Performance study on a single-screw compressor for a portable natural gas liquefaction process

Accepted Manuscript Performance study on a single-screw compressor for a portable natural gas liquefaction process Yong Li, Gongnan Xie, Sunden Bengt, Yuanwei Lu, Yuting Wu, Jiang Qin PII:

S0360-5442(18)30219-6

DOI:

10.1016/j.energy.2018.02.003

Reference:

EGY 12303

To appear in:

Energy

Received Date: 8 November 2017 Revised Date:

25 January 2018

Accepted Date: 1 February 2018

Please cite this article as: Li Y, Xie G, Bengt S, Lu Y, Wu Y, Qin J, Performance study on a singlescrew compressor for a portable natural gas liquefaction process, Energy (2018), doi: 10.1016/ j.energy.2018.02.003. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.

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Revised-EGY-17-05781

Performance Study on a Single-Screw Compressor for a

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Portable Natural Gas Liquefaction Process

Yong Li 1, 2, Gongnan Xie 1, *, Sunden Bengt 3, Yuanwei Lu 4, *,

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Yuting Wu 4, Jiang Qin 5, *

School of Marine Science and Technology, Northwestern Polytechnical University, P.O. Box 24, Xi'an, 710072, China.

2

School of Mechanical Engineering, Northwestern Polytechnical Univerisity, P.O. Box 522, Xi’an, 710072, China.

Department of Energy Sciences, Lund University, P.O. Box 118, SE-22100 Lund, Sweden.

4

College of Environmental and Energy Engineering, Beijing University of Technology, Beijing

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3

100124, China.

School of Energy Science and Engineering, Harbin Institute of Technology, Harbin 150001,

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5

*

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Heilongjiang, China.

Corresponding author:

[email protected] (G. Xie); [email protected] (J. Qin)

Tel: 86-29-88492611;

Fax: 86-29-88495278

1

[email protected] (Y. Lu);

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Abstract: Portable natural gas liquefaction devices are widely used to exploit unconventional natural gas resources and one of key components is a compressor. In this paper, a new natural gas liquefaction process is simulated and optimized and a single-screw compressor with 200 mm

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diameter is designed to be applied in this process. To verify the performance of the single-screw compressor, an experimental system was set up and performance parameters were studied, including volumetric efficiency, specific power rate, shaft efficiency, irreversible loss rate and cooling

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efficiency of the lubricating oil. The influence of the lubricating oil on the volumetric efficiency was

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also observed. Furthermore, the impact of actual compressor performance on the liquefaction process was also analyzed. The results indicate that the performance of the compressor changes greatly when it operates deviating from its design condition. The maximum volumetric efficiency, shaft efficiency, irreversible loss rate and cooling efficiency of the lubricating oil are 0.91, 0.64,

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0.56 and 0.22, respectively. The minimum specific power rate is 6.5 kW·(m3·min-1)-1. In addition, the lubricating oil has a good effect of sealing so that the volumetric efficiency can be improved by

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10.5%. Considering the actual performance of the single-screw compressor, specific power

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consumption of the process increases by 43.3%.

Keywords: Portable natural gas liquefaction process; Single-screw compressor; Volumetric efficiency; Shaft efficiency; Specific power consumption

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Nomenclature the constants relating to the gas species

A

dimensionless attractive parameter

b

the constants relating to the gas species

B

dimensionless effective molecular volume in Peng-Robinson equation o state

cp

specific heat of the cold and hot fluids [kJ·kg-1·K-1]

CP

heat capacity flow rate [kW·K-1] binary interaction coefficient

n

rotational speed of compressor [r·min-1]

n1

polytropic process index of process A-B1

n2

polytropic process index of process A-B2

p

pressure [Pa, MPa]

P

power [kW]

q

flow rate [m3·min-1, kmol·h-1, kg·s-1]

Q

heat flow of the cold and hot fluids [kW]

R

molar gas constant, [8.314 J·mol-1·K-1]

Rg

gas constant [kJ·kg-1·K-1]

T

temperature [K]

v

molar volume [m3·mol-1]

V0

primitive volume of one groove [m3]

W

power [kW]

x

mole fraction

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compression factor groove number of the screw

