MXPA06011153A - Gapless screw rotor device - Google Patents

Abstract

A screw rotor (10) has a housing (12) with an inlet port (18) and an outlet port (20), a male rotor (14) with helical threads (34, 36), and a female rotor (16) with helical grooves (38, 40). The helical threads (34, 36) and helical grooves (38, 40) are designed to eliminate the blow hole leak pathway for multiple-pitch screw rotor products as well as single-pitch screw rotor products. The male rotor has a pair of helical threads with a phase-offset aspect, and the female rotor has a corresponding pair of helical grooves. The female rotor counter-rotates with respect to the male rotor and each of the helical grooves respectively intermeshes in phase with each of the helical threads. The phase-offset aspect of the helical threads is formed by a pair of teeth (42, 44) bounding a toothless sector (46).

Description

ROTOR PE DEVICE SCREWS WITHOUT HOLES PRIORITY APPLICATION This application claims priority of the Application of E.U.A. Do not. / 810,513, filed on March 27, 2004, under the U.S.C. 35 §§120.365 (c).

DECLARATION WITH RESPECT TO FEDERALLY SPONSORED RESEARCH OR DEVELOPMENT Does not apply FIELD OF THE INVENTION The present invention relates generally to rotary devices and, more particularly to screw rotors.

BACKGROUND OF THE INVENTION Screw rotors are generally known to be used in compressors, extenders and pumps. For each of these applications, a pair of screw rotors have threads and helical grooves that intertwine in a housing. For an extender, a pressurized gaseous work fluid enters the rotors, expands in volume as the work is obtained from at least one of the rotors, and is discharged at a lower pressure. For a compressor, work is placed on at least one of the rotors to compress the gaseous working fluid. Similarly, for a pump, work is placed on at least one of the rotors to pump the liquid. The working fluid, either gas or liquid, which enters through an inlet in the housing, is positively positioned within the housing, as the rotors rotate in the opposite direction, and exits through an outlet in the housing. accommodation. The rotor profiles define sealing surfaces between the rotors themselves between the rotors and the housing, thus sealing a volume for the working fluid in the housing. The profiles are traditionally designed to reduce leakage between the sealing surfaces, and special attention is paid to the interface between the rotors, where the threads and grooves of a rotor respectively intertwine with the grooves and threads of the other rotor. The interface between the rotors must be designed in such a way that the threads do not lock in the grooves, and this has typically resulted in gear-like profile designs, which have radially widening grooves and involute rocks seperately separated around of the circumference of the rotors. However, an involution for a gear tooth is designed primarily for strength and to avoid blocking as the gear interlocks with each other and are not necessarily optimal for the circumferential sealing of rotors within a housing. The performance characteristics of screw rotors depend on several factors, including thermodynamic efficiencies, volumetric efficiencies and mechanical efficiencies. Adiabatic efficiency is a type of parameter to evaluate the thermodynamic efficiency of a screw rotor system. Adiabatic efficiency is the proportion of adiabatic horsepower required to compress a certain amount of gas for the horsepower used in the compressor cylinder. The volumetric efficiency is the ratio of the current volume of the working fluid flowing through the screw rotor, such as in one complete revolution, to the geometric volume of the measured screw rotor, which is also measured for a complete revolution. Mechanical efficiencies can include the efficiencies of any gear course that can be used to keep the rotors in the proper phase with each other, bearings and seals. Although diabatic efficiency and volumetric efficiency are different performance parameters, a number of screw rotor characteristics can affect both of these efficiencies. For example, the tolerances for narrowing between the rotors and the distance can improve both the volumetric efficiency and the adiabatic efficiency of a particular rotor design. However, if the tolerances are too narrow for a given design, the volumetric efficiency can be improved while the adiabatic efficiency falls. This performance characteristic could be produced by the thermal expansion of the rotors, machining tolerances and even material properties of the rotors, which can result in intermittent contact with the rotors and sides of the housing or between the rotors themselves. In general, one of the best ways to improve thermodynamic efficiencies is to maintain close tolerances and minimize leakage paths between the rotors and the housing and between the rotors themselves. However, in the prior art screw rotors, the leak paths are inherent to the current design of the rotors, that is, the leaks can be reduced, but not eliminated. Such inherent leaks could occur even when the tolerances are perfected, that is, zero thermal expansion, perfect machining tolerances and a perfectly smooth finishing material. These leakage paths result in losses that adversely affect both the thermodynamic efficiency and the volumetric efficiency of the screw rotors. Accordingly, leakage paths are some of the most important losses to consider for the performance of screw rotors when screw rotors are being designed because these losses adversely affect both thermodynamic efficiency and volumetric efficiency. Even with this knowledge that the trajectories could be reduced to a minimum, the design methodology used for the screw rotors causes these trajectories as an inherent aspect of the traditional screw rotor profiles. In fact, it is a common belief of the designers, manufacturers and users of screw rotors that it is impossible to eliminate some of the leaks in a screw rotor system. For example, according to Mattai Compressors, Inc., on its website www.matteicompo.com/About/ScrewCompressors/, this belief is established concisely even as an international phase of this application that is presented in March 2005: "The technical problem is typical of the geometry of the screw compressors, the screw compressor suffers a problem known as a ventilation hole." - All screw manufacturers have tried to reduce the effect of the "screw hole". ventilation "analyzing and adapting new rotor profiles to create smaller openings at the critical point, although their complete elimination is impossible". Accordingly, in order to minimize leakage paths, it is common knowledge that the rotors could be sealed perfectly along the contact line, although a number of references of the prior art also teach that the contact line should be so cut as possible, that is, it should not extend to cusps on opposite sides of the housing. Examples of short contact lines are described and illustrated in the Patents of E.U.A. Nos. 2,486,770 and 3,245,612. However, there remains a need for better methodologies for the design of screw rotor profiles that are taken into account for the constraints of machining, thermal expansion and tolerances of materials, as well as mechanical efficiencies and that can also eliminate any path of inherent flow of the design procedure, even when they are currently considered impossible. An example of a machining limitation established in the prior art is the need for rough edges due to the concern that sharp edges tend to break, for example, U.S. Pat. No. 2,486,770. Once the leakage path problem is eliminated from the design methodology, that is, the rotor profiles of screws that do not inherently produce a leak path, the designer can balance all the performance characteristics of the rotors. . For example, a rotor design with no inherent leakage trajectory can be changed slightly to include a small orifice or leakage path to allow another aspect to improve the overall performance of the rotors at a given design point, ie, closer tolerances in Static state operation with thermal expansion. In comparison, when the leakage path remains an inherent characteristic of the rotor profiles, the designer must first minimize the flow path using more complex designs that are harder and more expensive to manufacture and then the changes to the design are limited by the complexity of design, machining and other manufacturing capabilities and thermal expansion requirements. Accordingly, a new design methodology is required that produces screw rotor profile shapes without any leakage paths. Additionally, it could also be advantageous if the sharp edge shapes that eliminate the leak paths and do not have a tendency to break could be designed and manufactured. Leakage trajectories are usually caused by internal leaks between the rotors and the housing and between the rotors themselves and result in volumetric losses and thermodynamic losses due to the recirculation of the working fluid within the rotors. For example, working fluid that is pressurized and leaks in a region of lower pressure of the rotors is produced to expand to the lower pressure state with a higher temperature due to entropy and then must circulate again through the rotors before being ejected. Consequently, the overall temperature of the complete rotor system, including the rotors and the working fluid, increases due to the gain in entropy. The internal leakage is detected in specific way in the following points: (1) the holes between the entrance port and / or exit port in the housing and the rotors, resulting in a less than complete capture or ejection for the fluid working through the rotors; (2) the holes between the outer periphery of each rotor and the inner surface of the housing through which the working fluid leaks around the upper base of a thread or the channel of a groove for the adjacent work volume, respectively; (3) the holes between the front and rear of the male interlacing rotor thread and the female rotor slot, through which the working fluid leaks from the pressurized side to the suction side; and (4) a hole formed in the front side of the rotor in the transition region, where the threads of the male rotor are intertwined with the grooves of the female rotor close to the cusp of the cylindrical holes and which generally form a tetrahedron ( or a triangular shape in two dimensions) that is defined by the shape of the hole between the interlacing thread and the groove and the cusp, and another hole formed similarly on the back side of the rotor, through which the working fluid leakage from a V-shaped work volume to an adjacent V-shaped work volume, that is, commonly referred to as a vent hole, and through which working fluid leaks from a pressurized region to a less pressurized region or a suction region. As stated above, the threads must provide seals between the rotors and the walls of the housing and between the rotors themselves, and in all designs prior to the present invention, there has been a transition of the seal around the circumference of the housing for sealing between the rotors. In this transition, a hole is formed between the interlock threads and the housing, causing leakage of working fluid through the hole in the sealing surfaces and resulting in less efficiency in the rotor system. A number of arched profile designs improve the sealing between the rotors and can reduce the hole in this transition region, although these profiles still retain the characteristic gear profile with narrowly spaced teeth around the circumference, resulting in a number of holes in the transition region that are respectively produced by each of the threads. Some pumps minimize the number of threads and grooves and can only have a single higher intensity thread for each of the rotors, although these threads have a wide profile around the circumferences of the rotors and generally result in larger holes in the rotor. transition region. So far, extenders, compressors and screw-rotor pumps have had similar fundamental cracks. In general, these allow the leakage paths between the work side, that is, expansion, compression or pumping to the side that should be sealed from the working side for the proper operation of the rotors, that is, not job. These rotor designs are commonly referred to as Roots type rotors and Lysholm type rotors. Krigar-type rotors, which are described in German Patent Nos. DE4121 and DE7116 for more than a century, have become less popular, and this may be possible due to the increase of Lysholm-type rotors in the 1930s and 1940's. . In an article entitled "A New Rotary Compressor" and written by Lysholm in the 1940's, Lisholm criticizes Krigar's design as unable to obtain any compression between the lobes with a two-thread / two-slot design (2x2 configuration). Although it is clear from the images of Krigar's design that they were definitively sealed issues, especially between threads and grooves, and Krigar seems to be more aimed at radial flow, Lysholm's conclusion is that Krigar's design could not performing any compression only with the 2x2 configuration is defective. Independently, the industry and the teachings have generally followed Lysholm and Roots with very little interest given to Krigar, except as a historical reference. Based primarily on the Lysholm concept, many screw rotor designs have tried to seal the male rotor with the female rotor and housing, although the prior art designs have either a leakage path between the rotors themselves, or a leakage path between the rotors and the housing, ie, which are cited above according to the prior art, the elimination of which is "impossible". In the past, the design of screw rotors have been based on profile designs that do not necessarily follow the mathematical formula, that is, the empirical design methodology, while other designs are based on particular curves or a combination of curves of parts smart, that is, the design methodology of the formula, such as lines, arcs, circles, squares, trapezoids, involutions, reverse involutions, parabolas, hyperbolas, cycloids, trochoids, epicloids, epitrocoids, hypocycloids, hypotrocoids, as well as other straight and arched lines, and still other designs combine formula design and empirical methodologies. However, regardless of the design methodology, empirical or formula or a combination thereof, previous designs and respective methods to create rotor profiles, either explicitly to teach or implicitly suggest and describe the creation of the profile for the corresponding thread and groove using the shortest path of fiow between the rotors, ie, the sealing region does not extend from the frontal cusp to the entire trajectory of the posterior cusp. Additionally, many of the methods of the prior art are based on and remain similar to traditional gear design methods. Some of the above designs have come close to a complete seal or may even have the ability to make a complete seal at an angle, see in particular, the U.S. Patent. No. 6,719,548. Even for these cornering rotors, some of the seals can only be found along the sealing lines, instead of the sealing areas. Additionally, because the rotor profiles are designed in accordance with traditional gear profile design methods, these rotors are usually limited in the types of arched lines that can be used to effect the seal. Without taking into account the three dimensions, the arched lines have typically been limited to epitrocoids, epicycloids, hypocycloids and other types of spirals, such as an Archimedean spiral. When the third dimension is taken into account for prior art design methodologies, it is usually limited to the standard helix angle definitions that have been developed for ordinary screws, i.e. fastening screws. This method fails to really take into account and does not take advantage of the third dimension. It is well known that for any screw rotor, the helix angle of the slots and threads vary depending on their depth. In particular, the upper base of the thread has a helix angle smaller than the root of the thread, and the depression of the groove has a helix angle greater than the crest of the groove. Therefore, using only a single helix angle for a rotor, such as the upper base, the root and any other single angle, even with a correction factor, have not taken into account the variations in the screw helices angles and the slot. In this sense, the known rotor rotor geometries are created using flat design methodologies for the rotor profiles instead of using a volumetric design methodology. Flat design methodologies fail to apply the helix angle function with respect to the radius, resulting in profiles with leakage trajectories discussed above. In one aspect, the flat design methods are unnecessarily restrictive because they only take advantage of the two-dimensional space to overcome the limitation that the threads should not be blocked in the slots. In another aspect, the flat design methods are not restrictive enough because when the profiles are expanded in three-dimensional space, the profiles have three-dimensional leakage trajectories. The extra degree of freedom provided by the three dimensions allows a volumetric design that avoids blocking while allowing the perfect seal between the male rotor and the female rotor and between the rotors and the housing, a perfect seal which is equivalent to the complete seal of the pistons. More generally, similar fundamental flaws in the prior art designs and their respective methodologies can be traced back to their failure to accommodate and utilize the additional degree of design freedom provided by the third dimension. This is the additional degree of design freedom of volumetric design methodologies that allow for an unlimited number of profile designs, which perform a complete seal without blocking the rotors and without the unnecessary restrictions of flat design methodologies. For many rotors of the prior art, the flow path can be found between the face of the threading and the housing. In particular, the thread and groove are designed with significant curvatures at their upper base edges and channels according to the standard design form for intermeshing gear teeth. Said rounded edges and channels may not be able to seal between the rotors and the housing, when the thread and groove begin interlacing with each other. As the thread and groove rotate away from their seals with the housing and in their interlocked positions with each other, the rounded edges produce a hole between the housing and the groove and / or the thread before the groove and thread actually a sealing line is interlocked and reformed. The hole between the housing for the groove and the thread seal can be of an order of magnitude greater than the tolerances for the seals between the rotors and the housing and the rotors themselves. In some designs, the hole may even be longer, such as in screw rotors having a different number of threads and slots, that is, they do not have the same number of threads as slots, and the loss in pressure in the low pressure side causes the thermodynamic efficiency to fall. Accordingly, the rotors must work harder to pump the same volume of air as compared to the rotors according to the present invention, which maintains the same order of magnitude of the seal tolerances when each respective thread and groove begins to interlock each other compared to the seal between the rotors and the housing and the rotors when they are in their fully interlaced positions. Additionally, by failing to take advantage of the third dimension in the design of the thread and groove, the prior art design methods have failed to optimize the basic screw rotor design or improve the efficiency of the screw rotor for its full potential. As discussed above, the prior art design methodologies generally use planar coordinates to define the thread and groove profiles, and the third dimension is only considered for the helix angle of the profiles. In an attempt to compensate for this inadvertent failure to take advantage of the third dimension, the prior art designs have become increasingly complex over the years without offering many improvements in the thermodynamic efficiency of the rotor system. As evidence of the failure to appreciate volumetric design methodologies as an alternative to traditional gear design methods combined with traditional fastener screw methods, these flat design methodologies increasingly lead to these more complex screw rotor designs., like machining and other manufacturing methods improved over the years and allowed for increased complexity. Additionally, these increasingly complex screw rotor profile designs, which need improved fabrication methods, support the conclusion that failure to take advantage of the third dimension has been an unintended failure because the volumetric design methodologies currently allow for much more simplified designs, which may be less complex to manufacture than profiles created using flat design methodologies.

