Patent Application
20240384720
This disclosure relates to positive displacement machines that convert energy; for example, positive displacement expanders comprising rotors that rotate in a single rotational direction to release a working fluid contained in operating chambers of the rotors. In another example, the rotors of the rotary positive displacement device may be rotationally reversible, acting as a compressor.
References herein to patent documents or other matter is not to be taken as a statement that the document or matter was known or that the information it contains was part of the common general knowledge as at the priority date of any of the claims.
Mechanical machines are devices that convert energy to achieve a desired objective. Energy is the ability to produce or create work, while work is the process in which the energy transfer takes place, for example, a force applied to an object makes it move in the direction of the force. An engine or electric motor connected to a compressor or pump may boost the energy of a fluid, for example, so that the fluid may travel down a pipeline. A fluid is comprised of a mixture of gases and/or liquids.
Potential and kinetic energy of a fluid may be reduced in order to provide useful work via an expander or liquid motor. In one example, an electric motor is used to drive a screw compressor, which continually fills a vessel with compressed air at a comparably elevated pressure. Those knowledgeable in the art understand that the air could be subsequently released to reversibly operate the screw compressor as a screw expander. This simple form of Compressed Air Energy Storage (CAES) could allow a facility owner to purchase comparably low cost electricity to drive the screw compressor when there is a surplus on the electrical grid and later release the high pressure air to produce higher price electricity when the price is high. Energy storage strategies help encourage wide-spread adoption of intermittent renewable power generation. Similarly, an electric motor may be used to pump liquid to a higher elevation. Some pumps may be reversibly operated as a “liquid motor” with the reversibly operated electric motor producing electricity as an electrical generator.
Two generic principles for gas compression involve the use of positive displacement devices or dynamic compression devices. Dynamic compression is accomplished from rotating blades on an impeller accelerating gas to a high velocity which is subsequently transformed into static pressure in a diffuser where it is decelerated. Dynamic expansion occurs in “turbines” where gas expands across blades to rotate blades at high velocity and low drag to produce useful work. Compression and expansion equipment can be incorporated on the same shaft, for example in a gas turbine. For example, a jet engine uses dynamic compression to pull in air, subsequent internal combustion of a fuel-air mixture and finally useful work is produced through a turbine to drive the compressor on the same shaft. Remaining kinetic energy in the exhaust produces forward thrust. Another example of dynamic compression is integrally geared centrifugal compressors. This ‘open flow’ type of compression/expansion can be well suited to applications that have high and near constant volumetric throughput with a relatively low and near constant compression/expansion ratio with only small variations in gas composition expected and little to no expected liquid formation in gas applications. A large number of blades and stages can be used with equipment in the 10 s or even 100 s of MW.
The applicants of U.S. patent Ser. No. 11/242,726 explain that “To use a single turbine and have adequate efficiency over an entire range of ambient conditions is problematic. It has been found that use of two or more turbines in series or parallel which are optimized for different ambient conditions addresses the problem.”
Dynamic compression devices can face additional challenges including “surge” where a reduction in volumetric throughput can lead to flow reversal. Very generally speaking, it is the applicant's understanding that dynamic compression/expansion equipment may be more complicated and therefore may be more expensive to build than positive displacement devices in many applications below 1-5 MW. There are exceptions such as centrifugal pumps and axial fans, both of which may be relatively simple devices that work well in applications with high volumetric throughput and low compression ratios. In applications in the range of approximately 10 to 90 MW or approximately 100 to 900 MW, it is the applicant's understanding that dynamic compression/expansion devices may often be scalable and therefore see a reduction in cost per MW. At this power range, positive displacement compression/expansion devices would typically require duplication of equipment.
Positive displacement compression/expansion devices allow a fluid to enter a chamber, then seal the chamber from the upstream/downstream piping, before allowing the chamber to be in fluid communication with the downstream piping. Accordingly, the upstream and downstream pressure can vary significantly across a single stage and relatively high pressure rated designs are available. Devices such as reciprocating piston-cylinder devices require valves and packing glands, both of which can be major sources of internal and external leakage, respectively. Leakage past packing glands is a major source of greenhouse gas emissions in the natural gas compression industry. Valves are designed to open/close at a specific operating pressure and as the actual operating pressure or gas composition changes, inefficiency and eventual failure may occur. The portion of volume in the chamber that cannot be evacuated is often referred to as the “clearance volume” and the relatively high pressure gas that is inefficiently recirculated is often referred to as the “recirculated volume”. The piston in reciprocating piston-cylinder devices does not come into contact with the ends of the cylinder. This characteristic and the geometry at the valves both introduce a significant clearance volume. In other words, the final compression “stroke” of the piston cannot evacuate all of the high pressure gas remaining in the cylindrical chamber before it is subsequently mixed with the lower pressure intake gas. A large clearance volume can dominate the recirculated volume, but internal leakage from higher pressure to lower pressure, for example at the gap between the piston and cylinder, may also contribute. Application engineers have hundreds of combinations of frame, cylinder and valve configurations to choose from for a given application.
For example, much of the natural gas pipeline network in Alberta and in North America ranges from 1000-1440 psig in pressure, with some upstream pipeline pressures being at a much lower pressure. Although natural gas production wells can initially start above pipeline pressure, and therefore a compressor isn't required initially, as the natural gas is released from the reservoir, the pressure at the surface typically eventually approaches near atmospheric pressure. When the pressure upstream of the compressor decreases, typically a producer must decide between installing new, larger, cylinders (i.e. with new valves) and/or investing in upstream compression to keep the current reciprocating compressor utilized well. Most natural gas compressors in service are driven by reciprocating natural gas engines which have very limited turn-down capabilities to efficiently adjust to the lower expected power requirement of the compressor. Some producers opt for inefficient bypass methods where recompression is required to keep the engine loaded up. In other cases, a significant investment is made towards (upstream) booster compression, with the intention of keeping the existing compression closer to the original design conditions, including volumetric throughput.
Rotary positive displacement devices such as screw compressors are also common since their widespread adoption approximately 50 years ago. Since the chamber itself moves past stationary porting, no valves are required, substantially increasing reliability. Some screw compressors may better accommodate varying pressure ratios by slowly adjusting the position of “slide gates”. The simple rotary motion of the screw rotors alleviates a lot of vibration concerns that reciprocating compressors may have associated with continuously accelerating and decelerating mass. Oil-flooded screw compressors/expanders allow higher pressure lubricating oil to bleed into the gas, allowing these devices to be zero-emissions in operation. Liquids have a much higher heat capacity than gases. This fluid mixture makes much higher compression/expansion ratios possible in a single stage compared to a reciprocating compressor/expander. Reciprocating equipment typically requires 2-3 stages of compression in many natural gas compression applications, with inter-cooling after each stage to manage the temperature increase of the gas. Oil-flooded screw compressors have been proven to reliably operate in applications with meaningful volumetric throughput that have upstream and/or downstream pressures ranging from 0 to 350 psig. Some of these combinations would produce excessively high gas temperatures in a single stage of compression if lubricating oil were not used to manage the heat of compression. Since the screw rotor shafts can only be supported by bearings at the ends of the shafts, excessive chamber pressure causes the shafts to bend which eventually causes contact between the rotors and housing and/or excessive internal leakages. Recirculated gas volumes introduced by the slide gates and high pressure axial port can also substantially reduce efficiency. This efficiency is particularly reduced when lubricating oil is not used to help seal, which is often the case in closed-loop cycles such as the Organic Rankine Cycle (ORC), which uses a refrigerant as the working fluid.
The ORC uses a heat exchanger to allow a source of geothermal or waste heat to fully boil the working fluid. The refrigerant remains in a vapor state both before and after expansion as it produces useful work, making both turbines or screw expanders examples of appropriate equipment. It is the applicant's belief that when above 1 MW of power may be generated, the higher efficiency and lower cost option is generally turbines, while below 1 MW, screw expanders may be more attractive, largely due to their low upfront costs. After the expander, heat is rejected to the environment to fully condense the refrigerant into a liquid so that it can be subsequently pumped up to high pressure to be boiled again. It will be appreciated that more power is thermodynamically available to be harnessed when there is a larger temperature differential between the heat source and heat sink (i.e. environment). A useful first calculation to perform is the Carnot efficiency equation using the temperatures in Kelvin:
This known equation represents the maximum efficiency, “η”, that a hypothetical heat engine may have in operation between the relatively higher temperature heat source, “Thot”, and the lower temperature heat sink, “Tcold”. In reality, the practically attainable efficiency will be noticeably lower, but this equation may be a good starting point to demonstrate the potential approximate relative benefit of attempting to track the variance in ambient temperature (i.e. “Tcold”) versus employing the industry standard of designing the system to run like it is always a hot summer day.
If a high grade waste heat source (i.e. >650° C.) is available, then rejecting heat at −40° C. instead of 30° C. would increase the Carnot efficiency from 67.2% to 74.7%. Since most of the waste heat sources are low-grade (i.e. <200-277° C.), with many geothermal brine type applications in Alberta, Canada reaching no more than 110° C. at the surface heat exchanger, 110° C. will be used as a second example. In this 110° C. source temperature case, the Carnot efficiency can be increased from 20.9% to 39.1% on a cold winter day. If ambient conditions range from −40° C. to 30° C. and heat exchangers are designed with a 10° C. temperature differential, for example, then the refrigerant would vary from −30° C. to 100° C. in the winter and 40° C. to 100° C. in the summer. The respective Carnot efficiencies would be calculated as 16.1% to 34.8%. Losses in the process and mechanical equipment need to be taken into account for more accuracy, but the difference in Carnot efficiency implies that somewhere in the ballpark of 2× more power could be generated on a cold winter day versus a warm summer day if efficient mechanical equipment existed in the prior art to handle the application range. At lower source temperatures, a larger variance in seasonal power generation could be expected.
Additional power may be generated with a larger temperature differential by using a larger pressure ratio in the power generation cycle. Turbines require additional rows of blades since each row of blades can only withstand a limited pressure differential. Some speculate that the efficiency would still be adequate if two turbines are used, since the turbines could be fed in series or parallel depending on the season, but even the theoretical performance seems to be unclear.
