RU2547211C2 - Dry screw compressor - Google Patents

Abstract

FIELD: machine building.

SUBSTANCE: invention relates to the dry screw compressor for gas, in particular for air, for use in pressure applications, for example, for movement of pellets or powders or in water treatment applications, where it is necessary to transport large air volumes for starting and conducting of aerobic reactions, and in vacuum applications, for example, in gas, smokes or steam discharge applications. The screw compressor, the rotational speed of which is below 80 m/s, contains a housing (4) with one inlet (10) for suction gas fluid medium and at least one outlet (11) for compressed fluid medium, at least one male rotor (2) and at least one male rotor (3), engaged and located inside the housing (4). Certain ratio between the length (Lm) and external diameter of the male rotor (2) and the swirl angle of the male rotor (3) is offered. The compressor has low working pressure.

EFFECT: invention is aimed at creation of the dry screw compressor, which can operate at low pressure, with an intense flow and with thermodynamic efficiency typical for such type of machines.

6 cl, 4 dwg

Description

FIELD OF THE INVENTION

The present invention relates to a dry compression screw compressor for gas, in particular air, for use in pressure applications (for example, when moving granules or powders, or when treating water where large volumes of air need to be moved to start and facilitate aerobic reactions) and vacuum applications (e.g. in gas, smoke or steam exhaust systems). In particular, a true dry compression screw compressor is used in applications with a low differential pressure of 1 to 3 bar and with a vacuum up to a threshold absolute pressure of 150 mbar.

BACKGROUND

As you know, in the field of application of low differential pressure (less than 1 bar), cam compressors are used. These are compressors in which two-cam rotors (usually two or three cams) with parallel axes are coupled together and synchronously rotate in opposite directions.

Nevertheless, these cam compressors, although they are structurally simple, economical and provide good flow, have a low thermodynamic efficiency.

Therefore, a screw compressor was developed that can operate at low pressure, provide an intensive flow and thermodynamic efficiency of an internal combustion engine, but whose design characteristics would be as close as possible to a cam compressor.

As is known, a conventional high-pressure screw compressor comprises at least a male rotor and at least a female rotor engaged with each other during rotation around respective axes and enclosed in a housing. Each of the two rotors has helical ribs that are meshed with the corresponding helical grooves of the other rotor. The section of both the male and female rotors shows a predetermined number of teeth corresponding to their ribs, as well as a predetermined number of cavities corresponding to their grooves. The housing has a gas inlet and an outlet (also called a “discharge outlet") for compressed gas. The suction gas is compressed between two moving rotors and fed to the outlet under the required pressure.

In addition, it is known that dry compression screw compressors, commonly referred to as “oil-free”, in contrast to oil injected compressors, are widely used where contaminants should be kept below a predetermined threshold percentage (usually very low).

In recent years, some manufacturers have proposed dry compression screw compressors for differential pressures from 3 to 10 bar, thereby adapting the technology of screw compressors with oil injection for high pressure applications (above 10 bar).

However, the production of such dry screw compressors is rather complicated and costly, since it is necessary to take into account the significant mechanical and thermal loads to which the rotors are subjected. In particular, in order to avoid excessive deflection under load, the ratio between the length and the outer diameter of the covered rotor is usually between 1.5 and 1.8, and this requirement greatly limits compressor performance and requires the installation of a gear-multiplier design in the compressor to start rotors with very high peripheral speeds, usually> 150 m / s.

By modifying the discharge outlets, compressors can also be used at a differential pressure of 1 to 3 bar. However, the disadvantage of these low-pressure compressors is that they are structurally as complex as high-pressure compressors.

OBJECT OF THE INVENTION

Thus, it is an object of the present invention to provide a dry compression screw compressor that can operate under low pressure, with an intensive flow and with a thermodynamic efficiency typical of this type of machine.

In particular, the main objective of the present invention is to provide a dry compression screw compressor with a low differential pressure (1 to 3 bar) and an intensive flow, which is structurally simple, economical and easy to maintain.

In addition, an additional object of the present invention is to provide a dry compression screw compressor, also suitable for use in a vacuum, up to a threshold absolute pressure of 150 mbar.

