US20230129424A1

US20230129424A1 - System and method to produce liquefied natural gas

Info

Publication number: US20230129424A1
Application number: US17/717,197
Authority: US
Inventors: Henry Edward Howard
Original assignee: Individual
Current assignee: Praxair Technology Inc
Priority date: 2021-10-21
Filing date: 2022-04-11
Publication date: 2023-04-27
Also published as: WO2023069139A1

Abstract

A small to mid-scale liquefied natural gas production system and method is provided. The disclosed liquefied natural gas production system employs at least one heat exchanger, three turbine/expanders and at least three refrigerant compression stages. The expansion ratio of one turbine/expander is appreciably lower than the expansion ratio of the other turbine/expanders such that the temperature of the exhaust stream from the turbine/expander with the lower expansion ratio is above the critical point temperature of the compressed natural gas containing feed stream but colder than about −15° C. The present system and method may be configured using either a single nitrogen-based expansion refrigerant circuit or two separate refrigerant circuits wherein the turbine/expander with the lowest expansion ratio is contained within a separate refrigeration circuit from the other two turbine/expanders with the higher expansion ratios.

Description

CROSS REFERENCE TO RELATED APPLICATIONS

This application claims the benefit of and priority to U.S. provisional patent application Ser. No. 63/270,186 filed Oct. 21, 2021.

TECHNICAL FIELD

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 at a plurality of refrigerant compression stages and employing either a single refrigerant circuit or two separate refrigerant circuits.

BACKGROUND

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 CO₂extraction 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 FIG. 1 , there exists a need to improve the thermal efficiency of such systems. The use of only two turbine/expanders and the condensing profile of natural gas result in meaningful divergences in the heat exchanger composite curves. Coupling the turbine/expanders to one or more compression stages via an integral gear machine or ‘compander’ further complicates efforts to improve thermal performance. Specifically, one cannot simply manipulate turbine expansion ratio, flow and thermal positioning independent a simultaneous consideration of the turbomachinery performance and any such additional turbomachinery-complexity would require a power reduction commensurately large to offset such additional capital.
Another limiting aspect of the conventional natural gas liquefaction system and process depicted in FIG. 1 is found with respect to the temperature levels served by each of the turbine/expanders. Since the cold turbine/expander provides the subcooling duty necessary to prevent any meaningful loss of product upon depressurization, the exit state is largely fixed by the cold-end delta temperature (CEDT) of the heat exchanger and the condition of saturation (which minimizes unit power consumption). The cold turbine/expander inlet state is defined by a narrow range of temperature in which the coldest portion of the composite curves can be made to roughly match. As the inlet temperature to the cold turbine/expander approaches the pseudo dew-point inflection temperature of natural gas, it becomes impossible for the warming exhaust flow to match the subcooling curve of natural gas. Given these considerations, and the parallel arrangement, the pressure ratios are largely fixed and/or limited by the cold turbine/expander operation.
The conventional two turbine/expander liquefaction system shown in FIG. 1 also exhibits a highly skewed distribution of refrigeration. Since the warm turbine/expander in such conventional natural gas liquefaction systems discharges below the critical point temperature of the natural gas (i.e. −82.6° C.), its flow absorbs much of the duty associated with precooling the refrigerant and natural gas flows as well as the duty of NG pseudo-condensation. In the conventional natural gas liquefaction system and process depicted in FIG. 1 , the warm turbine/expander accounts for about 69% of the recycle refrigerant flow and supplies about 83% of the delivered refrigeration. Consequently, the absorbed power of pinion #2 which couples the cold turbine/expander to a downstream compression stage is substantially higher than that of pinion #1 which couples the warm turbine/expander to an upstream compression stage. This arrangement complicates both the design of the turbomachinery as well as the ability of the process to fully utilize the capacity of any given ‘compander’ frame.
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 U.S. 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 Dubar based three turbine/expander liquefaction arrangement 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.