Screw expanders are believed to be the dominant choice for ORC applications below 1 MW. It is believed that the lower upfront cost of dry screw expanders is a significant selling feature, despite the efficiency being lower than for turbines, at least when comparing performance in a near constant operational range. In one example, a dry screw expander may be around 60-65% (isentropic) efficiency with a turbine being around 85% efficient in the same near constant operation range. Dry screw expander stages are operated in series or parallel. Sometimes just a single stage is required, which is typical in oil-flooded screw compressors/expanders.
In some applications, oil-flooding can improve efficiency by about 10%. However, dry screws seem to be preferred for ORC applications under 1 MW, perhaps partly due to the simplicity and upfront cost savings. Screws with a fixed pressure ratio can offer a relatively high efficiency when operating conditions match the design conditions. However, this lack of adjustability is often not very practical and efficiency penalties result when the chamber becomes in fluid-communication with gas in the adjacent piping either too early or too late. The introduction of slide gates adds sufficient flexibility for many applications. However, the recirculated gas volume that is introduced is a significant source of inefficiency. In any configuration, the screw rotor profile requires an axial port which also contributes to recirculated gas volume. Recirculated gas volume introduces large inefficiency at high pressure-ratios. This inefficiency can be inadequate when attempting to apply a screw expander in the Trilateral Flash Cycle (TFC) or Partial Evaporating Cycle (PEC), which require a relatively smaller chamber volume to be initially sealed in comparison to the Organic Rankine Cycle (ORC). In the ORC, the refrigerant is completely boiled, while in the TFC, none of the refrigerant is boiled; it is still 100% (saturated) liquid phase when expansion begins. The PEC is somewhere between ORC and TFC, with the refrigerant being partially liquid phase and partially vapor phase when expansion begins.
It should be noted that turbines are used in vapor applications where less than 10% by mass of liquids are expected, since these high velocity liquids impinging on the blades cause damage. Turbines are designed for a particular gas or combination of gases, and the efficiency can drop significantly if the gas composition changes. The Kalina Cycle makes use of a mixture of ammonium vapor and water vapor, which have substantially the same molecular weight. Therefore, turbines can maintain their efficiency with varying proportions of ammonium vapor and water vapor; this composition is varied to lower the boiling point and therefore more closely match the ambient temperature when condensing the refrigerant mixture. It is believed that equipment is selected accordingly to match conditions for a warm summer day, with little to no adjustments made for seasonal temperature variations. When compared to the Rankine Cycle, which is linked to the boiling point of pure water, the Kalina Cycle is understood to enable a higher temperature differential by condensing at a lower temperature. As suggested by the examples above, a higher Carnot efficiency is expected. It is believed that the additional complexity, but also high source temperature compatibility of the Kalina Cycle makes the cycle most competitive at high grade (i.e. >650° C.) waste heat sources where the Rankine Cycle would commonly be used. It should also be noted that rotary screw expanders are far less sensitive to gas composition, so they could work well with a variety of gas compositions.
Several power generation cycles have been proposed in the public domain. An exergoeconomic comparison can be done to weigh both thermodynamic and economic viewpoints. Some have speculated that the Trilateral Flash Cycle (TFC) could be the lowest cost per unit of power produced at source temperatures below 120° C. if there were a commercially available expander capable of achieving an 85% isentropic efficiency in the TFC. An example of an exergoeconomic comparison is provided in [https.//ideas.repec.org/a/eee/energy/v83y2015icp712-722.html, Accessed on 5 May 2023], titled “Exergoeconomic comparison of TLC (Trilateral Rankine cycle), ORC (organic Rankine cycle) and Kalina cycle using a low grade heat source”.
The authors of [https://www.sciencedirect.com/science/article/pii/S187661021931224X entitled “Trilateral Flash Cycle (TFC): A Promising Thermodynamic Cycle for Low Grade Heat to Power Generation”, Accessed on 5 May 2023] concluded that “TFC has about at least 50% more power generation potential than that of conventional ORC for the same heat source and heat sink conditions if the heat source temperature is just below 100° C. TFC even has the potential of power generation from the heat source below 80° C. where ORC is not economically viable.” It was also said that the “TFC system can utilize up to 70% of available thermal energy where ORC can only deal with 20%.” Also, “It is obvious that TFC system requires larger heat exchanger and condenser as it deals with larger amount of heat than that of ORC, however, it also requires more pumping power which basically reflects more initial investment which can be compensated by producing additional power.” Additionally, “TFC system can facilitate to cool down the system leaving hot water temperature close to ambient temperature and reduces overall wastage of energy.”
The ORC may be superior in situations where the heat source will be condensed from a vapor to a liquid because this provides the ORC refrigerant with a lot of heat energy at a comparatively high temperature to fully boil the refrigerant. Since so much heat energy is required to boil the refrigerant in the ORC, this is why the utilization of energy becomes so low at source temperatures below 90° C. where no phase change of the heat source is expected. In many locations in the world, temperatures of 90-120° C. can be obtained by drilling several kilometers down and pumping geothermal brine up. The Oil and Gas Industry has already done this in many locations, including in Alberta. In some examples, where oil reservoirs have been depleted over the past decades, mostly hot water is being brought to the surface by Electric Submersible Pumps (ESPs). Eventually the operational costs of supplying that electricity and maintaining that equipment will exceed the profit that can be made from producing the small amount of oil that comes up from the water, justifying the abandonment of the well. Despite the efficiency of the TFC being lower than the ORC, far more net power could be generated in these applications with the finite brine source because the TFC uses far more of the available heat. If the net power output from the TFC is sufficiently higher than the parasitic electrical load of the ESP(s), then it would make sense to continue to run the equipment even after all the oil had been depleted. This would imply that excess electricity could be sold to the grid, or used locally for applications like crypto-currency mining. Geothermal electricity production is a reliable source of renewable base-load power that could become important for countries that are attempting to minimize use of non-renewable sources of power, but are struggling with widespread adoption of intermittent renewable power sources. As an example, many ESP well applications could be in the 50-300 kW range with wells feeding into satellites that could have 1 MW or more of ESPs upstream.
Across many industries, Pressure Control Valves (PCVs) keep the system pressure below a desired upper limit safely or maintain a part of a circuit at a pressure set point. In Flow Control Valve (FCV) applications, a flow set point may be maintained, with the flow calculated based on the gas composition, temperature and/or pressure measurements. This calculation may need to be proven to be accurate enough to meet jurisdictional requirements; for example, Measurement Canada requirements in Canada. In either example, the Control Valve (CV) controls the flow of the fluid by changing its flow path based on the signal from the controller, which changes the temperature, pressure and/or flow. Gases may experience high velocity choked flow, which in some cases could lead to acoustical pulsations and/or premature wear. When the required gas pressure drop is high enough, the expected high temperature drop could lead to freezing. This is commonly mitigated by burning natural gas to provide heat for line heaters upstream of the control valve. In cryogenic gas plants, it is common for a Flow Control Valve (FCV) known as a Joule Thomson Valve or JT Valve to be placed in parallel to a turboexpander (turbine) for startup purposes. While the isenthalpic throttling process does not theoretically steal energy from the gas, the generation of useful work via a turboexpander does result in lower energy and temperature of the gas. Dynamic expansion devices such as turbines may not adequately replace a CV partly because the open-flow design does not allow the flow to be stopped completely. Positive displacement expansion devices known in the prior art also lack the capabilities to be a suitable replacement for CVs. For example, a control valve may require 0-100% volumetric capacity control adjustment over a wide range of expected pressures on the upstream and downstream sides.
In one example, around 40-400 W may be required at remote locations to allow for remote monitoring and to actuate valves. For example, a site may use radio, cellular and/or satellite connectivity to transmit critical data in real-time. Natural gas from wells may sometimes be vented to the atmosphere, for example, when it is being used to actuate valves because there is insufficient power to operate air compressors or electrically actuate valves. Although solar panels and batteries can be a solution at times, these may be an expensive and unreliable option, especially where theft and vandalism are occurring or in northern climates having reduced sunlight, especially in the winter. However, venting natural gas to the atmosphere contributes to increased greenhouse gas emissions, and there exists regulatory and financial incentives in some jurisdictions to reduce or eliminate greenhouse gas emissions. Thus, it is desirable to develop new mechanisms for remote monitoring and actuation of valves at remote wellsites while reducing greenhouse gas emissions.
Cold climates such as Canada face additional challenges when attempting to reduce greenhouse gas emissions because natural gas is commonly burned to produce heat. Electric vehicles, for example as manufactured by Tesla™, were initially deployed with electric resistive heating, with newer models now being released with “heat pump” technology. A “heat pump” itself actually consists of a compressor, which moves a refrigerant through a refrigeration cycle, including a heat exchanger to extract heat from the source. An issue encountered with electric resistive heating in electric vehicles was the significant amount of electrical energy that is required in cold climates to keep the cabin at a comfortable temperature. The significant amount of heat loss and subsequent electricity consumption forced consumers to charge their vehicles more frequently when traveling long distances. Even when it is −40° C. outside, the ambient environment is an abundant source of thermal energy. Since temperature can only flow from higher temperature to lower temperature, heat pump cycles have an internal refrigerant drop below ambient temperature. This refrigerant begins as a liquid-vapor mixture and the environment supplies heat energy to boil the refrigerant until it is a vapor mixture. The amount of heat energy supplied from the environment typically far exceeds the electrical energy required to run the subsequent (isentropic) compression process. Compression boosts the pressure of the refrigerant and brings the temperature above a comfortable cabin temperature. A heat exchanger is used to transfer latent heat from the condensation process of the refrigerant into the comparably cooler cabin air. The resulting high pressure, and higher temperature liquid-vapor refrigerant undergoes isenthalpic expansion via a throttling valve to complete the cycle. The Coefficient of Performance (COP) is defined as the ratio of useful heating (or cooling) provided over the input work required. As an example, if three times more heat is provided to the cabin in comparison to the electricity used to run the compressor, then a COP of 3 is calculated. This would imply that three times less electricity were required for the purpose of heating the cabin, resulting in significantly less electricity usage.
