FIELD OF THE INVENTION
The present invention relates in general to the control of the flow of pressurized fluids through industrial and commercial piping systems, and in particular to a dynamic variable device for dampening pressure and flow pulsations passing through these systems, especially to systems that include one or more reciprocating (piston-type) compressor cylinders with variable operating conditions.
BACKGROUND OF THE INVENTION
Reciprocating compressors typically include one or more pistons that “reciprocate” within closed cylinders. They are commonly used for a wide range of applications that include, but are not limited to, the pressurization and transport of air, natural gas, and other gases and mixtures of gases through systems that are used for gas transmission, distribution, injection, storage, processing, refining, oil production, refrigeration, air separation, utility, and other industrial and commercial processes. Reciprocating, compressors typically draw a fixed mass of gaseous fluid at a relatively lower pressure from a suction pipe and, a fraction of a second later, compress and transfer the fixed mass of fluid into a discharge pipe at a relatively higher pressure.
The intermittent mass transfer within reciprocating compressor systems produces complex time-variant pressure waves, commonly referred to as pulsations. The pulsations are affected by the compressor operating speed, temperature, pressure, and thermodynamic properties of the gaseous fluid, and the geometry and configuration of the reciprocating compressor and the system to which it is connected. For example, a reciprocating compressor cylinder that compresses gas on only one end of its piston, referred to as a single-acting compressor, produces pulsation having a fundamental frequency that is equal to the compressor operating speed. Similarly, a reciprocating compressor cylinder that compresses gas on both ends of its piston, referred to as a double-acting compressor, produces pulsation having a fundamental frequency that is equal to twice the compressor operating speed. In addition, the compressor cylinders and piping systems have individual acoustic natural frequencies that affect the magnitude and frequencies of the combined pulsations throughout the system.
These pressure pulsations travel as waves through an often complex network of connected pipes, pressure vessels, filters, separators, coolers and other system elements. They can travel for many miles until they are attenuated or damped by friction or other means that reduce the dynamic variation of the pressure.
The pulsations may excite system mechanical natural frequencies, cause high vibration, overstress system elements and piping, interfere with meter measurements, and affect compressor thermodynamic performance. These effects can severely compromise the reliability, performance and structural integrity of the reciprocating compressor and its connected system, as well as flow meters and other compressors connected to the system.
Therefore, effective reduction and control of the pressure and flow pulsations generated by reciprocating compressors, both upstream (i.e., the suction side) and downstream (i.e., the discharge side) of the compressor, is necessary for safe and efficient operation. Current technology involves creating a detailed model of the compressor and its system that is used to predict its pulsation behavior at the specified operating conditions, which are often variable. When such modeling predicts pulsations, associated shaking forces, and component stresses that are judged to be beyond safe limits, based on accepted industry guidelines, sound engineering analysis and/or practical experience, it is customary to employ a system of pulsation attenuation elements.
Common pulsation attenuation elements include pulsation bottles (expansion volumes, often containing internal baffles, multiple chambers and choke tubes), external choke tubes, additional pulsation bottles, and fixed orifice plates installed at specific locations in the both the suction and discharge side of each compressor cylinder. These prior art pulsation attenuation devices can be used singly or in combination to dampen the pressure waves and reduce the resulting forces to acceptable levels. These devices typically accomplish pulsation attenuation by adding resistance to the system. This added resistance causes system pressure losses and energy losses both upstream and downstream of the compressor cylinders. The pressure and energy losses typically increase as the frequency of the pulsation increases, and these losses add to the work that must be done by the compressor to move fluid from the suction line to the discharge line.
Of the aforementioned pulsation attenuation elements, fixed orifice plates are one of the most common elements employed. They have the advantages of relatively easy installation and low cost. They may be used at multiple locations throughout the system. The fixed orifices are typically thin metal sheets having a round hole of a specified diameter, located at the center of the flow channel (usually a pipe) cross-section. The orifice diameter is generally 0.5 to 0.7 times the inside diameter of the pipe in which it is installed. However, smaller and larger diameter ratios are sometimes used. The orifice plate is retained between two adjacent pipe flanges that are held together with multiple threaded fasteners and sealed with gaskets to prevent gas leaks. Once the flanges are installed the orifice plates remain fixed in place, and can only be removed or changed by safely stopping the compressor, completely venting all gas to atmospheric pressure, loosening all the threaded fasteners, removing the original orifice plate, installing a new orifice plate with new gaskets, re-assembling and tightening the threaded fasteners, purging the system to remove air, pressurizing the system with gas and restarting the compressor.
In the majority of applications, compressor operating conditions vary with time, with the variables being speed, suction pressure/temperature, discharge pressure/temperature, displacement, effective clearance volume, and even the gas composition. Operating condition variations may be gradual over time, but are more often intermittent, changing frequently to higher or lower levels as dictated by the demands of the application. Some applications, e.g., natural gas transmission and gas storage, have extreme variations in operating conditions over time. In fact, the majority of reciprocating applications require operation over a wide speed range of conditions as well as multiple flow rates that range from very low flows to very high flows.
Fixed orifice plates are effective in reducing pulsations over a narrow compressor operating range, however they cause an associated pressure drop that adds to the work and power consumption required by the compressor. The system pulsation control design is almost always a compromise between pulsation control and pressure drop or power penalty. For example, a very restrictive (low diameter ratio) fixed orifice plate may be required to adequately dampen pulsations at certain operating conditions. However, at other operating conditions, the pulsations might be acceptable with a less restrictive (larger diameter ratio) fixed orifice plate or possibly with no orifice plate at all. In addition, a fixed orifice plate that controls pulsations with a tolerable pressure drop and power penalty at some conditions, may cause excessive damping, pressure drop and power penalty at other conditions.
There are therefore multiple challenges when trying to achieve pulsation control with pulsation bottles and fixed orifice plates. A typical disclaimer by the pulsation control designer states that, “Orifice and choke tube diameters are selected to provide the optimum pulsation dampening and pressure drop over the entire operating range of the unit. Typically, the predicted pressure drop levels for the compressor will range from at or below American Petroleum Institute Specification No. 618 (API 618) allowable levels at normal and low flow conditions to above API 618 allowable levels at high flow conditions. Additionally, the pulsation dampening will be generally good at normal and high flow conditions, but may be marginal to poor at certain frequencies when operating at the minimum flow conditions.”
Although a fixed orifice plate having a specific diameter may be necessary and effective for pulsation control at one set or range of operating conditions, it may be unnecessary, ineffective, and/or the cause of unacceptably high pressure drop and associated power consumption at other ranges of operating conditions. Therefore, it would be advantageous to change one or more fixed orifice plate diameters as operating conditions change.
As noted above, fixed orifice plates are commonly placed between two mating flanges that are held together with multiple threaded studs and nuts and sealed with gaskets to prevent leakage of process gas to the atmosphere. Optionally, they may be permanently welded into the inside of the piping or other flow passage. Accordingly, the downtime, labor and lost production required for changing fixed orifice plates make this alternative impractical. As a result, compressor systems tend to run with higher pressure and power losses or with higher pulsation induced vibration, and associated risk, than would be optimal if the orifice size could be changed when dictated by operating conditions. In many cases the range of operating conditions has to be reduced or limited to restrict the operation of the compressor system.