SUMMARY OF THE INVENTION

The present invention may be characterized as a natural gas liquefaction system comprising: a refrigeration circuit and an integral gear machine. The refrigeration circuit includes: a natural gas liquefaction system, comprising: (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 or more 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) at least three refrigerant compression stages including an upstream refrigerant compression stage and a pair of downstream refrigerant compression stages arranged in parallel, wherein the three refrigerant compression stages are configured to compress the warmed recycle streams. The integral gear machine includes a drive assembly, a bull gear, and at least three pinions arranged to drive the at least three refrigerant compression stages and/or for receiving work produced by the at least three turbines/expanders. All three pinions are configured to be net absorbers of power from the drive assembly of the integral gear machine and the power is distributed to these three pinions in generally equal or roughly equal proportions of between 30% and 40% of the total power to each of the three pinions.
More specifically, the three or more turbines/expanders further comprise: a cold turbine/expander configured to expand a cold portion of the refrigerant stream and produce a cold exhaust that is also recycled to the upstream refrigerant compression stage; a first warm turbine/expander configured to expand a first warm portion of the refrigerant stream and produce a first warm exhaust to be recycled to the downstream 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 to be recycled to the downstream refrigerant compression stages.
The present invention may also be characterized as a natural gas liquefaction system comprising at least one heat exchanger configured to liquefy and subcool a compressed natural gas containing feed stream via indirect heat exchange with a nitrogen-based refrigerant stream from a first refrigeration circuit and a secondary refrigerant stream traversing a secondary refrigeration circuit. The first refrigeration circuit includes at least two turbine/expanders configured to expand portions of the nitrogen-based refrigerant stream to produce one or more 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. The first refrigeration circuit also includes at least two primary refrigerant compression stages including an upstream refrigerant compression stage and a serially arranged downstream refrigerant compression stage configured to compress the warmed recycle streams. The second refrigeration circuit includes at least one turbine/expander configured to expand portions of the secondary refrigerant stream to produce a secondary exhaust stream that is directed to the heat exchanger and exits the heat exchanger as a warmed secondary recycle stream. The second refrigeration circuit also includes at least one secondary refrigerant compression stage configured to compress the warmed secondary recycle stream. If used, the secondary refrigerant is a different composition than the nitrogen based refrigerant stream and preferably a natural gas or other refrigerant, including hydrocarbon based refrigerants.
When using the separate refrigeration circuits, the cold turbine/expander is configured to expand a cold portion of the nitrogen-based refrigerant stream and produce a cold exhaust that is also recycled to the upstream refrigerant compression stage. The first warm turbine/expander is configured to expand a first warm portion of the nitrogen-based refrigerant stream and produce a first warm exhaust to also be recycled to the upstream refrigerant compression stage. The second warm turbine/expander is configured to expand a second warm portion of the secondary refrigerant stream and produce a second warm exhaust to be recycled to the secondary refrigerant compression stage in the second refrigeration circuit.
In all embodiments, 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 warm turbine/expander. The second warm exhaust is above the critical point temperature of the natural gas containing feed stream and preferably less than about −15° C. Preferably, 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 turbine work used to produce 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 turbine work used to produce 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 turbine work.
In all embodiments, an integral gear machine comprising a drive assembly; a bull gear; and at least three pinions is configured to drive the plurality of refrigerant compression stages and for receiving work produced by the plurality of turbine/expanders. Also, the purified, compressed natural gas 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.

BRIEF DESCRIPTION OF THE DRAWINGS

It is believed that the claimed invention will be better understood when taken in connection with the accompanying drawings in which:

FIG. 1 shows a generalized schematic of the process flow diagram for a conventional two turbine and two refrigerant compression stage natural gas liquefaction process known in the prior art;

FIG. 2 shows a schematic of the process flow diagram for an embodiment of the present system and method for liquefied natural gas production using three turbine/expanders and three refrigerant compression stages using a single refrigerant and having two of the three refrigerant compression stages arranged in parallel; and

FIG. 3 shows a schematic of the process flow diagram for another embodiment of the present system and method for liquefied natural gas production using three turbine/expanders and three refrigerant compression stages having two separate refrigerant circuits.

DETAILED DESCRIPTION

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 containing feed being a purified, compressed natural gas feed stream at a pressure greater than the critical pressure of natural gas but it may originate from a source of methane containing biogas. As used herein, the term purified natural gas containing feed stream means a natural gas feed stream substantially free of heavy hydrocarbons, carbon dioxide, water, and other impurities. 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 containing 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. 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.

