Patent Application
20230114229
The present invention relates to production of liquefied natural gas (LNG), and more particularly, to a small or mid-scale liquefied natural gas production systems and methods using a nitrogen based refrigerant that employs at least three turbine/expanders and two or more refrigerant compression stages.
Demand for liquified natural gas production in applications related to energy infrastructure, transportation, heating, power generation is rapidly increasing. The use of liquified natural gas as a lower cost, alternative fuel also allows for a potential reduction in carbon emissions and other harmful emissions such as nitrogen oxides (NOx), sulphur oxides (SOx), and particulate matter which are generally recognized as detrimental to air quality. As a result of this demand, a trend has emerged for construction and operation of lower capacity liquified natural gas production systems built in regions where attractive sources of low cost natural gas or methane biogas are available and/or where there is a current demand for liquified natural gas, or the demand is expected to grow over time.
Small-scale to mid-scale liquified natural gas opportunities include various energy applications such as oil well seeding or boil-off gas re-liquefaction, integrated CO2 extraction and natural gas liquefaction, utility sector applications such as peak-shaving or emergency reserves, liquified natural gas supply at compressed natural gas filling stations, and transportation applications including marine transportation applications, off-road transportation applications, and even on-road fleet transportation uses. Other small-scale or mid-scale liquified natural gas opportunities might include liquified natural gas production from biogas sources such as landfills, farms, industrial/municipal waste and wastewater operations.
Most conventional small-scale or mid-scale liquified natural gas production systems target a production of between 100 mtpd and 500 mtpd of liquified natural gas and higher. Many of these liquefaction systems employ mechanical refrigeration or a nitrogen-based gas expansion refrigeration cycles to cool to the natural gas feed to temperatures required for natural gas liquefaction. Use of nitrogen-based gas expansion refrigeration cycles are the preferred technology for small scale applications due to simplicity, safety, ease of operation, turndown, dynamic responsiveness and maintenance.
The current market for such small scale natural gas liquefaction systems using nitrogen-based gas expansion refrigeration cycles is dominated by the sale of equipment. Even though many recent opportunities are driven by environmental considerations, minimizing the installed cost of such natural gas liquefaction systems is a dominant factor in the liquefaction process design. When designing natural gas liquefaction cycles and liquefaction systems, capital costs and operational efficiency must be balanced. Such design decisions are highly dependent on site-specific variables, including natural gas feed quality as well as the intended applications and transport of the liquified natural gas product.
In a conventional high-pressure natural gas liquefaction system employing nitrogen-based gas expansion refrigeration cycle with dual expansion, such as that shown in
Another limiting aspect of the conventional natural gas liquefaction system and process depicted in
The conventional two turbine/expander liquefaction system shown in
What is needed therefore, is a natural gas liquefaction system and process that provides a more equitable distribution of power to the individual pinions and which exhibits an outsized capitalized power benefit relative to the conventional two turbine/expander liquefaction systems with limited added capital expense.
Another natural gas liquefaction system that discloses a three turbine/expander based natural gas liquefaction cycle is disclosed in United States Pat. No. 5,768,912 (Dubar). In that prior art disclosure, three booster loaded nitrogen expanders are disposed in series and the resulting efficiencies of this three turbine/expander liquefaction system is less than ideal resulting in additional capital costs without the corresponding reduction in power and operating costs.
Thus, what is also needed are improvements in the overall design and performance of such natural gas liquefaction systems and processes with the objective of minimizing the heat exchange liquefaction inefficiencies while facilitating turbomachinery design. In this way, power consumption can be minimized. This goal of minimizing the heat exchange liquefaction inefficiencies is critical to achieving meaningful performance improvements.