In light of the above, there is therefore a need for a practical device that can change the effective orifice resistance to maintain acceptable pulsation control with minimal pressure drop and power consumption as operating conditions change. There is also a need for a device and a means that could quickly and easily change the effective diameter (or flow restriction) while the compressor is pressurized and operating. Such a device would enable the optimal and safe control of pulsations, while minimizing power consumption.
SUMMARY OF THE INVENTION
Accordingly, the present invention relates to a pulsation dampening apparatus which provides the ability to adjust its effective orifice size or restriction. The inventive pulsation dampening apparatus can also be referred to herein as a “dynamic variable orifice” (DVO). The invention provides a practical means of changing the effective orifice sizes to optimal values in response to changing compressor operating conditions. The DVO can be adjusted while the compressor is operating and pressurized, and allows a user to increase or decrease the effective orifice size or restriction. The orifice size of the DVO can be adjusted manually with a wrench or hand crank, or automatically with the assistance of an electrical, pneumatic or hydraulically powered actuator or motor. The power-assisted adjustment may be controlled by a human operator, or by an automatic control system programmed to automatically adjust the orifice size as operating conditions change.
A first aspect of the invention provides a pulsation dampening apparatus for providing a variable effective orifice size for a reciprocating compressor, the pulsation dampening apparatus comprising: (a) a fixed inner conical cage including a plurality of inner conical cage ports; (b) a rotatable outer conical cage including a plurality of outer conical cage ports; and (c) a central cylindrical port created by alignment of the inner conical cage and the outer conical cage about a central axis, wherein the inner conical cage and the outer conical ca have mating contours allowing the rotatable outer conical cage to slide over the fixed inner conical cage as it rotates about the central axis, rotation of the outer conical cage causing the plurality of inner conical cage ports and the plurality of outer conical cage ports to be selectively aligned, the relative alignment of the plurality of inner conical cage ports with the plurality of outer conical cage ports determining the effective orifice size of the apparatus.
A second aspect of the invention provides a pulsation dampening apparatus for providing a variable effective orifice size for a reciprocating compressor, the pulsation dampening apparatus comprising: (a) a fixed inner conical cage including a plurality of inner conical cage ports; (b) a rotatable outer conical cage including a plurality of outer conical cage ports; (c) a central cylindrical port created by alignment of the inner conical cage and the outer conical cage about a central axis, wherein the inner conical cage and the outer conical cage have mating contours allowing the rotatable outer conical cage to slide over the fixed inner conical cage as it rotates about the central axis, rotation of the outer conical cage causing the plurality of inner conical cage ports and the plurality of outer conical cage ports to be selectively aligned, the relative alignment of the plurality of inner conical cage ports with the plurality of outer conical cage ports determining the effective orifice size of the apparatus; and (d) a bevel gear drive including a shaft having rotatable gear teeth, the outer conical cage further including a flange including fixed gear teeth which engage the rotatable gear teeth, wherein rotation of the rotatable gear teeth causes the outer conical cage to be rotated, rotation of the outer conical cage causing a change in the orientation of the plurality of outer conical cage ports with respect to the plurality of inner conical cage ports, thereby allowing a adjustment of the apparatus to any desired effective orifice size.
The nature and advantages of the present invention will be more fully appreciated from the following drawings, detailed description and claims.
BRIEF DESCRIPTION OF THE DRAWINGS
The accompanying drawings illustrate embodiments of the invention and together with a general description of the invention given above, and the detailed description given below, serve to explain the principles of the invention.
FIG. 1 is a side cross-sectional view of a 3D representation of a conical embodiment of the apparatus of the invention;
FIG. 2 is an expanded cross-sectional view of the area within the square frame shown in FIG. 1;
FIG. 3 is an expanded cross-sectional view as viewed from the lop of a conical embodiment of the apparatus of the as having a plurality of plate ports;
FIGS. 4A through 4E show a series of top views inside the inner conical cage of the embodiment of FIG. 3, showing the flow passage openings in the plate ports as the inner conical cage is rotated from a fully closed (4A) to a fully open (4E) position;
FIG. 5 is a side cross-sectional view of a 3D representation of a flat, disc-like embodiment of the apparatus of the invention. FIG. 6 is an expanded cross-sectional view of the area within the square frame shown in FIG. 5;
FIG. 7 is an expanded cross sectional view as viewed from the top of a flat, disc-like embodiment of the apparatus of the invention having a plurality of plate ports;
FIGS. 8A through 8E show a series of top views inside the upper flat plate of the embodiment of FIG. 5, showing the flow passage openings in the plate ports as the upper flat plate is rotated from fully closed (8A) to fully open (8E);
FIG. 9 is an isometric representation of the suction system for a reciprocating compressor for the case study of FIGS. 11A and 11B;
FIG. 10 is an isometric representation of the discharge system for a reciprocating compressor for the case study of FIGS. 11A and 11B;
FIGS. 11A and 11B are a tabulation of data comparing a case study of a reciprocating compressor with common and optimal pulsation orifice sets operating at three different operating conditions;
FIG. 12 is a perspective view of one embodiment of a bevel gear drive according to the present invention;
FIG. 13 is an up close cross-sectional side view of the bevel gear drive according to the present invention;
FIG. 14 is a perspective view of one embodiment of the dynamic variable orifice of the present invention showing a motor mount for the inventive bevel gear drive retained by flange bolting;
FIG. 15 is a perspective view of the embodiment of FIG. 14, in which a positioning motor is attached to the motor mount to power the bevel gear drive;
FIG. 16 is a perspective cross-sectional side view of the bevel gear drive according to the present invention.
DETAILED DESCRIPTION OF THE INVENTION
The present invention relates to an apparatus for controlling/adjusting the effective orifice size or restriction of a pulsation control orifice for a reciprocating compressor. Termed a dynamic variable orifice apparatus or DVO, the inventive pulsation dampening apparatus provides a practical means for varying the effective orifice sizes to optimal values in response to changing operating conditions within the reciprocating compressor.
The DVO allows a user to control the pressure and flow pulsations generated by reciprocating compressors while the compressor is operating and pressurized. It can be adjusted manually with a wrench or hand crank, or with the assistance of an electrical, pneumatic or hydraulically powered actuator or motor. The power-assisted adjustment may be controlled by a human operator or by an automatic control system that is programmed to set the required orifice setting as operating conditions change.
One embodiment of a conical-shaped Dynamic Variable Orifice apparatus (DVO) of the present invention is shown in FIG. 1. The DVO can be installed as a complete assembly between adjacent bolted flanges, similar to a typical fixed orifice, except that the DVO assembly is significantly thicker than a typical flat plate fixed orifice. The bolted flanges are typically ANSI standard flanges; however, they may be other standard flanges or special non-standard flanges. A first gasket or seal 5 can be positioned between the top of the DVO assembly and the upper bolted flange 6. Similarly, a second gasket or seal 30 can be positioned between the bottom of the DVO assembly and the lower bolted flange 7. These two gaskets or seals prevent leakage of high pressure process gas to the atmosphere. The gaskets 5, 30 can be made of a malleable material that can be “crushed” (as is known in the art) by the force created by multiple threaded studs or bolts and nuts (not shown) that are tensioned by torque wrenches or mechanical means, or by the force created by other mechanical means (such as, but not limited to, clamps) in order to create a seal which prevents leakage of high pressure gas to the atmosphere.
