Patent Application
20200025202
The invention generally relates to fluid pumps, such as compressors and expanders.
Compressors have typically been used for a variety of applications, such as air compression, vapor compression for refrigeration, and compression of industrial gases. Compressors can be split into two main groups, positive displacement and dynamic. Positive displacement compressors reduce the compression volume in the compression chamber to increase the pressure of the fluid in the chamber. This is done by applying force to a drive shaft that is driving the compression process. Dynamic compressors work by transferring energy from a moving set of blades to the working fluid.
Positive displacement compressors can take a variety of forms. They are typically classified as reciprocating or rotary compressors. Reciprocating compressors are commonly used in industrial applications where higher pressure ratios are necessary. They can easily be combined into multistage machines, although single stage reciprocating compressors are not typically used at pressures above 80 psig. Reciprocating compressors use a piston to compress the vapor, air, or gas, and have a large number of components to help translate the rotation of the drive shaft into the reciprocating motion used for compression. This can lead to increased cost and reduced reliability. Reciprocating compressors also suffer from high levels of vibration and noise. This technology has been used for many industrial applications such as natural gas compression.
Rotary compressors use a rotating component to perform compression. As noted in the art, rotary compressors typically have the following features in common: (1) they impart energy to the gas being compressed by way of an input shaft moving a single or multiple rotating elements; (2) they perform the compression in an intermittent mode; and (3) they do not use inlet or discharge valves. (Brown, Compressors: Selection and Sizing, 3rd Ed., at 6). As further noted in Brown, rotary compressor designs are generally suitable for designs in which less than 20:1 pressure ratios and 1000 CFM flow rates are desired. For pressure ratios above 20:1, Royce suggests that multistage reciprocating compressors should be used instead.
Typical rotary compressor designs include the rolling piston, screw compressor, scroll compressor, lobe, liquid ring, and rotary vane compressors. Each of these traditional compressors has deficiencies for producing high pressure, near isothermal conditions.
The design of a rotating element/rotor/lobe against a radially moving element/piston to progressively reduce the volume of a fluid has been utilized as early as the mid-19th century with the introduction of the “Yule Rotary Steam Engine.” Developments have been made to small-sized compressors utilizing this methodology into refrigeration compression applications. However, current Yule-type designs are limited due to problems with mechanical spring durability (returning the piston element) as well as chatter (insufficient acceleration of the piston in order to maintain contact with the rotor).
For commercial applications, such as compressors for refrigerators, small rolling piston or rotary vane designs are typically used. (P N Ananthanarayanan, Basic Refrigeration and Air Conditioning, 3rd Ed., at 171-72.) In these designs, a closed oil-lubricating system is typically used.
Rolling piston designs typically allow for a significant amount of leakage between an eccentrically mounted circular rotor, the interior wall of the casing, and/or the vane that contacts the rotor. By spinning the rolling piston faster, the leakages are deemed acceptable because the desired pressure and flow rate for the application can be easily reached even with these losses. The benefit of a small self-contained compressor is more important than seeking higher pressure ratios.
Rotary vane designs typically use a single circular rotor mounted eccentrically in a cylinder slightly larger than the rotor. Multiple vanes are positioned in slots in the rotor and are kept in contact with the cylinder as the rotor turns typically by spring or centrifugal force inside the rotor. The design and operation of these type of compressors may be found in Mark's Standard Handbook for Mechanical Engineers, Eleventh Edition, at 14:33-34.
In a sliding-vane compressor design, vanes are mounted inside the rotor to slide against the casing wall. Alternatively, rolling piston designs utilize a vane mounted within the cylinder that slides against the rotor. These designs are limited by the amount of restoring force that can be provided and thus the pressure that can be yielded.
Each of these types of prior art compressors has limits on the maximum pressure differential that it can provide. Typical factors include mechanical stresses and temperature rise. One proposed solution is to use multistaging. In multistaging, multiple compression stages are applied sequentially. Intercooling, or cooling between stages, is used to cool the working fluid down to an acceptable level to be input into the next stage of compression. This is typically done by passing the working fluid through a heat exchanger in thermal communication with a cooler fluid. However, intercooling can result in some condensation of liquid and typically requires filtering out of the liquid elements. Multistaging greatly increases the complexity of the overall compression system and adds costs due to the increased number of components required. Additionally, the increased number of components leads to decreased reliability and the overall size and weight of the system are markedly increased.
For industrial applications, single- and double-acting reciprocating compressors and helical-screw type rotary compressors are most commonly used. Single-acting reciprocating compressors are similar to an automotive type piston with compression occurring on the top side of the piston during each revolution of the crankshaft. These machines can operate with a single-stage discharging between 25 and 125 psig or in two stages, with outputs ranging from 125 to 175 psig or higher. Single-acting reciprocating compressors are rarely seen in sizes above 25 HP. These types of compressors are typically affected by vibration and mechanical stress and require frequent maintenance. They also suffer from low efficiency due to insufficient cooling.
Double-acting reciprocating compressors use both sides of the piston for compression, effectively doubling the machine's capacity for a given cylinder size. They can operate as a single-stage or with multiple stages and are typically sized greater than 10 HP with discharge pressures above 50 psig. Machines of this type with only one or two cylinders require large foundations due to the unbalanced reciprocating forces. Double-acting reciprocating compressors tend to be quite robust and reliable, but are not sufficiently efficient, require frequent valve maintenance, and have extremely high capital costs.
Lubricant-flooded rotary screw compressors operate by forcing fluid between two intermeshing rotors within a housing which has an inlet port at one end and a discharge port at the other. Lubricant is injected into the chamber to lubricate the rotors and bearings, take away the heat of compression, and help to seal the clearances between the two rotors and between the rotors and housing. This style of compressor is reliable with few moving parts. However, it becomes quite inefficient at higher discharge pressures (above approximately 200 psig) due to the intermeshing rotor geometry being forced apart and leakage occurring. In addition, lack of valves and a built-in pressure ratio leads to frequent over or under compression, which translates into significant energy efficiency losses.
Rotary screw compressors are also available without lubricant in the compression chamber, although these types of machines are quite inefficient due to the lack of lubricant helping to seal between the rotors. They are a requirement in some process industries such as food and beverage, semiconductor, and pharmaceuticals, which cannot tolerate any oil in the compressed air used in their processes. Efficiency of dry rotary screw compressors are 15-20% below comparable injected lubricated rotary screw compressors and are typically used for discharge pressures below 150 psig.
Using cooling in a compressor is understood to improve upon the efficiency of the compression process by extracting heat, allowing most of the energy to be transmitted to the gas and compressing with minimal temperature increase. Liquid injection has previously been utilized in other compression applications for cooling purposes. Further, it has been suggested that smaller droplet sizes of the injected liquid may provide additional benefits.
In U.S. Pat. No. 4,497,185, lubricating oil was intercooled and injected through an atomizing nozzle into the inlet of a rotary screw compressor. In a similar fashion, U.S. Pat. No. 3,795,117 uses refrigerant, though not in an atomized fashion, that is injected early in the compression stages of a rotary screw compressor. Rotary vane compressors have also attempted finely atomized liquid injection, as seen in U.S. Pat. No. 3,820,923.
Published International Pat. App. No. WO 2010/017199 and U.S. Pat. Pub. No. 2011/0023814 relate to a rotary engine design using a rotor, multiple gates to create the chambers necessary for a combustion cycle, and an external cam-drive for the gates. The force from the combustion cycle drives the rotor, which imparts force to an external element. Engines are designed for a temperature increase in the chamber and high temperatures associated with the combustion that occurs within an engine. Increased sealing requirements necessary for an effective compressor design are unnecessary and difficult to achieve. Combustion forces the use of positively contacting seals to achieve near perfect sealing, while leaving wide tolerances for metal expansion, taken up by the seals, in an engine. Further, injection of liquids for cooling would be counterproductive and coalescence is not addressed.
Liquid mist injection has been used in compressors, but with limited effectiveness. In U.S. Pat. No. 5,024,588, a liquid injection mist is described, but improved heat transfer is not addressed. In U.S. Pat. Publication. No. U.S. 2011/0023977, liquid is pumped through atomizing nozzles into a reciprocating piston compressor's compression chamber prior to the start of compression. It is specified that liquid will only be injected through atomizing nozzles in low pressure applications. Liquid present in a reciprocating piston compressor's cylinder causes a high risk for catastrophic failure due to hydrolock, a consequence of the incompressibility of liquids when they build up in clearance volumes in a reciprocating piston, or other positive displacement, compressor. To prevent hydrolock situations, reciprocating piston compressors using liquid injection will typically have to operate at very slow speeds, adversely affecting the performance of the compressor.
U.S. Patent Application Publication No. 2013-0209299, titled “Compressor With Liquid Injection Cooling” discloses another rotary compressor with liquid injection cooling. The entire contents of U.S. Patent Application Publication No. 2013-0209299 are incorporated herein by reference in its entirety.
The presently preferred embodiments are directed to rotary compressor designs. These designs are particularly suited for high pressure applications, typically above 200 psig with pressure ratios typically above that for existing high-pressure positive displacement compressors.
One or more embodiments provides a compressor that includes: a casing with an inner wall defining a compression chamber; a drive shaft and rotor rotatably coupled to the casing for common rotation relative to the casing, the rotor having a non-circular profile; and a gate coupled to the casing for pivotal movement relative to the casing, the gate comprising a sealing edge, the gate being operable to move relative to the casing to locate the sealing edge proximate to the rotor as the rotor rotates such that the gate separates an inlet volume and a compression volume in the compression chamber.