The present invention may be characterized as a natural gas liquefaction system with a refrigeration circuit that includes, among other elements: (i) at least one heat exchanger configured to liquefy and subcool a compressed natural gas containing feed stream via indirect heat exchange with a refrigerant stream; (ii) three turbine/expanders configured to expand portions of the refrigerant stream to produce at least three exhaust streams that are directed to the at least one heat exchanger to liquefy and subcool the natural gas containing feed stream via indirect heat exchange and exit the at least one heat exchanger as one or more warmed recycle streams; and (iii) two or more refrigerant compression stages including an upstream refrigerant compression stage and a downstream refrigerant compression stage both configured to compress the warmed recycle streams. The three or more turbine/expanders further comprise a cold turbine/expander configured to expand a cold portion of the refrigerant stream and produce a cold exhaust that is recycled to the upstream refrigerant compression stage of the two or more refrigerant compression stages; a first warm turbine/expander configured to expand a first warm portion of the refrigerant stream and produce a first warm exhaust that is recycled to the upstream refrigerant compression stage of the two or more refrigerant compression stages; and a second warm turbine/expander configured to expand a second warm portion of the refrigerant stream and produce a second warm exhaust that is recycled to the downstream refrigerant compression stage of the two or more refrigerant compression stages. The expansion ratio of the secondary warm turbine/expander is lower than an expansion ratio of the cold turbine/expander and lower than an expansion ratio of the first warm turbine/expander.
The present invention may be characterized as a method to produce liquefied natural gas comprising the steps of: (a) receiving a purified, compressed natural gas containing feed stream; (b) liquefying and subcooling a purified, compressed natural gas containing feed stream in at least one heat exchanger via indirect heat exchange with one or more refrigerant streams to produce one or more lower pressure recycle streams and a higher pressure recycle stream; (c1) compressing the one or more lower pressure recycle streams in an upstream refrigeration compression stage to produce a compressed refrigerant stream; (c2) compressing the higher pressure recycle streams and the compressed refrigerant stream in a downstream refrigeration compression stage to produce a further compressed refrigerant stream; (d) cooling the further compressed refrigerant stream in the at least one heat exchanger; (e1) extracting a cold portion of the further compressed refrigerant stream from the at least one heat exchanger; (e2) extracting a first warm portion of the further compressed refrigerant stream from an intermediate location of the at least one heat exchanger; (e3) extracting a second warm portion of the further compressed refrigerant stream from a second intermediate location of the at least one heat exchanger; (f1) expanding the cold portion of the compressed refrigerant stream in a cold turbine/expander and produce a cold exhaust at a temperature colder than -145° C.; (f2) expanding the first warm portion of the compressed refrigerant stream in a first warm turbine/expander to produce a first warm exhaust at a temperature colder than about -90° C. and warmer than the cold exhaust;(f3) expanding the second warm portion of the compressed refrigerant stream in a second warm turbine/expander to produce a second warm exhaust at a temperature above the critical point temperature of the compressed natural gas containing feed stream and colder than about -15° C. and at an outlet pressure higher than the outlet pressures of the cold turbine/expander and the first warm turbine/expander; (g1) directing the cold exhaust and the first warm exhaust to the at least one heat exchanger as a refrigeration source and yield one or more lower pressure recycle streams; (g2) directing the second warm exhaust to the at least one heat exchanger as a refrigeration source and yield a higher pressure recycle stream; (h1) recycling the one or more lower pressure recycle streams to one or more upstream refrigerant compression stages in the plurality of compression stages; and (h2) recycling the higher pressure recycle stream to downstream refrigerant compression stages in the plurality of compression stages.
In the present system and method, the first warm turbine/expander has an expansion ratio of between 4.0 and 5.0 and is configured to produce the majority of the refrigeration, preferably over 45% of the refrigeration whereas the cold turbine/expander also has an expansion ratio of between 4.0 and 5.0 and is configured to produce less than 25% of the refrigeration. The second warm turbine/expander preferably has an expansion ratio of between 1.5 and 2.5 and is configured to produce between about 20% to 35% of the refrigeration.
The present system and method is also preferably configured with an integral gear machine having a drive assembly, a bull gear, and a plurality of pinions arranged to drive the refrigerant compression stages and/or for receiving work produced by the turbine/expanders. For example, the second warm turbine/expander and the one of the upstream compression stages or downstream compression stage may be operatively coupled to a first pinion of the integral gear machine, and the first warm turbine/expander and at least one of the upstream compression stages or downstream compression stage may be operatively coupled to a second pinion of the of the integral gear machine. The cold turbine/expander may be coupled to a third pinion of the integral gear machine by itself or operatively coupled via the third pinion preferably to at least one of the upstream compression stages.