Three (3) Turbine LNG Production System with Single Refrigeration Circuit

Turning to FIG. 2 , a schematic of the high-level process flow diagram for one embodiment of the present system and method for liquefied natural gas production using three turbine/expanders having a single nitrogen-based expansion refrigerant circuit is shown. The illustrated refrigerant circuit includes at least one heat exchanger, two aftercoolers; three turbine expanders, and three refrigerant compression stages wherein two of the refrigerant compression stages are arranged in parallel. The illustrated system also includes a three-pinion integral gear machine, a fuel gas circuit, and a post liquefaction conditioning circuit, having one or more expansion valves and a phase separator configured for separating nitrogen and other light gases from the liquefied and subcooled natural gas stream.
The purified and compressed natural gas containing feed 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 to the depicted natural gas liquefaction system.
As indicated above, the purified, compressed natural gas feed stream is liquefied and subcooled within the heat exchanger(s) via indirect heat exchange against one or more nitrogen-based refrigerant streams to form a subcooled and liquified natural gas stream. The subcooled and liquified natural gas stream is thereafter treated in the post liquefaction conditioning circuit where the subcooled and liquefied natural gas is reduced in pressure via one or more valves, or a liquid turbine (not shown), and phase separated using a phase separator to separate nitrogen vapor and other light gases. The resulting liquid natural gas stream constitutes the liquefied natural gas product.
The primary refrigeration source used in the illustrated natural gas liquefaction system 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 a plurality of refrigerant compression stages, namely an upstream refrigerant compression stage and two downstream refrigerant compression stages arranged in parallel with appropriate intercooling and/or aftercooling 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 stream is then further cooled in the at least one heat exchanger(s) and directed to one or more turbine/expanders configured to expand the compressed refrigerant streams to generate refrigeration.
While the embodiment of FIG. 2 depicts a single heat exchanger having multiple warming passages and multiple cooling passages. Alternatively, the at least one heat exchanger can include multiple heat exchangers or multiple heat exchange cores with a first heat exchanger, or first heat exchange core configured for liquefying the natural gas feed stream and a second heat exchanger or second heat exchange core configured for cooling other streams, such as pre-cooling a portion of the refrigerant stream or perhaps even pre-cooling the natural gas feed stream. Any such second heat exchanger or second heat exchange core would preferably achieve such pre-cooling with the exhaust stream from the second warm turbine, discussed below.
Specifically, a first portion of the compressed refrigerant stream is substantially cooled in the heat exchanger and directed to a cold turbine/expander as a cold portion of the refrigerant stream. A second portion of the compressed refrigerant stream is partially cooled and exits the heat exchanger at an intermediate warmer temperature as a first warm portion which is then directed to a first warm turbine/expander. A third portion of the compressed refrigerant stream is also partially cooled and exits the heat exchanger 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 is then directed to a second warm turbine/expander.
The cold turbine/expander is configured to expand the cold portion of the compressed refrigerant stream to produce a cold exhaust stream that is recycled back to the refrigerant compression stages via one or more of the plurality of warming passages in the heat exchanger(s). The partially cooled first warm portion of the compressed refrigerant stream is expanded in the first warm turbine/expander to produce a first warm exhaust stream that is also recycled to the one or more refrigerant compression stages via one or more of the plurality of warming passages in the heat exchanger(s). The partially cooled second warm portion of the compressed refrigerant stream is expanded in the second warm turbine/expander to produce a second warm exhaust stream that is also recycled to the one or more refrigerant compression stages via one or more of the plurality of warming passages in the heat exchanger(s).