One or more embodiments provides a compressor that includes: a casing with an inner wall defining a compression chamber, an inlet leading into the compression chamber, and an outlet leading out of the compression chamber; a drive shaft and rotor rotatably coupled to the casing for common rotation relative to the casing, the rotor having a non-circular profile; a gate coupled to the casing for movement relative to the casing, the gate comprising a sealing edge, the gate being operable to move relative to the casing to locate the sealing edge proximate to the rotor as the rotor rotates such that the gate separates an inlet volume and a compression volume in the compression chamber, the inlet and outlet being disposed on opposite sides of the sealing edge from each other; and an outlet manifold in fluid communication with the outlet, wherein the outlet is elongated in a direction parallel to a rotational axis of the drive shaft, wherein the outlet manifold defines an interior passageway, and wherein the passageway varies in cross-sectional shape between an entrance into the manifold and an exit out of the manifold, and wherein the outlet manifold comprises a plurality of vanes disposed in the interior passageway to direct the flow of working fluid through the outlet manifold.
One or more embodiments provides a compressor that includes: a casing with an inner wall defining a compression chamber, an inlet leading into the compression chamber, and an outlet leading out of the compression chamber; a rotor coupled to the casing for rotation relative to the casing; a gate movably coupled to one of the casing and rotor for movement relative to the one of the casing and rotor, the gate comprising a sealing edge, the gate being operable to locate the sealing edge proximate to the other of the casing and rotor as the rotor rotates; and a hydrostatic bearing arrangement disposed between (1) the gate and (2) the one of the casing and rotor to reduce friction when the gate moves during operation of the compressor.
One or more embodiments provides a compressor that includes: a compression chamber casing with an inner wall defining a compression chamber, an inlet leading into the compression chamber, and an outlet leading out of the compression chamber; a drive shaft and rotor rotatably coupled to the compression chamber casing for common rotation relative to the compression chamber casing; a gate coupled to the compression chamber casing for movement relative to the compression chamber casing, the gate comprising a sealing edge, the gate being operable to move relative to the compression chamber casing to locate the sealing edge proximate to the rotor as the rotor rotates such that the gate separates an inlet volume and a compression volume in the compression chamber, the inlet and outlet being disposed on opposite sides of the sealing edge from each other; and a gate positioning system coupled to the gate, the gate positioning system being shaped and configured to reciprocally move the gate during rotation of the rotor so that the sealing edge remains proximate to the rotor during rotation of the rotor.
According to various embodiments, the gate positioning system includes a cam shaft rotatably coupled to the compression chamber casing for rotation relative to the compression chamber casing, the cam shaft being spaced from the drive shaft, the cam shaft being connected to the drive shaft so as to be rotationally driven by the drive shaft, a cam rotatably coupled to the compression chamber casing for concentric rotation with the cam shaft relative to the compression chamber casing, a cam follower mounted to the gate for movement with the gate relative to the compression chamber casing, the cam follower abutting the cam so that rotation of the cam causes the cam follower and gate to move relative to the compression chamber casing.
One or more embodiments provides a compressor system that includes: a plurality of compressors. Each compressor may include a casing with an inner wall defining a compression chamber, an inlet leading into the compression chamber, and an outlet leading out of the compression chamber, a rotor rotatably coupled to the casing for rotation relative to the casing, and a gate coupled to the casing for movement relative to the casing, the gate comprising a sealing edge, the gate being operable to move relative to the casing to locate the sealing edge proximate to the rotor as the rotor rotates such that the gate separates an inlet volume and a compression volume in the compression chamber, the inlet and outlet being disposed on opposite sides of the sealing edge from each other. The system includes a mechanical linkage between the rotors of the plurality of compressors, the mechanical linkage connecting between the rotors such that compression cycles of the plurality of compressors are out of phase with each other.
One or more embodiments provides a compressor that includes: a casing with an inner wall defining a compression chamber, an inlet leading into the compression chamber, and an outlet leading out of the compression chamber; a drive shaft and rotor rotatably coupled to the casing for common rotation relative to the casing such that when the rotor is rotated, the compressor compresses working fluid that enters the compression chamber from the inlet, and forces compressed working fluid out of the compression chamber through the outlet; and a mechanical seal located at an interface between the drive shaft and casing where the drive shaft passes through the casing.
According to various embodiments, the mechanical seal includes: first, second, and third seals disposed sequentially along a leakage path between the drive shaft and casing rotor, a source of pressurized hydraulic fluid, and a hydraulic fluid passageway that connects the source to a space along the leakage path between the second and third seals so as to keep the space pressurized with hydraulic fluid.
One or more embodiments provides a non-circular seal for sealing an interface between two moving parts. The seal includes a non-circular structural base (e.g., comprising steel) having a closed perimeter; and a low friction sealing material (e.g., graphite or Teflon) bonded to the base.
One or more embodiments provides a compressor that includes: a casing with an inner wall defining a compression chamber, an inlet leading into the compression chamber, and an outlet leading out of the compression chamber; a rotor rotatably coupled to the casing for rotation relative to the casing such that when the rotor is rotated, the compressor compresses working fluid that enters the compression chamber from the inlet, and forces compressed working fluid out of the compression chamber through the outlet; a gate coupled to the casing for reciprocating movement relative to the casing, the gate comprising a sealing edge, the gate being operable to move relative to the casing to locate the sealing edge proximate to the rotor as the rotor rotates such that the gate separates an inlet volume and a compression volume in the compression chamber; and a mechanical seal located at an interface between the gate and casing. The mechanical seal includes: first, second, and third seals disposed sequentially along a leakage path between the gate and casing, a source of pressurized hydraulic fluid, and a hydraulic fluid passageway that connects the source to a space along the leakage path between the second and third seals so as to keep the space pressurized with hydraulic fluid.
According to various embodiments, the mechanical seal further includes a vent disposed between the first and second seals, the vent being fluidly connected to the inlet so as to direct working fluid that leaks from the compression chamber past the first seal back to the inlet.
According to various embodiments, the first, second, and third seals are all supported by a removable housing, such that the first, second, and third seals and housing can be installed into the casing as a single unit.
According to various embodiments, the mechanical seal comprises n sequential seals along the leakage path between the gate and casing, wherein 3≤n≤50, wherein n includes the first, second, and third seals, wherein one or more spaces between adjacent ones of the seals are filled with pressurized hydraulic fluid, and wherein one or more spaces between adjacent ones of the seals comprise a vent that is fluidly connected on the inlet.
These and other aspects of various non-limiting embodiments of the present invention, as well as the methods of operation and functions of the related elements of structure and the combination of parts and economies of manufacture, will become more apparent upon consideration of the following description and the appended claims with reference to the accompanying drawings, all of which form a part of this specification, wherein like reference numerals designate corresponding parts in the various figures. In one embodiment of the invention, the structural components illustrated herein are drawn to scale. It is to be expressly understood, however, that the drawings are for the purpose of illustration and description only and are not intended as a definition of the limits of the invention. In addition, it should be appreciated that structural features shown or described in any one embodiment herein can be used in other embodiments as well. As used in the specification and in the claims, the singular form of “a”, “an”, and “the” include plural referents unless the context clearly dictates otherwise.
All closed-ended (e.g., between A and B) and open-ended (greater than C) ranges of values disclosed herein explicitly include all ranges that fall within or nest within such ranges. For example, a disclosed range of 1-10 is understood as also disclosing, among other ranged, 2-10, 1-9, 3-9, etc.
Embodiments of the invention can be better understood with reference to the following drawings and description. The components in the figures are not necessarily to scale, emphasis instead being placed upon illustrating the principles of various embodiments of the invention. Moreover, in the figures, like referenced numerals designate corresponding parts throughout the different views.
To the extent that the following terms are utilized herein, the following definitions are applicable:
Balanced rotation: the center of mass of the rotating mass is located on the axis of rotation.
Chamber volume: any volume that can contain fluids for compression.
Compressor: a device used to increase the pressure of a compressible fluid. The fluid can be either gas or vapor, and can have a wide molecular weight range.
Concentric: the center or axis of one object coincides with the center or axis of a second object
Concentric rotation: rotation in which one object's center of rotation is located on the same axis as the second object's center of rotation.
Positive displacement compressor: a compressor that collects a fixed volume of gas within a chamber and compresses it by reducing the chamber volume.
Proximate: sufficiently close to restrict fluid flow between high pressure and low pressure regions. Restriction does not need to be absolute; some leakage is acceptable.
Rotor: A rotating element driven by a mechanical force to rotate about an axis. As used in a compressor design, the rotor imparts energy to a fluid.
Rotary compressor: A positive-displacement compressor that imparts energy to the gas being compressed by way of an input shaft moving a single or multiple rotating elements
Gate casing 150 is connected to and positioned below main casing 110 at a hole in main casing 110. The gate casing 150 is comprised of two portions: an inlet side 152 and an outlet side 154. Other embodiments of gate casing 150 may only consist of a single portion. As shown in
Referring back to
In the illustrated embodiment, the compressing structure comprises a rotor 500. However, according to alternative embodiments, alternative types of compressing structures (e.g., gears, screws, pistons, etc.) may be used in connection with the compression chamber to provide alternative compressors according to alternative embodiments of the invention.
Two cam followers 250 are located tangentially to each cam 240, providing a downward force on the gate. Drive shaft 140 turns cams 240, which transmits force to the cam followers 250. The cam followers 250 may be mounted on a through shaft, which is supported on both ends, or cantilevered and only supported on one end. The cam followers 250 are attached to cam follower supports 260, which transfer the force into the cam struts 230. As cams 240 turn, the cam followers 250 are pushed down, thus moving the cam struts 230 down. This moves the gate support arm 220 and the gate strut 210 down. This, in turn, moves the gate 600 down.
Springs 280 provide a restorative upward force to keep the gate 600 timed appropriately to seal against the rotor 500. As the cams 240 continue to turn and no longer effectuate a downward force on the cam followers 250, springs 280 provide an upward force. As shown in this embodiment, compression springs are utilized. As one of ordinary skill in the art would appreciate, tension springs and the shape of the bearing support plate 156 may be altered to provide for the desired upward or downward force. The upward force of the springs 280 pushes the cam follower support 260 and thus the gate support arm 220 up which in turn moves the gate 600 up.