FIGS. 1-4 depict various views and details of a conical-shaped Dynamic Variable Orifice apparatus (DVO) of the present invention, while FIGS. 5-8 depict various views and details of a flat, disc-shaped DVO. The conical embodiments of the DVO as shown in FIGS. 1-4 includes a rotatable inner conical cage 1 (FIGS. 1, 2), 51 (FIGS. 3, 4) and a fixed outer conical cage 2 (FIGS. 1, 2). 52 (FIGS. 3, 4). Viewing FIG. 3, wherein a plurality of ports for creating a plurality of openings 32, 33, 34, 35 between the rotatable inner conical cage and the fixed outer conical cage are shown, it can be appreciated that a common, central cylindrical port 31 is formed by the inner and outer cortical cages 51, 52 being aligned about a central axis A-A.
Also, looking at FIG. 1, it can be appreciated that the inner and outer conical cages 1, 2 have mating contours allowing the outer surface of the rotatable inner conical cage 1 to rotatably slide over and/or along the surface of the fixed outer conical cage 2 as it rotates about this central axis A-A. Rotation of the inner conical cage 1 relative to the fixed outer conical cage 2 causes their respective ports 9, 10 to be selectively aligned with one another. Therefore, the ports 9, 10 can be aligned in any configuration to create any size opening or effective orifice size for a pulsation control device within a reciprocating compressor.
In use, flow enters the large internal diameter of the inner conical cage and progresses through the smaller internal diameter of the central cylindrical port 31 (see FIG. 3) created by the inner conical cage 51 and the outer conical cage 52. Looking at FIG. 1, the relative alignment of the port 9 of the rotatable inner conical cage 1 with the port 10 of the fixed outer conical cage 2 determines opening area or the effective orifice size or restriction of the pulsation control device. Looking at FIG. 4, when the alignment between the two conical cages 51, 52 is out of line, such that the opening between the ports 59, 60 in the inner and outer conical cages 51, 52 is substantially closed, as shown in FIG. 4A, all of the flow must pass through the central cylindrical port 31. This minimum position creates the smallest effective orifice size and the highest resistance to flow.
In a typical application, the DVO apparatus would be designed to have a “built-in” Beta ratio, defined as the effective orifice size of the DVO divided by the internal diameter of the flow channel or pipe into which the DVO is placed. At the minimum position described above the built-in Beta ratio would be equivalent to 0.4. However, the DVO could be designed with a built-in minimum Beta ratio as low as about 0.3 or lower, and to as high as about 0.7 or higher. As shown in FIG. 4B, rotating the inner conical cage 51 in a first direction (e.g. clockwise) relative to the fixed outer conical cage 52, gradually increases the area of the openings 32, 33, 34, 35 (see also FIG. 3) between the plurality of cage ports 59, 60 of the inner and outer conical cages to permit flow to pass through the cage ports, as well as through the central cylindrical port 31. This increases the effective orifice size to a Beta ratio that is larger than the minimum built-in Beta ratio and reduces the resistance to flow. Further clockwise rotation of the inner conical cage, as shown in FIGS. 4C and 4D, further increases the plate port areas and the effective orifice size to larger and larger Beta ratios, further reducing the resistance to flow. In the limiting case, the maximum position occurs when the inner conical cage is rotated to a position where its ports are in line with the ports of the outer conical cage, causing the port areas to be fully open (see FIG. 4E). This maximizes the effective orifice size and Beta ratio and minimizes the resistance to flow. In a typical application, the DVO would be designed with a “built-in” maximum Beta ratio of about 0.7. However, the DVO could be designed with a built-in maximum Beta ratio of as high as about 0.9 to a low of about 0.5 or lower. Any configuration of the ports of the inner conical cage relative to the ports of the outer conical cage can be applied, thus providing any effective orifice size.
As can be seen in FIG. 1, the inner conical cage 1 is typically rotated in one direction within the outer conical cage 2 to reduce the effective orifice size, and in an opposite direction to increase the effective orifice size. Looking at FIG. 3, the inner conical cage 51 contains a flange 38 having gear teeth 17 located in a section of its lower rim that engage helical teeth 16 in a drive gear and shaft 15. As the drive gear and shaft 15 is rotated by a mechanical means, the helical teeth 16 engage the gear teeth 17 in the rim of the upper conical flange 39 to cause the inner conical cage 51 to be rotated so as to change the orientation of the ports 59 in the inner conical cage 51 with respect to the ports 60 in the outer conical cage 52. This allows the user to change the effective flow area or orifice area of the DVO while the compressor is operating and pressurized. An upper locator bushing 12 (see FIGS. 1 and 2) and a lower locator hushing 13 (see FIG. 1) align and position the inner conical cage 1 to the outer conical cage 2, providing radial and axial support and alignment, preventing vibratory motion of the inner conical cage, and maintaining a small clearance between the cages to prevent metal-to-metal contact, wear, and excessive resistance to rotation of the inner conical cage.
Looking at FIG. 3, one or more mechanical stops, markers or notches 18 may be located at another position in the rim 39 of the flange 38 of the inner conical cage 51. These notches 18 can be used to locate or measure the angular position of the inner conical cage 51 for orienting its ports 59 relative to the corresponding ports 60 in the outer conical cage .52, and therefore to adjust the flow area through the openings 32, 33, 34, and 35 of the plate ports between the inner and outer conical cages. As illustrated, port 59 of the inner conical cage lines up with port 60 of the outer conical cage 52 to form opening 34. Openings 32, 33 and 35 are also formed by the ports 59 and 60 (not labeled over these openings) of the inner and outer conical cages 51, 52, respectively. One embodiment of the invention utilizes one or more mechanical stops or markers 18, which may include, hut are not limited to, step changes in the flange lower rim diameter, or metal pins affixed to protrude radially from the flange lower rim, or shallow holes drilled radially into the flange lower rim. The location of the markers 18 may be measured by an electronic sensor(s) mounted in one or more sensor ports 23, 24, 25, 26, 27, which are located within the flange 3 of the outer conical cage 52 to determine the angular position of the inner conical cage 51 as it is being rotated to a new position by the, gear and shaft assembly 15.
In another embodiment, fixed mechanical stops or markers 18 embedded in the flange 38 of the inner conical cage (not shown) contact a pin, step or other mechanical means of limiting rotational travel of the inner conical cage 51 to a predetermined position. In this embodiment, the DVO is limited to positions corresponding to the limits imposed by fixed mechanical stops 18.