The purified, compressed natural gas containing feed stream is preferably at a pressure greater than the critical pressure of natural gas, and more preferably at a pressure between about 50 bar(a) and 80 bar(a). The refrigerant stream is a nitrogen-based refrigerant that preferably comprises more than about 80% nitrogen by volume.
It is believed that the claimed invention will be better understood when taken in connection with the accompanying drawings in which:
The design of high efficiency liquefaction processes that employ gas expansion to provide the refrigeration necessary to liquefy and subcool a purified and compressed natural gas containing feed stream is the result of a simultaneous considerations of heat transfer and turbomachinery within the system and/or process. The minimization of heat transfer irreversibility is achieved when the divergence of the warming and cooling composite curves (e.g. energy transferred vs temperature) is minimized. Process definition of flows, pressures and temperatures largely control the resulting composite curves. Turbomachinery efficiency is maximized when the head and flow characteristics of the process are consistent with experience-based optimums. These optimal designs are often characterized by established ratios of geometry, flow and head (Ns, Ds). Such considerations resulting from dimensional similarity are well known to the art of gas processing. See, for example, the publication entitled ‘How to Select Turbomachinery for your Application’ by Kenneth E. Nichols. These optimal turbomachinery conditions are a function of the type of machine under consideration.
In the present system and method, the use of a plurality of centrifugal turbomachines, and, in particular, three radial inflow turbines, find particular application. The present system and method requires or at least contemplates the natural gas feed being a purified, compressed natural gas feed stream at a pressure greater than the critical pressure of natural gas. As used herein, the term purified natural gas feed stream means a natural gas feed stream substantially free of heavy hydrocarbons, carbon dioxide, water, and other impurities and may even be a methane containing biogas. The subsequent and direct liquefaction of a sub-critical natural gas feed stream results in a composite curve divergence near the dewpoint of the mixture. Furthermore, liquefaction of natural gas at pressures lower than about 40 bar(a) generally results in a colder level of warm turbine/expander operation which in turn creates a meaningful penalty in terms of unit power consumption. To avoid this penalty, the natural gas feed stream is preferably at a pressure above the critical pressure of the natural gas feed stream, and more preferably between about 50 bar(a) and 80 bar(a).
Yet another advantageous feature of the present system and method to produce liquefied natural gas is the use of an integral gear machine comprising a drive assembly, a bull gear, and a plurality of pinions arranged or configured to drive two or more refrigerant compression stages and/or for receiving work produced by the three turbine/expanders. The shaft of the bull gear may also be connected via gears to the driver assembly. At least two of the plurality of pinions are net absorbers of power from the drive assembly, which can be an electric motor, a steam turbine, or even a gas turbine. Preferably, the integral gear machine is configured to distribute the power appropriately across the plurality of pinions, and more preferably is arranged or configured such that the power imparted to two pinions coupled to the refrigerant compression stages does not differ by more than 10%. An important aspect or advantage of this integral gear machine arrangements disclosed herein relates to the specific pairings of turbomachinery on the different pinions in a manner that optimizes the performance of the present liquefaction system and method.
The optimization of the turbomachinery starts with a consideration of turbine/expander efficiency. Any given process definition (e.g. Pressures, Temperatures, and Flows) that results in a feasible heat transfer (liquefaction) design also provides the necessary input, such as flow and head characteristics, that are necessary to define the non-dimensional characteristics (Ns, Ds) required to specify component turbine/expander rotational speed and diameter. It is well established that radial inflow turbines reach peak efficiency with U/Co (i.e. Rotor Tip Speed/Isentropic Spouting Velocity) values near 0.70. This ratio is also defined by the following equation [U/Co] = [NsDs]/154.