Due to the varying pressure angle between the cam followers 250 and cams 240, the preferred embodiment may utilize an exterior cam profile that differs from the rotor 500 profile. This variation in profile allows for compensation for the changing pressure angle to ensure that the tip of the gate 600 remains proximate to the rotor 500 throughout the entire compression cycle.
Line A in
A dual cam follower gate positioning system 300 is attached to the gate casing 150 and drive shaft 140. The dual cam follower gate positioning system 300 moves the gate 600 in conjunction with the rotation of the rotor 500. In a preferred embodiment, the size and shape of the cams is nearly identical to the rotor in cross-sectional size and shape. In other embodiments, the rotor, cam shape, curvature, cam thickness, and variations in the thickness of the lip of the cam may be adjusted to account for variations in the attack angle of the cam follower. Further, large or smaller cam sizes may be used. For example, a similar shape but smaller size cam may be used to reduce roller speeds.
A movable assembly includes gate struts 210 and cam struts 230 connected to gate support arm 220 and bearing support plate 156. In this embodiment, the bearing support plate 157 is straight. As one of ordinary skill in the art would appreciate, the bearing support plate can utilize different geometries, including structures designed to or not to perform sealing of the gate casing 150. In this embodiment, the bearing support plate 157 serves to seal the bottom of the gate casing 150 through a bolted gasket connection. Bearing housings 270, also known as pillow blocks, are mounted to bearing support plate 157 and are concentric to the gate struts 210 and the cam struts 230. In certain embodiments, the components comprising this movable assembly may be optimized to reduce weight, thereby reducing the force necessary to achieve the necessary acceleration to keep the tip of gate 600 proximate to the rotor 500. Weight reduction could additionally and/or alternatively be achieved by removing material from the exterior of any of the moving components, as well as by hollowing out moving components, such as the gate struts 210 or the gate 600.
Drive shaft 140 turns cams 240, which transmit force to the cam followers 250, including upper cam followers 252 and lower cam followers 254. The cam followers 250 may be mounted on a through shaft, which is supported on both ends, or cantilevered and only supported on one end. In this embodiment, four cam followers 250 are used for each cam 240. Two lower cam followers 252 are located below and follow the outside edge of the cam 240. They are mounted using a through shaft. Two upper cam followers 254 are located above the previous two and follow the inside edge of the cams 240. They are mounted using a cantilevered connection.
The cam followers 250 are attached to cam follower supports 260, which transfer the force into the cam struts 230. As the cams 240 turn, the cam struts 230 move up and down. This moves the gate support arm 220 and gate struts 210 up and down, which in turn, moves the gate 600 up and down.
Line A in
An embodiment using a belt driven system 310 is shown in
An embodiment of the present invention using an offset gate guide system is shown in
Reciprocating motion of the two-piece gate 602 is controlled through the use of an offset spring-backed cam follower control system 320 to achieve gate motion in concert with rotor rotation. Single cams 342 drive the gate system downwards through the transmission of force on the cam followers 250 through the cam struts 338. This results in controlled motion of the crossarm 334, which is connected by bolts (some of which are labeled as 328 ) with the two-piece gate 602. The crossarm 334 mounted linear bushings 330, which reciprocate along the length of cam shafts 332, control the motion of the gate 602 and the crossarm 334. The cam shafts 332 are fixed in a precise manner to the main casing through the use of cam shaft support blocks 340. Compression springs 346 are utilized to provide a returning force on the crossarm 334, allowing the cam followers 250 to maintain constant rolling contact with the cams, thereby achieving controlled reciprocating motion of the two-piece gate 602.
Alternative embodiments may use an alternate pole orientation to provide attractive forces between the gate and rotor on the top portion of the gate and attractive forces between the gate and gate casing on the bottom portion of the gate. In place of the lower magnet system, springs may be used to provide a repulsive force. In each embodiment, electromagnets may be used in place of permanent magnets. In addition, switched reluctance electromagnets may also be utilized. In another embodiment, electromagnets may be used only in the rotor and gate. Their poles may switch at each inflection point of the gate's travel during its reciprocating cycle, allowing them to be used in an attractive and repulsive method.
Alternatively, direct hydraulic or indirect hydraulic (hydropneumatic) can be used to apply motive force/energy to the gate to drive it and position it adequately. Solenoid or other flow control valves can be used to feed and regulate the position and movement of the hydraulic or hydropneumatic elements. Hydraulic force may be converted to mechanical force acting on the gate through the use of a cylinder based or direct hydraulic actuators using membranes/diaphragms.
As one of skill in the art would appreciate, these alternative drive mechanisms do not require any particular number of linkages between the drive shaft and the gate. For example, a single spring, belt, linkage bar, or yoke could be used. Depending on the design implementation, more than two such elements could be used.
In alternate embodiments, the outlet ports 435 may be located in the rotor casing 400 instead of the gate casing 150. They may be located at a variety of different locations within the rotor casing. The outlet valves 440 may be located closer to the compression chamber, effectively minimizing the volume of the outlet ports 430, to minimize the clearance volume related to these outlet ports. A valve cartridge may be used which houses one or more outlet valves 440 and connects directly to the rotor casing 400 or gate casing 150 to align the outlet valves 440 with outlet ports 435. This may allow for ease of installing and removing the outlet valves 440.
As discussed above, the preferred embodiments utilize a rotor that concentrically rotates within a rotor casing. In the preferred embodiment, the rotor 500 is a right cylinder with a non-circular cross-section that runs the length of the main casing 110.
The radii of the rotor 500 in one preferred embodiment can be calculated using the following functions:
According to an alternative embodiment, the radii of the rotor 500 is calculated as a 3-4-5-polynomial function.
In a preferred embodiment, the rotor 500 is symmetrical along one axis. It may generally resemble a cross-sectional egg shape. The rotor 500 includes a hole 530 in which the drive shaft 140 and a key 540 may be mounted. The rotor 500 has a sealing section 510, which is the outer surface of the rotor 500 corresponding to section II, and a non-sealing section 520, which is the outer surface of the rotor 500 corresponding to sections I and III. The sections I and III have a smaller radius than sections II creating a compression volume. The sealing portion 510 is shaped to correspond to the curvature of the rotor casing 400, thereby creating a dwell seal that effectively minimizes communication between the outlet 430 and inlet 420. Physical contact is not required for the dwell seal. Instead, it is sufficient to create a tortuous path that minimizes the amount of fluid that can pass through. In a preferred embodiment, the gap between the rotor and the casing in this embodiment is less than 0.008 inches. As one of ordinary skill in the art would appreciate, this gap may be altered depending on tolerances, both in machining the rotor 500 and rotor housing 400, temperature, material properties, and other specific application requirements.
Additionally, as discussed below, liquid is injected into the compression chamber. By becoming entrained in the gap between the sealing portion 510 and the rotor casing 400, the liquid can increase the effectiveness of the dwell seal.
As shown in
The rotor design provides several advantages. As shown in the embodiment of
The cross-sectional shape of the rotor 500 allows for concentric rotation about the drive shaft's axis of rotation, a dwell seal 510 portion, and open space on the non-sealing side for increased gas volume for compression. Concentric rotation provides for rotation about the drive shaft's principal axis of rotation and thus smoother motion and reduced noise.
An alternative rotor design 502 is shown in
The rotor surface may be smooth in embodiments with contacting tip seals to minimize wear on the tip seal. In alternative embodiments, it may be advantageous to put surface texture on the rotor to create turbulence that may improve the performance of non-contacting seals. In other embodiments, the rotor casing's interior cylindrical wall may further be textured to produce additional turbulence, both for sealing and heat transfer benefits. This texturing could be achieved through machining of the parts or by utilizing a surface coating. Another method of achieving the texture would be through blasting with a waterj et, sandblast, or similar device to create an irregular surface.
The main casing 110 may further utilize a removable cylinder liner. This liner may feature microsurfacing to induce turbulence for the benefits noted above. The liner may also act as a wear surface to increase the reliability of the rotor and casing. The removable liner could be replaced at regular intervals as part of a recommended maintenance schedule. The rotor may also include a liner. Sacrifical or wear-in coatings may be used on the rotor 500 or rotor casing 400 to correct for manufacturing defects in ensuring the preferred gap is maintained along the sealing portion 510 of the rotor 500.
The exterior of the main casing 110 may also be modified to meet application specific parameters. For example, in subsea applications, the casing may require to be significantly thickened to withstand exterior pressure, or placed within a secondary pressure vessel. Other applications may benefit from the exterior of the casing having a rectangular or square profile to facilitate mounting exterior objects or stacking multiple compressors. Liquid may be circulated in the casing interior to achieve additional heat transfer or to equalize pressure in the case of subsea applications for example.
As shown in
The drive shaft 140 is mounted to endplates 120 in the preferred embodiment using one spherical roller bearing in each endplate 120. More than one bearing may be used in each endplate 120, in order to increase total load capacity. A grease pump (not shown) is used to provide lubrication to the bearings. Various types of other bearings may be utilized depending on application specific parameters, including roller bearings, ball bearings, needle bearings, conical bearings, cylindrical bearings, journal bearings, etc. Different lubrication systems using grease, oil, or other lubricants may also be used. Further, dry lubrication systems or materials may be used. Additionally, applications in which dynamic imbalance may occur may benefit from multi-bearing arrangements to support stray axial loads.
Operation of gates in accordance with embodiments of the present invention are shown in
The gate 600 may include an optional tip seal 620 that makes contact with the rotor 500, providing an interface between the rotor 500 and the gate 600. Tip seal 620 consists of a strip of material at the tip of the gate 600 that rides against rotor 500. The tip seal 620 could be made of different materials, including polymers, graphite, and metal, and could take a variety of geometries, such as a curved, flat, or angled surface. The tip seal 620 may be backed by pressurized fluid or a spring force provided by springs or elastomers. This provides a return force to keep the tip seal 620 in sealing contact with the rotor 500.