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 and the first warm turbine expander are preferably between about 4.0 and 5.0. Using similar expansion ratios, the cold exhaust and the first warm exhaust 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 has an expansion ratio much less than the expansion ratio of the cold turbine/expander and first warm turbine/expander. Preferably, the second warm turbine/expander has an expansion ratio of between 1.5 and 2.5 and since the second warm exhaust is at a pressure greater than the cold exhaust and the first warm exhaust, it should be warmed in an independent passage of the heat exchanger(s).
Upon exiting the heat exchanger, the warmed cold turbine exhaust and the warmed first warm turbine exhaust are recycled as a lower pressure recycle stream to the upstream refrigerant compression stage where the lower pressure recycle stream is compressed with the resulting compressed recycle stream being aftercooled in the upstream aftercooler. The warmed second warm turbine exhaust is also recycled as a higher pressure recycle stream and is mixed with the aftercooled, compressed refrigerant stream exiting the upstream refrigerant compression stage. This mixed stream is then split into a first refrigerant stream and a second refrigerant stream. The first refrigerant stream is directed to one of the parallel downstream refrigerant compression stages, namely a first downstream refrigerant compression stage while the second refrigerant stream is directed to the other of the parallel downstream refrigerant compression stages, namely a second downstream refrigerant compression stage. The preferred split of the first refrigerant stream and second refrigerant stream is roughly between 35% and 45% of the flow is taken as the first refrigerant stream while 55% to 65% of the flow is taken as the second refrigerant stream. The first and second refrigerant streams are then further compressed in the respective downstream refrigerant compression stages, recombined into a further compressed recycle stream and subsequently cooled in the downstream aftercooler.
In the depicted embodiment, the cold exhaust is at a temperature colder than −145° C. while the first warm exhaust is at a temperature colder than −90° C. but warmer than the cold exhaust. The second warm exhaust is at a temperature above the critical point temperature of the compressed natural gas feed stream and warmer than the first warm exhaust and preferably colder than about −15° C. Also, the distribution of the compressed refrigerant stream between the cold portion, the first warm portion, and the second warm portion is such that the first warm turbine/expander is configured to produce over 45% of the turbine work used to produce the refrigeration for the natural gas liquefaction system. The cold turbine/expander is configured to produce less than 25% of the turbine work used to produce the refrigeration for the natural gas liquefaction system while the second warm turbine/expander is configured to produce between about 25% to 35% of the turbine work used to produce the refrigeration for the liquefaction system.
The first warm turbine/expander, the second warm turbine/expander, and the cold turbine/expander as well as the upstream refrigerant compression stage and the downstream refrigerant compression stages are operatively coupled to the integral gear machine. In particular, the first downstream refrigerant compression stage and the cold turbine/expander are operatively coupled to the same pinion of the integral gear machine, identified as the second pinion of the three pinion integral gear machine. Likewise, the second downstream refrigerant compression stage and the second warm turbine/expander are operatively coupled to the same pinion of the integral gear machine, shown as the first pinion. The first warm turbine/expander and the upstream refrigerant compression stage are coupled to yet a different pinion, shown as the third pinion of the integral gear machine.
A portion of the liquified and subcooled natural gas feed stream may be diverted to the fuel gas circuit. The fuel gas circuit includes one or more valves configured to expand the diverted portion of the liquified and subcooled natural gas stream to a pressure less than about 6.0 bar(a). The lower pressure fuel gas stream is then directed to the heat exchanger to subcool the purified, compressed natural gas stream with the warmed, low pressure fuel gas stream exits the heat exchanger near ambient temperature to be used or stored as fuel gas.