In a similar process to the heat pump cycle, a reversing valve can be used to provide Air Conditioning (AC) to the cabin. Warm cabin air rejects heat to evaporate/boil comparably cooler refrigerant, causing a liquid to vapor phase change. This vapor is boosted in pressure and temperature via isentropic compression before rejecting heat to the comparably lower temperature environment. As the high pressure and temperature refrigerant rejects heat to the environment, it becomes a liquid-vapor mixture. The temperature and pressure drop via isenthalpic throttling before the refrigerant enters the heat exchanger that functions to cool the cabin air. Minimizing wasted energy helps maintain electric vehicle range. U.S. Pat. No. 10,967,702 B2 is assigned to Tesla Inc. and titled “Optimal source electric vehicle heat pump with extreme temperature heating capability and efficient thermal preconditioning”. It is stated that “in the second heating mode, the vehicle thermal management system 300 of
The author of the article titled “How Tesla Heat Pump Works (vs HVAC Heat Pumps)” [https://aircondlounge.com/how-tesla-heat-pump-works/by Yu Chang Zhen, Accessed on 5 May 2023] states that “in a dire situation where the ambient temperature drop below freezing point, the compressor of the Tesla heat pump will switch to an inefficient mode or “lossy mode” to provide a direct one-to-one energy ratio heating to the cabin.” Also stated is that “this heating mode yields a COP of 1, just like an electric space heater. However it keeps the people inside the vehicle warm and toasty even when the outside temperature is 10 degrees Celsius below freezing point.”
Air source heat pump systems used to heat electric vehicles or buildings face a challenging variation in operating conditions. In the fall, spring and summer, ambient conditions may be somewhat close to desired interior temperatures, requiring a modest compression ratio and therefore high operational efficiency and COP in either heating or cooling mode. Cold winter days in Canada; especially those around −40° C., require very high compression ratios. Dynamic compression devices lack the required flexibility, while commercially available positive displacement technologies are quite inefficient at such high compression ratio applications. Screw compressors are valveless and offer some flexibility to adapt to the changing inlet pressure, but the reduction in volumetric throughput from internal recirculated gas volumes and internal leakages makes high compression ratio cases inefficient. There is often a need for low to medium duty cycle capacity where screw equipment can become increasingly less efficient. Slower speed operation allows a noticeably larger percentage of the flow to leak internally from higher pressure to lower pressure, thereby increasing the recirculated volume and decreasing the efficiency. In extreme scenarios, compressors can reach a limit where far too high of compression ratios are attempted compared to resulting recirculated volume. In this scenario, net flow from the high pressure to low pressure side is theoretically possible, contributing to substantial temperature build-up in the compressor.
In one example, the heat pump system in an electric vehicle could have a scroll compressor with shaft rotational speeds ranging from 500 RPM to 9,000 RPM. Scroll compressors can have near zero clearance volume and low recirculated volume, even at low operational speeds, which may demonstrate a substantially linear relationship between operational speed and both flow rate and power consumption. These qualities, combined with relatively quiet operation make them a common choice where the flow rate and implied duty cycle varies significantly. Compression occurs as gas becomes trapped in crescent-shaped pockets between the involute spiral shaped geometries. The volume of the compression chamber decreases as it reaches the center of the assembly, where it becomes in fluid communication with the gas in the high pressure piping via a discharge port. The size and geometry of the outlet port determines the amount of internal compression and there are practical limitations on the achievable compression ratio before this fluid communication occurs. When the desired compression ratio exceeds the limitations of the equipment, the pressure in the chamber can be significantly lower when it first becomes fluidly connected to the gas at the high pressure piping. This premature fluid communication with the chamber and gas in the compressor discharge piping is known as “under-compression” and is expected to introduce increasingly higher inefficiency when higher compressions ratios are attempted. There are practical compression limitations, which are mainly related to the size and geometry of the outlet port and the size, shape and wall thickness of the scrolls themselves. Scroll compressors may be able to generate up to 150 psig in a single stage unit. Scroll equipment is used in relatively low power heat pump, refrigeration and air conditioning applications, but in many other applications multiple units would be required, which may be prohibitively expensive. Scroll compressors have a relatively high upfront cost and are not easily repairable, due to the complex geometry of the components. Inlet valves are not required, but sometimes discharge valves are used. There is an efficiency penalty associated with the flow restriction introduced by adding a discharge valve, but in high compression ratio applications where the upstream and/or downstream pressure(s) vary significantly, larger efficiency penalties might be expected without the discharge valve.
Markets may exist, especially below 1 MW, for a highly efficient, reliable, rotary positive displacement device. Mechanical equipment experts may regard isothermal compressors and isentropic expanders as the most desirable equipment in most applications. The applicant of U.S. Pat. No. 5,674,053 stated that “a single stage compression of gas at ambient to 4000 psig would result in a gas temperature of over 600° C. This temperature exceeds the desired operating temperature of valves, seals and other components in the compressor. To avoid the use of exotic materials, it is often desirable to maintain the gas charge at substantially lower temperatures.” Also understood is the ability for high temperatures to be avoided in high compression ratios in a single stage when comparably high heat capacity liquids are introduced in the gas, which is commonly known as oil-flooding. It will be appreciated that high operational flexibility, low equipment costs, low mechanical overhead, high efficiency and high reliability may be desirable design objectives. It may also be desirable to provide a single device that is capable of a wide range of processes, including a range between near-isothermal processes to near-isentropic processes. The applicants of U.S. Pat. No. 10,975,869 stated that “Oil flooded screw compressors that attempted operation at discharge pressures much higher than 350 psig have experienced wear and/or other mechanical design issues that make such devices unreliable, and thus not widely adopted. Thus, the constraint for high discharge pressures with screw compressors has been mechanical design, and not a lack of pursuit or understanding of the appeal of very high compression ratios in combination with high discharge pressures in a single stage and with controlled temperatures.” Dry gas compression is understood to substantially lack liquid and solid entrainment and therefore multiple stages of compression are typically required. For example, a screw compressor may be capable of compressing atmospheric pressure “0 psig” gas to up to 350 psig by using three stages. It is understood that the gas is typically cooled with heat exchangers after each stage of compression. Oil-flooded screw compressors may be capable of compressing “0 psig” gas to up to 350 psig in a single stage of compression without producing excessive temperatures because the higher heat capacity of the liquid oil and amount used significantly impact the temperature that the gas reaches as it rejects heat to the oil. It is understood that the “pressure ratio” is defined as the ratio of the pressure at the high pressure side of a compressor to its pressure at the low pressure side, using stagnation pressures. Given typical ambient conditions this example may be calculated with a pressure ratio of around
Expansion is the reverse process of compression, using the same calculation of “pressure ratio”.
In North America, reciprocating compressors are widely used along much of the natural gas gathering pipelines when the natural gas is boosted above 350 psig (i.e. above the pressure rating that screw compressors have been proven to be reliable at). A relatively large number of gathering pipelines may feed into relatively fewer and larger diameter transmission pipelines, before many distribution pipelines are used to deliver natural gas to end users. As an example, many natural gas gathering and transmission pipelines in North America are rated for 1440 psig (i.e. pounds per square inch gauge) of pressure. In other examples, the pressure in a pipeline can exceed 1440 psig. The pressure of newly drilled natural gas wells can often exceed this maximum rating and therefore be reduced via a throttling valve, which is also known as a choke valve, before the gas may be safely incorporated. As the natural gas is expended from the well, the pressure decreases accordingly and can approach 0 psig or even less at the surface, eventually requiring a compressor to boost the pressure enough to maintain natural gas production. For example, the pressure within the gathering pipeline may range from 0 to 1440 psig. In another example, the pressure within a gathering pipeline may range from 0 to 1600 psig or more. For example, when a compressor is initially installed, a pressure ratio of just over 1 may be required, but over time this required pressure ratio could be 110 or more, which, to the knowledge of the Applicant, no prior art equipment has been shown to address successfully.
As an example, the pressure within a transmission pipeline may range from 400 psig to 1440 psig. As another example, the pressure at the end of a transmission pipeline may range from 400 psig to 900 psig and the pressure in a subsequent distribution pipeline may range from 0.25 psig to 100 psig, with generally lower pressures at the end of the respective pipelines. Higher pressures are associated with higher densities and therefore less volume is required in the pipelines and relatively smaller diameter pipelines may be used. A pressure differential is required to flow the gas through the pipeline at a desired flow rate, also accounting for friction that the gas experiences with the inner walls of the pipelines.
Reciprocating compressors have a piston component accelerating and decelerating in a cylindrical housing, making use of pressure-activated valves to allow gas to enter/exit chambers. These valves are custom designed for a particular gas composition and expected pressure and often experience premature failure as conditions change. While relatively expensive modifications may be possible, it is common practice for the natural gas to leak out of a reciprocating compressor where the piston rod rubs at the packing glands. Considerable clearance volume is required where the piston would otherwise slam into the stationary housing at the ends of the stroke and there is little to no capability to reliably introduce liquids for the same concern, since liquids are known to be substantially less compressible than gases and may not exit the limited flow area offered by the valves in time. The clearance volume significantly increases recirculated volume, which is gas that needs to be inefficiently recompressed altogether. Typical applications may include multiple stages of compression with inter-cooling and adjustments to the volumetric throughput are limited and inefficient. While screw compressors may offer improvements to many of these challenges, there still is a need for an improved device that may reliably operate with a pressure rating above 350 psig.
Similar challenges are understood to exist for the reverse (expansion) process, especially at power ratings below about 1 MW where turbines may become less cost-effective to employ. Many applications are in need of a more reliable, efficient and flexible compressor and/or expander; especially one that is capable of high pressure ratios in a single stage. However, one limitation of the prior art is that the pressure ratio that can be reliably and efficiently achieved is limited by significant recirculated gas volume and/or other geometry related constraints. In an extreme example, in relatively high pressure ratio cases, the volumetric throughput reduction associated with recompression of the recirculated volume could actually result in a negative net volumetric throughput. In other words, effort would be exhausted in causing the compressor to boost the pressure of a gas and the net result could be undesirable backflow of said gas. This extreme example demonstrates that there are certain known limitations to the pressure ratio that a compressor can practically achieve, which is understood to be significantly impacted by the recirculated volume of the compressor. High recirculated volume would significantly reduce expander efficiency as well, because this portion of flow may not noticeably contribute to useful work.