As such, effective process definition will dictate the speed and diameter necessary for the turbine/expander to operate at peak efficiency. With respect to gas compression, process definition dictates compression stage head and the associated turbine/expander on the same pinion dictates rotational speed which in turn results in a specific speed. The above calculation forms one part of the overall process optimization. More specifically, the optimization is an iterative process involving process definition, turbomachine pairing based upon the above calculation and finally a consideration of the integral gear machine pinion power and overall input power limitations.
Conventional small-scale and medium-scale liquified natural gas plants that use a nitrogen-based gas expansion as the primary source of refrigeration typically employ centrifugal recycle compression stages for the refrigerant that are typically driven by an integral gear machine contained within a common housing that includes a large diameter bull gear with several meshing pinions upon the ends of which the various compression impellers are mounted forming the plurality of refrigerant compression stages and expansion impellers of the turbine/expanders. The pinions may have differing diameters to best match the speed requirements of the coupled compression impellers. Each of the multiple compression impellers and turbine/expanders are typically contained within their own respective housings and collectively provide several stages of recycle compression and expansion , as desired.
Linde Inc., a member of the Linde Group of Companies, has also developed a portfolio of integral gear machines or single machines that combine compression stages and high efficiency radial inflow expanders having up to four pinions in what is referred to as an integral gear ‘bridge’ machine or BRIM. Linde’s ‘bridge’ machines are conventionally used in hydrogen/syngas plants as well as air separation plants and typically come in different frame sizes, for example between about 90 mm and 180 mm frame sizes. Design studies have examined applications of the Linde ‘bridge’ machines to operatively couple a plurality of radially inflow turbines and centrifugal refrigeration compression stages in a natural gas liquefaction system. The Linde ‘bridge’ machines come fully packaged or integrated with appropriate PLC controllers, control valves, safety valves, oil system, etc. and can be easily outfitted with intercoolers and/or aftercoolers. The hardware constraints and limitations of the Linde ‘bridge’ machines are typically a function of bull gear and driver assembly size. In general, the Linde ‘bridge’ machine drivers pertinent for the present system and method spans the range of about 4 MW to 20 MW with associated maximum pinion speeds in the range of 20,000 to 50,000 rpm. Furthermore, the maximum power imparted to any given pinion or any given turbine-compression stage pairing is preferably limited to less than 50% and in some cases to about 35% of the total ‘bridge’ machine driver power.
Turning to
The purified and compressed natural gas feed 12 substantially free of heavy hydrocarbons and other impurities and at a feed pressure that is greater than the critical pressure of natural gas (i.e. above 46 bar(a)), preferably at a pressure of between about 50 bar(a) and 80 bar(a) and more preferably at a pressure between about 60 bar(a) and 75 bar(a) is provided as a feed stream 14 to the depicted natural gas liquefaction system 10.
A first majority portion of the purified, compressed natural gas feed stream 16 is directed to the cooling passages in the heat exchanger(s) 20 where it is liquefied and subcooled via indirect heat exchange with refrigerant streams traversing the warming passages of the heat exchanger(s). A second minor portion of the purified, compressed natural gas feed stream 17 is diverted to the fuel gas circuit 18 comprising one or more valves 19 configured to expand the second minor portion of the purified, compressed natural gas feed stream 17 to a pressure less than about 6.0 bar(a)
As indicated above, the first major portion of the purified, compressed natural gas feed stream 16 is liquefied and subcooled within the heat exchanger(s) 20 via indirect heat exchange against one or more nitrogen-based refrigerant streams to form a subcooled and liquified natural gas stream 21. The subcooled and liquified natural gas stream 21 is thereafter treated in the post liquefaction conditioning circuit 23 where the subcooled and liquefied natural gas is reduced in pressure via one or more valves 27, or a liquid turbine (not shown), and phase separated using a phase separator 28 to separate nitrogen vapor and other light gases. The resulting liquid natural gas stream 29 constitutes the liquefied natural gas product.