Different types of contacting tips may be used with the gate 600. As shown in
Alternatively, a non-contacting seal may be used. Accordingly, the tip seal may be omitted. In these embodiments, the topmost portion of the gate 600 is placed proximate, but not necessarily in contact with, the rotor 500 as it turns. The amount of allowable gap may be adjusted depending on application parameters.
As shown in
Alternatively, liquid may be injected from the gate itself. As shown in
Preferred embodiments enclose the gate in a gate casing. As shown in
In alternate embodiments, the seals could be placed on the gate 600 instead of within the gate casing 150. The seals would form a ring around the gate 600 and move with the gate relative to the casing 150, maintaining a seal against the interior of the gate casing 150. The location of the seals may be chosen such that the center of pressure on the gate 600 is located on the portion of the gate 600 inside of the gate casing 150, thus reducing or eliminating the effect of a cantilevered force on the portion of the gate 600 extending into the rotor casing 400. This may help eliminate a line contact between the gate 600 and gate casing 150 and instead provide a surface contact, allowing for reduced friction and wear. One or more wear plates may be used on the gate 600 to contact the gate casing 150. The location of the seals and wear plates may be optimized to ensure proper distribution of forces across the wear plates.
The seals may use energizing forces provided by springs or elastomers with the assembly of the gate casing 150 inducing compression on the seals. Pressurized fluid may also be used to energize the seals.
The gate 600 is shown with gate struts 210 connected to the end of the gate. In various embodiments, the gate 600 may be hollowed out such that the gate struts 210 can connect to the gate 600 closer to its tip. This may reduce the amount of thermal expansion encountered in the gate 600. A hollow gate also reduces the weight of the moving assembly and allows oil or other lubricants and coolants to be splashed into the interior of the gate to maintain a cooler temperature. The relative location of where the gate struts 210 connect to the gate 600 and where the gate seals are located may be optimized such that the deflection modes of the gate 600 and gate struts 210 are equal, allowing the gate 600 to remain parallel to the interior wall of the gate casing 150 when it deflects due to pressure, as opposed to rotating from the pressure force. Remaining parallel may help to distribute the load between the gate 600 and gate casing 150 to reduce friction and wear.
A rotor face seal may also be placed on the rotor 500 to provide for an interface between the rotor 500 and the endplates 120. An outer rotor face seal is placed along the exterior edge of the rotor 500, preventing fluid from escaping past the end of the rotor 500. A secondary inner rotor face seal is placed on the rotor face at a smaller radius to prevent any fluid that escapes past the outer rotor face seal from escaping the compressor entirely. This seal may use the same or other materials as the gate seal. Various geometries may be used to optimize the effectiveness of the seals. These seals may use energizing forces provided by springs, elastomers or pressurized fluid. Lubrication may be provided to these rotor face seals by injecting oil or other lubricant through ports in the endplates 120.
Along with the seals discussed herein, the surfaces those seals contact, known as counter-surfaces, may also be considered. In various embodiments, the surface finish of the counter-surface may be sufficiently smooth to minimize friction and wear between the surfaces. In other embodiments, the surface finish may be roughened or given a pattern such as cross-hatching to promote retention of lubricant or turbulence of leaking fluids. The counter-surface may be composed of a harder material than the seal to ensure the seal wears faster than the counter-surface, or the seal may be composed of a harder material than the counter-surface to ensure the counter-surface wears faster than the seal. The desired physical properties of the counter-surface (surface roughness, hardness, etc.) may be achieved through material selection, material finishing techniques such as quenching, tempering, or work hardening, or selection and application of coatings that achieve the desired characteristics. Final manufacturing processes, such as surface grinding, may be performed before or after coatings are applied. In various embodiments, the counter-surface material may be steel or stainless steel. The material may be hardened via quenching or tempering. A coating may be applied, which could be chrome, titanium nitride, silicon carbide, or other materials.
Minimizing the possibility of fluids leaking to the exterior of the main housing 100 is desirable. Various seals, such as gaskets and o-rings, are used to seal external connections between parts. For example, in a preferred embodiment, a double o-ring seal is used between the main casing 110 and endplates 120. Further seals are utilized around the drive shaft 140 to prevent leakage of any fluids making it past the rotor face seals. A lip seal is used to seal the drive shaft 140 where it passes through the endplates 120. In various embodiments, multiple seals may be used along the drive shaft 140 with small gaps between them to locate vent lines and hydraulic packings to reduce or eliminate gas leakage exterior to the compression chamber. Other forms of seals could also be used, such as mechanical or labyrinth seals.
It is desirable to achieve near isothermal compression. To provide cooling during the compression process, liquid injection is used. In preferred embodiments, the liquid is atomized to provide increased surface area for heat absorption. In other embodiments, different spray applications or other means of injecting liquids may be used.
Liquid injection is used to cool the fluid as it is compressed, increasing the efficiency of the compression process. Cooling allows most of the input energy to be used for compression rather than heat generation in the gas. The liquid has dramatically superior heat absorption characteristics compared to gas, allowing the liquid to absorb heat and minimize temperature increase of the working fluid, achieving near isothermal compression. As shown in
The amount and timing of liquid injection may be controlled by a variety of implements including a computer-based controller capable of measuring the liquid drainage rate, liquid levels in the chamber, and/or any rotational resistance due to liquid accumulation through a variety of sensors. Valves or solenoids may be used in conjunction with the nozzles to selectively control injection timing. Variable orifice control may also be used to regulate the amount of liquid injection and other characteristics.
Analytical and experimental results are used to optimize the number, location, and spray direction of the injectors 136. These injectors 136 may be located in the periphery of the cylinder. Liquid injection may also occur through the rotor or gate. The current embodiment of the design has two nozzles located at 12 o'clock and 10 o'clock. Different application parameters will also influence preferred nozzle arrays.
Because the heat capacity of liquids is typically much higher than gases, the heat is primarily absorbed by the liquid, keeping gas temperatures lower than they would be in the absence of such liquid injection.
When a fluid is compressed, the pressure times the volume raised to a polytropic exponent remains constant throughout the cycle, as seen in the following equation:
P*V
n=Constant
In polytropic compression, two special cases represent the opposing sides of the compression spectrum. On the high end, adiabatic compression is defined by a polytropic constant of n=1.4 for air, or n=1.28 for methane. Adiabatic compression is characterized by the complete absence of cooling of the working fluid (isentropic compression is a subset of adiabatic compression in which the process is reversible). This means that as the volume of the fluid is reduced, the pressure and temperature each rise accordingly. It is an inefficient process due to the exorbitant amount of energy wasted in the generation of heat in the fluid, which often needs to be cooled down again later. Despite being an inefficient process, most conventional compression technology, including reciprocating piston and centrifugal type compressors are essentially adiabatic. The other special case is isothermal compression, where n=1. It is an ideal compression cycle in which all heat generated in the fluid is transmitted to the environment, maintaining a constant temperature in the working fluid. Although it represents an unachievable perfect case, isothermal compression is useful in that it provides a lower limit to the amount of energy required to compress a fluid.
Embodiments of the present invention achieve these near-isothermal results through the above-discussed injection of liquid coolant. Compression efficiency is improved according to one or more embodiments because the working fluid is cooled by injecting liquid directly into the chamber during the compression cycle. According to various embodiments, the liquid is injected directly into the area of the compression chamber where the gas is undergoing compression.
Rapid heat transfer between the working fluid and the coolant directly at the point of compression may facilitate high pressure ratios. That leads to several aspects of various embodiments of the present invention that may be modified to improve the heat transfer and raise the pressure ratio.
One consideration is the heat capacity of the liquid coolant. The basic heat transfer equation is as follows:
Q=mcpΔT
where Q is the heat, m is mass, ΔT is change in temperature, and cp is the specific heat. The higher the specific heat of the coolant, the more heat transfer that will occur.
Choosing a coolant is sometimes more complicated than simply choosing a liquid with the highest heat capacity possible. Other factors, such as cost, availability, toxicity, compatibility with working fluid, and others can also be considered. In addition, other characteristics of the fluid, such as viscosity, density, and surface tension affect things like droplet formation which, as will be discussed below, also affect cooling performance.
According to various embodiments, water is used as the cooling liquid for air compression. For methane compression, various liquid hydrocarbons may be effective coolants, as well as triethylene glycol.
Another consideration is the relative velocity of coolant to the working fluid. Movement of the coolant relative to the working fluid at the location of compression of the working fluid (which is the point of heat generation) enhances heat transfer from the working fluid to the coolant. For example, injecting coolant at the inlet of a compressor such that the coolant is moving with the working fluid by the time compression occurs and heat is generated will cool less effectively than if the coolant is injected in a direction perpendicular to or counter to the flow of the working fluid adjacent the location of liquid coolant injection.