A specific technical goal and these tasks are essentially solved by means of a dry compression screw compressor containing technical features declared in independent clause 1 and in additional clauses dependent on clause 1.

BRIEF DESCRIPTION OF THE DRAWINGS

Additional characteristics and advantages of the present invention will become more apparent from the following approximate and therefore non-limiting description of a dry compression screw compressor, illustrated in the accompanying drawings, in which:

Figure 1 is a longitudinal section of a dry screw compressor according to the present invention,

Figure 2 is a perspective view in longitudinal section of some parts related to a screw compressor, dry compression, which, in turn, relates to a screw compressor, dry compression, shown in figure 1,

Figure 3 is a section (not to scale) of the rotors used in the compressor according to figures 1 and 2, and

Figure 4 is a schematic (not to scale) perspective view of a male rotor used in a dry compression screw compressor according to the invention.

BEST MODE FOR CARRYING OUT THE INVENTION

As can be seen in the drawings, FIG. 1 shows a dry compression screw compressor for gas, in particular air, according to the invention.

Compressor 1 can be used both under pressure and in vacuum.

Compressor 1 comprises at least a male rotor 2 and at least a female rotor 3 engaged (FIGS. 1, 2, 3).

In the embodiment described and illustrated herein, a single male rotor 2 and a single female rotor 3 are provided, enclosed in a single housing 4.

In particular, the housing 4 is obtained by connecting two communicating cylinders (not shown) so that they form a single cavity 5 containing the rotors 2 and 3.

In an alternative embodiment (not shown), a plurality of mating pairs of male rotors 2 and female rotors 3 are provided.

As shown in FIG. 1, the female rotor 3 is secured with a key on the shaft 17 (having a rotation axis (01)), and the male rotor 2 is secured with a key on the shaft 18 (having a rotation axis (02)). In particular, the first axis of rotation (01) is located at a certain distance (I) (commonly known as the "center distance") from the second axis of rotation (02). The first axis of rotation (01) of rotation and the second axis of rotation (02) are parallel to each other.

Each of the rotors 2 and 3 has helical ribs engaged with helical grooves formed between the respective helical ribs of the other rotor 2 and 3. Thus, in section (Fig. 3) cams 6 (or teeth) and troughs 7 of the male rotor 2 are shown, meshed with the corresponding depressions 8 and cams 9 (or teeth) of the enclosing rotor 3.

In addition, figure 3 shows some dimensional parameters characterizing the profiles of the rotors 2 and 3. In particular, the outer circumference (Cef) of the female rotor 3 and the outer circle (Cem) of the male rotor 2 are visible.

In addition, as shown in FIG. 1, the length (L _m ) of the male rotor 2 corresponds to the length (L _f ) of the female rotor 3.

Paired profiles identical to those shown in Fig. 3 were described and proposed in international application PCT / IB2010 / 051416 addressed to the Applicant, the contents of which should be considered as part of this detailed description, since in combination with the geometric elements described hereinafter allows to maximize compressor performance and minimize gas leaks, usually occurring at the joints between the rotors, as well as between the rotors and their casing.

Substantially within the scope of the present invention and with reference specifically to FIG. 4, the “twist angle” (φ) is formed by the angle of a conventional helix 40 (described by a typical tooth head) formed between the OA segment connecting the axis (02) of the male rotor 2 with the helical a line 40 on the first end surface (nl) of the rotor 2, and a segment O'B 'also connecting the axis (02) with a helix line 40 on the second end surface (n2) of the rotor 2 opposite the first end surface (nl).

As also shown in figure 4, the rotor 2 contains three helical lines 30, 40, 50, parallel to each other, described by the heads of the respective teeth.

In addition, the term "length (Lm)" of the male rotor 2 defines the distance between the two end planes (n1), (n2), the term "step (Pz)" between the two helical lines 30 and 40 defines the distance between point B and point B1, and the term “helix angle (φ)” defines the angle between the tangent (r) to the helix 40 at any point (P) and the axis (02) of the male rotor 2.