As an example, heat pump and power generation cycles ideally match a higher pressure ratio as the temperature differential between the heat source and heat sink temperatures increases. However, no equipment is known in the prior art that may have a combination of low enough recirculated volume (i.e. including internal leakages) and other capabilities to properly address all of the pressure ratios that these cycles ideally desire. It should be noted that scroll equipment can be used in expansion when there is no valve at the high pressure side, but this equipment has not been shown to be appropriate enough to demonstrate widespread use in power generation cycles. Also, to the Applicant's knowledge, no equipment is known in the prior art that may efficiently and reliably accommodate the ideal range of pressure ratio adjustment in heat pump and power generation cycles. In order to address limitations in the prior art, it may be common practice to operate these cycles in suboptimal conditions, for example not generating excess power that may be available in the winter time. Air-source heat pump cycles in the prior art are believed to be capable of boosting ambient temperature air as low as around −15° C. up to around 20° C.-25° C. with acceptable efficiency and in colder ambient conditions, it may be typical for natural gas or electricity to be consumed to still achieve a desired temperature of 20° C.-25° C. for example. In one example, two scroll compressor stages may be used in an air-source heat pump cycle to achieve a COP of 1 at an ambient temperature of −25° while achieving a desired temperature of 20°-25° in a building space. In electric vehicle applications, there may only be enough space for one compressor stage, increasing the need for a compressor that can more efficiently handle the high compression ratios required in a single stage to boost the large temperature differentials. When electric vehicles are driven in cold climates (e.g. −15° C. to −40° C. ambient conditions), the expected driving range between battery charges may be significantly reduced accordingly as a comfortable cabin temperature of 20° C.-25° C. is maintained. A compressor with lower recirculated volume than is known in the prior art may be required in cohort with a wide range of on-the-fly, efficient, pressure ratio adjustability in the heat pump cycle to avoid this supplemental energy wastage.
Some companies state they have a gas expander that acts as power-generating (gas) pressure reducing valve, which is understood to produce power without substantially adjusting the volumetric throughput of the gas. To the best of the knowledge of the applicant, no gas expanders act as a power-generating gas control valve. An expander acting as a power-generating gas control valve, is understood to produce power while adjusting the volumetric throughput of the gas by 0-100%. While such functionality is desirable, to the knowledge of the Applicant, a power-generating gas control valve requires an expanded range of volumetric capacity control that is not disclosed in the prior art expanders. Control valve applications at remote sites along natural gas pipelines may vent natural gas to the atmosphere during operation if no other sources of electricity or pressurized fluid are present. Maintenance operations may sometimes involve releasing natural gas contained in piping and equipment to the atmosphere.
In one aspect of the present disclosure, what is disclosed herein may be particularly useful for applications below 1 MW, where positive displacement equipment is generally the dominant choice; however, this is not intended to be limiting and it will be appreciated that the present disclosure may also be utilized for applications at or exceeding 1 MW.
As indicated by the foregoing examples, there is a need to develop equipment that is capable of high pressure ratios (i.e. over 110:1 or higher as an example) in a single stage in combination with higher pressure ratings (i.e. over 350 psig as an example) and that controlling temperature may be done by introducing liquid into the gas. Also indicated is the practical limitation of sufficiently increasing the pressure ratio without a means of sufficiently reducing the recirculated volume and/or without a means of modifying limiting geometrical constraints. Also desirable is the ability to efficiently accommodate varying pressure ratios and/or ideally a range of 0-100% volumetric capacity control.
The novel Rotary Positive Displacement Device disclosed herein may, in some embodiments, be the first expander capable of efficiently producing power and the full functionality of a gas control valve over an expanded range of pressure ratios. A range of 0-100% volumetric capacity control may also be offered as well as reversible expansion/compression modes without changing internal structures. In control valve applications, this novel Rotary Positive Displacement Device generates power in expansion mode, provides its own seal when the equipment isn't in use and/or compression mode may be used to substantially evacuate fluid from the low pressure side before maintenance is done. The novel Rotary Positive Displacement Device disclosed herein may also reduce the recirculated volume by orders of magnitude beyond anything known in the prior art, which may result in efficiency improvements in different applications and improved pressure ratios and use in heat pump cycles and power generation cycles, including the Trilateral Flash Cycle (TFC) and/or Partially Evaporated Cycle (PEC) where prior art equipment may not have been successfully implemented. As an example, the novel Rotary Positive Displacement Device disclosed herein may be implemented in an air-source heat pump cycle with −40° C. ambient air to achieve interior temperatures of around 20-25° C., in one example, without supplemental energy, for example from the consumption of natural gas or electricity. This may significantly reduce the range loss that electric vehicles face in −40° C. ambient conditions. In other words, the novel rotary positive displacement equipment disclosed herein may be particularly suitable to integrate with the thermal management system and minimize electric vehicle range loss when the cabin requires heating. The novel Rotary Positive Displacement Device disclosed herein contains a minimal number of components and does not require components that are known in some of the prior art equipment to experience significant wear, such as valves or packing glands. Therefore, it is believed that the devices disclosed herein will operate reliably and safely with long component lifespan.
As another example, the novel Rotary Positive Displacement Device may, in some aspects, provide certain advantages in Compressed Air Energy Storage (CAES) applications with the same equipment being reversibly operated as both a compressor and expander to reduce costs and with the high pressure ratio capabilities and flexibility allowing for the efficient filling of a fixed volume storage vessel. High efficiency may also be important in this application as it is expected to improve the economics of such applications. While some exemplary applications have been disclosed, it shall be understood that the previous paragraphs are not an exhaustive list regarding what the novel Rotary Positive Displacement Device may be well suited for.
The compact device disclosed herein, in one example, is particularly advantageous for the ability to efficiently accommodate a large range of upstream and downstream pressures, working fluid composition and desired volumetric throughput in a single stage of expansion and/or compression. In one example, this device may be particularly advantageous at comparably increasing heat output in a heat pump cycle, while significantly reducing electricity usage to conserve range in electric vehicles and/or regulate interior building temperature. In one example, the rotary positive displacement device may generate power and/or useful work while also acting as a control valve, reliably generating power and/or compressed air at remote locations. The device may be particularly advantageous at increasing power generation capabilities in low grade waste heat recovery applications. In one example, the device may generate power in the Trilateral Flash Cycle (TFC) or Partially Evaporating Cycle (PEC) to further cool the heat source, as is desired, in one example, when hot water from an industrial process is being returned to an ecosystem.
The term substantially is used herein to indicate being largely, but not necessarily wholly that which is specified, nearly. The term frusto is used herein to define a section or segment of a geometric surface or shape. As an example, the term frusto-spherical surface is used to define a surface lying on a sphere, but not only a segment of the entire sphere. Therefore, the term frusto defines a portion of the surface of a solid herein. As an example, when referring to the outer surface of the rotors, the term “frusto-spherical” defines a surface lying on a sphere; said spherical surface has regions removed and does not form a continuous sphere. As further examples, a “frusto-conical” surface may lie on a cone and a “frusto-cylindrical” surface may lie on a cylinder.
The term “working fluid” is used herein to denote any combination of gases and/or liquids that are generally directed through the Rotary Positive Displacement Device disclosed herein. It is to be understood that while some solids may also be present in the “working fluid” the conveyance of solids is generally not the intent of the Rotary Positive Displacement Device. The term “compression” is defined herein as the act of increasing the pressure of the working fluid, which is understood to be any combination of gases and/or liquids. The term “compressor” is defined herein as a Rotary Positive Displacement Device that increases the pressure of a working fluid which is comprised of some amount of gas or vapor. The term “expansion” is defined herein as a Rotary Positive Displacement Device that decreases the pressure of a working fluid which is comprised of any combination of gases and/or liquids. The term “expander” is defined herein as a Rotary Positive Displacement Device that decreases the pressure of a working fluid which is comprised of some amount of gas or vapor. The term “pump” is defined herein as a Rotary Positive Displacement Device that increases the pressure of a working fluid which is substantially comprised of liquid. The term “liquid motor” is defined herein as a Rotary Positive Displacement Device that decreases the pressure of a working fluid which is substantially comprised of liquid. Therefore, the term “Rotary Positive Displacement Device”, capitalized to refer to the device(s) disclosed herein, may be used interchangeably with the terms “compressor”, “expander”, “pump” and “liquid motor” to denote the same device.
The term “adiabatic” as used herein, denotes any expansion or compression process involving substantially no transfer of heat to or from a gas contained in the working fluid, respectively.
The term “isothermal” as used herein, denotes any non-adiabatic compression process that derives energetic benefit, including increased efficiency, from the deliberate transfer of heat from a gas contained in the working fluid as well as any non-adiabatic expansion process involving a deliberate transfer of heat to a gas contained in the working fluid. It is to be understood that the deliberate transfer of heat to or from a gas contained in the working fluid may be used to modify the temperature of the gas accordingly, offering advantages in some applications.
The terms “low pressure” and “high pressure” as used herein, denotes any relative difference in working fluid pressure wherein “high pressure” is understood to be at a relatively higher pressure than “low pressure” without any implied limitation regarding the magnitudes of the pressures.
The terms “driver” and “generator” as used herein, may be used interchangeably when referring to “drivers” and “generators” which may be reversibly operated to drive or be driven. As one example, electrical motors are a driver that may be reversibly operated as a generator. Hydraulics, pneumatics and human power are other examples of “drivers and generators”. It is to be understood that use of the term “driver” does not limit use for example when relating to drivers that are not known to be reversibly operated such as engines that combust natural gas, diesel and/or gasoline fuel.
Several examples of the novel Rotary Positive Displacement Device are disclosed herein which include methods for constructing basic sawtooth rotor geometry, advanced sawtooth rotor geometry and customized sawtooth rotor geometries with varying numbers of lobes and lobe tip geometries. Pairs of rotors with frusto-spherical outer surfaces may mesh, with respectively interposing lobes and valleys, inside of the immediately adjacent frusto-spherical inner surfaces of gates and/or stationary housing components, forming chambers there between. Openings formed by gates and/or stationary housing components may be adjustable, for example if gates are used and circumferentially repositioned via actuator assemblies, selectively enabling conveyance of a working fluid through the Rotary Positive Displacement Device. The gates may be further positioned to substantially obstruct outer ports when zero or near zero volumetric throughput is desired. In some embodiments, the full range of 0-100% volumetric throughput capabilities may allow use as a control valve.