BRIEF DESCRIPTION OF THE INVENTION In general, the present invention provides a design methodology for generating thread and groove profiles, which take advantage of the three-dimensional geometry of the interlacing of rotors. In particular, the present invention has generally solved the problem of leak paths that have plagued the screw rotor designs for more than one hundred years. The present invention provides a design methodology which is based on the fundamental premise that the propeller angles of the screw rotors vary with respect to each other as their threads are joined and then separated with the slots and that they eliminate the screw hole. ventilation, the sealing region must extend completely from the front cusp of the housing to its posterior cusp. Accordingly, each screw rotor embodiment of the present invention can eliminate at least one vent hole cavity, while also maintaining a seal between the threads and slots regardless of the number of angles. It is an advantage of the present invention to maximize the thermodynamic efficiency and volumetric efficiency in a screw rotor system by various means, such as reduction holes, minimization of recirculation within the screw rotor housing, reduction of shock waves within the rotors of screw, reduction of entropy, and reduction of sliding friction between the male rotor and the female rotor. It is also an advantage of the present invention that it can be easily produced. The designs can be simple and still maintain a good sealing ratio. Accordingly, the present invention does not suffer from an excessively complicated design that is difficult to machine or otherwise fabricate. It is another advantage of the present invention that it can reduce and almost eliminate the arc return. It is still another advantage of the present invention that it can reduce the manufacturing cost of screw rotor compressors, and because of its decreased thermodynamic and volumetric efficiencies, it can also reduce the cost of ownership of screw rotor compressors. It is a further advantage that the present invention provides assembly in economical, efficient and fast manufacturing, and also reduces the cost of assembling components and packing costs of the product. It is still a further advantage of the present invention that the screw rotor system can be designed as a modular device that can be replaced with a cartridge type system or integrated in its entirety into a particular product. Up to the extent to which various components of the screw rotor system are manufactured separately, and subsequently transported to an assembler for the attachment of additional components and / or for the additional assembly in the final products, the modular aspects of the present invention improve efficiency and assembly economy. Compared to blade and turbine compressors, the present invention is much stronger, more economical and provides much more compact components. Accordingly, no prior design follows the design methodology of the present invention, which, as discussed above, can perform a complete seal regardless of the types of lines, straight or arched. The present invention can also effect a complete seal for multiple angled rotors. The new design method is even so robust that it produces geometries that can even make a complete seal in multiple areas simultaneously, including the areas between the male rotor and the female rotor as well as between the rotors and the housing.

Now that this design problem has been identified, it will be appreciated that by visualizing the threads and grooves in the third dimension and making accommodations for the third dimension in the design procedure, there is an additional degree of design freedom, which allows interlocking the screw rotors to be designed without leakage paths or other holes between the male rotor and the female rotor, and between the rotors and the housing, including the vent hole in the transition region as previously stated. Once the design problem is visualized in the third dimension, it becomes clear that there must be a way to eliminate the ventilation hole opening while maintaining the seals between the thread and the slot. Accordingly, the present invention teaches that, in order to eliminate the hole in the blow hole, the sealing region must extend completely from the front cusp of the housing to its posterior cusp. Finally, when design choices are translated back into a flat design methodology, the creation of designs becomes much less difficult than many of the flat design methodologies that are increasingly being suggested as the only way to increase efficiencies. . An example of the inventive method for designing complete families of the corresponding threads and slots of the present invention is also described herein. The new thread and groove design results in a high efficiency screw rotor system, which is therefore unknown in the prior art. The features of the present invention result in an advantage of the improved thermodynamic efficiency and improved volumetric efficiency of the screw rotor device. The tests on the prototype design show that the thermodynamic efficiency are similar to reaching more than 85% and can even exceed 90%. The present invention is seminal because its first screw rotor achieved these efficiencies over a wide range of rotor speeds. The additional features and advantages of the present invention, as well as the structure and operation of various embodiments of the present invention, are described in detail below with reference to the accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS The drawings accompanying the present description, which are incorporated in and form a part of the specification, illustrate the embodiments of the present invention and together with the description, serve to explain the principles of the present invention. In the drawings: Figure 1 illustrates an axial cross-sectional view of a screw rotor device according to the present invention; Figure 2A illustrates a detailed cross-sectional view of one embodiment of the screw rotor device taken along line 2-2 of Figure 1; Figure 2B illustrates a detailed cross-sectional view of another embodiment of the screw rotor device taken along line 2-2 of Figure 1; Figure 3 illustrates a detailed cross-sectional view of the screw rotor device taken along line 3-3 of Figure 1; Figure 4 illustrates a cross-sectional view of the screw rotor device taken along line 4-4 of Figure 1; and Figure 5 illustrates a schematic diagram of an alternative embodiment of the present invention. Figure 6A illustrates a detailed cross-sectional view of the screw rotor device taken along line 6-6 of Figure 2A. Figure 6B illustrates a detailed cross-sectional view of the screw rotor device taken along line 6-6 of Figure 2B. Figure 7A illustrates an axial cross-sectional view of another alternative embodiment of the screw rotor device according to the present invention. Figure 7B illustrates a cross-sectional view along the screw rotor device taken along the line 7B-B of Figure 7A.

Figures 8A-8D illustrate perspective views of another embodiment of the screw rotor device according to the present invention. Figure 9 illustrates an axial cross-sectional view of the screw rotor device according to the embodiment of the present invention in Figures 8A-8D. Figure 10A illustrates a cross-sectional view of the screw rotor device according to the embodiment of the present invention in Figures 8A-8D and 9. Figure 10B illustrates an elevation view of the screw rotor device of according to the embodiment of the present invention in Figures 8A-8D and 9 and with the rotors rotated 90 ° to show the sealing lines and the areas between the rotors themselves and between the rotors and the housing. Figure 10C illustrates an elevation view of the screw rotor device according to the embodiment of the present invention in Figures 8A-8D and 9, and with the rotors rotated 90 ° to show the sealing lines and areas as they exist between the rotors themselves and between the rotors and the housing. Figure 10D shows a detailed cross-sectional view of the screw rotor device according to the embodiment of the present invention in Figures 8A-8D and 9, and showing the ability of the present invention to eliminate the opening of the screw rotor. ventilation hole.

Figures 11A-11H show a series of cross-sectional views of the screw rotor device according to the embodiment of the present invention in Figures 8A-8D, 9 and 10, as the male and female rotors are they intertwine and seal. Figures 12 and 12A-12F show a cross-sectional view of the screw rotor device according to still another embodiment of the present invention together with a series of cross-sectional views of the screw rotor device as the rotors Male and female intertwine and seal. Figures 13A-13H show a series of cross-sectional views of the screw rotor device according to the embodiment of the present invention in Figures 7A and 7B as the male and female rotors interlock and seal. Figures 14 to 16, show a schematic representation of the rotor design procedure according to the present invention together with the families of screw rotor devices resulting from the rotor design procedure. Figure 17 shows a flow chart of the design procedure for making a family of screw rotor devices according to the present invention. Figure 18 shows the screw rotor device in a cooling / cooling cycle application.

