Patent Application
20050129561
Not Applicable.
1. Field of the Invention
This invention relates generally to rotor devices and, more particularly to screw rotors.
2. Description of Related Art
Screw rotors are generally known to be used in compressors, expanders, and pumps. For each of these applications, a pair of screw rotors have helical threads and grooves that intermesh with each other in a housing. For an expander, a pressurized gaseous working fluid enters the rotors, expands into the volume as work is taken out from at least one of the rotors, and is discharged at a lower pressure. For a compressor, work is put into at least one of the rotors to compress the gaseous working fluid. Similarly, for a pump, work is put into at least one of the rotors to pump the liquid. The working fluid, either gas or liquid, enters through an inlet in the housing, is positively displaced within the housing as the rotors counter-rotate, and exits through an outlet in the housing.
The rotor profiles define sealing surfaces between the rotors themselves between the rotors and the housing, thereby 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 given to the interface between the rotors where the threads and grooves of one rotor respectively intermesh with the grooves and threads of the other rotor. The meshing interface between rotors must be designed such that the threads do not lock-up in the grooves, and this has typically resulted in profile designs similar to gears, having radially widening grooves and tightly spaced involute threads around the circumference of the rotors. However, an involute for a gear tooth is primarily designed for strength and to prevent lock-up as teeth mesh with each other and are not necessarily optimum 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 one type of parameter to evaluate the thermodynamic efficiency of a screw rotor system. Adiabatic efficiency is the ratio of the adiabatic horsepower required to compress a given amount of gas to the actual horsepower expended in the compressor cylinder. Volumetric efficiency is the ratio of the actual volume of working fluid flowing through the screw rotor, such as in one complete revolution, to the geometric volume of the screw rotor measured, which is also measured for one complete revolution. Mechanical efficiencies can include the efficiencies of any gear train that may be used to keep the rotors in proper phase with each other, bearings, and seals.
Although adiabatic efficiency and volumetric efficiency are different performance parameters, a number of screw rotor features can affect both of these efficiencies. For example, tightening tolerances between the rotors and the housing can improve both the volumetric efficiency and the adiabatic efficiency of a given rotor design. However, if tolerances are too tight for a given design, the volumetric efficiency may be improved while the adiabatic efficiency drops. Such performance characteristic could be caused by thermal expansion of the rotors, machining tolerances, and even the material properties of the rotors, which can result in intermittent contact between the rotors and the sides of the housing or between the rotors themselves.
Generally, one of the best ways to improve thermodynamic efficiencies is by keeping tight tolerances and minimizing leak pathways between the rotors and the housing and between the rotors themselves. However, in prior art screw rotors, leak pathways are inherent in the actual design of the rotors, i.e., the leaks can be reduced but not eliminated. Such inherent leaks would occur even when the tolerances are perfected, i.e., zero thermal expansion, perfect machining tolerances, and a perfectly smooth finished material. These leak pathways result in losses that adversely affect both the thermodynamic efficiency and the volumetric efficiency of screw rotors.
Accordingly, leak pathways are some of the most important losses to consider for the performance of screw rotors when the screw rotors are being designed because these losses negatively affect both thermodynamic efficiency and volumetric efficiency. Even with this knowledge that leak pathways should be minimized, the design methodology used for screw rotors produces these pathways as an inherent aspect of traditional screw rotor profiles. In fact, it is a common belief by 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., at its web site www.matteicomp.com/About/ScrewCompressors/, this belief is concisely stated even as this application is being filed in March 2004: “The technical problem is typical of the geometry of screw compressors. All screw manufacturers have tried to reduce the effect of the ‘blow hole’ by analyzing and adapting new rotor profiles to create smaller openings at the critical point, but its complete elimination is impossible.” Accordingly, to minimize the leak pathways, it is common knowledge that the rotors should seal perfectly along the contact line, but a number of prior art references also teach that the contact line should be as short as possible, i.e., should not extend to cusps on opposite sides of the housing. Several embodiments of short contact lines are set forth in the applicant's patent application Ser. Nos. 10/283,421 (Pub. No. 2003/0077198) and U.S. application Ser. No. 10/283,422. However, there remains a need for better methodologies for designing screw rotor profiles that account for machining constraints, thermal expansion and material tolerances, as well as mechanical efficiencies, and that also eliminate any inherent leak pathway from the design process, even though it is presently considered impossible. One example of a machining constraint set forth in the prior art is the need for blunt edges because of the concern that sharp edges have a tendency to break, e.g., U.S. Pat. No. 2,486,770.