Three (3) Turbine LNG Production System with Separate Refrigeration Circuits

Turning to FIG. 3 , there is shown a schematic of the high-level process flow diagram for another embodiment of the present system and method for liquefied natural gas production using three turbine/expanders and three refrigerant compression stages. Many of the features, components and streams associated with the natural gas liquefaction system shown in FIG. 3 are similar or identical to those described above with reference to the embodiment of FIG. 2 and for sake of brevity will not be repeated here. The key differences between the natural gas liquefaction system shown in FIG. 3 compared to the natural gas liquefaction system described above with reference to FIG. 2 , is the addition of a separate and distinct refrigeration circuit that includes the second warm turbine/expander, one of the downstream refrigerant compression stages, a second downstream aftercooler, and dedicated cooling and warming passages in the at least one heat exchanger. Note that by separating the two downstream refrigerant compression stages into separate refrigeration circuits, the downstream refrigerant compression stages are no longer arranged in parallel in the embodiment of FIG. 3 .
The natural gas liquefaction system illustrated in FIG. 3 includes a first refrigerant circuit with the cold turbine/expander, the first warm turbine expander, the upstream refrigerant compression stage, a first downstream refrigerant compression stage, the at least one heat exchanger and two aftercoolers. The natural gas liquefaction system illustrated in FIG. 3 also includes a mixed service integral gear machine, a fuel gas circuit, and a post liquefaction conditioning circuit, similar to those described above with reference to FIG. 2 . The natural gas liquefaction system of FIG. 3 also includes a separate and distinct second refrigeration circuit that includes the second warm turbine/expander, one of the downstream refrigerant compression stages, a second downstream aftercooler, and dedicated cooling and warming passages in the at least one heat exchanger.
As indicated above, the purified and compressed natural gas feed is at a feed pressure that is greater than the critical pressure of natural gas and preferably at a pressure of between about 50 bar(a) and 80 bar(a). The primary refrigerant source in the first refrigeration circuit is preferably a stream that comprises more than about 80% nitrogen by volume while the secondary refrigerant source in the second refrigeration circuit has a different composition than the nitrogen-based refrigerant and is preferably a stream of natural gas or another hydrocarbon based refrigerant.
As detailed in FIG. 3 , the warmed cold exhaust and the warmed first warm exhaust exiting the heat exchanger are recycled as a lower pressure nitrogen recycle stream to the upstream refrigerant compression stage. The compressed refrigerant stream exiting the upstream refrigerant compression stage is then cooled in the upstream aftercooler. The aftercooled stream is then directed to the first downstream refrigerant compression stage where it is further compressed as further compressed recycle stream and aftercooled in downstream aftercooler. The cooled, further compressed nitrogen recycle stream is then directed to the at least one heat exchanger(s) where it is cooled to the appropriate inlet temperatures for the first warm turbine/expander and the cold turbine/expander.
The warmed second warm exhaust is also recycled as a higher pressure recycle stream and directed to the second downstream refrigerant compression stage where it is further compressed to form the secondary refrigerant stream which is then cooled in the third aftercooler. The secondary refrigerant stream is then partially cooled in the heat exchanger(s) and directed to the second warm turbine configured to expand the secondary refrigerant stream to generate refrigeration.
In this embodiment, the cold exhaust is also at a temperature colder than about −145° C. while the first warm exhaust is at a temperature colder than −90° C. but warmer than the cold exhaust. The second warm exhaust is at a temperature above the critical point temperature of the compressed natural gas feed stream and warmer than the first warm exhaust and preferably colder than about −15° C. In this embodiment the first warm turbine/expander is configured to produce over 60% of the turbine work used to produce the refrigeration for the natural gas liquefaction system while the cold turbine/expander is configured to produce less than 20% of the turbine work used to produce the refrigeration for the natural gas liquefaction system. The second warm turbine/expander in the separate refrigeration circuit is preferably designed or configured to produce between about 20% to 30% of the turbine work used to produce the refrigeration for the liquefaction system.
The first warm turbine/expander, the second warm turbine/expander, and the cold turbine/expander as well as the upstream refrigerant compression stage and the downstream refrigerant compression stages are operatively coupled to the integral gear machine. In particular, the first downstream refrigerant compression stage and the cold turbine/expander are operatively coupled to the same pinion of the integral gear machine, identified as the second pinion of the three pinion integral gear machine. Likewise, the first warm turbine/expander and the upstream refrigerant compression stage are coupled to yet a different pinion, shown as the third pinion of the integral gear machine. Lastly, the secondary refrigerant compression stage and the second warm turbine/expander of the second refrigeration circuit are operatively coupled to the same pinion of the integral gear machine, shown as the first pinion.

Examples of LNG Production

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 at 164.4° C. and 1.5 bar(a) 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 above with reference to FIG. 2 .
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 FIG. 2 . Similarly, Table 1B provides the process flow and refrigerant stream characteristics for this example using the same FIG. 2 embodiment of the three turbine/expander and two refrigerant compression stage natural gas liquefaction system.