Figure 19 shows the screw rotor device in a hydrostatic impeller application. Figure 20 shows the screw rotor device in a hydrodynamic impeller application. Figure 21 shows the screw rotor device in a compressor application and in an energy impeller application. Figure 22 shows the screw rotor device in a gas turbine engine application.

DETAILED DESCRIPTION OF THE INVENTION Referring to the accompanying drawings in which similar reference numerals indicate similar elements, Figures 1 and 9 illustrate a schematic cross-sectional view of a screw rotor device 10. The screw rotor device 10 it generally includes a housing 12, a male rotor 14 and a female rotor 16. The housing 12 has an inlet port 18 and an outlet port 20. The inlet port 18 is preferably located at the end of the gear 22 and the housing 2, and the outlet port 20 is located at the opposite end 24 of the housing 12. The male rotor 14 and the female rotor 16, respectively, rotate about a pair of substantially parallel axes 26, 28 within a pair of holes cylindrical 30, 32 extending between the ends 22, 24.

In the preferred embodiment, the male rotor 14 has at least one pair of helical threads 34, 36, and the female rotor 16 has a pair of corresponding helical grooves 38, 40. The female rotor 16 rotates in the opposite direction with respect to the male rotor 14 and each of the helical grooves 38, 40 respectively is entangled in phase with each of the helical threads 34 36. In this way, the working fluid flows through of the inlet port 18 and inside the screw rotor device 10 in the spaces 39, 41 delimited by each of the helical threads 34, 36, the female rotor 16 and the cylindrical bore 30 around the male rotor 14. It will be appreciated that the helical grooves 38, 40 also define the spaces that delimit the working fluid. The spaces 39, 41 are closed from the inlet port 18 as the helical threads 34, 36 and the helical grooves 38, 40 intertwine at the inlet port 18. As the female rotor 16 and the male rotor 14 they continue to rotate in the opposite direction, the working fluid is positively displaced towards the outlet port 20. The pair of helical threads 34, 36 have a phase compensation aspect which is particularly described with reference to Figures 2A, 2B and 3, which show the cross-sectional profile of the screw rotor device through line 2-2, the two-dimensional profiles are shown in the perpendicular plane of the rotation axes 26, 28. The aspect of phase compensation is also raised further on referring to Figure 7A, and also shown in the embodiments shown in Figures 10 to 16. The cross section of the helicoidal thread pair. oidales 34, 36 includes a pair of corresponding teeth 42, 44 delimiting an edentulous sector 46. The phase compensation of the helical threads 34, 36 is defined by the arc angle ß subtending the edentulous sector 46, which depends on the angle of the arc a of any of the teeth 42, 44. Particularly, for the helical phase compensation threads, the edentulous sector 46 has an arc of angle ß which is preferably equal to or greater than the arc angle to which it subtends from any of the teeth 42, 44. The phase compensation ratio referred to between the angle of the arc ß and the angle of the arc a is defined in particular by the following equation: Arc Angle ß > M * Arc angle a, M > 1 (1) As illustrated in Figures 2A, 2B, 10A, 12 and 13, the angle between the beam segment a and the beam segment ob, which subtends tooth 42, is the angle of the arc a. According to the definition of phase compensation provided above, the angle of the arc ß of the edentulous sector 46 extends from the beam segment ob to the beam segment a ', which could generally correspond to a multiplier (M) of the arc angle arc a. It is considered that the highest efficiencies can be obtained by the compensation multipliers of two or greater. In the preferred embodiment, the angle of the arc ß of the edentulous sector 46 which extends approximately five times the angle of the arc a to the ray segment or a ", which corresponds to a phase compensation multiplier of five (5). another two additional teeth can be potentially adjusted on opposite sides of the male rotor 4 between the teeth 42, 44. In order to balance the male rotor 14, it is preferable to have an equal radial spacing of the teeth. teeth because an odd number of teeth can also be equally spaced around the male rotor 14. Additionally, the number of teeth that can be adjusted around the male rotor 4 is not particularly limited by the preferred embodiment. the arc angle ß is proportionally larger than the angle of arc a according to the phase compensation multiplier. of the arc ß of the edentulous sector 46 can decrease proportionally to any decrease in the angle of the arc a of the teeth 42, 44, thus allowing more teeth to be added to the male rotor 14 while maintaining the phase compensation ratio . Whatever the number of teeth in the male rotor 14, the female rotor has a corresponding number of helical grooves. Accordingly, the helical grooves 38, 40 have a phase compensation aspect corresponding to those helical threads 34, 36. Therefore, the female rotor has the same number of helical grooves 38, 40 as the number of helical threads 34. , 36 in the male rotor, and the helix angle of the helical grooves 38, 40 is on the opposite side of the helix angle of the helical threads 34, 36. It will be appreciated that, by a rotor of given diameter, the helix angle of the grooves and threads really varies depending on their depths In particular, referring again to Figure 1, the upper base of the threads will have a lower helix angle than the root of the threads, and the depression of the grooves will have an angle of helix higher than the channel of the slots. In one embodiment, each of the helical grooves 38, 40 has a concave rear cut profile 48 and corresponds radially to the axial widths that narrow from the locations between the smaller diameter 50 (md) and the larger diameter 52 (MD ) towards the larger diameter 52 at the periphery of the female rotor 16. The concave rear cutting profile 48 includes the line segment jk extending radially between the smaller diameter 50 and the larger diameter 52 in a beam from the shaft 28, the I line segment extending radially between the smaller diameter 50 and the larger diameter 52, and a smaller diameter arc Ij extending circumferentially between the line segments jk, Im. The line segment jk is substantially perpendicular to! larger diameter 52 in the periphery of the female rotor 16 and the line segment Imn preferably has a radius Im combined with a straight segment mn. In particular, the radius Im is between the straight segment mn and the smaller diameter arc Ij and the straight segment mn intersecting the greater diameter 52 at an acute outer angle f which results in a rear cut angle f defined by the following equation (2). Rear cutting angle f = Right Angle (90 °) - External Angle f, (2) The rear cutting angle f and the substantially perpendicular angle on the opposite sides of the rear cutting concave profile 48 result in narrowing axial width radially on the periphery of the female rotor 16. In this mode of rear cutting, the helical grooves 38, 40 are opposite each other about the axis 28, such that the line segment jk for each of the pairs of helical grooves 38, 40 is directly in line with each other through the axis 28. consequently, in the back cut mode, the line segment kjxj'k 'is preferably straight. In the preferred embodiment of the present invention, the screw rotor device 10 operates as a screw compressor in a gaseous working fluid. Each of the helical threads 34, 36 may also include a distal labyrinth seal 54 and a sealing band 56 may also be wedged within the distal labyrinth seal 54. The distal labyrinth seal 54 may also be formed by a number of striations at the tip of the helical threads (not shown). When operating as a screw compressor, the screw rotor device 10 can also use a valve 58 that communicates operatively with the outlet port 20. As an example, a valve 58 is a time synchronizing plate 60 attached to and rotates with the male rotor 14 and is located between the male rotor 14 and the outlet port 20. As illustrated in particular in Figure 4, the time synchronizer plate 60 has a pair of disruptors 62, 64 that open from sequentially to the output port 20. Between the disruptors 62, 64, the time synchronizing plate 60 forms additional limits 66, 68 for the spaces 39, 41 respectively. As the male rotor 14 rotates in the opposite direction with the female rotor 16, the limits 66, 68 cause the volume in the spaces 39, 41 to decrease, and the working fluid pressure increases. Then, as the disruptors 62, 64 pass respectively on the outlet port 20, the pressurized working fluid is forced out of the spaces 39, 41 and the spaces 39, 41 continue to decrease in volume until the bottom of the respective helical threads 34, 36 pass over the outlet port. Figure 5 illustrates another embodiment of the screw rotor device which only has a helical thread 34 which intertwines with the corresponding helical groove 38 and preferably has a valve 58 at the outlet port 20. As illustrated in Figure 5, the valve 58 may be a reed valve 70 attached to the housing 12. In this single-threaded embodiment, the weight can be added to the male rotor 14 and the female rotor 16 to balance them. The helical groove 38 can have a concave rear cut profile 48 described above, and the male rotor 14 again rotate in the opposite direction with respect to the female rotor 16. The single thread mode also illustrates another aspect of the screw rotor device 10 of the present invention. In this embodiment, the length of the screw rotor device 10 is approximately a single angle of the helical thread 34 and the slot 38. The angle of a screw is generally defined as the distance from any point on a screw thread to a corresponding point on the next thread, measured parallel to the axis and on the same side of the axis. The particular screw rotor device 10 illustrated in Figure 5 has a single thread 34 and a corresponding slot 38. Accordingly, a single angle of the thread 34 and the slot 38 require a full 360 ° helical rotation of the thread 34 and the slot 38 correspondingly. The present invention is directed to screw rotor devices 10 having an identical number of threads and grooves (N), and the helical turn required to provide the single angle is merely defined by the number of threads and grooves (N = 1, 2, 3, 4, ...) according to equation (3) below. Helical rotation of the single angle = 360 N (3) Of course, it will be appreciated that even in the example in which the length of the device of the screw rotor 10 is a single angle, the length of the angle can be changed by means of the alteration of the helix angle of the threads and grooves. The length of the angle increases as the helix angle becomes steeper. The screw rotor device 10 illustrated in Figure 1 has a pair of threads 34, 36 and a pair of corresponding helical grooves 38, 40, (N = 2). Therefore, a single angle of these rotors could only require a 180 ° helical turn (36072). However, it is evident that the screw rotor device 10, as illustrated in Figure 1, has a length slightly greater than angles. Therefore, for the determined length of the rotors, the helix angle for the threads and grooves must be increased for the rotors having a single angle length. For example, Figures 7A and 7B illustrate a screw rotor device 10 having a pair of threads 34, 36 and a pair of corresponding helical grooves 38, 40 having a helical turn of 180 °. Accordingly, Figures 7A and 7B particularly illustrate rotor lengths having a unique angle of threads 34, 36 and slots 38, 40. Although it may be preferable, and in some cases even advantageous, to design the length of the rotor to approximately a single angle to design certain threads, it is not necessary to design a limitation for the screw rotors according to the present invention. The screw rotor device 10 illustrated in Figure 7A also incorporates phase compensation relationship in its design. The angle between the ray segment a and the ray segment ob, which subtends the teeth 42, is an angle of arc a. According to the definition of phase compensation provided above, the arc angle ß of the edentulous sector 46 extends from the ray segment ob to the ray segment a ', which could correspond to the multiplier (M) and the angle of arch a. As particularly illustrated in Figure 3, the helical thread 34 in this embodiment has a convex inner cutting profile 72 which is linked to the concave rear cutting profile 48 of the helical groove 38. The convex inner cutting profile 72 has a tooth segment 74 extending radially from the smaller diameter of the arch ab. The tooth segment 74 is subtended by the arc angle a and is further defined by equation (4) below, according to the arc angle? for the smaller diameter of arch ab. Arc angle to >; Arc angle? (4) The phase compensation ratio defined for a pair of threads can also be applied to the male rotor 14 with the single thread 34, such that the toothless sector 46 must have an arc angle β which is at least two times the arc angle a of the single helical thread 34. The circumference of the male rotor 14 is 360 °. Therefore, to design a rotor having a phase compensation multiplier of at least 2 and a single thread, the arc angle ß for the dented sector 46 must be at least 240 ° and the arc angle can be not be greater than 120 °. In a similar way, to design the rotor having a phase compensation multiplier of at least 2 with a pair of threads 34, 36, 60 ° is the maximum arc angle that this minimum compensation phase multiplier could satisfy. of two (2) and 30 ° could be the maximum arc angle that the multiplied phase compensation multiplier of five (5) could satisfy. For practical purposes, it is likely that only large diameter rotors can have a phase compensation multiplier of 50 (maximum arc angle of 3 °) and fabricate items that can limit higher multipliers. The male rotor 14 and the female rotor 16 each have a respective central shaft 76, 78. The shafts 76, 78 are rotatably mounted within the housing 12 through the bearings 80 and seals 82. The male rotor 14 and the Female rotor 16 are linked together through a pair of gears rotating in opposite direction 84, 86 which are respectively linked to shafts 76, 78. Central shaft 76 of male rotor 14 has an end extending out of the housing 12. When the screw rotor device 10 operates as a compressor, the shaft 76 is rotated causing the male rotor 14 to rotate. The male rotor 14 causes the female rotor 16 to rotate in the opposite direction through the gears 84, 86, and the helical threads 34, 36 intertwine with the slots 38, 40. As described above, the distal labyrinth seal 54 it aids the sealing between each of the helical threads 34, 36 in the male rotor 14 and the cylindrical bore 30 in the housing 12. In a similar manner, as particularly illustrated in Figure 3, the axial seals 88 can be formed in the housing 12 along the length of the cylindrical bore 32 to assist the sealing on the periphery of the female rotor 16. As the male rotor 14 and the female rotor 16 make the transition between the bonding with each other and the respective sealing around of the housing 12, a small hole 90 is formed between the male rotor 14, the female rotor 16 and the housing 12. The rotors 14, 16 are adjusted in the housing 12 with close tolerances between the rotors and the housing and the rotors themselves having close tolerances between the threads 34, 36 and the slots 38, 40. In particular, the upper base 120 of the threads 34, 36 and the larger diameter of the female rotor 52 are in a sealing relationship with the holes cylindrical members 30, 32 of housing 12, respectively. Additionally the upper base 120 of the threads 34, 36 are also in sealing relation with the depression and the base base 110 of the grooves 38, 40. As it is set forth in detail with respect to the following Figures 10 to 16, the relationship The seal can be in the form of a sealing line or a sealing area. Generally, the close tolerances that supply the lift to the sealing ratio are in the order of magnitude of approximately 0.0076 centimeters per 991 liters per minute of the screw rotor compressor system, although the tolerances could relax depending on the size of the rotor device. screw and the amount and index of working fluid that is compressed, pumped and expanded. For example, if the threads and slots are designed to displace 991 liters per minute for rotors with diameters of approximately 6.7 inches, a larger compressor that has similar threads and grooves could have a slightly larger tolerance while maintaining a thermodynamic efficiency that can be compared. As will be explained in detail below, there are also other factors that can affect the sealing tolerances for a particular screw rotor system, such as the application in which the screw rotor is to be used. It will also be appreciated that, depending on the application, the temperature range experienced by the rotors could vary, and the tolerances can be designated to account for the thermal expansion and contraction of the rotors, as well as the housing.