Once the leak pathway problem is eliminated from the design methodology, i.e., screw rotor profiles that do inherently produce a leak pathway, the designer can balance all of the rotors' performance characteristics. For example, a rotor design without any inherent leak pathway may be slightly changed to include a small gap or leak pathway to permit another aspect to improve the rotors' overall performance at a given design point, i.e., tighter tolerances at steady state operation with thermal expansion. In comparison, when the leak pathway remains an inherent feature of the rotor profiles, the designer must first minimize the leak pathway using more complex designs that are harder and costlier to manufacture and then changes to the design are limited by the complexity of the design, machining and other manufacturing capabilities and thermal expansion requirements. Therefore, a new design methodology that produces screw rotor profile shapes without any leak pathways is needed. Additionally, it would also be advantageous if sharp-edged shapes that eliminate leak pathways and do not have a tendency to break could be designed and manufactured.
Leak pathways are generally cause by internal leakage between the rotors and the housing and between the rotors themselves and result in volumetric losses and thermodynamic losses due to recirculation of the working fluid within the rotors. For example, working fluid that is pressurized and leaks into a lower pressure region of the rotors is caused to expand to the lower pressure state with a higher temperature due to entropy and then must recirculate through the rotors before being expelled. Therefore, the overall temperature of entire rotor system, including the rotors and the working fluid, is increased due to the gain in entropy. Internal leakage is detected specifically at the following points:
As discussed above, threads must provide seals between the rotors and the walls of the housing and between the rotors themselves, and in all designs before the present invention, there has been a transition from sealing around the circumference of the housing to sealing between the rotors. In this transition, a gap is formed between the meshing threads and the housing, causing leaks of the working fluid through the gap in the sealing surfaces and resulting in less efficiency in the rotor system. A number of arcuate profile designs improve the seal between rotors and may reduce the gap in this transition region but these profiles still retain the characteristic gear profile with tightly spaced teeth around the circumference, resulting in a number of gaps in the transition region that are respectively produced by each of the threads. Some pumps minimize the number of threads and grooves and may only have a single acme thread for each of the rotors, but these threads have a wide profile around the circumferences of the rotors and generally result in larger gaps in the transition region.
Until now, screw rotor expanders, compressors and pumps have had similar fundamental flaws. Generally, they allow for leak pathways between the working side, i.e., expansion, compression or pumping, to the side that should be sealed from the working side for proper operation of the rotors, i.e., non-working. 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. DE 4121 and DE 7116 from more than a century, have fallen out of favor, and this may possibly be due to the rise of the Lysholm-type rotors in the 1930's and 1940's. In an article entitled “A New Rotary Compressor” and written by Lysholm in the 1940's, Lysholm puts down the Krigar design as being unable to obtain any compression between the lobes with a two-thread/two-groove design (2×2 configuration). While it is clear from the images of the Krigar design that there definitely were sealing issues, especially between the threads and the grooves, and Krigar appears to be more directed to radial flow, the Lysholm conclusion that the Krigar design could not perform any compression with only the 2×2 configuration is flawed. Regardless, the industry and 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 attempted to seal the male rotor with the female rotor and the housing, but the prior art designs have either a leak pathway between the rotors themselves or a leak pathway between the rotors and the housing, i.e., which according to the prior art quoted above, 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 a mathematical formula, i.e., empirical design methodology, while other designs are based on particular curves or a combination of piecewise curves, i.e., formula design methodology, such as lines, arcs, circles, squares, trapezoids, involutes, inverse-involutes, parabolas, hyperbolas, cycloids, trochoids, epicycloids, epitrochoids, hypocycloids, hypotrochoids, as well as other straight and arcuate lines, and still other designs combine formula and empirical design methodologies. However, regardless of the design methodology, empirical or formula or a combination thereof, prior designs and respective methods for creating rotor profiles either explicitly teach or implicitly suggest and disclose creating the profile for the thread and corresponding groove using the shortest seal path between the rotors, i.e. the sealing region does not extend from the front cusp all the way to the back cusp. Additionally, many of the prior art methods are based on and remain similar to traditional gear design methods.
Some earlier designs have come close to a complete seal or may even be able to effect a complete seal in one pitch, see in particular co-pending U.S. application Ser. No. 10/283,422. Even for these single-pitch sealing rotors, some of the seals may only be along sealing lines, rather than sealing areas. Additionally, since the rotor profiles are designed according to the traditional gear profile design methods, these rotors are usually limited in the types of arcuate lines that can be used to effect the seal. Without accounting for the third dimension, the arcuate lines have typically been limited to epitrochoids, epicycloids, hypocycloids and other types of spirals, such as an Archimedean spiral.