TABLE 1A

					Net
FIG. 2		Power		Power	Power
Pinion#	Service #1	(kw)	Service #2	(kW)	(kw)

Pinion #1	N2 Comp #CB1	17886	N2 - Cold Turbine	−200	860
Pinion #2	N2 Comp #CB2	1119	N2 - 1^stWarm Turbine	−928	919
Pinion #3	N2 Comp #CB3	1545	N2 - 2^ndWarm Turbine	−496	1049

TABLE 1B

			Refrigerant
FIG. 2		Temp	Flow
Stream Description	Stream#	(° C.)	(% of Total)

CB1 Inlet	M06	33.50	64.0%
CB2 Inlet	M12	34.75	42.0%
CB3 Inlet	M12A	34.75	58.0%
Cold Turbine Inlet	M03	−114.75	18.9%
Cold Turbine Exhaust	M04	−166.90	18.9%
1^stWT Inlet	R02	−40.35	45.1%
1^stWT Exhaust	R03	−114.65	45.1%
2^ndWT Inlet	S02	9.47	36.0%
2^ndWT Exhaust	S03	−38.00	36.0%
Lower Pressure Recycle	M06	33.50	64.0%
High Pressure Recycle	S04	33.50	36.0%
Upstream Aftercooler	M10	36.00	64.0%
Downstream Aftercooler	M15	36.00	100.0%

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 40,000 rpm. Note that all other three pinions are net absorbers of power from the drive assembly of the integral gear machine and the power is distributed to these three pinions in generally equal or roughly equal proportions or 37.1%, 32.5% and 30.4%. Note, however, that the upstream refrigeration compression stage is designed to compress about 64% of the refrigerant and this compressed refrigerant is mixed or combined with the higher pressure recycle stream which contains the remaining 36% of the refrigerant. The downstream refrigerant compression stages arranged in parallel are 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 49.1% of the compressed refrigeration stream and expands the stream from an inlet pressure of 48.1 bar(a) to an outlet pressure of 11.68 bar(a) or an expansion ratio of 4.12. The cold turbine/expander, on the other hand receives only 18.9% of the compressed refrigeration stream and expands the stream from an inlet pressure of 47.75 bar(a) to an outlet pressure of 11.78 bar(a) or an expansion ratio of 4.05 while second warm turbine/expander receives about 36% of the compressed refrigeration stream and expands the stream from an inlet pressure of 48.3 bar(a) to an outlet pressure of 24.15 bar(a) or an expansion ratio of 2.0.
As indicated above, designs of small to mid-scale natural gas liquefaction cycles and liquefaction systems, there are numerous trade-offs between capital costs and operational efficiencies that must be made. The natural gas liquefaction system shown in FIG. 2 and operated in a manner similar to the example of Case 1 is among a very good compromise of thermal performance, capital cost and accommodation of turbomachinery constraints.
In another computer simulation, referred to as Case 2, 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 three turbine/expander and three refrigerant compression stage arrangement disclosed in FIG. 3 . Table 2A provides the work distribution for the example of Case 2 using the embodiment of the three pinion integral gear machine used in the three turbine/expander and three refrigerant compression stage system schematically depicted in FIG. 3 while Table 2B provides the process flow and refrigerant stream characteristics for the three turbine/expander and three refrigerant compression stage natural gas liquefaction system of FIG. 3 .

TABLE 2A

					Net
FIG. 3		Power		Power	Power
Pinion#	Service #1	(kw)	Service #2	(kW)	(kw)

Pinion #1	N2 Comp #CB1	1788	N2 - 1^stWarm Turbine	−928	860
Pinion #2	N2 Com3 #CB2	1713	N2 - Cold Turbine	−200	1513
Pinion #3	NG Comp #CB3	761	NG - 2^ndWarm Turbine	−365	396

TABLE 2B

			Refrigerant
FIG. 3		Temp	Flow
Stream Description	Stream#	(° C.)	(% of Total)

CB1 Inlet	M07	33.50	N2 - 100.0%
CB2 Inlet	M12	36.00	N2 - 100.0%
CB3 Inlet	M07A	32.51	NG - 100.0%
Cold Turbine Inlet	M03	−114.75	N2 - 29.5%
Cold Turbine Exhaust	M04	−166.90	N2 - 29.5%
1^stWT Inlet	R02	−40.35	N2 - 70.5%
1^stWT Exhaust Inlet	R03	−114.65	N2 - 70.5%
2^ndWT Inlet	S02	−4.84	NG - 100.0%
2^ndWT Exhaust	S03	−38.00	NG - 100.0%
Higher Pressure Recycle	S04	32.51	NG - 100.0%
Lower Pressure Recycle	M08	33.50	N2 - 100.0%
Upstream Aftercooler	M10	36.00	N2 - 100.0%
Downstream N2 Aftercooler	M15	36.00	N2 - 100.0%
Downstream NG Aftercooler	S01	36.00	N2 - 100.0%