Bearings may be placed in close proximity to the chambers to minimize deflection and increase pressure ratings to exceed 350 psig. Since the maximum deflection may be substantially cubically related to the axial distance between the bearings and the radial component of the pressure-induced load acting on the respective rotor, in one example, the pressure rating may be much larger than 350 psig. In one example, the pressure rating may be at or around 1440 psig, which may be useful in addressing applications along natural gas pipelines. In another example the pressure rating may be at or around 5000 psig, which may be useful in addressing markets where the working fluid is substantially hydrogen based. In one example there may be a single driver, which may be a reversible motor/generator or another Rotary Positive Displacement Device for example to provide capabilities of rotating the rotor shaft in either rotational direction to respectively facilitate both compression and expansion capabilities. In another example, motors/generators are coupled on both respective first/second rotor shaft to reduce or even eliminate torque transfer between immediately adjacent rotor surfaces, which may increase component lifespan in the absence of gearing and/or a U-joint for example. In yet another example, a plurality of motors/generators with differing rated speeds may interact with one or both rotor shafts to offer speed adjustment without requiring a costly Variable Frequency Drive (VFD) or transmission. The internal adjustability of the Rotary Positive Displacement Device may allow for such driver choices.
In one aspect of the present disclosure, it is desirable to provide a reliable, high efficiency expander and/or compressor with high operational flexibility, wherein in some embodiments, is capable of efficiently handling varying fluid compositions, pressures and desired flow rates.
In another aspect of the present disclosure, it is desirable to provide a high efficiency expander and/or compressor with low equipment cost and low mechanical overhead in some embodiments, as compared to existing expander and/or compressor devices.
In another aspect of the present disclosure, it is desirable to provide a high efficiency expander and/or compressor that, in some embodiments, may range from being near isothermal to near isentropic.
Through reference to the following description, accompanying drawings and the claims, these and other characteristics, along with advantages and features of the present invention disclosed herein, will become apparent. It is to be understood that the features of the various examples described herein may exist in various combinations and permutations without mutual exclusivity.
In some embodiments, a rotary positive displacement device comprises: a housing having a low pressure port and a high pressure port; a first rotor having a frusto-spherical outer surface, an axial surface, a first rotor shaft and a first rotational axis passing through the first rotor shaft, wherein the axial surface comprises at least one teardrop surface and at least one involute surface, the at least one teardrop surface and at least one involute surface together defining at least one lobe and a corresponding valley; a second rotor having a frusto-spherical outer surface, an axial surface, a second rotor shaft and a second rotational axis passing through the second rotor shaft, the axial surface comprises at least one teardrop surface and at least one involute surface, the at least one teardrop surface and at least one involute surface together defining at least one lobe and a corresponding valley, the second rotational axis intersecting with the first rotational axis; a high pressure opening extending between the first and second rotors and the high pressure port of the housing; and a low pressure opening extending between the first and second rotors and the low pressure port of the housing, wherein the low pressure opening is configured to selectively communicate with one or more chambers of the at least two chambers. The second rotor is configured to intermesh with the first rotor such that at least two chambers are separated by the axial surfaces of the first and second rotors, each chamber of the at least two chambers having a variable volume as the first and second rotors rotate about their respective rotational axes. A lobe tip of each rotor is in contact with, so as to form a seal against, the corresponding teardrop surface of the other rotor. The involute surface of each rotor is in contact with, so as to form a seal against, the corresponding involute surface of the other rotor. The high pressure opening comprises a lower edge, the lower edge positioned along an outer diameter of the frusto-spherical outer surface of the second rotor, wherein the lower edge is positioned between the second rotor shaft and the at least one valley of the first rotor.
In some embodiments, each involute surface of the first and second rotors includes an involute curve at the outer diameter of the first and second rotors, the involute curve having a distal point adjacent to the corresponding lobe tip, the distal point having a distance measured between the distal point and the corresponding shaft of the rotor, wherein the said measured distance is greater than a distance measured between any other point on the involute curve and the corresponding shaft. Each involute curve has a substantially horizontal slope at the distal point, the substantially horizontal slope being tangent to the involute curve and the substantially horizontal slope lying in a plane that is substantially perpendicular to the corresponding rotational axis of the first and second rotors.
In some embodiments, the lobe tip of each lobe of each of the first and second rotors is a rounded lobe tip, the rounded lobe tip comprising a spline curve. In some embodiments, the spline curve of each lobe tip varies at each diameter of each rotor of the first and second rotors.
In some embodiments, the at least one involute surface of each axial surface of each rotor includes a radial recess extending along the at least one involute surface, the radial recess extending to the outer diameter of the rotor.
In some embodiments, the high pressure opening is formed in the housing of the device. In some embodiments, the low pressure opening is formed in the housing of the device.
In some embodiments, the high pressure opening comprises an aperture in a repositionable high pressure gate, the high pressure gate sandwiched between the intermeshed first and second rotors and the housing, wherein the aperture of the high pressure gate may be selectively circumferentially positioned relative to the position of the at least two chambers formed between the intermeshed first and second rotors. In some embodiments, the corresponding low pressure opening is formed in the housing of the device.
In some embodiments, the low pressure opening comprises an aperture in a repositionable low pressure gate, the low pressure gate sandwiched between the intermeshed first and second rotors and the housing, wherein the aperture of the low pressure gate may be selectively circumferentially positioned relative to the position of the at least two chambers formed between the intermeshed first and second rotors. In some embodiments, the low pressure gate comprises two or more low pressure apertures.
In some embodiments, the low pressure opening is an aperture in a repositionable low pressure gate and the high pressure opening is an aperture in a repositionable high pressure gate, the low pressure and high pressure gates each sandwiched between the intermeshed first and second rotors and the housing, wherein each aperture of the low pressure and high pressure gates may be independently and selectively circumferentially positioned relative to the position of the at least two chambers formed between the intermeshed first and second rotors.
In some embodiments, at least one of the high pressure gate and the low pressure gate are selectively circumferentially positionable via an actuator. The actuator may, in some embodiments, drive a worm drive, wherein at least one of the high pressure gate and the low pressure gate comprises a worm wheel of the worm drive.
In some embodiments, the high pressure opening comprises the aperture in the repositionable high pressure gate and an abutting aperture in the housing, wherein the aperture in the repositionable high pressure gate comprises adjustable leading and trailing edges and the abutting aperture in the housing comprises stationary leading and trailing edges and the lower edge, and wherein the configuration of the high pressure opening may be changed by selectively positioning the aperture of the repositionable high pressure gate relative to the stationary abutting aperture in the housing.
In some embodiments, the corresponding low pressure opening is formed in the housing of the device.
In some embodiments, the low pressure gate is selectively circumferentially positioned so as to align the low pressure aperture with each chamber of the at least two chambers such that each chamber will pass the low pressure aperture and be sealed by the low pressure gate at a selected volumetric capacity of each chamber.
In some embodiments, the high pressure gate is selectively circumferentially positioned so as to align the high pressure aperture with each chamber of the at least two chambers such that each chamber will pass the high pressure aperture and the volume in each chamber of the at least two chambers will be maintained at a selected volume ratio.
In some embodiments, at least one of the high pressure gate and the low pressure gate is circumferentially repositioned so as to selectively form a seal between the at least two chambers formed between the intermeshed first and second rotors and the corresponding high pressure port or low pressure port.
In some embodiments, at least one of the first and second rotors is a driven rotor, wherein the shaft of the driven rotor is driven by a driver. The driver may be, in some embodiments, selected from a group comprising: an electric motor, an electric motor controlled with a variable frequency drive (VFD), hydraulics, pneumatics, a second rotary positive displacement device.
In some embodiments, the lower edge of the high pressure opening substantially lies in a plane that is substantially perpendicular to the first rotational axis.
In some embodiments, the lower edge of the high pressure opening is positioned substantially between the at least one lobe tip of the first rotor and the at least one valley of the second rotor.
Like reference characters in the drawings generally refer to the same parts throughout the different views.
This disclosure includes several examples of rotary positive displacement devices within the scope of the present disclosure. In some aspects, some embodiments may have high operational flexibility, high power to mass ratio and/or comparatively low cost. The devices disclosed herein, in one example, provides an expander that may be reversibly operated as a compressor without any change to the structures. The devices disclosed herein, in another example, may be operated as a pump that may be reversibly operated as a liquid motor.
To aid in description and reduce the length of the text, in some instances, specific examples of a component include an alphabetic suffix denoting the specific example of a more generic example. For instance, the lobe tips 138A/138B may denote the lobe tips on the respective “idler” and “driver” rotors and a sawtooth rotor contains a lobe tip 138. In some examples, a specific label is shown in the drawing, and the generic label is used in the specification indicative of the specific example and equivalents.
Both positive displacement and dynamic displacement expansion and/or compression devices are discussed in the background section. For the purposes of this disclosure, it is to be understood that while liquids may be substantially incompressible, an increase or decrease in pressure across the device may be referred to as compression or expansion, respectively. Several addressable markets are discussed in the background section which do not represent all of the potential markets that the novel Rotary Positive Displacement Device 100 (See
The basic, advanced and customizable sawtooth rotor profiles comprise of teardrop surfaces with instructions outlined below. The construction of a teardrop base curve, prior to offsetting, is described in the prior art. With reference to
Axes 1 and 2 intersect at the intersect point 32, with a first axis (Cartesian coordinate) system 20 comprising of axis 12 indicating the x direction, axis 14 indicating the Y direction and axis 16 indicating the Z direction. Axis 1 is collinear with the Z axis, “16z”. Axis 2 is set at a prescribed angle “α” (i.e. aka “alpha angle” or “p”) from the Z axis, “16z”. The alpha angle may be an acute angle with a magnitude less than 90° or
radians. Reference axis 26 is set at a prescribed angle “a” (i.e. aka “reference angle”) from axis 1. The reference axis 26 contains the reference point 30. The reference angle may be an obtuse angle with a magnitude greater than 90° or
radians and less than 180° or π radians. As axes 1 and 2 rotate by an equal amount, “θ”, the reference point 30 defines the arc path 28. This arc path 28 is circumferentially positioned around the first axis 1, but with respect to the rotating axis 24, the path of the reference point 30 defines base curve 33, with points lying on an outer sphere 34. Reference axis 26 may be used to define the base curve 33, but when a shape, for example a conical frustum with radius, “ρ”, from reference axis 26, with defining surface 38, is moved with reference axis 26 the offset surface 42 is created.