When the third dimension is accounted for in prior art design methodologies, it is typically limited to standard helix angle definitions that have been developed for ordinary screws, i.e., fastening screws. Such an approach fails to truly account for and does not take advantage of the third dimension. It is well known that for any screw rotor, the helix angle of the grooves and threads vary depending on their depth. In particular, the top land of the thread has a lesser helix angle than the root of the thread, and the trough of the groove has a greater helix angle than the ridge of the groove. Accordingly, merely using a single helix angle for a rotor, such as the top land, the root, or any other single angle, even with a correction factor, has not accounted for the variations in the helix angles of the thread and the groove. In this way, the known screw rotor geometries are created using planar design methodologies for the rotor profiles rather than using a volumetric design methodology.
The planar design methodologies fail to apply the function of the helix angle with respect to the radius, resulting in the profiles with leak pathways discussed above. In one aspect, the planar design methods are unnecessarily restrictive because they only take advantage of two-dimensional space to overcome the limitation that the threads must not lock-up in the grooves. In another aspect, the planar design methods are not restrictive enough because when the profiles are expanded into three-dimensional space, the profiles have three-dimensional leak pathways. The extra degree of freedom provided by the third-dimension allows for a volumetric design that prevents lock-up while permitting perfect sealing between the male rotor and female rotor and between the rotors and the housing, a perfect seal which is equivalent to the complete seal of pistons. More generally, similar fundamental flaws in the prior art designs and their respective methodologies can be traced back to their failure to accommodate for and use the additional degree of design freedom provided by the third dimension. It is the additional degree of design freedom of volumetric design methodologies that permits an unlimited number of profile designs which effect a complete seal without locking up the rotors and without the unnecessary restrictions of the planar design methodologies.
For many prior art rotors, the leak pathway can be found between the face of the thread and the housing. In particular, the thread and groove are designed with significant curvatures at their top land edges and ridges according to the standard manner of designing meshing gear teeth. Such rounded edges and ridges cannot possibly seal between the rotors and the housing when the thread and groove begin meshing with each other. As the thread and groove rotate away from their seals with the housing and into their meshing positions with each other, the rounded edges produce a gap between the housing and the groove and/or the thread before the groove and thread actually mesh and reform a sealing line. The gap between the housing to groove and thread seal can be an order of magnitude greater than the tolerances for the seals between the between the rotors and the housing and the rotors themselves. In some designs, the gap can be even larger, such as in screw rotors that have a different number of threads and grooves, i.e. not the same number of threads as grooves, and the loss in pressure to the low pressure side causes the thermodynamic efficiency to drop. Therefore, the rotors must work harder to pump the same volume of air as compared with rotors according to the present invention which can maintain the same order of magnitude in the seal tolerances when each thread and respective groove begin meshing with each other as compared to the seal between the rotors and the housing and the rotors when in their fully intermeshed 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 screw rotor efficiencies to their 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 merely considered for the helix angle of the profiles. In an attempt to compensate for this unwitting failure to take advantage of the third dimension, the prior art designs have increasingly become more complex over the years without offering much improvement 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 planar design methodologies increasingly led to these more complex screw rotor designs as machining and other manufacturing methods improved over the years and permitted the increasing complexity. Additionally, these increasingly complex screw rotor profile designs, which need such improved manufacturing methods, support the conclusion that the failure to take advantage of the third dimension has been an unwitting failure because volumetric design methodologies actually permit much more simplified designs which can be less complex to manufacture than profiles created using the planar design methodologies.
Generally, the present invention provides a design methodology for generating thread and groove profiles which take advantage of the three-dimensional geometry of intermeshing rotors. In particular, the present invention has generally solved the problem of leak pathways that have plagued screw rotor designs for over one hundred years. The present invention provides a design methodology that is based on the fundamental premise that the helix angles of screw rotors vary with respect to each other as their threads join and then separate with the grooves and that to eliminate the blow-hole, the sealing region must extend completely from the housing's front cusp to its back cusp. Accordingly, each screw rotor embodiment of the present invention can eliminate at least one blow-hole gap, the front side, the back side or both, and this is is the first screw rotor device that eliminates the blow hole gap while also maintaining the seal between the thread and the groove regardless of the number of pitches.
It is an advantage of the present invention to maximize the thermodynamic efficiency and the volumetric efficiency in a screw rotor system by several means, such as reducing gaps, minimizing recirculation within the screw rotor housing, reducing shock waves within the screw rotors, reducing entropy, and reducing sliding friction between the male rotor and the female rotor. It is also an advantage of the present invention that it is readily producible. The designs can be rather simple and still maintain a good sealing relationship. Therefore, the present invention does not suffer from an overly complicated design that is difficult to machine or to otherwise manufacture. It is another advantage of the present invention that it can reduces and nearly eliminate backlash. It is yet another advantage of the present invention that it can reduce the cost of manufacturing screw rotor compressors and, due to its increased thermodynamic and volumetric efficiencies, it can also reduce the cost of ownership for screw rotor compressors. It is a further advantage that the present invention provides economy, efficiency and speed of assembly in manufacturing, and also reduces the cost of component assembly and the packaging costs of the product. It is yet a further advantage of the present invention in that the screw rotor system can be designed as a modular device that can be replaced with a cartridge-type system or completely integrated into a particular product. To the extent that various components of the screw rotor system are manufactured separately, and then shipped to an assembler for fixation of additional components &/or for further assembly into final products, the modular aspects of the present invention improve the efficiency and economy of assembly. In comparison to bladed compressors and turbines, the present invention is much stronger, more economical, and provides more compact components.