In the Case 2 simulation, the nitrogen-based or primary refrigeration circuit has the cold turbine/expander on the second pinion as it is paired with the first downstream refrigeration compression stage while the first warm turbine/expander on the first pinion is paired with the upstream refrigeration compression stage. In the natural gas based or secondary refrigeration circuit, the second warm turbine/expander on the third pinion is paired with the natural gas refrigeration compression stage.
Note all three pinions are net absorbers of power from the drive assembly of the integral gear machine but unlike the Case 1 example, the power distribution to the three pinions is not uniform. Specifically, the second pinion coupling the cold turbine/expander and the first downstream refrigerant compression stage absorbs 55% of the power while the first pinion coupling the first warm turbine/expander and the upstream refrigerant compression stage absorbs 31% of the power and the third pinion coupling the natural gas second warm turbine/expander and the natural gas compression stage absorbs just over 14% of the power.
The first warm turbine/expander is configured to expand the nitrogen based or primary refrigerant stream from an inlet pressure of 48.10 bar(a) to an outlet pressure of 11.68 bar(a) or an expansion ratio of about 4.12. The cold turbine/expander also expands the primary refrigerant stream from an inlet pressure of 47.75 bar(a) to an outlet pressure of 11.78 bar(a) or an expansion ratio of 4.05. In the secondary refrigeration circuit, the second warm turbine/expander expands the natural gas or secondary refrigerant stream from an inlet pressure of 67.10 bar(a) to an outlet pressure of 39.5 bar(a) or an expansion ratio of about 1.7.
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.

Claims

What is claimed is:

1. A natural gas liquefaction system, comprising:

a refrigeration circuit comprising: (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 or more 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) at least three refrigerant compression stages including an upstream refrigerant compression stage and a pair of downstream refrigerant compression stages arranged in parallel, wherein the three refrigerant compression stages are configured to compress the warmed recycle streams;

an integral gear machine comprising a drive assembly, a bull gear, and at least three pinions arranged to drive the at least three refrigerant compression stages and/or for receiving work produced by the at least three turbines/expanders;

wherein the three or more turbines/expanders further comprise: (i) a cold turbine/expander configured to expand a cold portion of the refrigerant stream and produce a cold exhaust that is also recycled to the upstream refrigerant compression stage; (ii) a first warm turbine/expander configured to expand a first warm portion of the refrigerant stream and produce a first warm exhaust to be recycled to the downstream refrigerant compression stages; and (iii) a second warm turbine/expander configured to expand a second warm portion of the refrigerant stream and produce a second warm exhaust to be recycled to the downstream refrigerant compression stages;

wherein an 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 warm turbine/expander.

2. The natural gas liquefaction system of claim 1, wherein the second warm exhaust is above the critical point temperature of the compressed natural gas containing feed stream and less than about −15° C.

3. The natural gas liquefaction system of claim 1, wherein the first warm turbine/expander is configured with an expansion ratio of between 4.0 and 5.0 and is further configured to produce over 50% of the turbine work used to produce refrigeration for the natural gas liquefaction system.

4. The natural gas liquefaction system of claim 1, wherein the cold turbine/expander is configured with an expansion ratio of between 4.0 and 5.0 and is further configured to produce less than 20% of the turbine work used to produce refrigeration for the natural gas liquefaction system.

5. The natural gas liquefaction system of claim 1, wherein the second warm turbine/expander is configured with an expansion ratio of between 1.5 and 2.5 and is further configured to produce between about 20% to 35% of the turbine work used to produce refrigeration for the natural gas liquefaction system.

6. The natural gas liquefaction system of claim 1, wherein the first warm turbine/expander and the upstream compression stage are operatively coupled to a first pinion of the at least three pinions, and the cold turbine/expander and one of the pair of downstream compression stages are operatively coupled to a second pinion of the at least three pinions, and the second warm turbine/expander and another of the pair of downstream compression stages are operatively coupled to a third pinion of the at least three pinions.