Also, the material for the rotors and the housing can be selected such as the sealing distances, or tolerances, does not vary substantially throughout the range of operation of the thyroid rotor system. For example, the materials may have a similar thermal expansion module or they may be selected in such a way that they can achieve an optimum seal at the particular design point or in a region of operation in the static state. As discussed above, the preferred embodiment of the screw rotor device 10 is designed to operate as a compressor. The screw rotor system 10, can also be used as an extender. When acting as an extender, the gas has a higher pressure than the ambient pressure entering the thyme rotor device 10 through the outlet port 20, valve 58 is optional. The gas pressure forces the rotation of the male rotor 14 and the female rotor 16. As the gas expands in the working spaces 39, 41, it is drawn through the end of the shaft 76 which extends outside the housing 12. The pressure in the spaces 39, 41 decreases as the gas moves to the inlet port 18 and exits the ambient pressure at the inlet port 18. The screw rotor device 10 can operate with gaseous working fluids and it can also be used as a pump for a liquid working fluid. To pump liquids, a valve can also be used to prevent fluid from flowing back into the rotor. Figures 6A and 6B illustrate a detailed cross-sectional view of the helical grooves and helical threads of Figures 2A and 2B, respectively. These views illustrate the differences between the cusp profile of the threads 92, which may include one or more involution curves, and another characteristic of the present invention, a reinforcing thread profile 94. Among the smaller diameter 50 and the larger diameter 52 of the female rotor, the thread cusp profile 92 of the helical groove 38 includes a concave line 96 and a substantially straight line 98 opposite therefrom. The reinforcing thread profile 94 also includes a concave line 96 although it is particularly defined by a diagonal straight line 100. In the male rotor, the thread cusp profile 92 of the helical thread 34 is also between the major and minor diameters and they include a pair of opposite convex curves. In comparison, the reinforcing thread profile 94 has a diagonal straight line 102 that is parallel to and in close tolerance with the corresponding diagonal straight line 100 in the helical groove 38. In the particular example illustrated in Figure 6B, a convex curve 104 is opposite the diagonal straight line 102. Figures 7A and 7B, illustrate in particular the thyme rotor device 10 according to various aspects of the present invention, including the parallel diagonal straight lines 100, 102 of the thread profile of reinforcement 94, helical phase compensation threads 34, 36, and single angle design of male and female rotors 14, 16 within housing 12. With respect to the particular example illustrated in Figure 7B, the thread profile reinforcement 94 includes an opposite concave curve 104 of diagonal straight line 102. It will be appreciated that the benefits of the present invention can be achieved with manufacturing tolerances, such as in the parallel diagonal straight lines 100, 102. In particular, the tolerances in the parallel diagonal straight lines 100, 102 can allow a small radius of curvature between the diagonal lines and the greater and smaller diameters and an extremely slight divergence in the parallelism. It will be appreciated that manufacturing tolerances may vary depending on the type of material being used, such as metals, ceramics, plastics and compositions thereof, and depending on the manufacturing process, such as machining, extrusion, casting, and combinations thereof. Figures 8A to 11-H illustrate an embodiment of the present invention which, as in the embodiments discussed above, significantly reduces the vent opening, and as discussed below with respect to this embodiment, to the currently held belief, the present invention can even eliminate the vent opening in its entirety. As discussed above, prior designs have failed to create a complete seal except for a single angle full seal design, ie, the reinforcing thread design established and claimed in the US Patent Application. No. 10 / 283,422 co-pending. However, the present invention eliminates the vent hole opening as well as other internal leaks that reduce the thermodynamic efficiency and volumetric efficiency of the screw rotor devices. In particular, in addition to the blow hole orifice described above, the present invention can eliminate or significantly reduce the following forms of internal leakage: (1) the orifices between the inlet port and / or the outlet port in the housing and the rotors, resulting in a better than complete capture or ejection of the working fluid through the rotors; (2) the holes between the outer periphery of each rotor and the inner surface of the housing, through which the working fluid leaks around the upper base of a thread or the channel of a groove for an adjacent work volume , respectively; and (3) holes between the front and back of the thread of the male rotor and the groove of the female rotor which interlock, through which the working fluid leaks from the pressurized side to the suction side. To ensure that persons of ordinary skill in the art will appreciate the expansive scope of the present invention, it should be understood that although multi-angle screw rotor designs had the ability to significantly reduce or eliminate the three previous forms of leakage and the single-angle, reinforced screw rotor designs had the ability to significantly reduce or eliminate the vent opening, no previously known screw compressor design had the ability to eliminate or reduce significantly all these internal leaks simultaneously and without limitation. Although it is true that the designs of threaded support rotor could significantly reduce or eliminate the opening of the ventilation hole, its elimination is achieved at the price that the complete seal will only work for a single angle, although not due to the ventilation hole opening. Instead, when the back-screw rotor designs are used for multi-angle screw rotors, the hole between the front and the back of the male rotor thread and the interlacing female rotor slot could then produce a significant leak from the high pressure side of the screw rotor to the low pressure or suction side of the rotor. The holes between the rotors themselves and between one or more of the screw rotors and the housing, such as the hole 90 illustrated in Figure 3 and discussed below, can be visualized as leakage paths. A leakage path can be viewed generally as any current tube between the male rotor and the female rotor, or between one or more of the rotors and the housing, which extend from the front side and the rear side or on one side, between a region of higher pressure and a region of lower pressure. To define the current tube, this can be formed by a group of continuous holes with an effective diameter exceeding an order of magnitude greater than a sealing tolerance defined for the screw rotor system. For example, a sealing tolerance can be based on the distance between the upper base of the thread and the lower base of the slot. Alternatively, the seal tolerance 106 may be based on the distance between one or more of the rotors and the housing. It does not matter that the actual distance of the seal tolerance is based only on that which the reference makes sense for the particular use of the screw rotor design. For the present invention, an approximation of thermodynamic efficiency of 85% has already been observed and it is expected that a thermodynamic efficiency of 90% can be achieved. The thermodynamic efficiency of 85% to 90% should still be achieved in accordance with the modalities described in the present description when a positive displacement of the working fluid is controlled using a valve, such as the lamellae valve discussed above. As an example of the different tolerances for different applications, the screw rotor system 10 illustrated in Figure 9, could be used as a fluid measuring system or in a hydraulic system which does not activate too fast and / or generates much hot. In such a system, the sealing tolerances must be zero (0), or as close to zero (0) as physically possible with the machining and other manufacturing techniques and are required to allow thermal expansion of the rotors, such as when the screw rotor system is used as an internal combustion engine. For another system, such as the adiabatic compressor or extender, the tolerances can be a little more relaxed. As previously stated, the tolerances can also vary depending on the size of the screw rotor system, being narrower for smaller systems and the relaxation of larger systems. Generally, the sealing tolerance for the present invention, between the helical thread and the helical groove, may be set as a defined number, such as less than or equal to 0.0076 centimeters or 0.0025 centimeters or some other small distance. Still more generally, the sealing tolerance can be based on the ratio of the rotor diameters of the screw rotor system 10, such as a rule that the sealing tolerance is not greater than 1/1, 000 or 1 / 10,000. of the diameter of the male rotor. More generally, the seal tolerance can be based on any geometrical proximity, which can be defined by the distance between the rotors themselves, the rotors and the housing and any other distance that is relevant to the sealing conditions. Depending on the geometrical proximity that is selected, the sealing tolerance can be defined by the geometrical proximity itself or can be based on it, such as a sealing tolerance, which is within an order of magnitude of proximity geometric It will be appreciated that the hole 90 in the embodiment illustrated in Figure 3A and set forth above is within a sealing tolerance that is within an order of magnitude of the distance between the upper base of the thread and the lower surface of the slot. It will also be appreciated that the hole 90 in the embodiment illustrated in Figure 10D is even smaller than that in Figure 3A and that the hole can be completely removed, designing the cusp of the housing to be exactly at the point where the thread intersects. the groove, as discussed in detail below with respect to the thread and groove that seal on one or both of the cusps (SR-6 and SR-7). Accordingly, there is no leakage path or current tube to be shown in the present invention. However, leakage trajectories are already well defined in the art and are understood by those skilled in the art. For example, leakage paths are set forth in detail in the U.S. Patent. No. 5,533,887, which is incorporated herein by reference. The particular structure and method of the present invention are set forth with reference to the features illustrated in particular form in Figure 10C. As discussed above, the female rotor 16 has a larger diameter and a helical groove 38. The groove recedes from the larger diameter to a bottom base 110, or through it, located between the conductive side 112 and the tracking side 114. , which are shown respectively as the bottom side and the top side in the illustration. The main side and the retrospective side respectively include a main channel 116 and a retrospective channel 18 in the larger diameter. The male rotor 14 has a smaller diameter and a helical thread 36, which as previously stated, is entangled in a rotary fashion in phase with the helical groove. The helical thread extends from the smaller diameter to an upper base 120 located between a main face 122 and a retrospective face 124, which are respectively shown as the bottom face and the top face in the illustration. The main face and the retrospective face include a leading edge 126 and a retrospective edge 128, respectively. The housing 12 has a front cusp 130 along its front side FS and a rear cusp 132 along its rear side.