Accordingly, no earlier design follows the design methodology of the present invention which, as discussed below, can effect a compete seal regardless of the types of lines, straight or arcuate. The present invention can also effect a complete seal for multiple pitched rotors. The new design method is even so robust that it produces geometries that can even effect a complete seal multiple areas simultaneously, including 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 viewing the threads and grooves in the third dimension and making accommodations for the third dimension in the design process, there is an additional degree of design freedom which permits intermeshing screw rotors to be designed without leak pathways or other gaps between the male rotor and the female rotor and between the rotors and the housing, including the blow-hole at the transition region and discussed above. Once the design problem is viewed in the third dimension, it becomes clear that there should be a way to eliminate the blow-hole gap while maintaining the seals between the thread and groove. Accordingly, the present invention teaches that, to eliminate the blow-hole gap, the sealing region should extend completely from the housing's front cusp to its back cusp. Finally, when the design choices are again translated into planar design methodology, the creation of the designs becomes much less difficult than many of the planar design methodologies that are increasingly being suggested as the only way to increase the efficiencies.
Also disclosed herein is an example of the inventive method for designing entire families of the present invention's threads and corresponding grooves. The new thread and groove design results in a high-efficiency screw rotor system which is heretofore unknown in the prior art. The features of the invention result in an advantage of improved thermodynamic efficiency and improved volumetric efficiency of the screw rotor device. Tests on the prototype design show that the thermodynamic efficiency are likely to reach greater than 85% and may even exceed 90%. The present invention is seminal because it is the first screw rotor to achieve these efficiencies over a wide range of rotor speeds.
Further 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.
The accompanying drawings, 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 invention. In the drawings:
Referring to the accompanying drawings in which like reference numbers indicate like elements,
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 corresponding pair of helical grooves 38, 40. The female rotor 16 counter-rotates with respect to the male rotor 14 and each of the helical grooves 38, 40 respectively intermeshes in phase with each of the helical threads 34, 36. In this manner, the working fluid flows through the inlet port 18 and into the screw rotor device 10 in the spaces 39, 41 bounded 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 spaces bounding the working fluid. The spaces 39, 41 are closed off from the inlet port 18 as the helical threads 34, 36 and helical grooves 38, 40 intermesh at the inlet port 18. As the female rotor 16 and the male rotor 14 continue to counter-rotate, the working fluid is positively displaced toward the outlet port 20.
The pair of helical threads 34, 36 have a phase-offset aspect that is particularly described in reference to
Arc Angle β≧N*Arc Angle α, M≧1 (1)
As illustrated in
For balancing the male rotor 14, it is preferable to have equal radial spacing of the teeth. An even number of teeth is not necessary because an odd number of teeth could also be equally spaced around male rotor 14. Additionally, the number of teeth that can fit around male rotor 14 is not particularly limited by the preferred embodiment. Generally, arc angle β is proportionally greater than arc angle a according to the phase-offset multiplier. Accordingly, arc angle β of the toothless sector 46 can decrease proportionally to any decrease in the arc angle α of the teeth 42, 44, thereby allowing more teeth to be added to male rotor 14 while maintaining the phase-offset relationship. Whatever the number of teeth on the male rotor 14, the female rotor has a corresponding number of helical grooves. Accordingly, the helical grooves 38, 40 have a phase-offset aspect corresponding to that of the 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 on the male rotor, and the helix angle of the helical grooves 38, 40 is opposite-handed from the helix angle of the helical threads 34, 36. It will be appreciated that, for a given rotor diameter, the helix angle of the grooves and threads actually vary depending on their depth. In particular, referring back to
In one embodiment, each of the helical grooves 38, 40 has a cut-back concave profile 48 and corresponding radially narrowing axial, widths from locations between the minor diameter 50 (md) and the major diameter 52 (MD) towards the major diameter 52 at the periphery of the female rotor 16. The cut-back concave profile 48 includes line segment jk radially extending between the minor diameter 50 and the major diameter 52 on a ray from axis 28, line segment lm radially extending between the minor diameter 50 and the major diameter 52, and a minor diameter arc lj circumferentially extending between the line segments jk, lm. Line segment jk is substantially perpendicular to major diameter 52 at the periphery of the female rotor 16, and line segment lmn preferably has a radius lm combined with a straight segment mn. In particular, radius lm is between straight segment mn and minor diameter arc lj and straight segment mn intersects major diameter 52 at an acute exterior angle φ, resulting in a cut-back angle Φ defined by equation (2) below.
