US20230113326A1

US20230113326A1 - System and method to produce liquefied natural gas

Info

Publication number: US20230113326A1
Application number: US17/717,193
Authority: US
Inventors: Henry Edward Howard; Jeremiah J. Rauch
Original assignee: Individual
Current assignee: Praxair Technology Inc
Priority date: 2021-10-13
Filing date: 2022-04-11
Publication date: 2023-04-13
Also published as: WO2023063993A1

Abstract

A small to mid-scale liquefied natural gas production system and method is provided. The disclosed liquefied natural gas production system employs a nitrogen-based refrigerant, at least one heat exchanger, three turbine/expanders and two or more 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 −15° C.

Description

CROSS REFERENCE TO RELATED APPLICATIONS

This application claims the benefit of and priority to U.S. provisional patent application Ser. No. 63/255,124 filed Oct. 13, 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 two or more refrigerant compression stages.

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: (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/or 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 integral gear machine includes a drive assembly, a bull gear, and at least three pinions arranged to drive the two or more refrigerant compression stages and/or for receiving work produced by the at least three turbines/expanders.
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 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 that is also recycled to the upstream refrigerant compression stage, 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. Two of the at least three pinions are net absorbers of power from the drive assembly and the power imparted to the two of the at least three pinions does not differ by more than about 10%.
In one embodiment, the second warm turbine/expander and the upstream refrigerant compression stage are operatively coupled to a first pinion of the at least three pinions of the integral gear machine while the first warm turbine/expander and the downstream compression stage are operatively coupled to a second pinion of the at least three pinions. The cold turbine/expander is operatively coupled to a third pinion.
In an alternate embodiment of the natural gas liquefaction system the refrigerant compression stages comprise a first upstream refrigerant compression stage, a second upstream refrigerant compression stage arranged in parallel with the first upstream refrigerant compression stage, and a downstream refrigerant compression stage arranged in series with and downstream of the first and second upstream refrigerant compression stages. In such alternate embodiment, the second warm turbine/expander and the first upstream refrigerant compression stage are operatively coupled to the first pinion of the at least three pinions while the first warm turbine/expander and the downstream refrigerant compression stage are operatively coupled to the second pinion and the cold turbine/expander and the second upstream refrigerant compression stage are operatively coupled to the third pinion.
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) and may originate from a source of methane containing biogas. The refrigerant stream is preferably a nitrogen-based refrigerant that 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 two refrigerant compression stages;

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 with two of the three refrigerant compression stages arranged in parallel; and

FIG. 4 shows a schematic of yet another embodiment of the present system and method for liquefied natural gas production using multiple heat exchangers.