The sawtooth rotor profiles disclosed herein make use of a teardrop shaped base curve 33 formed from the reference axis 26 positioned equidistant from reference axes and 2, at an angle a=alpha/2+90° from axis 1.
Sharp Lobe Tip Basic Sawtooth Geometry—One Lobe with Helical Involute
With reference to
It should be appreciated that there may be more than one method to produce varying sawtooth geometries, including, but not limited to, methods involving mathematics and varying software, including Computer Animated Design (CAD) software. It will be appreciated that any rotors and devices falling within the scope of the claims of the present disclosure are intended to be included, regardless of any specific method or software utilized to produce the geometry of the rotors and devices.
Parameter “R” is the radius of the imaginary sphere that the base curves lie on. Referring to
It will be appreciated that rotating the vector v while keeping the Cartesian coordinate system fixed is mathematically equivalent to keeping the vector v fixed while rotating the Cartesian coordinate system. Four vector rotations can be applied to vector v to create the base curve “c” (See
Finally rotate the resulting vector by −t about the Z axis, giving the full matrix multiplication written out:
When a=p/2+π/2 radians, the base curve, c, becomes the teardrop base curve, which can be simplified as follows:
Various mathematical manipulations are done in this disclosure with equations that have points lying on a unit sphere where R=1 in the below equation.
It is to be understood that if the parameter R is missing in a teardrop or involute equation, that the X, Y and Z terms may be multiplied by a factor of R to scale to any desired radius. Therefore, the teardrop equation may be re-written below, which is implied to have points existing on a unit sphere where R=1:
As seen in
A sharp lobe tip sawtooth geometry is comprised of intersecting teardrop and involute curves. For a basic sawtooth geometry, the Cartesian parameterization of a Spherical Helix equation is manipulated to plot the involute base curve 114 seen in
The parameter “t” is typically used in parametric equations, but in the case of the involute curve 114, it does not represent the rotational position, theta (θ), seen in
It is to be understood that a Total Sum of Squares (TSS) approach may be used, minimizing the result of the below TSS equation at points 46′/106 and points 116/112.
While the TSS is being minimized at t=0 on both the teardrop equation and helical involute equation, the second desired location of coincidence is also included. It is to be understood that the value of R is equivalent in the teardrop equation and helical involute equation. It is good practice to set R=1 in both equations and also verify that all points of both equations indeed lie on the unit sphere, satisfying the basic equation shown below:
In this example, the input variables remain unchanged for varying values of R. As seen in
It will be appreciated by persons skilled in the art how to create surfaces from base curves. For example, a boundary surface may be created with curves 118 and 118′ that would have points coincident with points on lines 120, 122 and 124. Alternatively, the radial line 120 could be swept along the path of 118/118′ until it reaches it's starting position. Since curves 118/118′ are on spheres concentric to the intersect point 32, lines 120, 122 and 124 are all equal length. It is also to be understood that an infinitesimally small value for the radius, R, used to generate sawtooth rotor profile 118′ may be approximated as a point. However, the resulting sharp point lobe tip 138 of the rotor geometry 132 shown in
It is to be understood that the sharp lobe tip sawtooth geometry 132 seen in
The second rotor 150 (i.e. See
Maintaining a front view of the first rotor 144, it may be seen in
The sawtooth assembly 164 seen in
It is commonly understood that a video consists of a series of frames/images that are sequentially displayed. The function of a compressor was explained by studying
The views seen in
As can be seen in
Round Lobe Tip Basic Sawtooth Geometry—One Lobe with Helical Involute
An offsetting method is required to prepare round lobe tip sawtooth geometry. The below method was discovered.
First, a normal vector is obtained as a function of parameter t along the teardrop equation, T(t). It is to be understood, that if a variable, “u”, is an infinitesimally small value, then the parameter t in the teardrop equation may be replaced with t+u or t−u without resulting in a significant deviation from the intended offset curve. It is also to be understood that every point in the teardrop equation, T(t), can also represent a vector of length R that originates from the intersect point 32 at the origin. In this manner, T(t+u) and T(t−u) may appear to be nearly co-linear, but they are not. It will be understood that the cross product of two vectors will produce a vector that is orthogonal to each of the two vectors per the right-hand rule. Using the below mathematics will produce a normal vector, N, that is normal to the teardrop curve at any value of t.
It is to be noted that ‘u’ is an infinitesimally small positive value for the offset teardrop base curve 170 to be larger than the teardrop base curve 102 (See
A variable, ‘b’ is the desired offset angle. In the examples shown herein,
As seen in
Keeping R=1 will keep all of the points of the offset teardrop base curve 170 on the unit sphere as follows:
This equation in plotted in
Next, a vector rotation of “−e” about the Z axis is applied to the offset teardrop base curve 170 until the point 174 lies on the X-Y Plane as shown in
Plotting on a unit sphere, the shifted offset teardrop equation is written as follows:
Those skilled in the art would appreciate how to connect the offset involute base curve 184 and teardrop base curve 170′ with a tangent arc to produce the constant radius circular arc 200, trimming as appropriate. Another methodology is required to be able to vary the radius of the circular arc between the offset involute base curve 184 and teardrop base curve 170′, if desired. For example, see
The intersection of a plane with a sphere result in the formation of a circle and a portion of the circumference of a circle may be referred to as an arc. In
In the example seen in
In this form, Nx, Ny and Nz represent the respective X, Y and Z components of the vector normal to the plane. The length of this normal vector is not important, so for convenience, the respective Nx, Ny and Nz terms may be replaced with the X, Y and Z coordinates of point 198. The respective X, Y and Z coordinates of any point on the plane may be used for Xp, Yp and Zp, but it is intuitive to select point 174 of the shifted offset teardrop base curve 170′ (See
A vector may be defined as the subtraction of one point relative to another. In this example, a normal unit vector may be defined using the X, Y and Z coordinates of a point 206 on circular arc 200 minus the X, Y and Z coordinates of point 208 at the end of line 178. It is to be understood that a table of normal unit vectors may be plotted as the point 206 is moved from the location of point 196 along the circular arc 200 to the point 202. In this manner, the point 208 may be offset by an angle of b to produce the circular arc 200. It is to be understood that the value of b could be varied at varying positions of point 206 to produce a circular arc 200 with varying radius. It is to be understood that a plurality of points can be defined in this manner and continuously pieced together with continuity. The terms “spline” or “spline curve”, used interchangeably herein are used herein to define the construction of a curve which closely follows a sequence of points. Mathematically, a spline curve may consist of a piecewise combination of a plurality of polynomials. In one example, the spline curve may consist of circular arcs and line segments. In another example, the spline curve may consist of a single circular arc. In the below equation, the respective X, Y and Z coordinates of point 208 on the unit sphere are represented by Gx, Gy and Gz.
It is to be understood that points from the teardrop base curve 102 in
The sawtooth assembly 164 seen in
Some differences between the round lobe tip geometry example seen in
Constant Radius Round Lobe Tip Basic Sawtooth Geometry—One Lobe with Helical Involute
It is to be understood that the offset angle b could vary at every value of the spherical radius, R. In one example it may be shown how the lobe tip itself could have a constant radius. As seen in
A sawtooth geometry with only one Lobe is expected to be the most efficient in many applications because the pressure in the teardrop volume 168 in
Round Lobe Tip Basic Sawtooth Geometry—Three Lobe with Helical Involute
As discussed above, the one lobe geometry is simple and efficient, but the asymmetry does introduce dynamic imbalance that should be addressed in high horsepower applications. That is, relatively large rotors spinning relatively fast may produce a significant load on the bearings if balancing techniques are insufficient. Having multiple lobes that are substantially circularly symmetric may significantly reduce or even eliminate the need to apply balancing techniques known by persons skilled in the art. Despite being less efficient, using more than one lobe may increase the volumetric throughput noticeably for a given rotor size and operational speed.
As previously discussed, both the involute base curve 114 and offset involute base curve 184 in
Preferably, a geometry with multiple lobes will demonstrate circular symmetry, with identical teardrop and involute pairs that are knitted together. It is to be understood that a vector rotation about the Z-axis may be done to move coincident Points 186 and 188 to the location where the approximate midpoint of the involute should occur. In
The 3 lobe round tip solution with helical involutes is plotted in
Sharp Lobe Tip Advanced Sawtooth Geometry—One Lobe with Advanced Involute
It will be appreciated that any working fluid that is not evacuated at the end of a compression stroke may introduce a noticeable efficiency penalty. Pressurized liquids may contain gas in solution that is subsequently vented at lower pressure and work is required to re-compress that gas. Pressurized gases may introduce a relatively higher efficiency penalty given the requirement to re-compress that gas. Similar comments can be made for the reverse (expansion) process. Volumes that cannot be evacuated are often referred to as the “clearance volume” and the relatively high pressure gas that is inefficiently recirculated is often referred to as the “recirculated volume”. As discussed in later sections, the type of sawtooth geometry used at the lobe tip region may be, in preferred embodiments, particularly useful in reducing clearance volume. In particular, the advanced and customizable sawtooth geometries may beneficially facilitate orders of magnitude lower clearance and recirculated volume as compared to the known prior art.
In the applicant's opinion, an issue with the known prior art solutions is the significant recirculated volume that may be expected. For example, one prior art concept disclosed in U.S. Pat. No. 10,975,869, shown in FIG. 32 of U.S. Pat. No. 10,975,869, involved introducing an additional component affixed to the outer diameter of a frusto-spherical rotor, wherein all of the gas would be required to pass through the relatively small hole positioned at the valley of a rotor. In the applicant's opinion, this introduces a significant clearance volume which is likely to be largely gas filled, contributing to a significant efficiency penalty. Furthermore, any reduction to the cross-sectional flow area of the hole, for example to reduce the clearance volume, may increase the efficiency penalty related to the pressure drop across said restriction, likely resulting in a lower rated operational speed and power density. In another prior art concept, also disclosed in U.S. Pat. No. 10,975,869 at FIGS. 130 to 135 of U.S. Pat. No. 10,975,869, a large hole was introduced at the inside diameter of a rotor valley. While there is some uncertainty regarding how much of this clearance volume would be gas filled in oil-flooded applications, porting at the inside diameter of a rotor may require a certain minimum size of rotor, which is not believed to be practical to address the wide range of applications discussed herein. Advantageously, in one aspect of the present disclosure, the Advanced Sawtooth geometry described below will, in later sections, be used with efficient porting at the outside diameter of the rotors, without introducing any additional components affixed to the rotor and without introducing any substantial clearance volume.