It was experimentally established that the ratio between the length (Lm) and the outer diameter (Dm) of the covered rotor 2 (see also FIG. 4) should be greater than or equal to 2 in order to maximize the performance of the compressor and, thus, together with the mating profiles rotor cams provide intense gas flows. Preferably, the ratio (Lm) / (Dm) is from 2 to 3. In this context, the outer diameter (Dm) refers to the outer circumference (Cem) of the male rotor 2 (FIG. 3).

In addition, it was found that in order to maximize compressor performance, all other things being equal, the maximum twist angle (φ) should be 300 °, in fact with an increase in the twist angle (φ) and with the same length (Lm) and the same diameter ( Dm) and the same tooth profile of the male rotor 2, the overlap between the teeth of the two rotors 2 and 3 as a result increases, with a subsequent decrease in the overall performance of the compressor 1.

In addition, the quantities (Lm), (Pz) and the angles (tp), (φ) are geometrically connected with each other.

Therefore, by calculating it is possible to establish the optimal values of the parameters (Lm), (Dm), (Pz), (φ) to determine the optimal value of the "angle (φ) of the twist", providing maximum gas flow at a reduced peripheral speed of the covered rotor 2 and at a reduced pressure.

Preferably, the number of cams 6 of the male rotor 2 is different from the number of cams 9 of the female rotor 3. In particular, the number of cams 6 of the female rotor 2 is less than the number of cams 9 of the female rotor 3 by at least one. For example, in the embodiment described and illustrated here, the number of cams 6 of the male rotor 2 corresponds to three, while the number of cams 9 of the female rotor 3 corresponds to five. In another embodiment (not shown), the number of cams 6 of the male rotor 2 corresponds to four, while the number of cams 9 of the female rotor 3 corresponds to six.

The two rotors 2 and 3 are held in relative position by means of a synchronizing gear train formed by two gears 20a and 20b of a known type (FIG. 1).

Obviously, to ensure proper operation of compressor 1, the gear ratio between the synchronizing gears 20a and 20b must be equal to the ratio existing between the number of teeth of the two rotors 2 and 3.

Preferably, the drive shaft is a shaft 17 on which the enclosing rotor 3 is secured with a key, since there are more teeth on it, so that the rotation of this shaft 17 corresponds to filling more gaps and, more specifically, a larger volume moved by the compressor 1.

As shown in more detail in FIG. 2, the housing 4 has an inlet 10 for sucking in the gas fluid, flowing it according to arrow (F1) and at least one outlet 11 (or pressure outlet) for compressing the fluid flowing according to arrow (F2). The outlet 11 forms an opening 12 made in the housing 4.

In compressor 1, bearings of a known type are used. In particular, radial loads are perceived by the first radial ball bearing assembly 19a located adjacent to the inlet 10, and also the second cylindrical ball bearing assembly 19b located adjacent to the outlet 11. On the other hand, axial loads are perceived by the third oblique ball bearing assembly 19c, located adjacent to the bearings of the second node 19b.

In the specific embodiment shown in FIG. 1, the compressor 1 is provided with an electric motor, the rotor of which is preferably mounted with a key on the shaft of the female rotor 3 to start its rotation around the first axis of rotation (01). Preferably, the motor 16 is a permanent magnet motor. Preferably, the permanent magnet motor 16 is a water-cooled motor. Alternatively, an air-cooled permanent magnet motor can be used.

As noted above, the motor 16 is preferably mounted with a key on the shaft 17 of the female rotor 3, that is, it is aligned with the shaft 17.

When speed changes of rotors 2 and 3 are not required, compressor 1 can be connected to an electric motor (not shown) via a “belt” drive (not shown).

The operation of a dry compression screw compressor according to the present invention is described below.

Gas (for example, air) is absorbed by the compressor 1 and through the inlet 10 enters the housing 4 (Fig.1, 2). During rotation, the helical ribs of the female rotor 3 engage with the helical grooves of the male rotor 2 and vice versa. In embodiments where there is no contact between the rotors 2 and 3, the proper gear ratio / multiplication factor between the rotors 2 and 3 is driven by the synchronizing gears 20a and 20b.