As shown in
As shown in
According to various embodiments, coolant injection occurs during only part of the compression cycle. For example, in each compression cycle/stroke, the coolant injection may begin at or after the first 10, 20, 30, 40, 50, 60 and/or 70% of the compression stroke/cycle (the stroke/cycle being measured in terms of volumetric compression). According to various embodiments, the coolant injection may end at each nozzle shortly before the rotor sweeps past the nozzle (e.g., resulting in sequential ending of the injection at each nozzle (clockwise as illustrated in
As shown in
A further consideration is the location of the coolant injection, which is defined by the location at which the nozzles inject coolant into the compression chamber. As shown in
As one skilled in the art could appreciate, the number and location of the nozzles may be selected based on a variety of factors. The number of nozzles may be as few as 1 or as many as 256 or more. According to various embodiments, the compressor includes (a) at least 1, 2, 3, 4, 5, 6, 7, 8, 9, 10, 15, 20, 30, 40, 50, 75, 100, 125, 150, 175, 200, 225, and/or 250 nozzles, (b) less than 400, 300, 275, 250, 225, 200, 175, 150, 125, 100, 75, 50, 40, 30, 20, 15, and/or 10 nozzles, (c) between 1 and 400 nozzles, and/or (d) any range of nozzles bounded by such numbers of any ranges therebetween. According to various embodiments, liquid coolant injection may be avoided altogether such that no nozzles are used. Along with varying the location along the angle of the rotor casing, a different number of nozzles may be installed at various locations along the length of the rotor casing. In certain embodiments, the same number of nozzles will be placed along the length of the casing at various angles. In other embodiments, nozzles may be scattered/staggered at different locations along the casing's length such that a nozzle at one angle may not have another nozzle at exactly the same location along the length at other angles. In various embodiments, a manifold may be used in which one or more nozzle is installed that connects directly to the rotor casing, simplifying the installation of multiple nozzles and the connection of liquid lines to those nozzles.
Coolant droplet size is a further consideration. Because the rate of heat transfer is linearly proportional to the surface area of liquid across which heat transfer can occur, the creation of smaller droplets via the above-discussed atomizing nozzles improves cooling by increasing the liquid surface area and allowing heat transfer to occur more quickly. Reducing the diameter of droplets of coolant in half (for a given mass) increases the surface area by a factor of two and thus improves the rate of heat transfer by a factor of 2. In addition, for small droplets the rate of convection typically far exceeds the rate of conduction, effectively creating a constant temperature across the droplet and removing any temperature gradients. This may result in the full mass of liquid being used to cool the gas, as opposed to larger droplets where some mass at the center of the droplet may not contribute to the cooling effect. Based on that evidence, it appears advantageous to inject as small of droplets as possible. However, droplets that are too small, when injected into the high density, high turbulence region as shown in
According to various embodiments, average droplet sizes of between 50 and 500 microns, between 50 and 300 microns, between 100 and 150 microns, and/or any ranges within those ranges, may be fairly effective.
The mass of the coolant liquid is a further consideration. As evidenced by the heat equation shown above, more mass (which is proportional to volume) of coolant will result in more heat transfer. However, the mass of coolant injected may be balanced against the amount of liquid that the compressor can accommodate, as well as extraneous power losses required to handle the higher mass of coolant. According to various embodiments, between 1 and 100 gallons per minute (gpm), between 3 and 40 gpm, between 5 and 25 gpm, between 7 and 10 gpm, and/or any ranges therebetween may provide an effective mass flow rate (averaged throughout the compression stroke despite the non-continuous injection according to various embodiments). According to various embodiments, the volumetric flow rate of liquid coolant into the compression chamber may be at least 1, 2, 3, 4, 5, 6, 7, 8, 9, and/or 10 gpm. According to various embodiments, flow rate of liquid coolant into the compression chamber may be less than 100, 80, 60, 50, 40, 30, 25, 20, 15, and/or 10 gpm.
The nozzle array may be designed for a high flow rate of greater than 1, 2, 3, 4, 5, 6, 7, 8, 9, 10, and/or 15 gallons per minute and be capable of extremely small droplet sizes of less than 500 and/or 150 microns or less at a low differential pressure of less than 400, 300, 200, and/or 100 psi. Two exemplary nozzles are Spraying Systems Co. Part Number: 1/4HHSJ-SS12007 and Bex Spray Nozzles Part Number: 1/4YS12007. Other non-limiting nozzles that may be suitable for use in various embodiments include Spraying Systems Co. Part Number 1/4LN-SS14 and 1/4LN-SS8. The preferred flow rate and droplet size ranges will vary with application parameters. Alternative nozzle styles may also be used. For example, one embodiment may use micro-perforations in the cylinder through which to inject liquid, counting on the small size of the holes to create sufficiently small droplets. Other embodiments may include various off the shelf or custom designed nozzles which, when combined into an array, meet the injection requirements necessary for a given application.
According to various embodiments, one, several, and/or all of the above-discussed considerations, and/or additional/alternative external considerations may be balanced to optimize the compressor's performance. Although particular examples are provided, different compressor designs and applications may result in different values being selected.
According to various embodiments, the coolant injection timing, location, and/or direction, and/or other factors, and/or the higher efficiency of the compressor facilitates higher pressure ratios. As used herein, the pressure ratio is defined by a ratio of (1) the absolute inlet pressure of the source working fluid coming into the compression chamber (upstream pressure) to (2) the absolute outlet pressure of the compressed working fluid being expelled from the compression chamber (downstream pressure downstream from the outlet valve). As a result, the pressure ratio of the compressor is a function of the downstream vessel (pipeline, tank, etc.) into which the working fluid is being expelled. Compressors according to various embodiments of the present invention would have a 1:1 pressure ratio if the working fluid is being taken from and expelled into the ambient environment (e.g., 14.7 psia/14.7 psia). Similarly, the pressure ratio would be about 26:1 (385 psia/14.7 psia) according to various embodiments of the invention if the working fluid is taken from ambient (14.7 psia upstream pressure) and expelled into a vessel at 385 psia (downstream pressure).
According to various embodiments, the compressor has a pressure ratio of (1) at least 3:1, 4:1, 5:1, 6:1, 8:1, 10:1, 15:1, 20:1, 25:1, 30:1, 35:1, and/or 40:1 or higher, (2) less than or equal to 200:1, 150:1, 125:1, 100:1, 90:1, 80:1, 70:1, 60:1, 50:1, 45:1, 40:1, 35:1, and/or 30:1, and (3) any and all combinations of such upper and lower ratios (e.g., between 10:1 and 200:1, between 15:1 and 100:1, between 15:1 and 80:1, between 15:1 and 50:1, etc.).
According to various embodiments, lower pressure ratios (e.g., between 3:1 and 15:1) may be used for working fluids with higher liquid content (e.g., with a liquid volume fraction at the compressor's inlet port of at least 0.5, 1, 2, 3, 4, 5, 6, 7, 8, 9, 10, 15, 20, 25, 30, 35, 40, 50, 60, 70, 75, 80, 85, 90, 91, 92, 93, 94, 95, 96, 97, 98, and/or 99%). Conversely, according to various embodiments, higher pressure ratios (e.g., above 15:1) may be used for working fluids with lower liquid content relative to gas content. However, wetter gases may nonetheless be compressed at higher pressure ratios and drier gases may be compressed at lower pressure ratios without deviating from the scope of various embodiments of the present invention.
Various embodiments of the invention are suitable for alternative operation using a variety of different operational parameters. For example, a single compressor according to one or more embodiments may be suitable to efficiently compress working fluids having drastically different liquid volume fractions and at different pressure ratios. For example, a compressor according to one or more embodiments is suitable for alternatively (1) compressing a working fluid with a liquid volume fraction of between 10 and 50 percent at a pressure ratio of between 3:1 and 15:1, and (2) compressing a working fluid with a liquid volume fraction of less than 10 percent at a pressure ratio of at least 15:1, 20:1, 30:1, and/or 40:1.
According to various embodiments, the compressor efficiently and cost-effectively compresses both wet and dry gas using a high pressure ratio.
According to various embodiments, the compressor is capable of and runs at commercially viable speeds (e.g., between 450 and 1800 rpm). According to various embodiments, the compressor runs at a speed of (a) at least 350, 400, 450, 500, 550, 600, and/or 650 rpm, (b) less than or equal to 3000, 2500, 2000, 1800, 1700, 1600, 1500, 1400, 1300, 1200, 1100, 1050, 1000, 950, 900, 850, and/or 800 rpm, and/or (c) between 350 and 300 rpm, 450-1800 rpm, and/or any ranges within these non-limiting upper and lower limits. According to various embodiments, the compressor is continuously operated at one or more of these speeds for at least 0.5, 1, 5, 10, 15, 20, 30, 60, 90, 100, 150, 200, 250 300, 350, 400, 450, and/or 500 minutes and/or at least 10, 20, 24, 48, 72, 100, 200, 300, 400, and/or 500 hours.
According to various embodiments, the outlet pressure of the compressed fluid is (1) at least 200, 225, 250, 275, 300, 325, 350, 375, 400, 425, 450, 475, 500, 600, 700, 800, 900, 1000, 1250, 1500, 2000, 3000, 4000, and/or 5000 psig, (2) less than 6000, 5500, 5000, 4000, 3000, 2500, 2250, 2000, 1750, 1500, 1250, 1100, 1000, 900, 800, 700, 600 and/or 500 psig, (3) between 200 and 6000 psig, between 200 and 5000 psig, and/or (4) within any range between the upper and lower pressures described above.
According to various embodiments, the inlet pressure is ambient pressure in the environment surrounding the compressor (e.g., 1 atm, 14.7 psia). Alternatively, the inlet pressure could be close to a vacuum (near 0 psia), or anywhere therebetween. According to alternative embodiments, the inlet pressure may be (1) at least −14.5, −10, −5, 0, 5, 10, 25, 50, 100, 150, 200, 250, 300, 350, 400, 450, 500, 550, 600, 700, 800, 900, 1000, 1100, 1200, 1300, 1400, and/or 1500 psig, (2) less than or equal to 3000, 2000, 1900, 1800, 1700, 1600, 1500, 1400, 1300, 1200, 1100, 1000, 900, 800, 700, 600, 500, 400, and/or 350, and/or (3) between −14.5 and 3000 psig, between 0 and 1500 psig, and/or within any range bounded by any combination of the upper and lower numbers and/or any nested range within such ranges.