It was demonstrated that the formation of the teardrop surface 128 is corollary to the geometric interaction of the first/second rotors 144/150. Therefore, the same equation is still used and repeated below for clarity.
It was discovered that applying a vector rotation of 2t about the +Z axis to the teardrop equation produces an advanced involute base curve 118 for a one-lobe geometry. The mathematics are presented below:
Plotting on a unit sphere, the advanced involute equation is written as follows:
The helical involute base curve 114 seen in
As seen in
For even further clarity, this horizontal slope is defined as follows. Plane 308 in
Unlike the advanced involute base curve 114B, the helical involute base curve 114A does not have a horizontal slope near the start point 106A. Point 106A on plane 308 and in cohort with point 312A forms straight line 310A. Straight line 310A is at an angle 314A from plane 308. In this example of
Round Lobe Tip Advanced Sawtooth Geometry—One Lobe with Advanced Involute
As discussed, the use of sharp lobe tip geometry may not be practical for some applications. The offsetting methods used in foregoing examples may be applied as follows:
In
That is, the sawtooth rotor profiles 118C/D seen in
The trimmed portion of the advanced offset involute curve 184′ can be seen in
The prior examples aid in demonstrating that the interaction of lobe tip and teardrop surfaces of meshed rotors is independent of the interaction of the respective involute surfaces. This disclosure provides for construction of sawtooth rotor geometries using helical and advanced involutes. The offsetting methods disclosed herein may be used to construct offset involute curves from sharp involute curves by adjusting the variable ‘b’ to be a function of the parameter ‘t’ or the X, Y and/or Z component equations. Using this method, the helical offset involute curve 184 in
If it is desired for the involutes of meshing rotors to form a seal, it may, in some embodiments, require maintaining the ‘midpoint’ location after offsetting. By way of reminder, the ‘midpoint’ location is not defined as a value of the parameter ‘t’ since this may change after offsetting. As seen in
In
Before trimming, the advanced offset involute curve 184A and the advanced offset involute curve 184E are initially the same curve with midpoint 188A at 180 degrees. In this example, the midpoint 188B of helical offset involute curve 184B/B′ is at 135 degrees, which is coincidentally halfway between what is used for a conventional 1 and 2 lobe geometries. The construction of the helical offset involute curve 184B/B′ does not enforce tangency at the offset teardrop base curve 170A. Rather, a value of parameter ‘t’ or a coordinate position of the advanced offset involute curve 184A is selected. In this example, the advanced offset involute curve 184A is arbitrarily plotted up to a value of t=0.015, defining the starting point 204B of the helical offset involute curve 184B. A TSS approach may be used to satisfy all constraints, including the location of midpoint 188B.
The advanced offset involute curve 184C in
The sawtooth rotor profile 118A in
A person skilled in the art may have many variables to consider when customizing sawtooth rotor geometry to be optimized for an application. For example, in some large horsepower applications where dynamic loads are expected to be significant even after dynamic balancing a one-lobe sawtooth geometry, it may be advantageous to select sawtooth rotor geometry with a plurality of lobes, despite the lower expected efficiency introduced from having a relatively higher pressure introduced in the teardrop volume 168 (See
In smaller horsepower applications, the power density and efficiency are typically desirable parameters for optimization. Finding the practical maximum for the alpha angle and the operational speed is likely to maximize the power density while also reducing manufacturing costs. However, in some applications, maximizing the efficiency may provide an attractive payback period. When both rotors are rotating, it may be demonstrated how chambers may naturally become in fluid communication with a low pressure and high pressure side for compression or expansion operations. This valve-less operation is expected to maximize the reliability of the equipment, which for many operations may be amongst the most important considerations. The rate at which a given chamber is increasing or decreasing in volume is evaluated in cohort with the cross-sectional flow area and/or relative ease of the flow path. The performance at relative rotational positions may be evaluated, for example, with Computational Fluid Dynamic (CFD) studies and/or calculations including the effective hydraulic diameter; however, performance may be evaluated by other methods, as would be known to a person skilled in the art. In compressor mode, it is typically desirable for the flow area to be as large as practical during the intaking process and then have this fluid communication with the intake be naturally closed off as quickly as possible relative to the changes in chamber volume. Some amount of pressure differential is required to drive the flow into the chamber. Efficiently intaking with minimal pressure loss (i.e. compared to the intake pressure) is expected to boost volumetric throughput and power density since more gas was trapped in the chamber. When the chamber substantially reaches the desired discharge pressure it may be desirable for the chamber to be in fluid communication with the downstream/discharge piping with the flow area opening up as quickly as possible relative to the subsequent changes in chamber volume. That is, the chamber volume will continue to decrease and additional power is required to cause the gas to over-pressure. When evaluating the efficiency of sawtooth rotor geometries, it may be important to study how ‘quickly’ the flow area opens up naturally. Significant over-pressuring can be dangerous and while valveless equipment can very easily adjust to open the chamber up early to mitigate these effects, this can be a noticeable efficiency penalty. In one aspect, the shape of the rotor lobe may play an important role in determining how quickly the chamber opens up and therefore the appropriate operational speed range for the expected range of pressures and fluid composition.
There are other considerations to weigh. For example, increasing the steepness of the rotor lobe is likely to increase the recirculated volume somewhat. In the example provided in
Minimizing recirculated volume may be important in high volume ratio applications. While a high pressure ratio implies a high required volume ratio, a high volume ratio may also be associated with multi-phase applications. For example, the Trilateral Flash Cycle (TFC) or Partially Evaporated Cycle (PEC) involve transition from a fully liquid working fluid or partially liquid working fluid (respectively) into a liquid-vapor mixture. Liquids may be much higher density than gases and therefore a large volume ratio may be required even if the pressure ratio is modest. The substantially horizontal slope of the rotor lobes seen in
Several bottom view examples of the sawtooth assembly 164 are provided in
Thus, the position of leading edge 220, in some embodiments, is configured to be a stationary leading edge 220 and is configured such that a higher volume ratio than is reasonably efficient is not obtained during operation. That is, it may be preferred for the leading edge 220 to permit flow 222B no earlier than leading edge 218 permits flow 222B. In the example seen in
The high pressure opening 216 seen in
In
In
In
In
In the Applicant's opinion, relatively large instantaneous flow rates may warrant relatively slower speed operation and/or relatively large holes in prior art equipment compared to the small recesses 230/232 that are only responsible for addressing the start and end of the stroke where relatively low instantaneous flow rates are expected. That is, when the sawtooth rotor shafts are spinning at a constant velocity, the rate of change of the chamber volume decreases significantly at end of the stroke.
As seen in
It is to be understood that the sawtooth geometry 132 seen in
Still referring to
It is also understood that if the working fluid has gas present, then for a rotary positive displacement device to be efficient, the chamber should experience a noticeable change in volume while it is substantially sealed from both the high pressure opening 216 and low pressure opening(s) 236A/B/C. This ideal change in volume is dependent on a number of variables, including the gas composition, solid and/or liquid entrainment and the upstream/downstream pressures and temperatures. It may be typical for the position of the high pressure opening 216 and low pressure opening(s) 236A/B/C to be optimally set to allow for a chamber to change volume while sealed therebetween by a “volume ratio” that corresponds to the upstream and downstream pressures, forming a “pressure ratio”. In compression, the chamber increases in pressure as the volume decreases and there is an efficiency penalty if that chamber becomes in fluid communication with the high pressure opening 216 too early or too late. Similar efficiency penalties are expected in expansion mode where additional flow is required to produce the desired amount of power. It is also understood that strain gauges (not shown) affixed to rotational shafts may be used to determine torque and therefore power at a given rotational speed. Therefore, in this manner the control system may facilitate adjustment to the optimal shapes of these openings 216 and 236/A/B/C on-the-fly by minimizing the ratio of
in compression and maximizing the ratio of
in expansion. Often, only a certain amount of power is available or required and these adjustments may allow the equipment to still successfully operate within torque capabilities of attached equipment such as a motor or generator. It will be appreciated that a motor consumes electricity to power a compressor and an expander produces electricity via a generator, which may be a motor operating in reverse. In some embodiments, the sawtooth assembly 164 may be capable of considerable volumetric capacity control and pressure ratio flexibility such that substantially similar power and/or flow may be maintained over a wide range of gas compositions and upstream/downstream pressures and temperatures. It may also be possible for a chamber to be in fluid communication with both the high pressure opening 216 and low pressure opening(s) 236/A/B/C simultaneously, for example Chamber 166A in
For further clarity, any combination of housing and gate components may be used to form the low pressure opening(s) 236A/B/C and high pressure opening 216. The low pressure gate is understood to have a low pressure aperture which may be circumferentially repositioned to align the low pressure aperture with each chamber of the at least two chambers such that each chamber will pass the low pressure aperture and be sealed by the low pressure gate at a selected volumetric capacity and maintain a selected pressure ratio. The high pressure gate is understood to have a high pressure aperture which may be circumferentially repositioned to align the high pressure aperture with each chamber of the at least two chambers such that each chamber will pass the low pressure aperture and be sealed by the high pressure gate at a selected volumetric capacity and maintain a selected pressure ratio (or selected volume ratio). In other words, the high pressure gate and/or the low pressure gate may be used for volumetric capacity control and to maintain a selected pressure ratio (or selected volume ratio). In one example, in compression or expansion mode it may be preferred to use the low pressure gate for volumetric capacity control and subsequently circumferentially reposition the high pressure gate to maintain a selected pressure ratio (or selected volume ratio) because it may be easier to maximize volumetric capacity. However, it is to be understood that the high pressure gate may be used for volumetric capacity and the low pressure gate may be used to maintain a selected pressure ratio (or selected volume ratio).
The following are illustrative examples of how the devices and methods disclosed herein may be utilized in practical applications. These examples are provided for illustrative purposes only, and are not intended to be limiting.