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Z1

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Z

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k

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a

Greek symbol α

cooling efficiency of lubricating oil

∆h

specific enthalpy difference [kJ·kg-1]

ε

specific power rate [kW·(m3·min-1)-1]

η

efficiency

κ

isentropic exponent

φ

NG liquefaction ratio

ω

specific power consumption [kWh·kmol-1]

Subscripts

3

a

actual

ad

adiabatic

con

consumption

el

electricity

g

gas

i

component i

j

component j

l

irreversible loss

m

molar/mass

mo

motor

net

net power

NG

natural gas shaft

th

theoretical

V

volume

1

inlet

2

oulet

Superscripts

irreversible compression process A-B1 without external cooling process

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ʹ

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s

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liquefied natural gas

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LNG

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1. Introduction In China, the consumption capacity of natural gas is 2090×108 m3, which is more than the total yielding capacity of 1460×108 m3 in 2015 [1]. In order to meet China's increasing demand,

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unconventional natural gas resources are necessary to be exploited. The unconventional natural gas refers to some gas resources remote from markets or pipelines, or other gas resources close to

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markets whose reserves are too small to be piped [2], so the economic benefit is poor if a stationary liquefaction plant is built. Whereby, the portable natural gas liquefaction devices have received its

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rapid evolution and attracted attention of investors for the development of unconventional natural gas resources, with advantages of low investment, simple and compact process, start-stop convenience, strong mobility and mature progress [2].

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A natural gas liquefaction device is energy and cost intensive. Therefore, one of the major issues in the liquefaction device is to reduce energy input requirements and total cost [3]. For portable natural gas liquefaction devices, another important aspect is to make their scale compact as

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far as possible. There are mainly three types of liquefaction processes: a cascade liquefaction

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process, a mixed refrigerant liquefaction process and an expander liquefaction process [4]. The cascade liquefaction process consumes less power than the others but this process is complex and always used in the large liquefaction plants, so the mixed refrigerant liquefaction process and the expander liquefaction process are always applied in the portable natural gas liquefaction processes [5]. Fazlollahi et al. [6] compared four different natural gas liquefaction processes which provided coolant for the energy-storing version of cryogenic carbon captureTM (CCC) and they found that the single mixed refrigerant cycle (SMR) process had the lowest energy

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ACCEPTED MANUSCRIPT demand, costs of utilities and highest heat exchanger efficiencies, the propane-pre-cooled mixed refrigerant cycle (C3-MR) process had lowest capital cost but highest operating cost. In addition, they applied a transient model of Aspen HYSYS to simulate the single mixed refrigerant cycle and

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found that the SMR process provided the most efficient overall design [7, 8]. However, the preparation of mixed refrigerant is very difficult and its ratio is associated with the pressure of natural gas and ambient temperature [9-11]. In addition, the leakage of mixed refrigerant will

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reduce the liquefied rate [12]. Recently, some researchers compared different natural gas

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liquefaction processes and deemed that the expander liquefaction process is the leading candidate applied in the portable natural gas liquefaction processes [13-15]. In this paper, the expander liquefaction process is adopted.

There is a lot of research on expander liquefaction process. Based on a thermodynamic

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optimization theory of heat exchangers, Chang et al. [16, 17] presented a methane liquefaction process and N2–C2–C3 liquefaction cycle. Wang et al. [18] used Aspen HYSYS to design and

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optimize a small skid-mounted expander liquefaction process and found that increasing the storage

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pressure, initial pressure and the content of methane in the source gas could reduce the specific power consumption. Yuan et al. [2] proposed a novel small-scale liquefaction process adopting single nitrogen expansion with carbon dioxide pre-cooling and deemed this process is suitable for the development of small gas reserves considering compact device, safety operation and simple capability. Song et al. [19] built a multi-objective optimization problem based on the Yuanʹs study and results indicated that this method can better realize the synthetical performance of the process. Gu Anzhong research group also simulated natural gas liquefaction processes including nitrogen

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ACCEPTED MANUSCRIPT expansion liquefaction process [20-22]. Gao et al. [23] adopted a nitrogen expansion cycle with propane pre-cooling to liquefy the coalbed methane with nitrogen. They optimized the parameters of the process and studied the effect of nitrogen content.