The primary refrigeration source used in the illustrated natural gas liquefaction system 10 is preferably a nitrogen-based gas expansion refrigeration circuit, that preferably includes refrigerant stream(s) that comprises more than about 80% nitrogen by volume. In such illustrated refrigeration circuit, the refrigerant is compressed in two serially arranged refrigerant compression stages, namely an upstream refrigerant compression stage 40 and a downstream refrigerant compression stage 50 with appropriate intercooling and/or aftercooling 221,222 used to offset the temperature increases caused by the heat of compression. Such aftercooling may be accomplished by way of indirect contact with air, cooling water, chilled water or other refrigerating medium or combinations thereof. The compressed refrigerant 55 is then further cooled in the at least one heat exchanger(s) 20 and directed to one or more turbine/expanders 70, 80, 90 configured to expand the compressed refrigerant streams to generate refrigeration.
The embodiments of
Specifically, a first portion of the compressed refrigerant stream 72 is substantially cooled in the heat exchanger and directed to a cold turbine/expander 70 as a cold portion of the refrigerant stream. A second portion of the compressed refrigerant stream 82 is partially cooled and exits the heat exchanger 20 at an intermediate warmer temperature as a first warm portion which is then directed to a first warm turbine/expander 80. A third portion of the compressed refrigerant stream 92 is also partially cooled and exits the heat exchanger 20 as a second warm portion of the compressed refrigerant stream having a temperature warmer than the intermediate warmer temperature. The second warm portion of the compressed refrigerant stream 92 is then directed to a second warm turbine/expander 90.
The cold turbine/expander 70 is configured to expand the cold portion of the compressed refrigerant stream 72 to produce a cold turbine exhaust stream 74 that is recycled as warmed stream 76 back to the refrigerant compression stages 40,50 via one or more of the plurality of warming passages in the heat exchanger(s) 20. The partially cooled first warm portion of the compressed refrigerant stream 82 is expanded in the first warm turbine/expander 80 to produce a first warm turbine exhaust stream 84 that is also recycled as warmed stream 86 to the one or more refrigerant compression stages 40,50 via one or more of the plurality of warming passages in the heat exchanger(s) 20. The partially cooled second warm portion of the compressed refrigerant stream 92 is expanded in the second warm turbine/expander 90 to produce a second warm turbine exhaust stream 94 that is also recycled as warmed stream 96 to the downstream refrigerant compression stages 50 via one or more of the plurality of warming passages in the heat exchanger(s) 20.
The inlet pressures of the three turbine/expanders are approximately equal but the outlet pressures are different. Specifically, the expansion ratio of the cold turbine/expander 70 and the first warm turbine expander 80 are preferably between about 4.0 and 5.0. Using similar expansion ratios, the cold turbine exhaust 74 and the first warm turbine exhaust 84 may be warmed in the heat exchanger using the same warming pressure. Alternatively, the cold exhaust and the first warm exhaust may be warmed in independent passages of the heat exchanger(s) and/or may be at different outlet pressures. An important and advantageous feature of the present system and method is that the second warm turbine/expander 90 has an expansion ratio much less than the expansion ratio of the cold turbine/expander 70 and first warm turbine/expander 80. Preferably, the second warm turbine/expander has an expansion ratio of between 1.5 and 2.5 and since the second warm exhaust 94 is at a pressure greater than the cold turbine exhaust 74 and the first warm turbine exhaust 84, it should be warmed in an independent passage of the heat exchanger(s) 20.
Upon exiting the heat exchanger 20, the warmed cold turbine exhaust stream 76 and the warmed first warm turbine exhaust stream 86 are recycled as a lower pressure recycle stream 42 to the upstream refrigerant compression stage 40 where the lower pressure recycle stream 42 is compressed to form stream 44 and then cooled in the upstream aftercooler 221 to yield stream 46. The warmed second warm turbine exhaust stream 96 is also recycled as a higher pressure recycle stream and is mixed with the compressed refrigerant stream 46 exiting the upstream refrigerant compression stage. This mixed stream 52 is then directed to the downstream refrigerant compression stage 50 where it is further compressed to form the compressed refrigerant stream 54 and subsequently cooled in the downstream aftercooler 222 to form stream 55 and further cooled in heat exchanger(s) 20.