In yet another embodiment, the lower rim 39 of the flange of the inner conical cage 51 may contain a notch (not shown) in the shape of a “v” groove, slot, hole or other geometric form. An external detent actuator (not shown), controlled by electrical, pneumatic or hydraulic or manual mechanical means, contains a pin that engages the “v” groove, slot, hole or other geometric form to prevent rotation of the inner conical cage 51. The pin can be withdrawn from such engagement with the “v” groove when it is necessary to rotate the inner conical cage 51 to a new position, and then reinserted when the new position is reached to hold the inner conical cage in the new position.
Looking at FIG. 1, the outer conical cage 2 contains an integral flange 3 which typically includes one or more extensions (e.g., 37). One extension may be used for mounting the external actuator or the electronic sensors (not shown). Another extension 37 may be used for mounting a pneumatic, electrical or hydraulic actuator or motor 11 or other means to rotate the drive gear and shaft assembly 15. The drive gear and shaft assembly 15 may be rotated in a clockwise or a counter-clockwise direction, either manually with a wrench engaging opposing flats on the shaft, or with a hand crank attached either permanently or temporarily to the shaft, or with an electrical, pneumatic or hydraulically powered actuator or motor that engages the drive end of the gear and shaft assembly 15.
As can be appreciated by viewing FIG. 3, the drive gear and shaft assembly 15 is held in position radially and axially by at least two bushings or bearings 19, 20. One bushing or bearing 20 is held within a cylindrical bore in the flange 3 of the outer conical cage and the other bearing or bushing 19 is held in place by a hearing holder 21 that is inserted into a cylindrical bore in the flange 3 of the outer conical cage. The bearing holder 21 may be secured by threads that engage it with threads in the cylindrical bore in the flange 3 of the outer conical cage or by bolts, snap ring or other mechanical means. A seal 29 prevents leakage of high pressure gas to the atmosphere. A rotary shaft seal 22, held in place by a retainer 36 prevents high pressure gas from leaking along the shaft and gear assembly 15 to the atmosphere.
As shown in FIG. 2, the flange 38 of the inner conical cage 1 is positioned within a shallow bore in the flange 3 of the outer conical cage 2. A top plate 4, connected to the flange 3 of the outer conical cage by three or more threaded cap screws (not shown), captures the flange 38 of the inner conical cage to position it axially in the assembly. A seal 8 prevents the leakage of high pressure gas through the joint between the top plate 4 and the flange 3 of the outer conical cage 2 to the atmosphere. A minimum of three to a maximum of about twelve spring-energized support pads 13 are used to provide a limited axial preload three which holds the inner conical cage 1 in an axial position against the bushings 12, 13 within the assembly, but permits rotation when needed to change the effective orifice area. The support pads 13 are constructed of corrosion resistant metallic bearing material, such as bronze, brass, tin-plated or lead-plated aluminum or brass, or composite sintered metals, or a non-metallic bearing material, such as filled-Teflon, PEEK, or other suitable material. A helical spring 14 under each support pad is compressed within the assembly to provide a suitable axial three that holds the inner conical cage in position, but permits rotation when it is necessary to rotate the inner conical cage to a different position to change the effective orifice area. A contaminant barrier 28 may be used to prevent the accumulation of dirt, rust, liquid or other contaminants in the gas stream from accumulating around the gear teeth 16, 17 (FIG. 3).
In a different embodiment (not shown) the functions of the contaminant barrier 28 and the support pads 13 may be combined into a single non-metallic ring that is compressed by multiple helical springs 14, or by a single wafer spring, or by other type of springs.
The flat, disc-like embodiment of the DVO as shown in FIGS. 5-8 includes a rotatable upper flat plate 201 and a fixed lower flat plate 202. Viewing FIG. 7, wherein a plurality of plate ports 259, 260 for creating a plurality of openings 232, 233, 234, 235 between the upper and lower flat plates are shown, it can be appreciated that a common central cylindrical port 231 is formed by the upper and lower flat plates 201, 202 being aligned about a central axis B-B.
Also, looking at FIG. 5, it can be appreciated that the upper and lower flat plates 201, 202 have mating contours allowing the rotatable upper plate 201 to rotatably slide over the fixed lower plate 202 as it rotates about this central axis B-B. Rotation of the upper plate 201 relative to the fixed lower plate 202 causes their respective plate ports 209, 210 to be selectively aligned with one another. Therefore, the plate ports can be aligned in any configuration to create any size opening area or effective orifice size for a pulsation control device within a reciprocating compressor.
In use, flow enters the large internal diameter of the upper flat plate 201 and progresses through the smaller internal diameter of the central cylindrical port 231 (see FIG. 7) created by the upper flat plate 201 and the lower flat plate 202. As shown in FIG. 7, the relative alignment of the upper plate ports 259 of the rotatable upper flat plate 201 with the lower plate ports 260 of the fixed lower fiat plate 202 determines opening area or the effective orifice size or restriction of the pulsation control device. When the alignment between the two windowed plates is out of line, such that the opening between the upper and lower plate ports 259, 260 in the upper and lower flat plates is substantially closed, as shown in FIG. 8A, all of the flow must pass through the central cylindrical port 231. This minimum position creates the smallest effective orifice size and the highest resistance to flow.
As shown in FIG. 8B, rotating the upper flat plate 201 in a clockwise direction relative to the lower flat plate 202, gradually increases the area of the openings 232, 233, 234, 235 (see also FIG. 7) between the plurality of plate ports of the upper and lower flat plates to permit flow to pass through the plate ports, as well as through the central cylindrical port 231. This increases the effective orifice size to a Beta ratio that is larger than the minimum built-in Beta ratio and reduces the resistance to flow. Further clockwise rotation of the upper flat plate, as shown in FIGS. 8C and 8D, further increases the plate port areas and the effective orifice size to larger and larger Beta ratios, further reducing the resistance to flow. In the limiting case, the maximum position occurs when the upper flat plate is rotated to a position where its plate ports are in line with the plate ports of the lower flat plate, causing the plate port areas to be fully open (see FIG. 8E). This maximizes the effective orifice size and Beta ratio and minimizes the resistance to flow.
The upper flat plate 201 is typically rotated in one direction with respect to the lower flat plate 202 to reduce the effective orifice size, and in an opposite direction to increase the effective orifice size. Looking at FIG. 7, the upper flat plate 201 contains a flange 238 having gear teeth 217 located in a section of its lower rim that engage helical teeth 216 in a drive gear and shaft 215. As the drive gear and shaft 215 is rotated by a mechanical means, the helical teeth 216 engage the gear teeth 217 in the rim of the upper flange 239 to cause the upper flat plate 201 to be rotated so as to change the orientation of the ports 259 in the upper flat plate with respect to the ports 260 in the lower flat plate 202. This allows the user to change the effective flow area or orifice area of the DVO while the compressor is operating, and pressurized. An upper locator bushing 212 (see FIGS. 5 and 6) and a lower locator bushing 213 (see FIG. 5) align and position the, upper flat plate 201 to the lower flat plate 202, providing radial and axial support and alignment, preventing vibratory motion of the upper flat plate, and maintaining a small clearance between the cages to prevent metal-to-metal contact, wear, and excessive resistance to rotation of the upper flat plate.