Cut-Back Angle Φ=Right Angle (90°)−Exterior Angle φ, (2)
The cut-back angle Φ and the substantially perpendicular angle at opposite sides of the cut-back concave profile 48 result in the radial narrowing axial width at the periphery of the female rotor 16. In this cut-back embodiment, the helical grooves 38, 40 are opposite from each other about axis 28 such that line segment jk for each of the pair of helical grooves 38, 40 is directly in-line with each other through axis 28. Accordingly, in the cut-back embodiment, 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 on a gaseous working fluid. Each of the helical threads 34, 36 may also include a distal labyrinth seal 54, and a sealant strip 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 may use a valve 58 operatively communicating with the outlet port 20. As one example, a valve 58 is a pressure timing plate 60 attached to and rotating with the male rotor 14 and is located between the male rotor 14 and the outlet port 20. As particularly illustrated in
The single-thread embodiment also illustrates another aspect of the screw rotor device 10 invention. In this embodiment, the length of the screw rotor device 10 is approximately one single pitch of the helical thread 34 and groove 38. The pitch 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
Single Pitch Helical Twist=360°/N (3)
Of course, it will be appreciated that even in the example in which the length of the screw rotor device 10 is a single pitch, the pitch length can be changed by altering the helix angle of the threads and grooves. The pitch length increases as the helix angle steepens. The screw rotor device 10 illustrated in
The screw rotor device 10 illustrated in
As particularly illustrated in
Arc Angle α>Arc Angle θ (4)
The phase-offset relationship defined for a pair of threads is also applicable to the male rotor 14 with the single thread 34, such that the toothless sector 46 must have an arc angle β that is at least twice the arc angle a of the single helical thread 34. The male rotor 14 circumference is 360°. Therefore, to design a rotor having a phase-offset multiplier of at least 2 and a single thread, arc angle β for the toothless sector 46 must at least 240° and arc angle α can be no greater than 120°. Similarly, for designing rotor having a phase-offset multiplier of at least 2 with the pair of threads 34, 36, 60° is the maximum arc angle α that could satisfy the such a minimum phase-offset multiplier of two (2) and 30° would be the maximum arc angle α that could satisfy the phase-offset multiplier of five (5). For practical purposes, it is likely that only large diameter rotors would have a phase-offset multiplier of 50 (3° maximum arc angle α) and manufacturing issues may limit higher multipliers.
The male rotor 14 and female rotor 16 each has a respective central shaft 76, 78. The shafts 76, 78 are rotatably mounted within the housing 12 through bearings 80 and seals 82. The male rotor 14 and female rotor 16 are linked to each other through a pair of counter-rotating gears 84, 86 that are respectively attached to the shafts 76, 78. The central shaft 76 of the male rotor 14 has one end extending out of the housing 12. When the screw rotor device 10 operates as a compressor, shaft 76 is rotated causing male rotor 14 to rotate. The male rotor 14 causes the female rotor 16 to counter-rotate through the gears 84, 86, and the helical threads 34, 36 intermesh with the helical grooves 38, 40.
As described above, the distal labyrinth seal 54 helps sealing between each of the helical threads 34, 36 on the male rotor 14 and the cylindrical bore 30 in the housing 12. Similarly, as particularly illustrated in
It will also be appreciated that, depending on the application, the temperature range experienced by the rotors could vary, and the tolerances can be designed to account for 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 that the sealing distances, or tolerances, do not vary substantially throughout the operating range of the screw rotor system. For example, the materials may have a similar modulus of thermal expansion or may be selected such that they reach an optimal seal at a particular design point or in an operating region at steady state condition.
As discussed above, the preferred embodiment of the screw rotor device 10 is designed to operate as a compressor. The screw rotor device 10 can be also be used as an expander. When acting as an expander, gas having a pressure higher than ambient pressure enters the screw rotor device 10 through the outlet port 20, valve 58 being optional. The pressure of the gas forces rotation of the male rotor 14 and the female rotor 16. As the gas expands into the spaces 39, 41, work is extracted through the end of shaft 76 that extends out of the housing 12. The pressure in the spaces 39, 41 decreases as the gas moves towards the inlet port 18 and exits into ambient pressure at the inlet port 18. The screw rotor device 10 can operate with a gaseous working fluid and may also be used as a pump for a liquid working fluid. For pumping liquids, a valve may also be used to prevent the fluid from backing into the rotor.