Many companies may require relatively small amounts of power at remote locations, for example to facilitate remote monitoring and to actuate valves. In one example, 200 W of power may be required year-round. Compressors are used along natural gas pipeline networks to convey natural gas to end users. For example, utility companies may throttle the pressure down to less than 100 psig upstream of cities for safer distribution to residential and commercial buildings. In Alberta, Canada it may be typical for the pressure upstream of these sites to approach 900 psig in the summertime, when usage is lower, and be drawn down to around 400 psig in the winter time when usage is higher. Since 200 W is such a low power requirement, it may be typical to divert only a small percentage of the flow, as required, to a parallel pathway for that power generation. In this scenario there may be limitations to how much flow may be diverted without interfering with the control system that is throttling the main gas stream. A higher efficiency expander is expected to reduce flow requirements accordingly. Using these numbers as a demonstrative example, let's say that the operational specification is to be able to produce 200 W of power from 900 psig being expanded down to around 60 psig to 100 psig and also produce 200 W of power from 400 psig being expanded down to around 60 psig to 100 psig. The pressure ratio of 900 psig being expanded down to 60 psig is higher than 400 psig being expanded down to 100 psig.
Thermodynamically, if the flow rates were equivalent, it is theoretically possible to produce more power in the higher pressure ratio example. However, the flow rate, and therefore power associated with positive displacement devices depends on the amount of working fluid that may be sealed. For example, if an inflexible prior art device were used, or if the low pressure openings 236A/B/C seen in
The shape/position of the high pressure opening 216 was varied in
While the examples provided do not quantitatively imply any limitation of the equipment, it is to be understood that the sawtooth assembly 164 may have additional forms of volumetric capacity control flexibility. It will be appreciated that the full range of 0-100% volumetric capacity control may be desired in many (power-generating) control valve applications. In-line/series operation may require the expander to allow passage of variable amounts of flow. If the working fluid composition and upstream/downstream pressures and temperatures would produce more power than the electrical generator is rated for, then it may be desirable to operate less efficiently, increasing volumetric throughput accordingly. This may be accomplished by reducing the volume ratio that is sealed as a chamber passes between the high pressure opening 216 and the low pressure opening(s) 236A/B/C up to the extreme scenario, for example as seen in
relative to known prior art. This may reduce leakage at low speeds accordingly. Since the sawtooth assembly 164 may provide for reversible compressor/expander operation without modifications to internal structures, if an unacceptable amount of flow is passing through the equipment, compression mode may be periodically turned on to counter-act the internal leakage. This compression mode functionality may also be desirable for companies to substantially evacuate working fluid from the low pressure side of the sawtooth assembly 164, for example if maintenance is required. In such scenarios it may be common practice to vent or flare the natural gas contained at the low pressure side, so companies may avoid or significantly reduce such economic penalties with the sawtooth assembly 164.
As an additional safety precaution, the upstream and downstream sides of the sawtooth assembly 164 may be sealed to substantially prevent internal leakage when the equipment is not in service. Looking to
The component assemblies and components seen in
Another example may be seen in
In the example provided in
It is to be understood that in the examples provided herein, the low pressure gate 256 and high pressure gate 260 may be rotationally positioned from 0° to 360° rotational positions or even angles exceeding 360°. In some applications it may be desirable to quickly ramp to substantially 0% volumetric throughput. Typically, if the rotational inertia of the motor components and first/second rotors 144/150 is relatively low, it may be sufficiently fast for the driver itself to provide braking horsepower until the rotating components come to rest. In one example, the driver is a motor, which is understood to be reversibly operated as a generator. If a motor were used to rotate components fixed to the first/second rotors 144/150 in a sense to provide compression work, then simply stopping the input electrical current or even reversing the direction may quickly cause the rotating components to come to rest. If a generator were producing electricity when a sudden stop is desired, then the electrical current may be reversed to attempt to provide compression work until the rotating components come to rest. If the sufficient volumetric throughput reduction may be provided quick enough from the actions of the motor/generator and attached components, then the actuation of the gates 256/260 may not need to be relatively fast-acting. Often it is desirable for control systems to control relatively slow-acting actuation components, which may include reducing the measurement speed and therefore cost of required sensors and/or impact of inertial effects and measurement lag.
The position of a gate may be impacted by the net loading in the circumferential direction, which is tangent to the motion. The high pressure gate 260 seen in
A side view of the Rotary Positive Displacement Device 100 may be seen in
As an example, a remote power application was disclosed where the upstream/downstream pressures that natural gas utility companies may often have at their remote sites may vary substantially from season to season. While more frequent adjustments are expected than a seasonal time scale, this example does serve the purpose of demonstrating that systems that continuously re-adjust are expected to waste a lot more energy. In one example, continuously adjusting systems may require more than one adjustment every minute even if there were no notable changes in gas composition and/or upstream/downstream pressure and temperature for example. In the examples provided herein, electrical actuation is discussed, but it is to be understood that other forms of actuation, including using hydraulics or gases are known in the art and are applicable. It should also be noted that since the Rotary Positive Displacement Device 100 does not require any pressure-activated valves, this helps to facilitate the wide capabilities, but also may make it more robust should the operational strategy include intentionally opening a chamber early or late. Such an operational strategy may reduce the operational life of pressure-actuated valves, for example in reciprocating compressors.
Looking back to the example provided for natural gas utility companies that may wish to have 200 W of power at remote sites for remote monitoring, small reductions in the efficiency of the Rotary Positive Displacement Device 100 may not be concerning provided that only a relatively small amount of flow is still diverted to the expander. In this example, if the gates 256/260 were positioned accordingly for upstream/downstream pressures of 900 psig and 100 psig, respectively and the downstream pressure were to drop to 60 psig, this may not influence the power output significantly enough to cause concern. That is, the position of the gates 256/260 may correspond to a particular volumetric flow rate and pressure ratio while sealed. The notable difference in operation is that a given chamber may be closer to 100 psig than 60 psig when it becomes in fluid communication with the downstream piping. Essentially, a certain amount of potential energy in the gas was wasted. In other words, 200 W of power can be produced with less volumetric flow rate of the gas if the pressure ratio is higher, but to do this both gates 256/260 may need to be repositioned. If a Variable Frequency Drive (VFD) were employed or a multi-speed transmission with an engine, this rotational speed is another variable that may be adjusted accordingly.
In the example provided in
It should however be noted that when the working fluid contains a fuel and/or combustible gas, the use of electrical motors/generators may require hazardous certification approval(s) and design considerations to ensure that the ambient air is not likely to be mixed with the working fluid and that any resulting explosion may be contained in the housing of the equipment. While the temperatures may not typically be high enough to facilitate auto-ignition, the electrical motors/generators are one potential ignition source. Looking to
Instead of using electrical motors, it is to be understood that other types of drivers may be used. For example, it will be appreciated that magnetic couplings may also be used to facilitate other types of off-the-shelf input/output drivers while maintaining a hermitic or semi-hermitic seal of the Rotary Positive Displacement Device 100, if desired. Alternatively, dynamic seals, for example mechanical seals, may be used. In either sense, since the Rotary Positive Displacement Device 100 is expected to be capable of varying compression/expansion scenarios, one possible example is using a Rotary Positive Displacement Device 100 to drive another Rotary Positive Displacement Device 100. It will be appreciated that compressors and expanders sharing a common shaft in turbine applications is known, and therefore no further illustrations need to be provided in this regard as it will be appreciated that first/second motor assemblies 248/266 may be modified and/or removed as required to facilitate this example.
It is to be understood that the high pressure port 270 may form a perimeter of a fluid pathway that may be fluidly connected to piping and/or tubing.
It is also be understood that the low pressure port 272 may form a perimeter of a fluid pathway that may be fluidly connected to piping and/or tubing.
It is also to be understood that seals, fasteners and pins may be added to the Rotary Positive Displacement Device 100 as required. In the example provided in
It is to be understood that a working fluid may comprise of a combination of one or more gases and/or one or more liquids. While solid entrainment may often not be desirable, in part because it may lead to accelerated wear, it is sometimes a reality. For example, even pipeline grade natural gas contains some amount of solids. As another example, air compression applications often still contain residual dust in the air. The amount of solids in the gas may be meaningful in some scenarios. For example, isothermal or near isothermal compression processes may require significantly less work in comparison to isentropic compression processes. It should also be understood that the higher heat capacity of liquids and/or solids may desirably reduce the work required to compress the gas while also reducing the discharge temperature.
It should be understood that even before any attempt at optimization, the Rotary Positive Displacement Device 100 may have substantial competitive advantages over currently employed technologies. Chamber volumes are cubically related to the (frusto-spherical) rotor diameter in contrast to the chamber volumes of a screw profile being related to the screw rotor diameter squared and linearly related to the length. This may allow for much smaller space usage and therefore a higher power density when comparing equivalent operational speeds. The high ratio of volume to linear sealing perimeter may substantially minimize volumetric inefficiencies from internal leakage, especially if the gaps between the rotor outer diameter and housing inner diameter can also be minimized. The long screw rotors are supported by bearings at the ends of the shafts. This mechanical limitation of screw rotors is not well suited to withstanding relatively larger pressure-induced loads expected from higher pressure ratings. Over time, including during upset conditions, high loading conditions, for example pressures exceeding 350 psig, may cause the screw rotors to separate and undesirably contact adjacent (stationary) housing components and/or cause unacceptable high internal leakages. While some individuals may compare aspects of the Rotary Positive Displacement Device 100 core technology to the prior art of a screw rotor, it is clear that there are considerable differences. Looking to
Comparisons are made to screw compressors since they are a rotary positive displacement technology with widespread global dominance that are well understood by compressor experts that have managed to displace a substantial amount of the reciprocating compressor market that was previously dominant below 350 psig. Screw rotor equipment has been successfully demonstrated to operate in reversible compressor/expander modes and the limited internal adjustability capabilities and lack of valves have been found to be quite attractive. Advantageously, the novel Rotary Positive Displacement Device 100 disclosed herein is expected to have capabilities that may address a wide range of markets and fulfill the requirements for a wide range of different applications. Discoveries disclosed herein may represent a step-change improvement over the prior art and lead to a new class of Rotary Positive Displacement Devices 100 being adopted in widespread commercial applications.
This application claims the benefit of U.S. Provisional Application No. 63/502,302 filed on May 15, 2023 and entitled “Rotary Positive Displacement Device”, which is incorporated herein by reference.
| Number | Date | Country | |
|---|---|---|---|
| 63502302 | May 2023 | US |