According to various embodiments, the outlet temperature of the working fluid when the working fluid is expelled from the compression chamber exceeds the inlet temperature of the working fluid when the working fluid enters the compression chamber by (a) less than 700, 650, 600, 550, 500, 450, 400, 375 350, 325, 300, 275, 250, 225, 200, 175, 150, 140, 130, 120, 110, 100, 90, 80, 70, 60, 50, 40, 30, and/or 20 degrees C., (b) at least −10, 0, 10, and/or 20 degrees C., and/or (c) any combination of ranges between any two of these upper and lower numbers, including any range within such ranges.
According to various embodiments, the outlet temperature of the working fluid is (a) less than 700, 650, 600, 550, 500, 450, 400, 375, 350, 325, 300, 275, 250, 225, 200, 175, 150, 140, 130, 120, 110, 100, 90, 80, 70, 60, 50, 40, 30, and/or 20 degrees C., (b) at least −10, 0, 10, 20, 30, 40, and/or 50 degrees C., and/or (c) any combination of ranges between any two of these upper and lower numbers, including any range within such ranges.
The outlet temperature and/or temperature increase may be a function of the working fluid. For example, the outlet temperature and temperature increase may be lower for some working fluids (e.g., methane) than for other working fluids (e.g., air).
According to various embodiments, the temperature increase is correlated to the pressure ratio. According to various embodiments, the temperature increase is less than 200 degrees C. for a pressure ratio of 20:1 or less (or between 15:1 and 20:1), and the temperature increase is less than 300 degrees C. for a pressure ratio of between 20:1 and 30:1.
According to various embodiments, the pressure ratio is between 3:1 and 15:1 for a working fluid with an inlet liquid volume fraction of over 5%, and the pressure ratio is between 15:1 and 40:1 for a working fluid with an inlet liquid volume fraction of between 1 and 20%. According to various embodiments, the pressure ratio is above 15:1 while the outlet pressure is above 250 psig, while the temperature increase is less than 200 degrees C. According to various embodiments, the pressure ratio is above 25:1 while the outlet pressure is above 250 psig and the temperature increase is less than 300 degrees C. According to various embodiments, the pressure ratio is above 15:1 while the outlet pressure is above 250 psig and the compressor speed is over 450 rpm.
According to various embodiments, any combination of the different ranges of different parameters discussed herein (e.g., pressure ratio, inlet temperature, outlet temperature, temperature change, inlet pressure, outlet pressure, pressure change, compressor speed, coolant injection rate, etc.) may be combined according to various embodiments of the invention. According to one or more embodiments, the pressure ratio is anywhere between 3:1 and 200:1 while the operating compressor speed is anywhere between 350 and 3000 rpm while the outlet pressure is between 200 and 6000 psig while the inlet pressure is between 0 and 3000 psig while the outlet temperature is between −10 and 650 degrees C. while the outlet temperature exceeds the inlet temperature by between 0 and 650 degrees C. while the liquid volume fraction of the working fluid at the compressor inlet is between 1% and 50%.
According to one or more embodiments, air is compressed from ambient pressure (14.7 psia) to 385 psia, a pressure ratio of 26:1, at speeds of 700 rpm with outlet temperatures remaining below 100 degrees C. Similar compression in an adiabatic environment would reach temperatures of nearly 480 degrees C.
The operating speed of the illustrated compressor is stated in terms of rpm because the illustrated compressor is a rotary compressor. However, other types of compressors may be used in alternative embodiments of the invention. As those familiar in the art appreciate, the RPM term also applies to other types of compressors, including piston compressors whose strokes are linked to RPM via their crankshaft.
Numerous cooling liquids may be used. For example, water, triethylene glycol, and various types of oils and other hydrocarbons may be used. Ethylene glycol, propylene glycol, methanol or other alcohols in case phase change characteristics are desired may be used. Refrigerants such as ammonia and others may also be used. Further, various additives may be combined with the cooling liquid to achieve desired characteristics. Along with the heat transfer and heat absorption properties of the liquid helping to cool the compression process, vaporization of the liquid may also be utilized in some embodiments of the design to take advantage of the large cooling effect due to phase change.
The effect of liquid coalescence is also addressed in the preferred embodiments. Liquid accumulation can provide resistance against the compressing mechanism, eventually resulting in hydrolock in which all motion of the compressor is stopped, causing potentially irreparable harm. As is shown in the embodiments of
Alternative embodiments may include an inlet located at positions other than shown in the figures. Additionally, multiple inlets may be located along the periphery of the cylinder. These could be utilized in isolation or combination to accommodate inlet streams of varying pressures and flow rates. The inlet ports can also be enlarged or moved, either automatically or manually, to vary the displacement of the compressor.
In these embodiments, multi-phase compression is utilized, thus the outlet system allows for the passage of both gas and liquid. Placement of outlet 430 near the bottom of the rotor casing 400 provides for a drain for the liquid. This minimizes the risk of hydrolock found in other liquid injection compressors. A small clearance volume allows any liquids that remain within the chamber to be accommodated. Gravity assists in collecting and eliminating the excess liquid, preventing liquid accumulation over subsequent cycles. Additionally, the sweeping motion of the rotor helps to ensure that most liquid is removed from the compressor during each compression cycle by guiding the liquid toward the outlet(s) and out of the compression chamber.
Compressed gas and liquid can be separated downstream from the compressor. As discussed below, liquid coolant can then be cooled and recirculated through the compressor.
Various of these features enable compressors according to various embodiments to effectively compress multi-phase fluids (e.g., a fluid that includes gas and liquid components (sometimes referred to as “wet gas”)) without pre-compression separation of the gas and liquid phase components of the working fluid. As used herein, multi-phase fluids have liquid volume fractions at the compressor inlet port of (a) at least 0.5, 1, 2, 3, 4, 5, 6, 7, 8, 9, 10, 15, 20, 25, 30, 35, 40, 50, 60, 70, 75, 80, 85, 90, 91, 92, 93, 94, 95, 96, 97, 98, 99, and/or 99.5%, (b) less than or equal to 99.5, 99, 98, 97, 96, 95, 94, 93, 92, 91, 90, 85, 80, 75, 70, 60 ,50, 40, 35, 30, 25, 20, 15, 10, 9, 8, 7, 6, 5, 4, 3, 2, 1, and/or 0.5%, (c) between 0.5 and 99.5%, and/or (d) within any range bounded by these upper and lower values.
Outlet valves allow gas and liquid (i.e., from the wet gas and/or liquid coolant) to flow out of the compressor once the desired pressure within the compression chamber is reached. The outlet valves may increase or maximize the effective orifice area. Due to the presence of liquid in the working fluid, valves that minimize or eliminate changes in direction for the outflowing working fluid are desirable, but not required. This prevents the hammering effect of liquids as they change direction. Additionally, it is desirable to minimize clearance volume. Unused valve openings may be plugged in some applications to further minimize clearance volume. According to various embodiments, these features improve the wet gas capabilities of the compressor as well as the compressor's ability to utilize in-chamber liquid coolant.
Reed valves may be desirable as outlet valves. As one of ordinary skill in the art would appreciate, other types of valves known or as yet unknown may be utilized. Hoerbiger type R, CO, and Reed valves may be acceptable. Additionally, CT, HDS, CE, CM or Poppet valves may be considered. Other embodiments may use valves in other locations in the casing that allow gas to exit once the gas has reached a given pressure. In such embodiments, various styles of valves may be used. Passive or directly-actuated valves may be used and valve controllers may also be implemented.
In the presently preferred embodiments, the outlet valves are located near the bottom of the casing and serve to allow exhausting of liquid and compressed gas from the high pressure portion. In other embodiments, it may be useful to provide additional outlet valves located along periphery of main casing in locations other than near the bottom. Some embodiments may also benefit from outlets placed on the endplates. In still other embodiments, it may be desirable to separate the outlet valves into two types of valves—one predominately for high pressured gas, the other for liquid drainage. In these embodiments, the two or more types of valves may be located near each other, or in different locations.
The coolant liquid can be removed from the gas stream, cooled, and recirculated back into the compressor in a closed loop system. By placing the injector nozzles at locations in the compression chamber that do not see the full pressure of the system, the recirculation system may omit an additional pump (and subsequent efficiency loss) to deliver the atomized droplets. However, according to alternative embodiments, a pump is utilized to recirculate the liquid back into the compression chamber via the injector nozzles. Moreover, the injector nozzles may be disposed at locations in the compression chamber that see the full pressure of the system without deviating from the scope of various embodiments of the present invention.
According to various embodiments, some compressed working fluid/gas (e.g., natural gas) that has been compressed by the compressor is recirculated back into the compression chamber via the injector nozzles along with coolant to better atomize the coolant (e.g., similar or identical to how snow-making equipment combines a liquid water stream with a compressed gas stream to achieve increase atomization of the water).
One or more embodiments simplify heat recovery because most or all of the heat load is in the cooling liquid. According to various embodiments, heat is not removed from the compressed gas downstream of the compressor. The cooling liquid may cooled via an active cooling process (e.g., refrigeration and heat exchangers) downstream from the compressor. However, according to various embodiments, heat may additionally be recovered from the compressed gas (e.g., via heat exchangers) without deviating from the scope of various embodiments of the present invention.
As shown in
The high pressure working fluid exerts a large horizontal force on the gate 600. Despite the rigidity of the gate struts 210, this force will cause the gate 600 to bend and press against the inlet side of the gate casing 152. Specialized coatings that are very hard and have low coefficients of friction can coat both surfaces to minimize friction and wear from the sliding of the gate 600 against the gate casing 152. A fluid bearing can also be utilized. Alternatively, pegs (not shown) can extend from the side of the gate 600 into gate casing 150 to help support the gate 600 against this horizontal force. Material may also be removed from the non-pressure side of gate 600 in a non-symmetrical manner to allow more space for the gate 600 to bend before interfering with the gate casing 150.