Looking at FIG. 7, one or more mechanical stops, markers or notches 218 may be located at another position in the rim 239 of the flange 238 of the upper flat plate 201. These notches 218 can be used to locate or measure the angular position of the upper flat plate 201 for orienting its ports 259 relative to the corresponding ports 260 in the lower flat plate 202 to adjust the flow area through the openings 232, 233, 234, and 235 of the plate ports between the upper and lower flat plates. As illustrated, ports 259 of the upper flat plate line up with the ports 260 of the lower flat plate to form opening 233. Openings 232, 234 and 235 are also formed by the ports 259 and 260 (not labeled over these openings) of the upper and lower flat plates. One embodiment of the invention utilizes one or more markers, which may include, but are not limited to, step changes in the flange lower rim diameter, or metal pins affixed to protrude radially from the flange lower rim, or shallow holes drilled radially into the flange lower rim. The location of the marker(s) may be measured by an electronic sensor(s) mounted in one or more sensor ports 223, 224, 225, 226, 227 located within the flange 203 of the fixed lower flat plate 202 to determine the angular position of the rotatable upper flat plate 201 as it is being rotated to a new position by the gear and shall assembly 215.
In another embodiment, fixed mechanical stops embedded in the flange 238 of the upper flat plate (not shown) contact a pin, step or other mechanical means of limiting rotational travel of the upper flat plate 201 to a predetermined position. In this embodiment, the DVO is limited to positions corresponding to the limits imposed by fixed mechanical stops.
In yet another embodiment, the lower rim 239 of the flange of the upper flat plate 201 may contain a notch (not shown) in the shape of a “v” groove, slot, hole or other geometric form. An external detent actuator (not shown), controlled by electrical, pneumatic or hydraulic or manual mechanical means, contains a pin that engages the “v” groove, slot, hole or other geometric form to prevent rotation of the upper flat plate 201. The pin can be withdrawn from such engagement with the “v” groove when it is necessary to rotate the upper flat plate 201 to a new position, and then reinserted when the new position is reached to hold the upper flat plate in the new position.
Looking at FIG. 5, the lower flat plate 202 contains an integral flange 203 which typically includes one or more extensions (e.g. 237). One extension may be used for mounting the external actuator or the electronic sensors (not shown). Another extension 237 may be used for mounting a pneumatic electrical or hydraulic actuator or motor 211 or other means to rotate the drive gear and shaft assembly 215. The drive gear and shaft assembly 215 may be rotated in a clockwise or a counter-clockwise direction, either manually with a wrench engaging opposing flats on the shaft, or with a hand crank attached either permanently or temporarily to the shaft, or with an electrical, pneumatic or hydraulically powered actuator or motor that engages the drive end of the gear and shaft assembly 215.
As can be appreciated by viewing FIG. 7, the drive gear and shaft assembly 215 is held in position radially and axially by at least two bushings or bearings 219, 220. One bushing or bearing 220 is held within a cylindrical bore in the flange 203 of the lower flat plate and the other bearing or bushing 219 is held in place by a bearing holder 221 that is inserted into a cylindrical bore in the flange of the lower flat plate. The bearing holder 221 may be secured by threads that engage it with threads in the cylindrical bore in the flange 203 of the lower flat plate or by bolts, snap ring at other mechanical means. A seal 229 prevents leakage of high pressure gas to the atmosphere. A rotary shaft seal 222, held in place by a retainer 236 prevents high pressure gas from leaking along the shaft and gear assembly 215 to the atmosphere.
As shown in FIG. 6, the flange 238 of the upper flat plate 201 is positioned within a shallow bore in the flange 203 of the lower flat plate 202. A top plate 204, connected to the flange 203 of the lower flat plate by three or more threaded cap screws (not shown), captures flange 238 of the upper flat plate to position it axially in the assembly. A seal 208 prevents the leakage of high pressure gas through the joint between the top plate 204 and the flange 203 of the lower flat plate 202 to the atmosphere. A minimum of three to a minimum of about twelve spring-energized support pads 213 are used to provide a limited axial preload force which holds the upper flat plate 201 in an axial position against the bushings 212, 213 within the assembly, but permits rotation when needed to change the effective orifice area. The support pads 213 are constructed of corrosion resistant metallic bearing material, such as bronze, brass, tin-plated or lead-plated aluminum, or brass, or composite sintered metals, or a non-metallic bearing material, such as filled-Teflon, PEEK, or other suitable material. A helical spring 214 under each support pad is compressed within the assembly to provide a suitable axial force that holds the upper flat plate in position, but permits rotation when it is necessary to rotate the upper flat plate to a different position to change the effective orifice area. A contaminant barrier 228 may be used to prevent the accumulation of dirt, rust, liquid or other contaminants in the gas stream from accumulating around the gear teeth 216, 217 (FIG. 7).
Case Studies: The following case studies provide insight into the problems faced with the current use of prior art fixed orifice plates, and provides a quantification of the risks or disadvantages associated with having fixed orifice diameters versus the benefits or advantages of variable orifice diameters.
The compressor in this case study is a common is industrial reciprocating compressor that is commonly used throughout the natural gas compression industry. The compressor has four “throws” oriented in a horizontally opposed arrangement with two throws on each horizontal side of the crankcase. A common four-throw crankshaft with a 5.5 in. stroke drives each of the four throws. The compressor is driven through a flexible coupling by a natural gas reciprocating engine rated at 1680 horsepower at 1200 rpm. About 180 horsepower is consumed to drive auxiliary equipment, leaving 1500 horsepower available for driving the compressor at the 1200 rpm maximum rated speed. The engine and compressor can operate at continuous speeds of 900 to 1200 rpm. A double acting compressor cylinder having a bore diameter of 8.75 in. is mounted on each of the four compressor throws, and the system is configured such that the four cylinders operate in parallel.
The compressor is applied in an application that collects gas from multiple gas wells and pressurizes it for transport through a pipeline for processing and eventually to sales. Over the life of the application, the inlet, or suction, pressure will vary with time as individual gas wells come on and off line in an often unpredictable manner. In addition, the suction pressure will trend to lower levels over longer periods of time as the gas wells mature and production volumes and pressures decline. In order to accommodate the wide range of operating conditions within the rated limits of the compressor and the gas engine driver, the operating speed, suction pressure, volumetric clearance and number of active compressor ends have to be varied, often by means of automatic controls. This type of application is very common, and the design of an optimal pulsation control system is not only very challenging, it can be impossible to design a single fixed system that satisfactorily accommodates the entire operation range that is specified for the application. In this case the end user provided a total of eighteen different operating conditions that defined the wide range over which the system was required to operate.