7. The natural gas liquefaction system of claim 1, wherein all three pinions are net absorbers of power and the power is distributed to these three pinions in generally equal or roughly equal proportions.

8. The natural gas liquefaction system of claim 1, wherein the compressed natural gas containing feed stream is a methane containing biogas feed stream.

9. The natural gas liquefaction system of claim 1, wherein the compressed natural gas containing feed stream is at a pressure greater than the critical pressure of natural gas.

10. The natural gas liquefaction system of claim 1, wherein the compressed natural gas containing feed stream is at a pressure between about 50 bar(a) and 80 bar(a).

11. The natural gas liquefaction system of claim 1, wherein the one or more refrigerant streams comprise more than about 80% nitrogen by volume.

12. The natural gas liquefaction system of claim 1, wherein the driver assembly is an electric motor, a steam turbine, or a gas turbine.

13. The natural gas liquefaction system of claim 1, further comprising a phase separator configured for separating nitrogen and other light gases from the liquefied and subcooled natural gas stream.

14. A natural gas liquefaction system, comprising:

at least one heat exchanger configured to liquefy and subcool a compressed natural gas containing feed stream via indirect heat exchange with a nitrogen-based refrigerant stream and a secondary refrigerant stream;

a first refrigeration circuit comprising at least two turbine/expanders configured to expand portions of the nitrogen-based refrigerant stream to produce one or more 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 at least two primary refrigerant compression stages including an upstream refrigerant compression stage and a serially arranged downstream refrigerant compression stage, wherein the refrigerant compression stages are configured to compress the warmed recycle streams;

a second refrigeration circuit comprising at least one turbine/expander configured to expand portions of the secondary refrigerant stream to produce one or more secondary 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 a warmed secondary recycle stream; and at least one secondary refrigerant compression stage configured to compress the warmed secondary recycle stream; and

an integral gear machine comprising a drive assembly; a bull gear; and at least three pinions arranged to drive the at least two primary refrigerant compression stages, the at least one secondary refrigerant compression stage, and for receiving work produced by the turbines/expanders in the first refrigeration circuit and the second refrigeration circuit;

wherein the two or more turbines/expanders in the first refrigeration circuit further comprise: a cold turbine/expander configured to expand a cold portion of the nitrogen-based refrigerant stream and produce a cold exhaust that is also recycled to the upstream refrigerant compression stage; and a first warm turbine/expander configured to expand a first warm portion of the nitrogen-based refrigerant stream and produce a first warm exhaust to be recycled to the upstream refrigerant compression stage;

wherein the at least one turbine/expander in the second refrigeration circuit further comprises: a second warm turbine/expander configured to expand a second warm portion of the secondary refrigerant stream and produce a second warm exhaust to be recycled to the secondary refrigerant compression stage; and

15. The natural gas liquefaction system of claim 14, wherein the second warm exhaust is above the critical point temperature of the compressed natural gas containing feed stream and less than about −15° C.

16. The natural gas liquefaction system of claim 14, wherein the first warm turbine/expander and the cold turbine/expander are each configured with an expansion ratio of between 4.0 and 5.0 and wherein the second warm turbine/expander is configured with an expansion ratio of between 1.5 and 2.5.

17. The natural gas liquefaction system of claim 14, wherein the first warm turbine/expander and the upstream compression stage are operatively coupled to a first pinion of the at least three pinions, and the cold turbine/expander and the downstream compression stage in the first refrigeration circuit are operatively coupled to a second pinion of the at least three pinions, and the second warm turbine/expander and the secondary refrigerant compression stage are operatively coupled to a third pinion of the at least three pinions.

18. The natural gas liquefaction system of claim 14, wherein the compressed natural gas containing feed stream is at a pressure greater than the critical pressure of natural gas and between about 50 bar(a) and 80 bar(a).

19. The natural gas liquefaction system of claim 14, wherein the nitrogen-based refrigerant comprise more than about 80% nitrogen by volume and the secondary refrigerant has a different composition than the nitrogen-based refrigerant.