The helical thread is connected to the smaller diameter of the male rotor through its root portion 134. To show the sealing relationships of the present invention, Figure 10C uses the symbols A, B and C to refer to the seal on the front side of the screw rotor and A ', B' and C to refer to the seal on the back side of the screw rotor. It will be appreciated that the top and bottom of the screw rotor are relative to its placement and are only used for simplicity of reference in relation to the drawing. Generally, in accordance with the direction of travel shown in Figures 10A and 0B, the retrospective upper portions and the portions of the bottom are the major portions. Of course, if the direction of the rotors is inverted, the upper portions could then be the main portions and the portions of the bottom could then be retrospective portions.

Also, Figure 10C uses alpha-numeric reference codes and other symbols to uniquely identify the following sealing regions (SR), which may also be referred to as the sealing relationships; SR-1: 1f (CC ') - upper base seals with bottom base (1st SR). SR-2: A / CB - retrospective channel seals, at least partially, along the retrospective side (2nd SR). SR-3: AC - retrospective edge seals with retrospective side (3rd SR).

SR-4: A7CB '- main face seal with main channel (4th SR). SR-5: A'C - main edge seals with main side (5th SR). SR-6 / SR-7: T- triple seal between channel, edge and cusp A (6th.

SR) and cusp A '(7th SR). SR-8: || - larger diameter of the female rotor seals with cylindrical hole (8th SR). SR-9: || - Top base seals with cylindrical hole (9th SR). SR-10 || - larger diameter female rotor diameters with smaller male rotor diameter (10th SR), which includes the seal between the groove channel and the root portion of the thread. XB-superior and XB 'of the fund. SR-11 / SR-12: = - the seal of the ends of the housing with the respective ends of the rotor (11th SR and 12th SR). As summarized in the above listing and particularly illustrated in Figure 10C, the sealing relationships are described in detail below. The first seal ratio SR-1 has a center, the interlocking sealing area defined by the geometries of the upper base and the lower base. The second seal ratio SR-2 has a front, the inner sealing line defined by the retrospective edge geometries and the retrospective side. The fourth sealing relationship SR-4 has an outer, later sealing line defined by the geometries of the main face and the main channel. The fifth sealing ratio SR-5 has a rear, inner sealing line defined by the geometries of the leading edge and the main side. The outer front sealing line and the inner, front sealing line define boundaries of a sealing area that is intertwined between the retrospective side and the retrospective side and intersected at a common frontal sealing point according to a sixth relationship of SR-6 sealed defined by the retrospective intersecting edge, the retrospective channel and the frontal cusp. The outer, back seal line and the inner, posterior seal line define the boundaries of a back interlaced sealing area between the main face and the main side and intersect at a common posterior sealing point in accordance with the seventh aspect ratio. sealed SR-7 defined by the intersection of the main edge, the main channel and the cusp. The eighth sealing ratio SR-8 has a first peripheral sealing area defined by the geometries of the larger diameter of the female rotor and the cylindrical bores. The ninth seal ratio SR-9 has a second peripheral sealing area defined by the geometries of the upper base and the cylindrical bores. The tenth SR-10 sealing ratio has a central non-interlaced sealing area defined by the geometries of the larger diameter of the female rotor and the smaller diameter of the male rotor, and includes the seal between the groove channel and the root portion of the thread. As with most screw rotor compressors, the ends of the female rotor and male rotor are in a sealing relationship with the ends of the housing, ie the eleventh sealing ratio SR-11 and the twelfth ratio of sealed SR-12. It will be appreciated that a number of these sealing regions are sealing areas, while others may be sealing lines, depending on the particular selection of design variables for the rotors, which are discussed below. The creation and progression of these seals as the male and female rotors are intertwined, is illustrated in Figures 11 A-11 H. These illustrations show a series of cross-sectional views of the screw rotor device, and the regions of Seals are shown and described with reference to these. Even after the thread 36 and the slot 38 begin to seal, there is a seal between the larger diameter of the female rotor and the smaller diameter of the male rotor. On the front side of the screw rotors 14, 16, the upper part of the thread 124 begins to seal the upper part of the right slot 114 in the front cusp 130 and, as the rotors continue to interlock, the sealing continues along the top of the groove for the full length from the larger diameter of the female rotor to its smaller diameter (points A and C, illustrated respectively in Figure 10A). On the rear side of the screw rotors, the bottom part of the slot 12 begins to seal the bottom of the thread 122 at its root 134 (point B illustrated in Figure 10A), and as the rotors continue to interlock, they continue sealing more of the root until the bottom of the thread begins to seal along the bottom of the groove and is ultimately sealed along the entire bottom of the groove from the smaller diameter of the female rotor to its larger diameter ( points A 'and C, illustrated respectively in Figure 10A). the intermediate points that cover the upper and lower part of the grooves are also sealed respectively with the intermediate points that cover the top and bottom of the threads. The bottom of the slot completes the sealing of the bottom of the slot at the rear cusp 132. As discussed in detail below, with respect to the illustrations in Figures 14 to 16 and 17, all of these stamps can be designed in the family of screw rotors according to the present invention, and incorporating all these seals in a screw rotor system, all the leaks raised above, including the opening of the vent hole can be reduced simultaneously to within the specified tolerances, also raised above. With the reinforcing thread rotor designs (see Figures 7A and 7B), the vent opening can still be eliminated, although the full seal is limited to a small angle because, with the multi-angle rotors, a hole 134 extends between the retrospective side and the retrospective face of the thread (see Figure 13E) which could cause significant leaks from the high pressure side of! Screw rotor system beside low pressure or suction screw screw system. In accordance with the designs of other unsupported thread embodiments of the present invention, the hole between the retrospective side of the slot and the retrospective face of the thread does not exist, even though the screw rotors are multi-angle designs. Generally speaking, the back thread designs have a one-sided sealing relationship, ie, between the main side 112 of the slot 38 and the main face 122 of the thread 36, where the other designs have a relationship two side seal between the main side 112 of the slot 38 and the main face 122 of the thread 36 and between the retrospective side 114 of the slot 38 and the retrospective face 124 of the thread 36. The sealing ratio of two sides can be defined in a particular way by the first sealing ratio SR-1, the second sealing ratio SR-2, the third sealing ratio SR-3, the fourth sealing ratio SR-4, and the fifth sealing ratio SR-5. In this sense, no leakage path is provided through its two-side sealing relationship. An illustration of this two-sided seal 136 is shown particularly for the multi-angle rotors 138, 140 in Figure 10B. In particular, there is a main axial seal 142 between the main face of the thread and the main side of the groove and a retrospective axial seal 144 between the retrospective face of the thread and the retrospective side of the groove, and these sealing regions can be sealing areas. For compressor applications, the main face / main side seal may be more important than the retrospective / retrospective side seal because the retrospective face seal complies with and "disappears" on the end seal as it is applied. completes the compression stroke (see Figure 10B). However, the retrospective side / retrospective side seal can be especially useful if it is desired to maintain a preliminary compression of the working fluid, i.e., even before the thread is sealed with the groove.

Although similar groove shapes are shown in the prior art screw rotors and similar thread forms are shown in other prior art screw rotors, not only were said threads and grooves never combined in a rotor system of the prior art. screw, but none of these references of the prior art has ever suggested that they could be combined with the thread of the other references. In fact, none of these prior art designs were based on the present design method. Accordingly, the threads and slots of all these prior art screw rotors fail to meet the structural features described and claimed by the thread and groove of the present invention. Additionally, prior art references fail to describe cooperative relationships between the threading, slot and cusps of the housing, as described and claimed by the present invention. Finally, none of the references of the prior art describes or suggests the design method of the present invention. In fact, as stated in the previous background section of the invention, the prior art actually suggests that it is not possible to have any design procedure, or resulting design, which eliminates the opening of the vent hole. The design method of the present invention is established schematically in the illustrations of Figures 14 to 16, and is established as a flow chart 17. To obtain a visual photograph of the process, Figure 15 is particularly useful for understanding the inventive design procedure. Generally, the retrospective edge of the upper base 1 and the leading edge 2, respectively, define the retrospective side of helical groove T and the main side 2 'as the helical thread is intertwined with the helical groove. To eliminate the hole in the blow hole on the front side of the screw rotor device, the retrospective groove channel and the retrospective thread edge intersect at the front cusp 130, ie, within the defined seal tolerance by the rotors. Similarly, to eliminate the vent opening in the rear side of the screw rotor device, the main channel of the slot and the leading edge of the thread intersect at the rear cusp 132. Finally, the retrospective channel of the groove 3 and the main channel 4, respectively, define the retrospective root portion of the thread 3 'and the main root portion 4', and the intermediate points covering the bottom side of the groove 3"and the upper side 4", respectively define the intermediate points covering the bottom face of the thread 3" 'and the top face 4"'. During the removal of the vent hole opening on the front side and the rear side of the housing, it will be appreciated that the thread profile has discontinuities between its upper base and its upper and lower faces, i.e., the retrospective and main, respectively, for the type of compressor or pump of the application. The discontinuity of the main edge is located at the point of the main edge where the main line and the largest diameter arc intersect. The discontinuity of the retrospective edge is located at the retrospective edge point where the retrospective line and the largest diameter arc intersect. According to this visual image of the design procedure, it will be appreciated that the lines of the cross section profile of the thread between the upper base and the root can be formed from any type of line, including straight lines, concave lines, lines convex, involutions, reverse involutions, parabolas, hyperbolas, cycloids, trochoids, epicycloids, hypotrocids, continuous straight lines and arched lines and any combination of these in continuous lines of intelligent pieces. Figures 15 and 16, illustrate other thread and groove designs that can form complete families of screw rotor profiles. Figure 15, takes the retrospective line of groove and main line of Figure 14 and converts them into a main line of the thread and the retrospective line, that is, inverting them, to show that the same design procedure can be used in reverse and will result in slot sides that are a reverse take of the slot sides in Figure 14. Figure 16 shows in shaded lines, the main slot line and the retrospective line of the initial design stage, ie , before using the intermediate points covering the bottom side of the groove 3"and the upper side 4", respectively, to define the intermediate points that cover the face of the bottom of the thread 3"'and the top face 4" ' After completing this final step, the solid lines show that the main lines of the threads and the retrospective lines, that is to say, that correspond respectively to the main lines of the slot and the retrospective lines, become more arched. However, for machining purposes, it is still possible to change the design to a group of straight line segments, or even other arched sections, while still remaining within the design tolerances for the particular application and the rotor family. Figure 16 also shows how the families of curves can also be based on smaller diameters different from the male and female rotors, even when the larger diameters remain constant. The design procedure of the present invention is now described with reference to the flow diagram in Figure 17, and as indicated above, the visual image of the process is described with reference to Figure 14: (a) defining the larger diameters of male and female rotor and their amount of superpositioning (200), that is, define a pair of larger intersecting circles 52 'each having a center or and a larger diameter 52 such that each of the circles covers only its own center and centers are separated less than a sum of a half of the largest diameters; (b) defining the upper base 120 of the tooth 42 in one of the larger circles that are straightened 52 '(210); (c) identifying the pair of sides 1 ', 2', which are withdrawn radially from the other circle to a bottom base 110 (220); the sides are defined by the trajectory of the upper base 120 'when the circles 52' rotate in phase with each other by equal angular quantities; and the sides include a pair of intermediate line segments 3", 4" that are withdrawn from a pair of circumferential channels 3, 4 to the base bottom 110; (d) identifying the tooth pair of root sections 3 ', 4'; the root sections are defined respectively by the trajectories of channels 3? ,4? when these circles rotate in phase with each other through equal angular quantities and identify a pair of teeth of radially extending line segments 3 '", 4'" (230); radially extending line segments are defined by the trajectories of intermediate line segments of the slots 3"?, 4"? when the circles rotate in phase each other by equal angular quantities. Due to these design conditions, it will be appreciated that the threads and grooves may be designed in accordance with the present invention in such a way that they have minimal negative repercussions. In particular, many designs for screw rotors have pressure angles as large as 30 °, which results in a significant amount of negative repercussions. In comparison, the present invention allows designers to create complete families of screw rotors with minimal negative repercussions, such as with pressure angles less than half of 30 °, including families with 0 ° pressure angle and no negative impact.