FIG. 9 is an isometric drawing of the suction piping and pulsation control system that was designed for this application. The supply line 101 to the compressor cylinders 106, 107, 108, 109 splits into two branches 102, 103. Each branch feeds a three-chambered suction pulsation bottle 104, 105 that bridges the suction of two cylinders 106, 107 and 108, 109 on one side of the compressor. A fixed pulsation dampening orifice or Suction Restrictive Orifice (SRO-1) is placed between the pipe flange 110 and the inlet connection flange 111 on the suction pulsation bottle 104. The flow goes through the fixed orifice (SRO-1) into the first of the three chambers inside the three-chambered suction pulsation bottle 104. The first chamber 112 is connected to each of two other chambers 113, 114 by internal pipes (not shown) that serve as choke tubes to create volume-choke-volume acoustic filters. Each of the other two internal chambers 113, 114 is centered over a compressor cylinder 106, 107 and connected to the cylinder suction flange with a short riser pipe. Fixed pulsation dampening orifices (SRO-2) are placed between the riser flange and the cylinder suction flange for each cylinder. An identical configuration is used on the opposite side of the compressor for the other two cylinders.
FIG. 10 is an isometric drawing of the discharge piping and pulsation control system that was designed for this application. Fixed pulsation dampening orifices or Discharge Restrictive Orifices [DRO-1] are placed between a three-chambered discharge pulsation bottle riser flange and the cylinder discharge flange for each cylinder. Each of the short risers feeds into a separate internal chamber that is centered below a compressor cylinder. Each internal chamber is connected to an end chamber of the three-chambered discharge pulsation bottle by an internal pipe that serves as a choke tube to create a volume-choke-volume acoustic filter. A fixed pulsation dampening orifice [DRO-2] is placed between the outlet connection flange 115 on the discharge pulsation bottle 117 and the pipe flange 116. The flow goes through the fixed orifice [DRO-2] into a pipe branch 118 that joins a branch from an identical configuration on the opposite side of the compressor to a common outlet or discharge line 119.
As is customary practice, the compressor and piping system was modeled and analyzed over the range of operating conditions to determine the pulsations throughout the system. For the sake of brevity, the results of analyzing only three of the eighteen specified operating conditions are presented in FIGS. 11A and 11B.
Case 1 is a 1200 rpm operating point with all four cylinders in double acting mode, but with volumetric clearance added to each head or lower cylinder end to reduce the capacity to a rate of 86.5 million standard cubic feet per day (MMSCFD).
Case 3 is a 1084 rpm operating point with three of the four cylinders in single acting mode (i.e., suction valves removed or disabled to allow gas to bypass them, leaving only the crank or frame end of the cylinder able to compress gas) and with the fourth cylinder in double acting mode, but with volumetric clearance added to the head or lower end of that cylinder to reduce capacity to a rate of 58.0 MMSCFD.
Case 8 is a 1200 rpm operating point with all four cylinders in double acting mode with no volumetric clearance added to the head or lower cylinder end for a capacity of 149.9 MMSCFD. This provides maximum capacity from the compressor.
As is customary with the current state of the art, a common set of fixed pulsation control orifices was selected for all operating conditions. The common set consists of 5.50 in. diameter orifices for [SRO-1], 3.75 in, diameter orifices for [SRO-2], 350 in, diameter orifices for [DRO-1], and 4.25 in. diameter orifices for [DRO-2].
The data in FIGS. 11A and 11B shows that a common set of fixed pulsation control orifices is far from optimal. The set was selected to provide best overall performance at Operating Case 1, which is the highest power condition of the cases shown. With the common set of fixed orifices, the suction (from the suction header to the compressor suction flange) and discharge (from the compressor discharge flange to the discharge header) pressure drops are 1.96% and 1.93%, respectively. The suction and discharge pulsations are controlled to 1.9% and 1.3% of the line pressure, respectively, and the associated power consumed by the suction and discharge pressure drops is 2.60%.
A more optimal set for Operating Case 1 controls the suction and discharge pulsations to 2.2% and 1.4%, respectively, which were acceptable for that case. The larger diameter orifices in the optimal set resulted in suction, and discharge pressure drops of 1.53% and 0.99%, respectively, with an associated power consumption of 1.69%. The savings translates to $7.35 in driver fuel cost per day, based on a fuel cost of $3.50/MMBTU. If the compressor were to operate at this operating condition all the time, with the assumption of the industry norm of 96% availability, use of the optimal orifice set would result in annual fuel savings of $2,575.44.
Operating Case 3 provides an example of a different issue that occurs with the use of a common set of fixed pulsation control orifices. Case 3 is a low flow condition in which three of the four cylinders are operated in single acting mode. Single acting cylinder operation generally creates a more difficult pulsation control challenge. Power losses with the common set are 1.45%; however, the pulsation control is not adequate. Suction and discharge pulsations with the common set are 11.8% and 5.8%, respectively. These are unacceptably high and result in a high risk of pulsation related vibration, meter measurement problems and other safety and reliability problems upstream of, within and downstream of the compressor system. An optimal set of pulsation control orifices for Operating Case 3 result in suction and discharge pulsations of 7.2% and 5.6%, respectively. Although these are still higher than would be preferred, they are substantially better than the common orifice set and they represent the best practical alternative for this operating condition without more drastic redesign of the system. The resulting power consumption increases to 2.60%, however that is a reasonable premium for reducing the risk of pulsation related reliability problems.
Operating Case 8 provides an example of another problem associated with using a common set of fixed pulsation control orifices in a compressor that must operate over a wide range of flow conditions. At Operating Case 8, the common orifice set controls suction and discharge pulsations to 0.5% and 0.2%, respectively. This exceptional pulsation control comes with a significant power cost, however, for this low pressure ratio operating case, as the resulting power consumption is 11.06%. A more optimal set of pulsation control orifices for Operating Case 8 results in a power consumption of 3.02%. Suction and discharge pulsations remain very low, even with the larger optimal larger diameter orifice set. The power savings translates to $58.86 in driver fuel cost per day, based on a fuel cost of $3.50/MMBTU. If the compressor were to operate at this operating condition all the time, with the assumption of the industry norm of 96% availability, use of the optimal orifice set would result in annual fuel savings of $20,624.12.
In the foregoing Case Study, without the benefit of the present invention, the options are limited to: (1) restricting the compressor operation to a limited operating range, i.e., a low flow of about 60 MMSCFD to a high flow of about 80 MMSCFD with the use of the common set of fixed plate orifices, or (2) to frequently stop the compressor, vent the system to atmospheric pressure, physically unbolt ten sets of bolted flanges to change the fixed orifice plates to sets that are more optimal for the intended operation, reassemble the ten sets of flanges, purge the system to remove air, pressurize the system again, and then restart the compressor.
Option (1) could result in flow being limited by as much as 69.9 MMSCFD, or the difference between the desired 149.9 MMSCFD maximum capacity and the 80 MMSCFD limit imposed on the unit due to use of the fixed orifices. Based on a $3.50/MMBTU gas price, this lost production opportunity would be nearly $14,000 per day. Option (2) is generally not a practical alternative because of its high cost, its labor intensity, the environmental impact from the more frequent venting of gas that contains methane (a green house gas) and volatile organic compounds from the system to the atmosphere, and the fact that flow conditions are not always predictable, or controllable, which could pose a risk to operational safety. Assuming, however, that such a change could be tolerated and the fixed orifice plates could be changed out in one 24 hour period, based on a wellhead natural gas price of $3.50/MMBTU, the typical lost production alone would be at least $12,000 for the unit in this case study. This does not include the cost of labor and material required for changing the orifice plates.