To ensure that persons of ordinary skill in the art will appreciate the expansive scope of the present invention, it should be understood that while the prior art multi-pitch screw rotor designs were able to significantly reduce or eliminate the three forms of leakage above and the single-pitch, buttress-thread rotor designs were able significantly reduce or eliminate the blow-hole gap, no heretofore known screw compressor design has been able to eliminate or significantly reduce all of these internal leakages simultaneously and without limitation. While it is true that the buttress-thread rotor designs could significantly reduce or eliminate the blow-hole gap, its elimination came at the price that the complete seal would only work for a single-pitch, but not because of the blow-hole gap. Instead, when the buttress-thread rotor designs are used for multi-pitch screw rotors, the gap between the front and back of the intermeshing male rotor thread and female rotor groove could then cause significant leakage from the high pressure side of the screw rotor to the low pressure or suction side of the rotor.
Gaps between the rotors themselves and between one or more of the screw rotors and the housing, such as gap 90 illustrated in
As an example of different tolerances for different applications, the screw rotor system 10 illustrated in
Generally, the sealing tolerance for the present invention, between the helical thread and the helical groove, can be set as a defined number, such as less than or equal to 0.003″ or 0.001″ or some other small distance. Even more generally, the sealing tolerance can be based on a ratio of the rotor diameters of the screw rotor system 10, such as a rule that the sealing tolerance being no greater than {fraction (1/1,000)} or {fraction (1/10,000)} of the male rotor diameter. Most generally, the sealing tolerance can be based on any geometric proximity which can be defined by the distance between the rotors themselves, the rotors and the housing, or any other distance that is relevant to sealing conditions. Depending on the geometric proximity that is selected, the sealing tolerance may be defined by the geometric proximity itself or can be based thereon, such as a sealing tolerance which is within an order of magnitude of the geometric proximity.
It will be appreciated that the gap 90 in the embodiment illustrated in
The particular structure and process of the present invention is discussed with reference to features particularly illustrated in
To show the sealing relationships of the present invention,
As summarized in the listing above and particularly illustrated in
The creation and progression of these seals, as the male and female rotors intermesh, is illustrated in
As discussed in detail below, with regard to the illustrations in
According to the designs of the other non-buttress thread embodiments of the present invention, the gap between the trailing side of the groove and the trailing face of the thread does not exist, even when the screw rotors are multiple-pitch designs. Generally speaking, the buttress thread designs have a single-sided sealing relationship, i.e. between the leading side of the groove and the leading side of the face, whereas the other designs have a double-sided sealing relationship between the leading side of the groove and the leading side of the face and between the trailing side of the groove and the trailing side of the face. The double-sided sealing relationship can be particularly defined by the first sealing relationship, the second sealing relationship, the third sealing relationship, the fourth sealing relationship, and the fifth sealing relationship. In this way, no leak pathway is provided through this double-sided sealing relationship. An illustration of this double-sided sealing 136 is particularly shown for multiple-pitch rotors 138, 140 in
Although similar groove shapes appear to be shown in prior art screw rotors and similar thread shapes appear to be shown in other prior art screw rotors, not only were such threads and grooves never before combined in a single screw rotor system, none of these prior art references ever even suggested that such grooves should be combined with the thread of the other references. In fact, none of these prior art designs were based on the present design method. Therefore, the threads and grooves of all of these prior art screw rotors fail to satisfy the structural features disclosed and claimed for the thread and groove of the present invention. Additionally, the prior art references fail to disclose the cooperative relationships between the thread, groove and cusps of the housing, as disclosed and claimed by the present invention. Finally, none of the prior art references disclose or suggest the design process of the present invention. In fact, as discussed in the Background of the Invention section above, the prior art actually suggests that it is not possible to have any design process, or resulting design, which eliminates the blow-hole gap.
The design process of the present invention is schematically set forth in the illustrations of
In eliminating the blow-hole gap on the front side and the back side of the housing, it will be appreciated that the thread profile has discontinuities between its top land and its top and bottom faces, i.e. trailing and leading faces, respectively, for the compressor or pump type of application. The leading edge discontinuity is located at the leading edge point where the leading line and the major diameter arc intersect. The trailing edge discontinuity is located at the trailing edge point where the trailing line and the major diameter arc intersect. According to this visual image of the design process, it will be appreciated that the thread's cross-sectional profile lines between the top land and the root can be formed from any type of line, including straight lines, concave lines, convex lines, arcs, involutes, inverse-involutes, parabolas, hyperbolas, cycloids, trochoids, epicycloids, epitrochoids, hypocycloids, hypotrochoids, continuous straight lines and arcuate lines, and any combination thereof in piecewise-continuous lines.