The large horizontal forces encountered by the gate may also require additional considerations to reduce sliding friction of the gate's reciprocating motion. Various types of lubricants, such as greases or oils may be used. These lubricants may further be pressurized to help resist the force pressing the gate against the gate casing. Components may also provide a passive source of lubrication for sliding parts via lubricant-impregnated or self-lubricating materials. In the absence of, or in conjunction with, lubrication, replaceable wear elements may be used on sliding parts to ensure reliable operation contingent on adherence to maintenance schedules. These wear elements may also be used to precisely position the gate within the gate casing. As one of ordinary skill in the art would appreciate, replaceable wear elements may also be utilized on various other wear surfaces within the compressor.
The compressor structure may be comprised of materials such as aluminum, carbon steel, stainless steel, titanium, tungsten, or brass. Materials may be chosen based on corrosion resistance, strength, density, and cost. Seals may be comprised of polymers, such as PTFE, HDPE, PEEK™, acetal copolymer, etc., graphite, cast iron, carbon steel, stainless steel, or ceramics. Other materials known or unknown may be utilized. Coatings may also be used to enhance material properties.
As one of ordinary skill in the art can appreciate, various techniques may be utilized to manufacture and assemble embodiments of the invention that may affect specific features of the design. For example, the main casing 110 may be manufactured using a casting process. In this scenario, the nozzle housings 132, gate casing 150, or other components may be formed in singularity with the main casing 110. Similarly, the rotor 500 and drive shaft 140 may be built as a single piece, either due to strength requirements or chosen manufacturing technique.
Further benefits may be achieved by utilizing elements exterior to the compressor envelope. A flywheel may be added to the drive shaft 140 to smooth the torque curve encountered during the rotation. A flywheel or other exterior shaft attachment may also be used to help achieve balanced rotation. Applications requiring multiple compressors may combine multiple compressors on a single drive shaft with rotors mounted out of phase to also achieve a smoothened torque curve. A bell housing or other shaft coupling may be used to attach the drive shaft to a driving force such as engine or electric motor to minimize effects of misalignment and increase torque transfer efficiency. Accessory components such as pumps or generators may be driven by the drive shaft using belts, direct couplings, gears, or other transmission mechanisms. Timing gears or belts may further be utilized to synchronize accessory components where appropriate.
After exiting the valves the mix of liquid and gases may be separated through any of the following methods or a combination thereof: 1. Interception through the use of a mesh, vanes, intertwined fibers; 2. Inertial impaction against a surface; 3. Coalescence against other larger injected droplets; 4. Passing through a liquid curtain; 5. Bubbling through a liquid reservoir; 6. Brownian motion to aid in coalescence; 7. Change in direction; 8. Centrifugal motion for coalescence into walls and other structures; 9. Inertia change by rapid deceleration; and 10. Dehydration through the use of adsorbents or absorbents.
At the outlet of the compressor, a pulsation chamber may consist of cylindrical bottles or other cavities and elements, may be combined with any of the aforementioned separation methods to achieve pulsation dampening and attenuation as well as primary or final liquid coalescence. Other methods of separating the liquid and gases may be used as well.
In the illustrated embodiment, the coolant injectors 1190 direct coolant directly into the compression chamber 1020. However, according to one or more alternative embodiments, coolant injector(s) 1190 may additionally and/or alternatively inject coolant into the working fluid in the inlet manifold 1140 before the working fluid or coolant reach the compression chamber. Such an alternative may reduce manufacturing costs and/or reduce the amount of power required to inject the coolant.
As shown in
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The vanes 1160a and valve 1180 extend completely across the flow path of compressed fluid (e.g., into the page as shown in
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According to various embodiments, the use of multiple columns of bearings 1310, 1320, 1330, 1340 may facilitate fine tuning of the resistors 1410 of one column (or bearings within one column) relative to other column(s) to accommodate for varying conditions along the length of the gate 1110. For example, if the hydrostatic pressure causes the sleeve 1360 to bow out in the middle, the middle column of bearings 1310, 1320, 1330, 1340 can be tuned down to decrease flow to those larger gaps and increase flow to the end columns where the gaps are tighter and contact between the gate and sleeve would first be made.
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As already known, hydrostatic bearings work by using two flow resistors. In this embodiment, the first flow resistor is a flow resistor valve 1410 inline prior to the bearing 1310, 1320, 1330, 1340, which is held constant during operation. The bearing pad 1310c, 1320c, 1330c, 1340c itself is the second flow resistor. The resistance of the bearing pad 1310c, 1320c, 1330c, 1340c changes and is dependent on the gap between the gate 1110 and the bearing pad itself 1310c, 1320c, 1330c, 1340c. If this gap decreases the pressure in the bearing pad 1310c, 1320c, 1330c, 1340c and the pocket grooves 1310b, 1320b, 1330b, 1340b will go up and similarly if the gap increases the pressure in the pad 1310c, 1320c, 1330c, 1340c and the pocket grooves 1310b, 1320b, 1330b, 1340b will go down. The gap will change due to loads created by the cantilever pressure force on the gate 1110.
According to various embodiments, the flow resistor valve 1410 can be replaced by a set flow resistor or an annulus in the respective passageway 1400 that behaves similarly to the bearing pad resistor. An annulus can be designed into the bearing pad 1310c, 1320c, 1330c, 1340c that allows flow to pass through it with a resistance that is dependent on the gap. Typically the annulus is placed on the opposite surface of the bearing pad to which it is hydraulically connected. To be clear, lubricant would flow through the annulus on one side of the bearing and then flow to its respective bearing pad on the opposite side. Thus, according to various embodiments, the bearings 1310, 1320, 1330, 1340 comprise self-compensating bearings with flow resistors built into the opposing bearings. For example, the flow resistor valve 1400 for the bearing 1310 may be built into the opposite bearing 1330 so that flow to the bearing 1310 is reduced when the bearing 1330 gap is reduced. This may prevent excess hydraulic fluid flow through bearings 1310, 1320, 1330, 1340 with large gaps (because the gap on the opposing bearing is small) or permit larger flow rates to bearings 1310,1320, 1330, 1340 that have higher loads. Bearings 1320, 1340 oppose each other and can work in the same manner. This type of self-compensating hydrostatic bearing is described in U.S. Pat. No. 7,287,906, the entire contents of which are incorporated herein by reference.
As shown in
As used herein, the directional terms “upper” and “lower” with respect to bearings 1310, 1330, 1320, 1340 are defined along the direction of reciprocating movement of the gate 1110, and not necessarily along a gravitational up/down direction (though gravitational up/down aligns with the gate 1110's up/down reciprocating direction according to various embodiments).
According to various embodiments, the hydrostatic bearing arrangement 1300 creates a fluid film gap between the gate 1110 and casing 1010 on the inlet side 1020a of the compression chamber 1020, which may prolong the useful life of the gate 1110 and/or casing 1010 by reducing or eliminating wearing contact between the gate 1110 and casing 1010, and/or reduce the forces required to move the gate 1110 along its reciprocating path.
According to various alternative embodiments, the hydrostatic bearing is used on a rotary vane compressor in which the vanes rotate with and reciprocate relative to the rotor instead of the casing. In such embodiments, a hydrostatic bearing such as the bearing 1300 is disposed between the rotor and gate, rather than between the casing and gate.
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According to various embodiments, the surface of the gate 1110 and/or sleeve 1360 (or a coating thereon) is matted or otherwise constructed so as to create turbulence within the oil flow, thereby increasing the shear force of the oil as it forces its way through the gaps and increases the hydrostatic bearing pressure.
According to alternative embodiments, the hydrostatic bearing arrangement 1300 is replaced with a hydrodynamic bearing arrangement, which provides hydraulic liquid (e.g., oil) to an interface between the gate body 1440 and sleeve 1360. The hydrodynamic bearing relies on relative movement between the gate body 1440 and sleeve 1360 to cause the hydraulic fluid to pressurize and/or lubricate the intersection.
As shown in
Each of the two mechanical seals 1500 includes face seals 1510, 1520, a radial shaft seal 1550, a vent 1560, and hydraulic packing 1590. As shown in
According to various embodiments, the seals 1510, 1520 are retained in their grooves 1040b even when the wear surface of the seals 1510, 1520 (e.g., the graphite portion of the seals 1510, 1520 ) is worn through. For example, as shown in
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The operation of the mechanical seal 1500 is described with reference to
According to various embodiments, the mechanical seal 1500 provides an axially-compact seal that results in lower moment loads on the compressor's bearings.
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Although the seal 1500 is described as including various structures in the illustrated embodiment, the seal 1500 may include greater or fewer structures without deviating from the scope of the present invention. For example, one or more of the seals 1510, 1520, 1550 may be omitted without deviating from the scope of the present invention.
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The operation of the mechanical seal 5200 is described with reference to
According to various embodiments, the seal 5200 may be modified by adding or removing various seals. For example, the compressor 5150 includes one more seal between the compression chamber and the vent than is included in the compressor 1000. In particular, in the compressor 5150, four seals are disposed between the compression chamber 1020 and the vent 5290 (i.e., the seals 1520, 1510, 5240, 5260), while the illustrated compressor 1000 has three such seals (i.e., seals 1520, 1510, 1550). However, according to alternative embodiments greater or fewer such seals may be disposed between the compression chamber and vent without deviating from the scope of various embodiments. For example, one or more of the seals 1520, 1510, 5240, 5260 may be omitted. Alternatively, additional seals like the seals 5240, 5260 may extend between the collar 5220 and the face 5210a of the casing 5210 to further reduce leakage from the compression chamber 1020, and the collar 5220 and faces 5210a,b may be radially expanded to provide space for such additional seals, preferably without axially elongating the overall mechanical seal. Additionally and/or alternatively, the seal 5200 may be modified by adding a radial seal (e.g., like the seal 1550) between the casing 5210 and shaft 1030 along the leakage path between the seals 1510, 5240. Additionally and/or alternatively, the vent 5290 may be disposed along the leakage path between different ones of the seals 1520, 1510, 5240, 5260. For example, the vent may alternatively be disposed in the leakage path between the inner face seal 5240 and the outer face seal 5260.