In the depicted embodiment, the cold turbine exhaust stream 74 is at a temperature colder than -145° C. while the first warm turbine exhaust stream 84 is at a temperature colder than -90° C. but warmer than the cold turbine exhaust stream. The second warm turbine exhaust 94 is at a temperature above the critical point temperature of the compressed natural gas feed stream 14 and warmer than the first warm turbine exhaust stream 84 and preferably colder than about -15° C. Also, the distribution of the compressed refrigerant stream between the cold portion 72, the first warm portion 82, and the second warm portion 92 is such that the first warm turbine/expander 80 is configured to produce over 45%, and more preferably over 50% of the refrigeration for the natural gas liquefaction system 10. The cold turbine/expander 70 is configured to produce less than 25%, and more preferably less than 20% of the refrigeration for the natural gas liquefaction system 10 while the second warm turbine/expander 90 is configured to produce between about 20% to 35% of the refrigeration for the liquefaction system 10 .
The first warm turbine/expander 80, the second warm turbine/expander 90, and the cold turbine/expander 70 as well as the upstream refrigerant compression stage 40 and the downstream refrigerant compression stage 50 are operatively coupled to the integral gear machine 25. In particular, downstream refrigerant compression stage 50 and the first warm turbine/expander 80 are operatively coupled to the same pinion on the bull gear 26 of the integral gear machine 25, identified as the second pinion 32 of the three pinion integral gear machine. Likewise, the upstream refrigerant compression stage 40 and the second warm turbine/expander 90 are operatively coupled to the same pinion of the integral gear machine 25, shown as the first pinion 31. The cold turbine/expander 70 is coupled to yet a different pinion, shown as the third pinion 33 of the integral gear machine 25.
Turning to
Similar to the embodiment shown in
As indicated above, the primary refrigeration source is preferably a nitrogen-based gas expansion refrigeration circuit, that preferably includes refrigerant stream(s) that comprises more than about 80% nitrogen by volume. In the embodiment depicted in
As shown in
The warmed second warm turbine exhaust stream is also recycled as a higher pressure recycle stream 96 and is mixed with the compressed refrigerant streams exiting the first and second upstream refrigerant compression stages, preferably downstream of the upstream aftercooler 221. This mixed stream 52 is then directed to the downstream refrigerant compression stage 50 where it is further compressed to form the compressed refrigerant stream 54 which is then cooled in the downstream aftercooler 222. The cooled, compressed refrigerant stream 55 is then further cooled in the heat exchanger(s) 20 and directed to one or more turbine/expanders configured to expand the compressed refrigerant streams to generate refrigeration for the natural gas liquefaction system 10.
A number of computer simulations were run to characterize the performance of the present natural gas liquefaction system and processes. In one such computer simulation, referred to as Case 1, a natural gas liquefaction system designed to produce 175 metric tonnes per day of liquefied natural gas from a compressed, purified natural gas feed stream at a pressure of about 68 bar(a) and a temperature of about 30° C. was evaluated using the arrangement disclosed with reference to
Table 1A provides the work distribution in this example using the embodiment of the three pinion integral gear machine used in the three turbine/expander and two refrigerant compression stage system schematically depicted in
In the Case 1 simulation, the speed of the cold turbine/expander is the variable that constrains the process cycle and, in this example, approaches a speed of about 45,000 rpm. Note the integral gear machine or ‘bridge’ machine receives the work from the cold turbine/expander on the third pinion as it is unpaired with any refrigeration compression stage. The other two pinions are net absorbers of power from the drive assembly of the integral gear machine and the power is distributed to these two pinions in generally equal or roughly equal proportions. Note, however, that the upstream refrigeration compression stage is designed to compress over 71% of the refrigerant and this compressed refrigerant is mixed or combined with the higher pressure recycle stream which contains the remaining 29% of the refrigerant. The downstream refrigerant compression stage is thus designed to further compresses the entire refrigerant stream.