It will also be appreciated that, upon completion of the screw rotor system design, the inner sides of the housing are generally defined by the shape of an eighth figure in close tolerance with the circles 240. As illustrated in Figure 9, the entry and exit can be in the form of a wedge shape. In particular, the entrance can be a trapezoid, and the exit can be a triangular side port, that is, with a generally V-shaped. As it was stated with respect to the modalities set forth with respect to Figures 1 to 7, the port The output can be a circumferential end port or a V-shaped circumferential end port. Similarly, the input port can be a circumferential end port or a W-shaped circumferential end port. create the third dimension for the screw rotors, at least one helix angle needs to be selected 250. As stated above, the helix angle can be varied along the length of the rotors, resulting in this way a screw rotor compressor angle variable. Also, the major and minor diameters can be varied along the length of the rotors, thus resulting in a tapered screw rotor compressor. Still in greater detail of the design procedure, the first major circle of the rotor is defined. The first largest circle of the rotor has a larger first diameter. The second major circle of the rotor is also defined, so that it intersects with the first major circle of the rotor at a pair of intersection points. The second larger circle of the rotor has a second larger diameter, and less than one half of the second larger diameter extends into the first larger circle of the rotor. Less than one half of the first larger diameter extends into the second larger rotor circle, and the second larger rotor circle shares a single tangential point with a smaller first rotor circle centered within the first larger rotor circle. The first larger rotor circle shares another unique tangential point with a second smaller rotor circle centered within the second larger rotor circle. A first point is now selected in the first major rotor circle, and the point defines a first line segment that is withdrawn radially inwardly from the second major rotor point to the second minor rotor point. In particular, the first line segment is defined by the path of the first point as it progresses from the second larger rotor circle to the second smaller rotor circle when the first larger rotor circle and the second larger rotor circle rotate in phase with each other by equal angular quantities. Similarly, a second point in the first major rotor circle is selected and circumferentially separated from the first point, and the point defines a second line segment that is withdrawn radially inwardly from a second major rotor point circumferentially spaced to a second minor point of circumferentially spaced rotor. The second line segment is defined by the path of the second point as it progresses from the second major circle of the rotor to the second smaller circle of the rotor, when the first major circle of the rotor and the second largest circle of the rotor rotate in phase between yes by equal angular quantities. Additionally, the second major circumferentially spaced rotor point and the second minor rotor point are circumferentially spaced from the second major point of totoral second minor rotor point, respectively. Now we identify a pair of first rotor root line segments that extend from the smaller circle of the rotor to a pair of intermediate points. An intermediate point is located between the first smaller circle of the rotor and the first point in the first larger circle of the rotor and another intermediate point is located between the first smaller circle of the rotor and the second point in the first larger circle of the rotor. The intermediate points are circumferentially spaced from each other, and the first rotor root line segments are defined by the trajectories of the second major point of the rotor and the second major point of the rotor circumferentially spaced when the first circle of the largest rotor and the second circle The larger rotor rotates in phase with each other by equal angular quantities. Finally, to complete the profile of the thread, it is preferable to use a pair of circumferentially spaced first rotor line segments extending respectively between the pair of first rotor line segments of the rotor and the first point and the second point in the rotor. first largest circle of the rotor. During the design of profiles of the screw rotor devices, it will be appreciated that the upper base of the thread is preferably an arc instead of merely being a point in the larger diameter of the male rotor. This preference may be more important, because a point may tend to cause the Bernoulli effect, causing the upper base of the thread and the base of the groove bottom to act as a convergence-divergence nozzle. Due to the pressure differentials, this effect could even result in a supersonic flow through said nozzle, causing shock waves, which are not adiabatic and increase the entropy in the flow, thereby increasing the flow of temperature and the reduction of thermodynamic efficiency. From a close examination of the embodiments of the present invention, it will be apparent that, in the embodiments illustrated in Figures 10 1 2, the largest diameter of the female rotor is approximately equal to the smaller diameter of the male rotor, wherein in the embodiments illustrated in Figures 1 to 7, the larger diameter of the female rotor is not equal to the smaller diameter of the male rotor. Examining the procedure for designing all these rotor modalities, as stated above with reference to the illustrations in Figures 14 to 16 and the flow diagram in Figure 17, it will be appreciated that all embodiments are only different rotor families, designed in accordance with the present invention. Therefore, if these diameters are the same or different, it may be more important based on the application in which the screw rotor system 10 will be used, rather than any mere design choice. This selection could be important for particular applications because when the larger diameter of the female rotor seals with the smaller diameter of the male rotor (||), the rotors may be close together causing a friction between them, and the friction of rotation (the same diameters) is smaller than the sliding friction (different diameters).

By reducing the friction in the screw rotor system, the static state temperature of the rotors and the flow moving through the rotors can be kept lower than when there is a greater sliding friction friction between the rotors. This could be important in a refrigeration application or some other cooling application, in which the air or other working fluid is run through one or more screw rotors to cool the working fluid. In Figure 18, there is illustrated an application example that cools the working fluid, in which, a screw rotor device 10 operates as a compressor 154 for incoming working fluid and the screw rotor device 10 operates as an extender 156. After leaving the compressor output port, the working fluid is preferably passed through a fluid conduit 158 to an intermediate cooler 160 or other type of thermodynamic processor, such as a heat exchanger , and then, the working fluid enters the extender through its input section. The working fluid can also be circulated selectively by a control valve 162 through a recirculation path 164. Additionally, the compressor and expander can be mechanically linked through a drive shaft 166, which could also include gears. . Said mechanical link between the devices 10, 10, could reduce the static energy requirement of the compressor by more than 50%. In particular, the work that is extracted from the extender can be passed back to the compressor through the mechanical link. Therefore, with an extender that operates at or above a thermodynamic efficiency of 85%, most of the expansion energy is available to help activate the compressor. It will be appreciated that when the compressor and the extender are linked together in this manner, it is possible for the units to be integrated in a single housing 12. Of course, it will also be appreciated that the multiple stages of the compressors and / or extenders can be used for certain super cooling work fluids. The screw rotor system can also be used in many other applications. For example, screw rotors can be used in many types of hydrostatic power systems 168 and hydrodynamic power systems 170. A hydrostatic power system is proposed with reference to Figure 19, followed by a coupling of the system modalities hydrodynamic energy, which are raised with reference to Figures 20 and 21. Hydrostatic drive transmission systems are generally known to independently energize vehicle wheels 172 around an axle 174, offering infinitely variable speed control, a transition smooth from forward to reverse, precise steering control and hydrostatic braking. In some applications, the hydrostatic drive can also function as the primary braking system. Generally, hydrostatic drive systems are closed loop systems which receive their power supply from a source of pressurized fluid 176. In the present embodiment, the screw rotor system 10 according to the present invention, could be used to the hydrostatic drive motors 178, as well as the motor 180 that creates the source of pressurized fluid. In comparison with the hydrostatic impeller, the hydrodynamic impeller converts as much energy into the compressed working fluid as possible and then disperses the working fluid used. A couple of examples illustrated in a general manner by Figure 20, show how pressurized water 182 can be used as the working fluid. It will be appreciated that this pressurized water may come from a municipal water supply 184, through a piping system or it can be pumped directly from a well 186 or it can be stored in a local warehouse with the machine being turned on. In the hydrodynamic application, the water energizes the screw rotor system 10, which is linked through a drive shaft 166, which can include gears, to the work device 188. As the water passes through of the screw rotor device, the rotor extracts the energy and pours the water under low pressure out of the housing. A control valve 162 will probably be required for many applications, such as those applications that run intermittently, while possibly only a safety shut-off valve is used for a continuous operation system. A particular use that is within the scope of the present invention is the use of blades and other tools as the working device. In the case where the blades are for a garbage disposal or garbage disposal (drainage / housing of blades 190 shown), high pressure water (working fluid) energizes the crusher and water at low pressure (fluid used) it is dissipated in the drain or other receptacle where the garbage is being crushed and / or chopped. For a kitchen dishwasher application, high pressure water, preferably comes from the standard cold water supply of the dishwasher, and it will be appreciated that low pressure water is dispersed in the drain could be useful to wash the garbage towards the drain while the water at high pressure that is used to energize the crusher / chopper. Similarly, for a lawn mower (blade housing 192 shown) the high pressure water that (working fluid) energizes the handles and the water at low pressure (fluid used) is dispersed in the portion of the lawn that has just be cut. For the hydrodynamic lawn mower, high pressure water, preferably comes from a standard external wrench, although to energize larger blinding machines, a tank tank could be used to extract the water and the screw rotor compressor could be used to create the source of pressurized water. Once the water pressure is converted to energy for the blades, water can be pumped into the lawn. Another dynamic application is the use of screw rotor devices on a lathe 194 or other machining equipment. In this case, the working fluid is pressurized air. Therefore, in order to extract the energy from the air and thus energize the tool, the air expands inside the screw rotor system 10. As the air expands, its temperature decreases. Therefore, during the spring and summer months, the cooler expanded air can be used to cool the machining facilities, and during the autumn and winter months, the coldest air can be discharged through a valve to the outside. In the last application particularly raised for the present invention, a gas turbine engine includes linked rotor compressors 154-166-154, a burner section 196, an extender 156 and a nozzle 198. The linked rotor compressors are multiple stages of the compressors 10, 10, which are used to super compress the air before it is burned and subsequently expanded. In view of the above, it should be noted that the various advantages of the present invention are achieved and achieved. The modalities were chosen and described in order to better explain the principles of the present invention and its practical application in order to allow other experts in the field to better utilize the present invention in various modalities and with various modalities as are suitable for the particular use contemplated. Since various modifications could be made in the constructions and methods described herein and illustrated without departing from the scope of the present inventionIt is the intention that all matters contained in the foregoing description or shown in the accompanying drawings be construed as illustrative rather than limiting. For example, although the preferred embodiments of the present invention describe rotors having substantially parallel axes, the axes do not necessarily have to be parallel. Additionally, the method for designing the profiles of the screw rotor according to the present invention is not limited to any particular coordinate system. For example, a Cartesian coordinate system, that is, rectangular (x, y, z) or an angular coordinate system, that is, cylindrical (r, F, x) could be used to define the profiles. Other coordinate systems could also be used, such as a polar coordinate system, although it will be appreciated that some coordinate systems can unnecessarily add complexity to the design procedure. Additionally, the various applications raised herein are illustrative of the wide range of applications where the present invention may be useful. In particular, it will be appreciated that for the application of the internal combustion engine of the screw rotor system 10, a fuel inlet 108 could be used to deliver the fuel in one of the spaces 39, 41. It will also be appreciated that, for this embodiment, the flow could probably be moving in the opposite direction from that shown in Figure 9, and that the zero orifice fluid measurement application of the screw rotor system 10 may not have said fuel inlet, although said port or inlet It could be useful for a pressure gauge and / or a temperature gauge to measure the operating state of the device. Additionally, as illustrated in Figure 12C, the inertial energy of the rotors can be changed by providing the cuts 146 along the length of the rotors, and one or more of these cuts could also be used as the trajectories 148 for that a fluid that does not work flows through the rotors and the screw rotor system. Of course, it will also be appreciated that stages of compressors and / or multiple extenders can be used to supercool certain working fluids. Finally, in addition to the "stacking" screw rotors, that is, they mechanically link the motors of the screw rotor devices, the thread and slot can have a variable angle along the axial length of the rotors. and the rotors can be tapered. Accordingly, the spirit and scope of the present invention should not be limited by any of the exemplary embodiments described above, although it should be defined solely in accordance with the following claims appended thereto and their equivalents.