During the longitudinal passage through the housing 4, the gas is compressed between the "turns" of the two rotating rotors 2 and 3, and thus reaches the outlet 11.

The first embodiment, in which the hole 12 is made in the side surface of the housing 4, is used for “intermediate” compression ratios R, for example, between 1 and 4, and in the second embodiment, the hole 12 is located near the end of the housing 4 (on the plane (n1 1), this last solution is chosen for “high” compression ratios (R), for example, having a value from 4 to 10. Both embodiments can be equipped with forming means (not shown) that determines the actual size of the hole 12, corresponding to t ebuemoy degree (R) compression.

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

In particular, the ratio between the length and the outer diameter of the male rotor (greater than or equal to two) is ensured by a low differential pressure (from 1 to 3 bar) or a threshold absolute pressure of 150 mbar for use in vacuum.

Accordingly, the choice of the profile geometry and the operation of the compressor by means of the shaft 10 of the female rotor allows to maximize the performance of the compressor with the same length of the rotors, which allows to achieve the desired flow rate at a peripheral speed of the male rotor 2 less than 80 m / s

In addition, the geometry of the profiles of the two interlocked rotors allows you to get a shorter contact line between the rotors with better sealing, which reduces gas leakage.

In addition, since the compressor operates at peripheral speeds of the male rotor of less than 80 m / s, the peripheral speed of the female rotor is even lower, and thus the rotor of the electric motor can be mounted with a key directly on the shaft of the female rotor (i.e. without placing boost gears), which allows you to get a structurally simple, compact and having high energy efficiency compressor. In this case, the multiplication factor of the synchronizing gears of the rotors is applied, which corresponds to the ratio between the number of cams of the female rotor and the number of cams of the female rotor (in the described embodiment, it corresponds to 5/3 = 1.66667). This avoids the use of gear multipliers integrated in the compressor, which ultimately gives an advantage in terms of simplicity of design, compactness, cost and noise level.

In addition, the energy efficiency of the compressor is also ensured by the use of a permanent magnet motor, characterized by low consumption over a wide speed range. In particular, this type of permanent magnet motor is more efficient than the three-phase asynchronous electric motor used in the prior art, especially at reduced speeds. In addition, the use of a water-cooled permanent magnet motor reduces the size and weight of the motor, which allows it to be located directly on the shaft of the enclosing rotor using radial compressor bearings.

And finally, the optimization of energy efficiency is also achieved thanks to the discharge outlet, the dimensions of which vary depending on the required compression ratio, which contributes to the creation of an extremely flexible and modular compressor.

Claims (6)

1. Screw compressor (1) dry compression with a male rotor (2), the peripheral speed of which is less than 80 m / s, containing a housing (4) having an inlet (10) for suction of a gas fluid and at least one outlet (11) for compressed fluid, at least one male rotor (2) and at least one female rotor (3) meshed and located inside the housing (4), characterized in that the ratio between the length (Lm) and the outer diameter of the male the rotor (2) is greater than or equal to two, while the angle (φ) of the twist of the covered p the opening (2) is less than or equal to 300 °, and that it has a low working pressure, characterized by 1) low differential pressure from 1 to 3 bar and / or 2) absolute pressure up to 150 mbar in vacuum applications. 2. The compressor (1) according to claim 1, characterized in that the drive shaft is a shaft (17), on which the enclosing rotor (3) is fixed by means of a key. 3. The compressor (1) according to claim 2, characterized in that it further comprises an electric motor (16), functionally acting on the shaft (17) of the enclosing rotor (3) to start its rotation around the first axis of rotation (01). 4. The compressor (1) according to claim 3, characterized in that the rotor of the electric motor (16) is fixed by means of a dowel on the shaft (17) of the enclosing rotor (3). 5. The compressor (1) according to claim 3 or 4, characterized in that the electric motor (16) is a permanent magnet motor. 6. The compressor (1) according to claim 1, characterized in that the outlet (11) forms a hole (12) made in the housing (4), while the actual size of the hole (12) is configured to change the forming means to obtain a given degree (R) compression.

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