a permanent magnet motor. Preferably, the permanent magnet motor 16 is of the kind cooled by water circulation. As an alternative, it can be used a permanent magnet motor of the air- cooled kind.

As previously stated, the motor 16 is preferably keyed on the shaft 17 of the female rotor 3 , namely it is aligned with said shaft 17.

When no speed variation of rotors 2, 3 is required, the compressor 1 can be coupled to an electric motor (not shown) by means of a "belt and pulley" drive (not shown) .

The operation of the dry screw compressor according to the present invention is described hereinafter.

The gas (e.g. air) is taken in by the compressor 1 and, through the inlet 10, enters into the casing body 4 (figures 1, 2) . During the rotation, the screw-shaped ribs of the female rotor 3 mesh together •with the screw-shaped grooves of the male rotor 2 , and vice versa. In the embodiments with no contact between the rotors 2 , 3 the correct transmission/multiplication ratio between the rotors 2, 3 is actuated by means of the synchronization gears 20a, 20b.

By longitudinally crossing the casing body 4, the gas is compressed between the "coils" of the two rotating rotors 2, 3, thus reaching the outlet 11.

A first embodiment, wherein the opening 12 is arranged on the side surface of the casing body 4, is used for "intermediate" compression ratios R, e.g. comprised between 1 and 4; whereas, in a second embodiment, the opening 12 is arranged in correspondence to an end of trie casing body 4 (on the plane (πl) ; see figure 1); this last solution is chosen for "high" compression ratios (R), e.g. comprised between 4 and 10. Both embodiments can be provided with shaping means (not shown) defining the actual dimension of opening 12 corresponding to the desired compression ratio (R) .

The aforesaid description clearly shows the characteristics of the dry screw compressor according to the present invention, as well as its advantages.

In particular, the ratio between the length and the external diameter of the male rotor (higher than or equal to two) is made possible by low differential pressures (comprised between 1 bar and 3 bars) or by the threshold absolute pressure of 150 mbars for under vacuum applications .

Moreover, the choice of the profile geometry and the operation of the compressor by means of the shaft of the female rotor allow to maximize the compressor capacity, with rotors of the same length, thus allowing to reach the requested high flow at a peripheral speed of the male rotor 2 lower than 80 m/s .

Furthermore, the geometry of the profiles of the two coupled rotors allows to get a shorter contact line between the rotors with a better seal, thus reducing the blow by.

Moreover, thanks to the fact that the compressor works at peripheral speeds of the male rotor lower than 80 m/s, the peripheral speed of the female rotor is even lower, and therefore the rotor of the electric motor can be directly keyed on the shaft of the female rotor (namely with no interposition of multiplying gears), thus obtaining a compressor which is structurally simple, compact and having a higher energetic efficiency. This makes use of the multiplying ratio of synchronization gears of the rotors, corresponding to the ratio between the number of lobes of the female rotor and the number of lobes of the male rotor (in the described embodiment, it corresponds to 5/3=1.66667). This avoids the utilization of toothed-wheel multipliers integrated in the compressor, with a resulting advantage in structural simplicity, encumbrance, cost and noise.

Furthermore, the energetic efficiency of the compressor is also provided by the use of a permanent magnet motor, characterized by low consumptions over a large range of speeds. In particular, this kind of permanent magnet motor has higher efficiencies than the three-phase asynchronous electric motor used in the known art, especially at reduced speeds. Among other things, the use of a water-cooled permanent magnet motor allows a reduction in size and weight of the motor, thus allowing its direct arrangement of the shaft of the female rotor, exploiting the radial bearings of the compressor.

Finally, the optimization of the energetic efficiency is also obtained thanks to the use of a delivery outlet whose size varies according to the desired compression ratio, thus producing an extremely versatile and modular compressor.