Claims (33)

NOVELTY OF THE INVENTION CLAIMS

1. - A screw rotor device for the positive displacement of a working fluid, comprising: a housing comprising a first end, a second end, an exit port and a pair of cylindrical augers extending between said first end and said second end, said pair of cylindrical holes comprising a front cusp extending along a length of a front side of said housing and a rear cusp extending along a length of a rear side of said housing; a female rotor comprising a larger diameter and a helical groove retracting from said main diameter, said helical groove comprises a bottom base located between a main side and a retrospective side, said female rotor being rotatably mounted in said housing with said larger diameter in close tolerance with one of said for cylindrical drills to form a first peripheral sealing area; and a male rotor comprising a smaller diameter and a helical thread extending from said smaller diameter, said male rotor being rotatably mounted in said housing, said helical thread comprises an upper base placed between a main face and a retrospective face, a cross-sectional profile of said helical groove comprising respectively an upper base line located between a main line and a retrospective line, said main line and said retrospective line being selected from a group consisting of a concave line, a straight line, a convex line, and any combination thereof, said upper base being in close tolerance with another of said pair of cylindrical holes to form a second peripheral sealing area, said helical thread is intertwined in phase and in close tolerance with said helical groove to form an interlocking sealing area, said area is The interlacing portion extends continuously from a posterior region at a posterior intersection between said main face, said main side and said posterior cusp to a frontal region at a frontal intersection between said retrospective face, said retrospective side and said frontal cusp.

2. The screw rotor device according to claim 1, further characterized in that said rear intersection does not include a rear vent hole between said helical thread, said helical groove and said rear cusp and wherein said front intersection does not include a front vent hole between said helical thread, said helical groove and said front cusp.

3. The screw rotor device according to claim 1, further characterized in that said interlocking sealing area additionally comprises a central sealing area between said upper base and said base of the bottom.

4. - The screw rotor device according to claim 1, further characterized in that said main face and said retrospective face additionally comprise a main edge and a retrospective edge, respectively, and wherein said interlocking sealing area additionally comprises an area of main seal between said leading edge and said main side and a retrospective sealing area between said retrospective edge and said retrospective side.

5. The screw rotor device according to claim, further characterized in that said interlocking sealing area additionally comprises a main sealing area, a retrospective sealing area and a central sealing area connecting said sealing area main, wherein said main face and said retrospective face, wherein said main face and said retrospective face additionally comprise a leading edge and a retrospective edge, respectively, wherein said main side and said retrospective side additionally comprise a main channel and a retrospective channel, respectively, wherein said central sealing area is formed between said upper base and said base of the bottom, wherein said main sealing area is formed between said main side and said main face in a first region defined by said main edge intersecting said main side in a second region defined by said main channel intersecting said main face, and in a third region that is extends between said first region and said second region, and wherein said retrospective sealing area is formed between said retrospective face and said retrospective side.

6. The screw rotor device according to claim 1, further characterized in that said female rotor and said male rotor further comprise a plurality of grooves and threads, respectively, said plurality of grooves and threads being identical in number and interlacing in phase to each other, wherein a cross-sectional profile of said male rotor comprises a tooth, an adjacent tooth and an edentulous sector between said tooth and said adjacent tooth, said tooth being subtended by a first arc angle and said edentulous sector comprising a second arc angle proportional to said first arc angle by means of a phase compensation multiplier.

7 .- The screw rotor device according to claim 1, further characterized in that said female rotor and said male rotor each comprise a rotation axis centrally located within one of said pair of cylindrical bores, where said larger diameter of said female rotor rotates in close tolerance to said smaller diameter of said male rotor to form a central non-interlaced sealing area therebetween, and wherein the positive displacement of the working fluid between said inlet port and said port of The output of said housing is produced by said female rotor and said male rotor with a thermodynamic efficiency of at least 85%.

8. - The screw rotor device according to claim 1, further characterized in that said close tolerance between said thread that is interlocked with said slot is within the same order of magnitude as the at least one of said close tolerance between said rotor female and said housing and said close tolerance between said male rotor and said housing.

9. The screw rotor device according to claim 1, further characterized in that said close tolerance between said thread that is interlocked with said slot is less than or approximately equal to the at least one of said close tolerance between said rotor female and said housing and said close tolerance between said male rotor and said housing.

10. The screw rotor device according to claim 1, further characterized in that it additionally comprises a recirculation path for the working fluid from said outlet port to said inlet port and external to the positive displacement of the working fluid between said entry port and said exit port within said accommodation.

11. A screw rotor system for positive displacement of a working fluid, comprising: a housing comprising a first end, a second end, an inlet port, an outlet port and a pair of cylindrical bores that are extend between said first end and said second end, said pair of cylindrical augers comprise a first cusp extending along a length of a first side of said housing and a second cusp extending along a length of one second side of said housing; a female rotor comprising a larger diameter and a helical groove that recedes from said larger diameter, said helical groove comprises a bottom base located between a main side and a retrospective side, said main side and said retrospective side comprise a main channel and a retrospective channel in said main diameter, said main diameter being in a first sealing relationship with one of said pair of cylindrical augers; and a male rotor comprising a smaller diameter and a helical thread extending from said smaller diameter and intertwining in rotatable in phase with said helical groove within said housing, said smaller diameter having a second sealing relationship with said diameter of said female rotor, said helical thread comprising an upper base in a third sealing relationship with another of said pair of cylindrical augers, wherein said upper base is also in a fourth sealing relationship with said bottom base, wherein said base The upper base is located between a main face and a retrospective face, said retrospective face being in a fifth sealing relationship with said retrospective channel, said retrospective face additionally comprises a retrospective edge in a sixth sealing relationship with said retrospective side, wherein said retrospective channel and said retrospective edge are in a seventh relation n sealing each other and with said first height of said pair of cylindrical holes.

12. The screw rotor device according to claim 11, further characterized in that said main face of said helical thread additionally comprises a main edge, said main face being in an eighth sealing relation with said main channel, said leading edge being in a ninth sealing relationship with said main side, and wherein said main channel and said leading edge are in a tenth sealing relationship with each other and with said second cusp of said pair of cylindrical augers.

13. The screw rotor device according to claim 12, further characterized in that said main edge and said retrospective edge of said helical thread respectively define said main side and said retrospective side of said helical groove as said helical thread that it intertwines with said helical groove, and wherein said main channel and said retrospective channel of said helical groove respectively define a main root portion on said main face and a retrospective root portion on said retrospective face of said helical thread.

14. The screw rotor device according to claim 11, further characterized in that said sealing ratios each comprise a sealing tolerance defined by a geometric proximity between at least one of said female rotor and said male rotor, said female rotor and said housing, and said male rotor and said housing, and wherein said helical thread and said helical groove delimit a space within said cylindrical bores, seal the working fluid inside said housing, and the transition between the bond each other and sealing around said housing while maintaining said sealing of working fluid in said space.

15. The screw rotor device according to claim 14, further characterized in that said upper base of said helical thread separates said main face from said retrospective face by a minimum upper base distance and wherein said minimum upper base distance it is at least an order of magnitude greater than said sealing tolerance between said helical thread and said helical groove.

16. The screw rotor device according to claim 14, further characterized in that said sealing tolerance is not greater than at least one of an order of magnitude greater than said geometric proximity, 0.0072 centimeters and 1/1, 000 of said male rotor diameter.

17. The screw rotor device according to claim 11, further characterized in that said helical thread is intertwined with said helical groove in a double sealing relationship where a leakage path between said male rotor and said rotor is not provided. female from said first side to said second side of said housing, wherein said leakage path is a current tube with an effective diameter greater than said minimum upper base distance.

18. The screw rotor device according to claim 11, further characterized in that said first sealing relationship comprises a first peripheral sealing area defined by the geometries of said larger diameter and said cylindrical bores, wherein said second ratio of The sealing comprises a central non-interlaced sealing area defined by the geometries of said larger diameter and said smaller diameter, wherein said third sealing relationship comprises a second peripheral sealing area defined by the geometries of said upper base and said cylindrical bores, in wherein said fourth sealing relationship comprises a central interlaced sealing area defined by the geometries of said upper base and said bottom base, wherein said fifth sealing relationship comprises an exterior sealing line defined by the geometries of said retrospective face and said retrospective channel, where said sixth re The sealing portion comprises an inner sealing line defined by the geometries of said retrospective edge and said retrospective side, wherein said outer sealing line and said inner sealing line define boundaries of a first interlocked sealing area between said retrospective face and said sealing region. retrospective side and intersect at a common sealing point according to said seventh sealing relationship defined by the intersection of the retrospective edge, the retrospective channel and the first cusp.

19. - The screw rotor device according to claim 11, further characterized in that said helical thread is intertwined with said helical groove in close proximity to said first cusp in a triple sealing relationship, wherein a ventilation hole is not provided between said helical thread, said helical groove and said first cusp and wherein said helical thread is intertwined with said helical groove in close proximity to said second cusp in a second triple sealing relationship wherein a vent hole is not provided between said helical thread , said helical groove and said second cusp.