In the foregoing Case Study, the use of a dynamic variable orifice (DVO) of the present invention at each of the ten orifice locations would provide a practical means of expanding the compressor flow range from a low flow of 40 MMSCFD to a high flow of 120 MMSCFD while achieving effective pulsation control and reasonable pressure drop associated power consumption.
Although the preferred application of the present invention as explained herein is for the dampening of pulsations within reciprocating compressor systems, there are other applications for the present invention. These can include, but are not limited to, any compressor, pump, metering or piping systems containing a gaseous fluid, liquid, or bi-phase fluid where pulsation dampening is required, or where variable flow control is necessary or beneficial.
IMPROVEMENTS: The following paragraphs describe new features and improvements relating to the inventive pulsation dampening apparatus or DVO (dynamic variable orifice) illustrated above. These new features result from further development, testing and value engineering done subsequent to the filing of parent U.S. application Ser. No. 14/602,515. In particular, a novel radially-positioned bevel gear drive has been developed for adjusting the effective orifice size of the DVO. While achieving the same result as the original gear drive through a different means, the bevel gear drive described below and illustrated in new FIGS. 12-16 can reduce the cost, complexity and limitations of the inventive pulsation dampening device. More specifically, the updated bevel gear drive includes a rotating pinion gear and shaft mechanism that can fit between adjacent flange bolt holes. The bevel gear drive can be bolted to any flange size, thus overcoming the limitations of the original “worm” type gear drive and shaft assembly illustrated, e.g., in FIGS. 1 and 3 and described in paragraph [0045] above.
The original gear drive and shaft assembly was found to be a limiting feature when using the inventive DVO pulsation dampening apparatus in large, high pressure applications requiring more substantial flanges and larger flange sizes. For example, looking at FIG. 3, it can be appreciated that two (2) circular bolt holes (unlabeled) are located within the flange 3 of the outer conical cage 52, on either side of (or above and below) the gear and shaft assembly 15. A rotary shaft seal 22, held in place by a retainer 36, prevents high pressure gas from leaking along the shaft and gear assembly 15 to the atmosphere. As the number and/or the diameter of the bolt holes gets larger in the larger flange sizes, the spacing between these two (2) circular bolt holes (unlabeled) located in the flange 3 of the Outer conical cage 52 becomes smaller, leaving insufficient thickness between adjacent holes to fit the shaft assembly 15.
As a result, when the initial worm gear drive and shaft assembly 15 shown in FIGS. 1 and 3 is used to adjust the effective orifice size in applications requiring flange sizes for piping measuring 12 inches in diameter or larger, perceptible interference can occur between the gear and shaft assembly 15 and the two (2) circular bolt holes (unlabeled) located within the flange 3 of the outer conical cage 52, on either side of (or above and below) the gear and shaft assembly 15. The interference is a result of the proximity of the cylindrical bore for the shaft 15 (which is drilled between the two bolt holes in the flange 3) to the bolt holes, such that there is insufficient thickness of flange material between the cylindrical bore for the shall 15 and the bolt holes. During bolting of the flanges this flange material may be distorted or otherwise altered such that gas under pressure may leak through. As the flange size increases to about a 14-inch pipe size, the geometric arrangement is such that no flange material remains between the cylindrical bore for the shaft 15 and the bolt holes.
For example, when using standard ANSI/ASME B16.5-2017 flanges, indicated for piping larger than 12 inches in diameter, the spacing between the two adjacent bolt holes is too small for a straight, tangentially-positioned worm gear drive shaft 15 of adequate size to fit between the bolt holes. For reference, standard ANSI/ASME B16.5-2017 flanges have a specified number of bolts, size (diameter) and bolt circle diameter for each pipe diameter and class. The class relates generally to the pressure rating of the flange in pounds per square inch gauge pressure (psig). Smaller flanges such as an 8-inch class 600 ANSI/ASME B16.5-2017 flange, have twelve (12) bolt holes, each 1.25 inches in diameter, on a bolt circle that is 13.75 inches in diameter. Similarly, a small (but more robust) 8-inch class 900 ANSI/ASME 816.5-2017 flange has twelve (12) bolt holes, each 1.50 inches in diameter, on a bolt circle that is 15.50 inches in diameter.
While “small” flanges, such as the class 600 and class 900 flanges noted above for receiving 8-inch pipes, typically do not present an interference problem for the worm gear type shaft 15 of FIG. 3, larger flanges require more bolt holes, and thus less spacing between adjacent bolt holes. As a result, there is insufficient thickness between the bolt holes located within the flange 3 of the outer conical cage 52 to fit the tangentially-received shalt 15 (noted above, see FIG. 3). For example, large flanges for piping measuring 12 inches in diameter (or larger) typically have twenty (20) or more bolt holes, such as the 12-inch class 900 ANSI/ASME B16.5-2017 flange, which has twenty (20) bolt holes, each 1.50 inches in diameter, on a bolt circle that is 21.00 inches in diameter, or the 14-inch class 900 ANSI/ASME B16.5-2017 flange which has twenty (20) bolt holes, each 1.62 inches in diameter, on a bolt circle that is 22.00 inches in diameter.
Thus, while the initial warm gear drive and shaft assembly 15 described above and illustrated, e.g., in FIGS. 1 and 3 can be useful in many applications, the physical size and pressure rating of the flanges to which this gear drive can be applied is limited. More specifically, due to the amount of interference caused in larger applications (i.e. applications using flanges for piping measuring 12 inches in diameter or larger), it was determined that the original gear drive disclosed herein should not be used in large, high pressure applications requiring more substantial flanges and larger flange sizes. Thus, a solution was needed for the problem of using the inventive apparatus in such large, high pressure applications. Value engineering, which is a systematic method employed to improve the function vs. cost ratio of a product by examining and improving its function, has resulted in an embodiment of the inventive pulsation dampening apparatus described herein which utilizes a radially-positioned bevel gear drive to adjust the effective orifice size of the apparatus in large, high pressure applications.
In contrast to the original worm gear drive and shaft assembly 15 described above, the inventive bevel gear drive can be positioned radially about the outer conical cage. Looking at FIGS. 12-16, this bevel gear drive includes rotating pinion gear teeth 302 which can engage a section of fixed gear teeth 304. The fixed set of gear teeth 304 are located on the flange 310 of the outer conical cage 306. Movement of the gear teeth 302, 304 enables the selective rotation of the outer conical cage 306 relative to the fixed inner conical cage 308. As illustrated in FIG. 12, the fixed set of gear teeth 304 can be bolted 346 or otherwise mechanically secured to the flange 310 of the outer conical cage 306. As illustrated in FIG. 13, a hushing 324 holds the shaft 322 in position within the body 326 which is attached with multiple cap screws to the flange 310 of the inner conical cage 308. The bushing 324 holds the shaft 322 in position within a plug 330. Two elastomeric o-rings 328 seal the gear end of the shaft 322 within the bore of the bushing 324. The plug 330 is shown threaded into a mating hole that has been machined into the body 326, which is attached to the flange of the inner conical cage 308, and can be sealed with another elastomeric o-ring, such as a plug o-ring 332.