The design process of the present invention is now described with reference to the flowchart in
Given these design conditions, it will be appreciated that the threads and grooves can be designed according to the present invention such that they have minimal backlash. In particular, many designs for screw rotors have pressure angles as high as 30° which results in a significant amount of backlash. In comparison, the present invention allows designers to create entire families of screw rotors with minimal backlash, such as with pressure angles less than half of 30°, including families with 0° pressure angle and no backlash.
It will also be appreciated that, in completing the screw rotor system design, the interior sides of the housing are generally defined in the shape of a figure-eight in close tolerance with the circles 240. As illustrated in
Of course, to create the third dimension for the screw rotors, at least one helix angle needs to be selected 250. As discussed above, the helix angle can be varied along the length of the rotors, thereby resulting in a variable pitch screw rotor compressor. Also, the major and minor diameters can be varied along the length of the rotors, thereby resulting in a tapered screw rotor compressor.
As yet more detail into the design process, the first rotor major circle is defined. The first rotor major circle has a first major diameter. The second rotor major circle is also defined such that it intersects with the first rotor major circle at a pair of intersection points. The second rotor major circle has a second major diameter, and less than one half of the second major diameter extends into the first rotor major circle. Less than one half of the first major diameter extends into the second rotor major circle, and the second rotor major circle shares a single tangential point with a first rotor minor circle centered within the first rotor major circle. The first rotor major circle shares another single tangential point with a second rotor minor circle centered within the second rotor major circle.
A first point is now selected on the first rotor major circle, and the point defines a first line segment receding radially inward from the second rotor major point to the second rotor minor point. In particular, the first line segment is defined by the path of the first point as it progresses from the second rotor major circle to the second rotor minor circle when the first rotor major circle and the second rotor major circle rotate in phase with each other by equal angular amounts. Similarly, a second point on the first rotor major circle and circumferentially spaced from the first point is selected., and the point defines a second line segment receding radially inward from a circumferentially-spaced second rotor major point to a circumferentially-spaced second rotor minor point. The second line segment is defined by the path of the second point as it progresses from the second rotor major circle to the second rotor minor circle when the first rotor major circle and the second rotor major circle rotate in phase with each other by equal angular amounts. Additionally, the circumferentially-spaced second rotor major point and second rotor minor point are circumferentially spaced from the second rotor major point the second rotor minor point, respectively.
A pair of first rotor root line segments that extend from the first rotor minor circle to a pair of intermediate points are now identified. One intermediate point is situated between the first rotor minor circle and the first point on the first rotor major circle and the other intermediate point is situated between the first rotor minor circle and the second point on the first rotor major circle. The intermediate points are circumferentially spaced from each other, and the first rotor root line segments are defined by the paths of the second rotor major point and the circumferentially-spaced second rotor major point when the first rotor major circle and the second rotor major circle rotate in phase with each other by equal angular amounts. Finally, to complete the profile for the thread, it is preferable to use a pair of circumferentially-spaced first rotor line segments that respectively extend between the pair of first rotor root line segments and the first point and the second point on the first rotor major circle.
In designing profiles of the screw rotor devices, it will be appreciated that the top land of the thread is preferably an arc rather than merely being a point on the major diameter of the male rotor. This preference can be rather important because a point may tend to cause the Bernoulli effect, causing the top land of the thread and the bottom land of the groove to act as a converging-diverging nozzle. Due to pressure differentials, such an effect could even result in supersonic flow through such a nozzle, producing shock waves which are non-adiabatic and increase the entropy in the flow, thereby increasing the flow temperature and reducing the thermodynamic efficiency.
From a close examination of the embodiments of the present invention, it will be apparent that, in the embodiments illustrated in
This selection could be important to particular applications because when the female rotor major diameter seals with the male rotor minor diameter (), the rotors may be so close as to cause friction therebetween, and rolling friction (same diameters) is less than sliding friction (different diameters). By reducing the friction in the screw rotor system, the steady state temperature of the rotors and the flow traveling through the rotors can be kept lower than when there is the higher friction of sliding friction between the rotors. This could be important in a refrigeration application or some other cooling application in which air or another working fluid is being run through one or more screw rotors to cool the working fluid.
An example of an application that cools the working fluid is illustrated in
Such a mechanical linkage between the devices 10, 10, could reduce the steady-state power requirement of the compressor by more than 50%. In particular, the work that is extracted out of the expander can be passed back to the compressor through the mechanical linkage. Therefore, with an expander operating at or above a thermodynamic efficiency of 85%, most of the expansion energy is available to help run the compressor. It will be appreciated that when the compressor and the expander are linked together in this manner, it is possible for the units to be integrated into a single housing 12. Of course, it will also be appreciated that multiple stages of compressors and/or expanders can be used to super-cool certain working fluids.