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The above papers paid more attention on simulations and optimizations of the liquefaction processes. However, the actual influences of the key facilities for the liquefaction process were studied rarely. Compressors are important and their actual performances affect the efficiencies of

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natural gas liquefaction processes directly [24]. Therefore, it is necessary to conduct more

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additional studies about performances and influence factors of compressors.

Single-screw compressors are being highly favored because they have many advantages such as good force balance, long life, high compression ratio, high volumetric efficiency, high efficiency at part load, simple structure, low noise [25, 26]. In order to improve the performances of the

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single-screw compressors, more studies were focused on meshing pair [27-32]. Meanwhile, there is some research works studying the performance of single-screw compressors or single-screw

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expanders. Wang et al. [33] proposed Multicolumn Envelope Meshing Pair (MEMP) to reduce the

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wear of a meshing pair. The application of this new type of meshing pair changed the thermodynamic performance of a single-screw refrigeration compressor (SSRC) and the SSRC was proved to have high volume displacement and volumetric efficiency but needed slightly more shaft power. The maximum increment of shaft power and volumetric efficiency of the SSRC with MEMP is 5.06% and 3.7%, respectively, in all working conditions. Ziviani et al. [34] adapted a single-screw compressor as an expander in an organic Rankine cycle (ORC) for low-grade waste heat recovery. In the same way, the single-screw compressor was converted into a single-screw

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ACCEPTED MANUSCRIPT expander and its performance was tested. For examples, Li et al. [35] studied the performance of a single-screw expander applied in a small-scale pressure recovery system and reported that the maximum volumetric efficiency, isentropic efficiency, overall efficiency, and the lowest

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air-consumption were 83.5%, 66.4%, 62.2%, and 44.1 kg·(kW·h)-1, respectively. Zhang et al. [36] studied a single-screw expander applied in an organic Rankine cycle (ORC) system and reported that the maximums of volumetric efficiency, adiabatic efficiency and total efficiency of the

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single-screw expander were 90.73%, 73.25% and 57.88%, respectively. Wang et al. [37] studied the

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performance of single screw expanders based on gap adjustment and found that the prototype for which the gap between gaterotor and shell is 0.04 mm and the gap between screw and shell is 0.05 mm had the best overall performance. The power output was about 4.5 kW, the gas consumption

efficiency was about 60%.

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rate was about 65 kg·kWh-1, the maximum volumetric efficiency was about 66%, and the shaft

However, the actual performance of single-screw compressors combined with the simulation

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of natural gas liquefaction processes has been studied rarely. In addition, N2 and CH4 are always

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used as a refrigerant in the previous studies. In this article, a new expander liquefaction process based on an inverse Brayton cycle is proposed and air is used as a refrigerant. This process was simulated and optimized by Aspen HYSYS. According to the optimization results, a new single-screw compressor used in the presented liquefaction process was designed and its performance was tested. Finally, the simulation results of the compressor have been corrected based on the experimental data and the effect of the single-screw compressorʹs actual performance on the liquefaction process was also studied.

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2. Simulation details Aspen HYSYS is a process simulation software which has an abundant database, including more than 1800 substances and 16000 binary interaction coefficients. Due to its wide suitability and

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good precision, it has been used by many researchers for the natural gas liquefaction process simulation [38-41]. In this paper, the simulation and optimization of the natural gas liquefaction

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process is conducted by Aspen HYSYS and Peng-Robinson equation is used to calculate the physical parameters in the process.

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2.1. Feed gas parameters and initial values

Initially, the simulation discussed in this paper requires several specifications. The feed gas composition and its pressure can directly affect its liquefaction process. The mole fraction of the

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feed gas is referred from the references [15, 24] and its initial pressure is 0.1 MPa. The temperature-entropy (T-S) diagram of feed gas at different pressure is shown in Fig. 1. The higher

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the pressure of the feed gas is, the more the temperature curve is similar to a straight line, thus it will match well with the temperature curve of the refrigerant (air). However, more power will be