20. The screw rotor device according to claim 11, further characterized in that said female rotor and said male rotor further comprise a plurality of grooves and threads, said plurality of grooves and threads are identical in number and interleaved in phase each other, wherein a cross-sectional profile of said male rotor comprises a tooth, an adjacent tooth and an edentulous sector between said tooth and said adjacent tooth, said tooth being subtended by a first arc angle and said edentulous sector comprising a tooth. second arc angle proportional to said first width angle by a phase compensation multiplier, wherein said first sealing area and said third sealing area extend from said first side of said housing to said second side of said housing, and wherein the positive displacement of the working fluid between said inlet port and said outlet port This housing is produced by said female rotor and said male rotor with a thermodynamic efficiency of at least 85%.

21. A screw rotor product for positive displacement of a working fluid, comprising: a housing comprising an inlet port, an outlet port and a pair of cylindrical augers, said pair of cylindrical bores comprise a front cusp which extends along a length of a front side of said housing and a rear cusp extending along a length of a rear side of said housing: a female rotor comprising at least one retracting helical groove from a channel in a larger diameter to a bottom base, wherein said female rotor is rotatably mounted within said housing; and a male rotor comprising at least one helical thread extending from a root in a smaller diameter to an upper base, wherein said male rotor is rotatably mounted within said housing and rotates in the opposite direction with respect to said female rotor, wherein said helical thread is intertwined in phase with said helical groove and defines an interlocking sealing area extending continuously between said helical thread and said helical groove from said posterior cusp to said front cusp, and wherein at least one of a front vent hole and a rear vent hole is not provided between said helical thread, said helical slot, said front cusp and said rear cusp, and wherein said larger diameter of said female rotor and said smaller diameter said male rotor define a non-intersecting sealing area extending between said male rotor and said rotor hem from a seal on said channel and said root.

22. The screw rotor product according to claim 21, further characterized in that said helical thread and said helical groove each comprise multiple angles in said length of said housing and wherein said non-interlacing sealing area extends. between the adjacent interlocked sealing areas, from said channel and said root to an adjacent channel and an adjacent root.

23. The screw rotor product according to claim 21, further characterized in that it additionally comprises: a source of pressurized driving fluid; a fluid conduit in fluid communication between said source of impeller fluid and said inlet; a drive shaft in mechanical communication with at least one of said male rotor and said female rotor; and at least one of a tool holder, a blade and a wheel operatively connected to said drive shaft.

24. The screw rotor product according to claim 23, further characterized in that said source of pressurized driving fluid is selected from the group of a source of compressed air and a source of pressurized water.

25. The screw rotor product according to claim 24, further characterized in that said drive shaft is connected to at least one of said blade of a hydrodynamic garbage disposal, said blade of a hydraulic hydrodynamic lawn mower, and said tool holder of a lathe.

26. The screw rotor product according to claim 21, further characterized in that it additionally comprises at least one of a drive shaft, an energy input shaft, a fluid conduit, a source of compressed air, a nozzle , a valve, a fuel inlet, a wheel and a thermodynamic processor, wherein said female rotor and said rotor are arranged in at least one of a fluid communication and a mechanical communication with at least one of said axis impeller, said energy input shaft, said fluid conduit, said source of compressed air, said nozzle, said valve, said fuel inlet and said thermodynamic processor in a positive displacement machine configuration selected from the group consisting of a compressor , an extender, a motor, a pump, a hydrostatic drive, a hydraulic motor, a positive-drive motor, a hydraulic pump, a motor internal combustion, an axial flow injection machine and any equivalent rotary piston application, and wherein a cross-sectional profile of said male rotor further comprises a tooth, an adjacent tooth and an edentulous sector therebetween, said tooth being subtended by a first arc angle and said edentulous sector comprises a second arc angle which is at least twice said first arc angle.

27. The screw rotor product according to claim 21, further characterized in that said female rotor and said male rotor additionally comprise a plurality of grooves and threads, said plurality of grooves and threads are identical in number and are intertwined in phase with each other in a plurality of thread-groove pairs, wherein a cross-sectional profile of said male rotor comprises a tooth, an adjacent tooth and an edentulous sector between said tooth and said adjacent tooth, said tooth being subtended by a first arc angle and said edentulous sector comprises a second arc angle proportional to said first arc angle by means of a phase compensation multiplier.

28. The screw rotor product according to claim 27, further characterized in that said thread-slot pairs delimit a plurality of spaces that do not communicate inside said cylindrical bores, seal the working fluid inside said housing, and they make the transition between the interlacing one another and the seal around said housing while maintaining said sealing of the working fluid in said spaces that do not communicate.

29. The screw rotor product according to claim 27, further characterized in that said helical threads and said helical grooves are comprised of helical threads of variable angle and helical grooves of variable angle, respectively, wherein said angle varies axially with said length of said housing.

30. - The product of roíor de lomillo according to claim 27, further characterized in that the positive displacement of the working fluid between said inlet port and said outlet port of said housing is produced by said female rotor and said male rotor with an efficiency thermodynamics of at least 85%.

31.- A procedure for designing a screw rotor device, comprising the steps of: (a) defining a pair of intersecting major circles, said for intersecting larger circles respectively comprising a pair of centers and a pair of larger diameters, wherein each of said pair of intersecting major circles only encompass a respective one of said pair of centers, and wherein said pair of centers is separated less than a sum of one half of said pair of larger diameters; (b) defining a circumferential upper base of a tooth located in one of said pair of intersecting major circles; (c) defining a pair of sides that are radially withdrawn from the other of said pair of larger circles to a base of the bottom, wherein said sides are defined by a first path of said upper base when said pair of larger intersecting circles rotate in phase with each other by equal angular quantities, wherein said sides comprise a pair of intermediate line segments that are removed from a pair of circumferential channels to said base of the bottom; (d) identifying a pair of root sections, wherein said pair of root sections are respectively defined by a second path of said pair of circumferential channels when said pair of major intersecting circles rotate in phase with each other by equal angular amounts; (e) identifying a pair of line segments extending over said tooth, wherein said extending line segments are defined by a third path of said pair of intermediate line segments when said pair of larger intersecting circles rotate in phase with each other by equal angular quantities; (f) selecting at least one of said male rotor circle and said female rotor circle; and (g) defining a profile in two-dimensional coordinates relative to said selected rotor circle, said profile comprising a pair of points in said selected rotor circle, a first line segment radially inwardly offset from said rotor circle selected to a diameter less, a second line segment connecting one of said pair of points to said first line segment, and a third line segment connecting another of said pair of points to said first line segment, wherein said pair of points is located in said selected rotor greater diameter and comprises a first arc angle, wherein said first line segment comprises a pair of end points, said pair of end points comprises a second arc angle smaller than said first arc angle, wherein said second line segment connects one of said pair of points to one of said pair of end points, and wherein said third line segment connects another of said pair of points to another of said pair of end points.

32. - The method according to claim 31, further characterized in that steps (a) - (e) further comprise the steps of: defining a first larger rotor circle, said first larger rotor circle comprising a first larger diameter; defining a second larger rotor circle that is interfered with said first larger rotor circle at a pair of intersection points, said second larger rotor circle comprises a second larger diameter and wherein less than one half of said second larger diameter extends in said first larger circle of the rotor and wherein less than one half of said first greater diameter extends in said second larger rotor circle; selecting a first point in said first major rotor circle; identifying a first line segment that is withdrawn radially inwardly from a second major rotor point to a second minor rotor point, wherein said first line segment is defined by a first path of said first point as it progresses from said second larger rotor circle to said second smaller rotor circle when said first larger rotor circle and said second larger rotor circle rotate in phase with each other by equal angular quantities; selecting a second point in said first larger rotor circle, said second point being circumferentially spaced from said first point; identifying a second line segment that retracts radially inward from a second major rotor point circumferentially spaced apart to a second minor point of circumferentially separated rotor, wherein said second line segment is defined by a second path through said second point as it progresses from said second larger rotor circle to said second smaller rotor circle when said first larger rotor circle and said second larger rotor circle rotate in phase with each other by equal angular quantities, and wherein said second larger separated rotor point. circumferentially and second minor rotor point are circumferentially spaced from said second major rotor point and said second minor rotor point, respectively; identifying a pair of first rotor root line segments extending from said first smaller rotor circle to a pair of intermediate points, wherein one of said pair of intermediate points is located between said first smaller rotor circle and said first point in said first major circle of rotor and another of said pair of intermediate points is located between said first smaller circle of rotor and said second point in said first major circle of rotor, said pair of intermediate points being circumferentially separated from each other, and in wherein said first rotor root line segments are defined by a pair of trajectories of said second major rotor point and said second larger rotor point circumferentially separated when said first larger rotor circle and said second larger rotor circle rotate in phase each other by equal angular quantities; and identifying a pair of circumferentially separated first rotor line segments extending respectively between said pair of first rotor root line segments and said first point and said second point in said first larger rotor circle.

33. The method according to claim 31, further characterized in that steps (a) - (e) further comprise the steps of: defining a first sealing relationship comprising a center, interlocking sealing areas in accordance with the geometries of a top base and a bottom base, defining a second sealing relationship comprising a front outer sealing line in accordance with the geometries of a retrospective face and a retrospective channel, defining a third sealing relationship comprising a line of interior front sealing according to the geometries of a retrospective edge and a retrospective side, defining a fourth sealing relationship comprising a subsequent, exterior sealing line according to the geometries of said main face and said main channel, defining a fifth relation of sealing comprising a subsequent sealing line, interior according to the geometries of said edge p rincipal and said main side; define a sixth sealing relationship by means of an intersection of the retrospective edge, the retrospective channel and the front cusp, where a front, outer sealing line and a front, inner sealing line define the boundaries of a sealing area that is intertwined, frontal between said retrospective face and said retrospective side and intersect at a common frontal sealing point; define a seventh sealing relationship according to the intersection of the leading edge, main channel and posterior cusp; wherein said posterior, exterior sealing line and said inner, posterior sealing line define the boundaries of an interlocking, posterior sealing area between said main face and said main side and intersect at a common posterior sealing point; defining an eighth sealing relationship comprising a first peripheral sealing area in accordance with the geometries of the larger diameter of the female rotor and said cylindrical bores, defining a ninth sealing relationship comprising a second peripheral sealing area in accordance with the geometries of said upper base and said cylindrical boreholes, defining a tenth sealing ratio comprising a non-interlaced sealing area, central according to the geometries of said larger diameter of female rotor and said smaller diameter of male rotor. 34.- The method according to claim 31, further characterized in that steps (a) - (g) further comprise the steps of: selecting at least one helix angle; and defining at least one of a retrospective thread face and a retrospective slot side, respectively in accordance with at least one of said defined slot profile and said defined tooth profile, wherein said retrospective face and said retrospective side produce an interlocking sealing area extending from one of said pair of intersection points to another of said pair of intersection points. The method according to claim 31, further characterized in that it further comprises the steps of selecting at least one helix angle and defining an eighth figure in close tolerance with said intersecting major circles. 36.- The method according to claim 31, further characterized in that it further comprises the step of repeating steps (a) - (e) to create at least one of a plurality of threads, another identical screw rotor device and a family of screw rotors. 37.- The method according to claim 31, further characterized in that it additionally comprises the step of copying a profile resulting from steps (a) - (e) and creating at least one of a plurality of threads, another device of identical screw rotor and a family of variable sizes of screw rotors.