The screw rotor system can also be used in many other applications. For example, the screw rotors can be used in many types of hydrostatic power systems 168 and hydrodynamic power systems 170. A hydrostatic power system is discussed with reference to
In comparison to the hydrostatic drive, hydrodynamic drive converts into work as much of the energy in the compressed working fluid as possible and then dispels the spent working fluid. A couple of examples generally illustrated by
One particular use that is within the scope of the present invention is the use of blades and other tools as the working device. For example, the blades could be for a garbage crusher or for a lawn mower. In the case where the blade is for a garbage crusher or garbage chopper (drain/blade housing 190 shown), the high pressure water (working fluid) powers the crusher and the low pressure water (spent fluid) is dispelled into the drain or other receptacle where the garbage is being crushed and/or chopped. For a kitchen sink application, the high pressure water preferably comes from the standard cold water supply of the sink, and it will be appreciated that the low pressure water that is dispelled into the drain would be useful for washing the garbage down the drain while the high pressure water is used to power the crusher/chopper. Similarly, for a hydrodynamic lawn mower (blade housing 192 shown), the high pressure water (working fluid) powers the blades and the low pressure water (spent fluid) is dispelled onto the portion of the lawn that has just been cut. For the hydrodynamic lawn mower, the high pressure water preferably comes from a standard outside faucet, although for larger powered mowers, a reservoir tank could be used to haul the water and a screw rotor compressor could be used to create the pressurized water source. Once the water's pressure is spent to power the blade, the water can be dumped onto the lawn.
Another dynamic application is the use of the screw rotor devices in a milling machine 194 or other such tooling equipment. In this case, the working fluid is pressurized air. Therefore, to extract the energy from the air and thereby power the tool, the air is expanded within the screw rotor system 10. As the air expands, its temperature drops. Therefore, during spring and summer months, the colder expanded air can be used to cool the machining facility, and during the fall and winter months, the colder air can be dumped through a valve to the outside.
In the last application particularly discussed for the present invention, a gas turbine engine includes linked-rotor compressors 154-166-154, a burner section 196, an expander 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 then expanded.
In view of the foregoing, it will be seen that the several advantages of the invention are achieved and attained. The embodiments were chosen and described in order to best explain the principles of the invention and its practical application to thereby enable others skilled in the art to best utilize the invention in various embodiments and with various modifications as are suited to the particular use contemplated. As various modifications could be made in the constructions and methods herein described and illustrated without departing from the scope of the invention, it is intended that all matter contained in the foregoing description or shown in the accompanying drawings shall be interpreted as illustrative rather than limiting. For example, although the preferred embodiments of the present invention describes rotors having substantially parallel axes, the axes do not necessarily need to be parallel. Additionally, the method for designing screw rotor profiles according to the present invention is not limited to any particular coordinate system. For example, a Cartesian coordinate system, i.e., rectangular (x, y, z), or an angular coordinate system, i.e., cylindrical (r, Φ, x) could be used to define the profiles. Other coordinate systems may also be used, such as a polar coordinate system, although it will be appreciated that some coordinate systems may unnecessarily add complexity to the design process. Additionally, the several applications discussed herein are illustrative of the wide range of applications where the present invention can be useful. In particular, it will be appreciated that for the internal combustion engine application of the screw rotor system 10, a fuel inlet 108 would be used to deliver the fuel into one of the spaces 39, 41. It will also be appreciated that, for this embodiment, the flow would likely be moving in the opposite direction from that which is illustrated in
This application is a continuation-in-part of U.S. application Ser. No. 10/283,421, filed on Oct. 29, 2002 which is a continuation-in-part of U.S. application Ser. No. 10/013,747, filed on Oct. 19, 2001 and issued as U.S. Pat. No. 6,599,112 on Jul. 29, 2003. This application is also related to the subject matter in co-pending U.S. application Ser. No. 10/283,422, filed on Oct. 29, 2002, which is hereby incorporated by reference into the present invention disclosure. This application is also related to the subject matter in co-pending U.S. application Ser. No. 10/764,195, patent application filed on Jan. 23, 2004, Docket No. 71044-006CIPN2, which is also a continuation of U.S. application Ser. No. 10/283,421 and is also hereby incorporated by reference into the present invention disclosure.
| Number | Date | Country | |
|---|---|---|---|
| Parent | 10283421 | Oct 2002 | US |
| Child | 10810513 | Mar 2004 | US |
| Parent | 10013747 | Oct 2001 | US |
| Child | 10283421 | Oct 2002 | US |