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The gate supports 2130 mount to the gate 2050 to drive the reciprocating motion of the gate 2050. As shown in
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During operation of the compressor 2000, the drive shaft 2030 rotationally drives the pulley 2080, which rotationally drives the belt 2100, which rotationally drives the pulley 2095, which rotationally drives the shaft 2090, which rotationally drives the cams 2110. Rotation of the cams 2110 drives the cam followers 2120, gate support 2130, and gate 2050 upwardly toward the rotor 2040 against the spring bias of the springs 2140. The cams 2110 are shaped and the belt 2100 and pulleys 2080, 2095 are timed so that the gate positioning system 2060 maintains the seal 2050b of the gate 2050 proximate to (e.g., within 5, 4, 3, 2, 1, 0.5, 0.3, 0.1, 0.05, 0.04, 0.03, 0.02, 0.01, 0.005, 0.004, 0.003, 0.002, and/or 0.001 mm of) the rotor 2040 as the rotor 2040 rotates during operation of the compressor 2000. The gate-positioning system 2060 therefore generally works in a similar manner as the gate positioning system illustrated in
In the gate-positioning system 2060 according to various non-limiting embodiments, a mass of the reciprocating components (e.g., the gate 2050, gate supports 2130, cam followers 2120, portions of the springs 2140 and retainers 2150) is kept relatively low to reduce the forces needed to drive such reciprocation. According to various embodiments, such reduction in reciprocating mass may facilitate higher compressor 2000 operational speeds (in terms of RPMs) and/or smaller springs 2140 and other structural components of the system 2060.
In the illustrated embodiment, the cam shaft 2090 is belt-driven via the pulleys 2080, 2095 and belt 2100. However, according to alternative embodiments, the cam shaft 2090 may be driven by any other suitable mechanism for transferring rotation from the drive shaft 2030 to the cam shaft 2090 (e.g., chain drive, gear drive, etc.) without deviating from the scope of various embodiments.
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Additionally and/or alternatively, as shown in
According to alternative embodiments, the hydraulic packing 2170 may be replaced with any other suitable seal (e.g., conventional hermetic seals that are designed to seal rotating shafts where there is a significant pressure differential between opposing sides of the seal) or eliminated altogether (e.g., if the gate 2050's seal is sufficient) without deviating from the scope of various embodiments.
According to an alternative embodiment, the casing 1010 and 2070 are axially extended to entirely enclose the pulleys 2080, 2095 and cam shaft 2090 such that only the main drive shaft 2030 of the compressor 2000 extends from the casing 2010, 2070, requiring a single mechanical seal like the seal 2170 between the drive shaft 2030 and elongated casing to hermetically seal the compressor 2000.
The inlet manifold 3500 of the compressor 3000 fluidly connects to the inlets of each sub-compressor 3000a, 3000b, 3000c. According to various embodiments, the working fluid inlets of the three sub-compressors 3000a, 3000b, 3000c fluidly connect to each downstream from the manifold 3500. Similarly, the compressed working fluid outlets of the three sub-compressors 3000a, 3000b, 3000c rejoin in the compressor's discharge manifold 3510. According to various embodiments, check-valves are disposed in each sub-compressor's discharge outlets upstream from where the discharge passageways join together.
According to various embodiments, check-valves are also disposed in each sub-compressor's inlet downstream from where the inlet flow path diverges toward respective sub-compressors 3000a, 3000b, 3000c (e.g., downstream or within the inlet manifold 3500) so as to discourage backflow from one chamber 3020a, 3020b, 3020c into another chamber 3020a, 3020b, 3020c during out-of-phase operation of the sub-compressors 3000a, 3000b, 3000c.
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While the illustrated compressor 3000 includes three sub-compressors 3000a, 3000b, 3000c, the compressor may include greater or fewer sub-compressors without deviating from the scope of various embodiments (e.g., n sub-compressors that operate out of phase by 360/n degrees from each other, where n is an integer greater than 1 and preferably less than 100 (e.g., 2, 3, 4, 5, 6, 7, 8, 9, 10)).
Alternatively, the multi-phase concept of the compressor 3000 may be implemented using three discrete compressors (e.g., any of the above discussed compressors such as the compressors 1000, 2000, 5150) by connecting their respective drive shafts (e.g., via direct co-axial mounting such that the compressors are axially spaced from each other along a common drive shaft, via gears, belts, etc.) such that the compressors 1000, 2000, 5150 are out of phase from each other in the same way that the above-discussed sub-compressors 3000a, 3000b, 3000c are out of phase with each other.
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The pivoting gate 4050 helps the gate 4050 to resist the pressure that builds up on the compressed fluid outlet 4160 side of the gate 4050 within the compression chamber 4020. As shown in
According to various embodiments, the gate 4050 and shaft 4052 may be integrally formed.
In the illustrated embodiment, a torsion spring 4140 urges the gate 4050 toward the rotor 4040. However, any other suitable force-imparting mechanism may alternatively be used without deviating from the scope of the present invention (e.g., a compression or tension spring mounted between the casing 4010 and a lever arm attached to the gate 4050 or shaft 4052 to impart torque on the shaft 4052 and gate 4050, a motor, magnets, etc.).
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According to various alternative embodiments, the linear bearings 5090 are replaced with alternative linear movement devices that permit the gate supports 5050 to move in the direction of the arrows 5100. For example, thermal growth can be accounted for by slightly undersizing the gate support 5050 relative to the linear bearings 5080. Additionally and/or alternatively, the linear bearings 5080 may be fitted into slotted holes in the gate casing 5075 such that the linear bearings 5080 can move axially (in the direction of the arrows 5100) if needed due to thermal growth while movement in a perpendicular direction (i.e., in the direction into the page as shown in
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According to various embodiments, the lower portion 6100 may include a sump for oil from the compressor's hydraulic and lubrication systems such that fluids reservoirs are provided within the casing 6010.
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According to alternative embodiments, the seals 6210, 6230, 6250 and grooves 6220, 6240 do not extend continuously around the gate 6080, but instead are formed by two sets of seals and grooves, one set being disposed on the inlet side of the gate 6080 and one set being disposed on the outlet side of the gate 6080.
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The operation of the seal 6200 is described with reference to
According to various alternative embodiments, additional seals like the seals 6210, 6230, 6250 and corresponding vents like the vents 6220, 6240 may be disposed along the leakage path between the first of such seals and the last of such seals, which results in a plurality of drain vents 6220 back to the inlet and/or a plurality of pressurized vents/grooves 6240, with seals separating the different ones of the vents/grooves 6220, 6240. According to various embodiments, the total number of such seals along the leakage path may comprise from 3 to 50 seals.
According to alternative embodiments, the first seal 6210 and vent 6220 may be eliminated so that the mechanical seal 6200 relies on the pressurized groove/vent 6240 to discourage leaks across the seal 6200. According to alternative embodiments, the third seal 6250 and vent/groove 6240 are eliminated, so that the mechanical seal 6200 relies on the vent 6220 to discourage further leakage past the seal 6230.
According to various embodiments, a flywheel may be added to one or both ends of the drive shaft 6020 to reduce torsional loads on the shaft 6020 during operation of the compressor 6000.
According to various embodiments, any of the components or features (e.g., hydrostatic bearing 1300, mechanical seal 1500, compression of multi-phase fluids, etc.) of any of the above-described compressors (e.g., compressors 1000, 2000, 3000, 4000, 5000, 5150, 6000) may be used in any of the other compressors described herein. For example, the discharge manifold 1160 may be mounted to the outlet side 154 of the gate casing 150 of the compressor illustrated in
The presently preferred embodiments could be modified to operate as an expander. Further, although descriptions have been used to describe the top and bottom and other directions, the orientation of the elements (e.g. the gate 600 at the bottom of the rotor casing 400) should not be interpreted as limitations on embodiments of the present invention.
While various of the above-described embodiments comprise a rotary compressor that relies on a rotor that is rigidly mounted to a drive shaft so that the rotor and drive shaft rotate together relative to the compression chamber, various of the above-discussed features may be used with other types of compressors (e.g., rolling piston, screw compressor, scroll compressor, lobe, liquid ring, and rotary vane compressors) without deviating from the scope of these embodiments or the invention. For example, the above discussed hydrostatic bearing arrangement 1300 can be incorporated into a variety of other types of compressors that use moving gates/vanes (e.g., rolling piston compressors, rotary vane compressors, etc.) without deviating from the scope of such embodiments or the invention.
While the foregoing written description of various embodiments of the invention enables one of ordinary skill to make and use what is considered presently to be the best mode thereof, those of ordinary skill will understand and appreciate the existence of variations, combinations, and equivalents of the specific embodiment, method, and examples herein. The invention should therefore not be limited by the above described embodiment, method, and examples, but by all embodiments and methods within the scope and spirit of the invention.
It is therefore intended that the foregoing detailed description be regarded as illustrative rather than limiting, and that it be understood that it is the following claims, including all equivalents, that are intended to define the spirit and scope of this invention. To the extent that “at least one” is used to highlight the possibility of a plurality of elements that may satisfy a claim element, this should not be interpreted as requiring “a” to mean singular only. “A” or “an” element may still be satisfied by a plurality of elements unless otherwise stated.
This application claims the benefit of U.S. Provisional Application Ser. No. 62/139,884, filed on Mar. 30, 2015, the content of which is hereby incorporated herein by reference in its entirety.
| Number | Date | Country | |
|---|---|---|---|
| 62139884 | Mar 2015 | US |
| Number | Date | Country | |
|---|---|---|---|
| Parent | 15563061 | Sep 2017 | US |
| Child | 16566657 | US |