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 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. 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.
LNG Production with 3 Turbine/Expanders and 2 Refrigerant Compression Stages
Turning to FIG. 2 , a schematic of the high-level process flow diagram for one embodiment of the present system 10 and method for liquefied natural gas production using three turbine/expanders and two refrigerant compression stages is shown. The illustrated system 10 also includes a refrigerant circuit with at least one heat exchanger 20 and two aftercoolers 221, 222, an integral gear machine 25, a fuel gas circuit 18, and a post liquefaction conditioning circuit 23, having one or more expansion valves 27 and a phase separator 28 configured for separating nitrogen and other light gases from the liquefied and subcooled natural gas stream 21.
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 FIGS. 2 and 3 depict a single heat exchanger 20 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 20 or first heat exchange core configured for liquefying the natural gas feed stream 14 and a second heat exchanger 120 or second heat exchange core configured for cooling other streams, such as a portion of the refrigerant stream (as illustrated in the embodiment of FIG. 4 ), or perhaps sub-cooling the liquefied natural gas stream, or perhaps even pre-cooling the natural gas feed stream.
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.
LNG Production with 3 Turbine/Expanders and 3 Refrigerant Compression Stages
Turning to FIG. 3 , there is shown a schematic of the high-level process flow diagram for another embodiment of the present system 10 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 particular reference to FIG. 2 , is the addition of a third refrigerant compression stage 40B and the operative coupling of the cold turbine/expander 70 and one of the refrigerant compression stages 40B to the third pinion 33 of the integral gear machine 25 such that all three pinions are net absorbers of power.
Similar to the embodiment shown in FIG. 2 , the natural gas liquefaction system 10 illustrated in FIG. 3 also includes a refrigerant circuit with at least one heat exchanger 20 and two aftercoolers 221, 222, an integral gear machine 25, a fuel gas circuit 18, and a post liquefaction conditioning circuit 23, having one or more expansion valves 27 and a phase separator 28 configured for separating nitrogen and other light gases from the liquefied and subcooled natural gas stream 21. The purified and compressed natural gas feed 14 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).
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 FIG. 3 the refrigerant streams are compressed using three refrigerant compression stages 40A, 40B, 50 with two refrigerant compression stages 40A, 40B of the three refrigerant compression stages arranged in parallel.
As shown in FIG. 3 , the warmed cold turbine exhaust 76 and the warmed first warm turbine exhaust 86 exiting the heat exchanger 20 are recycled as a lower pressure recycle stream 42A and 42B to a pair of upstream refrigerant compression stages 40A and 40B, respectively. The pair of upstream refrigerant compression stages include a first upstream refrigerant compression stage 40A and a second upstream refrigerant compression stage 40B arranged in parallel. The lower pressure recycle stream is split wherein between 60% and 70% of the lower pressure recycle stream 42A is compressed in the first upstream refrigerant compression stage 40A while the remaining portion of the lower pressure recycle stream 42B is compressed in the second upstream refrigerant compression stage 40B. The parallel streams exiting the first upstream refrigerant compression stage 40A and the second upstream refrigerant compression stage 40B are then cooled in the upstream aftercooler 221.
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.

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 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 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

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

Pinion #1	N2 Comp	1916	N2 - 2^ndWarm	−399	1517
	#CB1		Turbine
Pinion #2	N2 Comp	2511	N2 - 1^stWarm	−994	1517
	#CB2		Turbine
Pinion #3	—	—	N2 - Cold	−214	−214
			Turbine

TABLE 1B

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

CB1 Inlet	M07	33.48	71.1%
CB2 Inlet	M12	35.28	100.0%
Cold Turbine Inlet	M03	−113.47	21.5%
Cold Turbine Exhaust	M04	−166.90	21.5%
1^stWT Inlet	R02	−34.57	49.6%
1^stWT Exhaust	R03	−111.80	49.6%
2^ndWT Inlet	S02	22.84	28.9%
2^ndWT Exhaust	S03	−27.60	28.9%
Lower Pressure	M06	33.48	71.1%
Upstream Aftercooler	M10	36.00	71.1%
Downstream	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 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 FIG. 2 and operated in a manner similar to the example of Case 1 is among the best compromises 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 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 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

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

Pinion #1	N2 Comp	2408	N2 -2^ndWarm	−748	1660
	#CB1		Turbine
Pinion #2	N2 Com3	4321	N2 - 1^stWarm	−1820	2501
	#CB2		Turbine
Pinion #3	N2 Comp	1297	N2 - Cold	−360	937
	#CB3		Turbine

TABLE 2B

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

CB1 Inlet	M07	33.48	44.5%
CB2 Inlet	M12	35.21	100.0%
CB3 Inlet	M07A	33.48	23.9%
Cold Turbine Inlet	M03	−111.57	19.3%
Cold Turbine Exhaust	M04	−166.90	19.3%
1^stWT Inlet	R02	−34.99	49.1%
1^stWT Exhaust	R03	−114.00	49.1%
2^ndWT Inlet	S02	18.97	31.6%
2^ndWT Exhaust	S03	−28.80	31.6%
Higher Pressure	S04	33.50	31.6%
Lower Pressure	M06	33.48	68.4%
Upstream Aftercooler	M10	36.00	68.4%
Downstream	M15	36.00	100.0%