The distribution of the fully compressed refrigerant stream between the cold turbine/expander, the first warm turbine/expander, and second warm turbine/expander in this Case 1 example is such that the first warm turbine/expander is configured to receive almost 50% of the compressed refrigeration stream and expands the stream from an inlet pressure of 50.2 bar(a) to an outlet pressure of 11.68 bar(a) or an expansion ratio of 4.29. The cold turbine/expander, on the other hand receives over 21% of the compressed refrigeration stream and expands the stream from an inlet pressure of 49.85 bar(a) to an outlet pressure of 11.78 bar(a) or an expansion ratio of 4.23 .while second warm turbine/expander receives about 29% of the compressed refrigeration stream and expands the stream from an inlet pressure of 50.4 bar(a) to an outlet pressure of 24.08 bar(a) or an expansion ratio of 2.09.
As indicated above, designs of small to mid-scale natural gas liquefaction cycles and liquefaction systems, involve numerous trade-offs between capital costs and operational efficiencies. The natural gas liquefaction system shown in
In another computer simulation, referred to as Case 2, a natural gas liquefaction system designed to produce 320 metric tonnes per day of liquefied natural gas from a compressed, purified natural gas feed stream at a pressure of about 68 bar(a) and a temperature of about 30° C. was evaluated using the three turbine/expander and three refrigerant compression stage arrangement disclosed in
In the Case 2 simulation, which examines a higher production capacity, the cold turbine/expander on the third pinion as it is paired with one of the upstream refrigeration compression stages while the second warm turbine/expander on the first pinion is paired with the other upstream refrigeration compression stage. The first warm turbine/expander on the second pinion is paired with the downstream refrigeration compression stage and all three pinions are net absorbers of power from the drive assembly of the integral gear machine. In this Case 2 example, the power is distributed to the three pinions in a manner where the second pinion coupling the first warm turbine/expander and the downstream refrigerant compression stage absorbs 49% of the power while the first pinion coupling the second warm turbine/expander and one of the upstream refrigerant compression stages absorbs 32.6% of the power and the third pinion coupling the cold turbine/expander and another of the upstream refrigerant compression stages absorbs 18.4% of the power. In this high production capacity example of Case 2, the integral gear machine is configured to absorb near the maximum total absorbable power for the subject ‘bridge’ machine.
Note, however, that the upstream refrigeration compression stages arranged in parallel are configured to compress over 68% of the total refrigerant. Specifically, the first upstream refrigeration compression stage compresses about 65% of the lower pressure recycle stream and the second upstream refrigeration compression stage compresses about 35% of the lower pressure recycle stream exiting the heat exchanger. These compressed streams are combined and directed to the upstream aftercooler and the resulting cooled stream is mixed or further combined with the higher pressure recycle stream which contains the remaining portion of the refrigerant, nearly 32%. The downstream refrigerant compression stage is thus designed to further compresses the entire refrigerant stream.
The distribution of the fully compressed refrigerant stream between the cold turbine/expander, the first warm turbine/expander, and second warm turbine/expander in this Case 2 example is such that the first warm turbine/expander is configured to receive 49.1% of the compressed refrigeration stream and expands the refrigerant stream from an inlet pressure of 52.4 bar(a) to an outlet pressure of 11.68 bar(a) or an expansion ratio of about 4.5. The cold turbine/expander, on the other hand receives about 19.3% of the compressed refrigeration stream and expands the refrigerant stream from an inlet pressure of 52.05 bar(a) to an outlet pressure of 11.78 bar(a) or an expansion ratio of 4.42 .while second warm turbine/expander receives about 31.6% of the compressed refrigeration stream and expands the refrigerant stream from an inlet pressure of 52.6 bar(a) to an outlet pressure of 26.84 bar(a) or an expansion ratio of 1.96.
The natural gas liquefaction process using the three pinion and three turbine/expander arrangements discussed above with reference to
While the present natural gas liquefaction systems and methods have been described with reference to several preferred embodiments, it is understood that numerous additions, changes, and omissions can be made without departing from the spirit and scope of the present inventions as set forth in the appended claims.
This application claims the benefit of and priority to United States provisional pat. application serial number 63/255,116 filed Oct. 13, 2021, the disclosure of which is incorporated by reference
| Number | Date | Country | |
|---|---|---|---|
| 63255116 | Oct 2021 | US |