Note that in the original worm gear drive embodiment illustrated in FIGS. 1 and 3, the inner conical cage is rotated relative to a fixed outer conical cage, while the new bevel gear drive described above rotates the outer conical cage relative to fixed inner conical cage. This change between rotating a particular conical cage versus the other is a result of the different gear drive arrangement being, used, i.e. the tangentially-positioned worm drive seen best in FIG. 3 versus the radially-positioned gear drive shown in FIGS. 12-16. Comparing FIGS. 1 and 3 to new FIG. 13, it can be appreciated that the large extension 37 seen in FIGS. 1 and 3, which is required to extend the original gear drive and shaft assembly 15 beyond the flange where it can be accessed for actuation, is eliminated. Specifically, the extension 37 of FIGS. 1 and 3 is replaced as shown in FIG. 13 by a plug 330 which contains the bevel gear drive including the rotating pinion gear teeth 302 which can engage the fixed gear teeth 304 on the flange 310 of the outer conical cage 306.
The outer end of the shaft 322 of FIGS. 12-16 can have a hexagonal shape, or hex 333 machined onto the shaft 322 for the purpose of complimentarily engaging a manual wrench or the drive of a small actuator or motor. This hex end 323 can be used for turning the shaft 322 when needed to re-position the outer conical cage 306 relative to the inner conical cage 308. While the hex end 323 of the shaft 322 has a hexagonal shape as illustrated, the shaft 322 is not limited to this hex shape; the shaft end 323 may also be in the shape of a square, spline or other geometric shape as is known in the art for such fits.
Looking at FIG. 13, the plug 330 can include a coarse pitch thread 334 machined onto its outer diameter. This coarse thread 334 can have as few as 3 threads per inch up to as many as 7 threads per inch, and is engaged by the mating internal thread of an indicator sleeve 336. An indicator insert 338 engages the indicator sleeve 336 and is retained by two set screws 340 to form an indicator sleeve assembly (336, 338, 340). The inner diameter of the indicator insert 338 can include an internal (female) shape that fittingly engages the external (male) hex end 323 of the shaft 322. Thus, as the pinion gear teeth 302 are rotated by a rotating shaft 322, the indicator sleeve assembly 336, 338, 340 also rotates and moves along the thread 334 of the plug 330, thereby changing the indicator insert's 338 linear position relative to the plug, and thus relative to the shaft 322. The hex end 323 of the outer end of the shaft 322 can include markings or lines 342 inscribed or etched onto its surface. These markings 342 are placed along the linear path of the indicator insert as it moves along the shaft 322, and they can relate to the position of the ports (windows) of the rotatable outer conical cage 306 relative to the ports of the fixed inner conical cage 308. With proper design and assembly, this arrangement provides a visual indication to a user of the relative positions of the ports in the two conical cages, and accordingly, it indicates the Beta ratio setting of the DVO pulsation dampening apparatus.
As best seen in FIGS. 13 and 16, the new bevel gear drive can include an elastomeric o-ring 320, which can function as a spring. This spring 320 can provide an axial preload force sufficient to hold the rotatable outer conical cage 306 in an axial position within the assembly, while permitting rotation of the outer cage 306 when needed to change the effective orifice size. In the event of excessive vibration or other unintentional forces or torques occurring within the pressurized system, the spring 320 also provides sufficient friction to prevent the outer conical cage 306 from unintentionally rotating relative to the fixed inner conical cage 308.
When used in place of the initial worm gear drives multiple spring-energized support pads (see 13 in FIG. 2) and accompanying springs (14 in FIG. 2) to provide the limited axial preload force needed, the new gear drive embodiment using the o-ring spring 320 described above can result in a reduction of the requisite thickness of the inventive apparatus, therefore reducing the cost as well. Thickness is an important and limiting feature, because the inventive DVO pulsation dampening apparatus must often be mounted directly onto a compressor cylinder inlet (i.e., suction) or outlet (i.e., discharge) flange. Increased thickness may require an increase in the length of the “riser” needed between the compressor cylinder flange and the pulsation bottle. Space and length are typically significant design constraints in high pressure systems, having an effect on cost. Further, an increased riser length can affect the acoustic natural frequency of the riser, which can alter the nature and size of the pulsations being controlled by the apparatus.
As shown in FIGS. 16, a laser-etched indicator plate 344 can be secured to the outer perimeter of the DVO body 326 with screws, pins or adhesive means. The indicator plate 344 includes a tabulation of the Beta ratio corresponding to each marking or line 342 inscribed on the hex end 323 of the shaft 322. From the tabulation on the indicator plate 344 an operator can read the correct mark for the desired Beta ratio. If manually adjusting the DVO, the operator can use a wrench to rotate the rotatable pinion gear teeth 302 clockwise or counter-clockwise until the outer face of the indicator insert 338 in the indicator sleeve 336 aligns visually with the proper marking or line 342 on the hex end 323 of the shaft 322. As noted above, this gear drive embodiment is an alternative to the worm gear drive described for FIGS. 14, and can replace the electronic sensor ports 23-27, mechanical stops and notches 18, and detent actuator illustrated in FIG. 3 and described in paragraph [0046] above. The novel bevel gear drive assembly can simplify the construction and reduce the size and cost of the inventive apparatus.
With the development and use of the radially-positioned gear drive shown in FIGS. 12-16, the bulky and costly motor mounting extension 37 on the flange 3 of the DVO (see FIG. 1) is no longer required, although a version of it could still be used if necessary for a specific application. For typical applications, the less expensive and more versatile motor mount 314 retained by the flange bolting 318, see FIGS. 14 and 15, can be used when a positioning motor 316 is required. The motor mount 314 and motor 316 are optional and only required when an electrical, pneumatic or hydraulically powered actuator or motor is needed for automatic operation of the DVO. As noted above, the orifice size of the DVO can also be adjusted by manually rotating the gear drive with a wrench, or hand crank. Thus, operation of the inventive DVO gear drive may be controlled by a human operator, or by an automatic control system programmed to automatically adjust the orifice size as operating conditions change.
In addition to the development of the novel gear drive above for rotating the conical cages of the inventive DVO, a manufacturing method has been developed to reduce the cost and complexity of the fixed gear teeth 304 used to rotate the outer conical cage 306. This method initially forms the required gear teeth 304 as a complete circular ring, having gear teeth all the way around its top face. The circular ring is then cut into segments of appropriate length, and each cut gear teeth segment 304 is then bolted or secured by other mechanical means 346 onto the flange 310 of the outer conical cage 306 in the correct location and orientation. A complete circular ring can be cut into such segments to make three to five gear teeth segments 304, depending on the amount of cage rotation required. The amount of conical cage rotation needed for a certain application is a function of the size of the conical cages being used and the range of Beta ratios needed for that application.
While the present invention has been illustrated by the description of embodiments sand examples thereof, it is not intended to restrict or in any way limit the scope of the appended claims to such detail. Additional advantages and modifications will be readily apparent to those skilled in the art. Accordingly, departures may be made from such details without departing from the scope of the invention.
Patent Application
20180313343