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 FIGS. 2 and 3 will achieve a minimum power advantage of about 5% relative to the conventional natural gas liquefaction process using the two pinion and two turbine/expander arrangement shown in FIG. 1 . The actual power advantage realized by the three pinion and three turbine/expander arrangements compared to the conventional two pinion and two turbine/expander arrangement is very much dependent on the operating conditions and constraints imposed upon the liquefaction process as well as the turbomachinery efficiencies and hardware constraints associated with the respective systems but could readily meet or exceed a 9.0% power advantage, which more than compensates for the small increase in capital costs. The power reduction in the three pinion and three turbine/expander configurations is also accompanied by an increase in the heat exchanger UA of almost 50% which is evidence of tighter composite curves within the heat exchanger. However, the increase UA is unlikely to require a second heat exchanger core for small to mid-scale natural gas liquefaction systems.
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) 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; and

an integral gear machine comprising a drive assembly, a bull gear, and at least three pinions arranged to drive the two or more 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 of the two or more refrigerant compression stages; (ii) 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 (iii) 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; and

wherein two of the at least three pinions operatively couple the upstream refrigerant compression stage to one of the turbine/expanders and the downstream refrigerant compression stage to another of the turbine/expanders and the two of the at least three pinions are a net absorber of power from the drive assembly.

2. The natural gas liquefaction system of claim 1, wherein the power imparted to the two of the at least three pinions does not differ by more than 10%

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

4. The natural gas liquefaction system of claim 3, wherein the cold turbine/expander is operatively coupled to a third pinion of the at least three pinions.

5. The natural gas liquefaction system of claim 1, wherein the two or more refrigerant compression stages further comprise at least three refrigerant compression stages and wherein the upstream refrigerant compression stage further comprises a first upstream refrigerant compression stage and a second upstream refrigerant compression stage arranged in parallel.

6. The natural gas liquefaction system of claim 5, wherein the second warm turbine/expander and the first upstream refrigerant compression stage are operatively coupled to a first pinion of the at least three pinions, the first warm turbine/expander and the downstream refrigerant compression stage are operatively coupled to a second pinion of the at least three pinions, and the cold turbine/expander and the second upstream refrigerant compression stage are operatively coupled to a third pinion of the at least three pinions.

7. The natural gas liquefaction system of claim 6, wherein each pinion of the at least three pinions is a net absorber of power from the drive assembly and the total power consumption of the natural gas liquefaction system is distributed between the three pinions such that each pinion absorbs between 30% and 40% of the power from the drive assembly.

8. The natural gas liquefaction system of claim 1, wherein the three or more turbines/expanders collectively provide between about 25% and 50% of power required by the three or more refrigerant compression stages.

9. The natural gas liquefaction system of claim 1, 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 cold turbine/expander.

10. 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.

11. 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.

12. The natural gas liquefaction system of claim 1, wherein the second warm exhaust is less than about −15° C.

13. The natural gas liquefaction system of claim 1, wherein the refrigerant stream comprises more than about 80% nitrogen by volume.

14. 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.

15. 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).

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

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

18. The natural gas liquefaction system of claim 1, further comprising a fuel gas circuit comprising one or more valves configured to expand the liquified and subcooled natural gas stream to a pressure less than about 6.0 bar(a) and a fuel gas heat exchanger configured to cool the expanded natural gas stream to produce a fuel gas

19. The natural gas liquefaction system of claim 18, wherein the fuel gas heat exchanger is the same heat exchanger that liquefies and subcools the natural gas stream and the fuel gas exits the heat exchanger at or near ambient temperatures.

20. The natural gas liquefaction system of claim 1, wherein the at least one heat exchanger further comprises multiple heat exchangers or multiple heat exchange cores with a first heat exchanger or first heat exchange core configured for liquefying the natural gas containing feed stream and a second heat exchanger or second heat exchange core configured for either cooling the natural gas containing feed stream, cooling a warm portion of the refrigerant stream, or sub-cooling the liquefied natural gas stream via indirect heat exchange with one or more of the at least three exhaust streams.