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consumed more if the feed gas is pressurized, so here is an optimal pressure for feed gas before being liquefied to make the power consumption minimum. The optimal pressure of the feed gas is obtained in section 2.4. The temperature of feed gas is set to 20

based on the ambient

temperature. The molar flow rate of feed gas depends on the results of optimization. Based on the references [21, 22], the LNG storage pressure is set to 2.0 MPa. The air (no carbon dioxide) is used as the refrigerant in the process. The initial temperature and pressure are set to 20

and 0.1 MPa,

respectively. The flow rate is set to 16.06 kmol·h-1 (6.5 Nm3·min-1) based on the primitive volume 9

ACCEPTED MANUSCRIPT of the compressor we designed. Detailed specifications used here are listed in Table 1. Refrigerant (air) temperature after the coolers is set to 30

.The highest pressure is specified as 2.916 MPa based on the design. The adiabatic efficiency

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to 4

.The minimum approach temperature of the multi-stream heat exchanger is set

of the compressors and expanders are set to 0.8 and 0.6, respectively. To simplify the simulation process and analysis, the pressure drop of each air-cooler and the multi-stream heat exchanger is

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assumed to be 0 MPa.

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2.2. Process description

The actual natural gas (NG) liquefaction process is shown in Fig. 2. The process includes two parts: an air expansion refrigeration cycle and a liquefaction cycle. Nos.1-4 represents the air

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expansion refrigeration cycle. Firstly. the air is compressed in compressors and precooled in a multi-stream heat exchanger. Then the air is expanded in two expanders and forms the cryogenic gas. Finally, the cryogenic gas is used to cool the natural gas. In each compressor, there is a

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lubricating oil circulation system. The air and lubricating oil are injected into the compressor

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together. Then they are separated in the oil-gas separator after being compressed and cooled in the plate-fin heat exchangers, respectively. Finally, the lubricating oil flows into the compressor again. The details of work for unit 2 (the compressors and the lubricating oil circulation system) will be illustrated in section 3.2. The operation process of unit 4 (the expanders and the lubricating oil circulation system) is similar to the unit 2. Nos.6-8 represent the liquefaction cycle. Firstly, NG is compressed in a NG compressor. The purpose has been stated in section 2.1. Then NG1 is precooled by the cryogenic gas and throttled in a throttle valve. Finally, NG becomes LNG (liquefied natural

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NG. In the Fig. 2 and Fig. 3, other components are the same in addition to the compression process and the expansion process. The compression processes in the red frames (Figs. 2 & 3) are used to illustrate the difference. The actual model (Fig. 2) includes lubricating oil injection into the

in the plate-fin heat exchanger. Finally, high pressure air is at a low temperature. In

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cooled to 30

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compressor for cooling when the compressor is working. Then the compressed air continues to be

the simulation model (Fig. 3), the air is compressed in the compressor which is assigned an isentropic efficiency in the Aspen HYSYS and then the compressed air is cooled to 30

in the air

cooler. Finally, high pressure air is obtained at a low temperature. The actual compression process

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represents a polytropic process along with an external cooling process and the simulated compression process represents an adiabatic process. However, the purpose is the same, i.e., to

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obtain high pressure air at low temperature and the adiabatic efficiency given by the software can be

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corrected according to the actual model (experimental data). In addition, the flow rate in the simulation process may also change because of the actual leakage and it also can be corrected according to the actual model (experimental data). The influence on the liquefaction process was analyzed in section 5.2.

2.3. Phase equilibrium equations The determination of physical properties is the basis of the simulation [2]. In this simulation, Peng-Robinson (PR), a physical package embedded in the Aspen HYSYS, was chosen. The PR state

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ACCEPTED MANUSCRIPT equation applies functionality to some specific component-component interaction parameters, which can be used in the calculation of phase equilibrium [2, 6-7]. It is written by:

RT a − v − b v ( v + b ) + b ( v − b)

N

N

a = ∑∑ xi x j (ai a j )0.5 (1 − kij )

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i =1 j =1

1

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p=

N

b = ∑ xi bi

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i =1

2

where p is pressure, R is molar gas constant, T is temperature, v is molar volume, a and b are the constants relating to the gas species, x is mole fraction of a certain component, k is binary interaction

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coefficient.

Also, the PR state equation is expressed in terms of compressibility factor as follows

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Z 3 − (1 − B ) Z 2 + ( A − 2 B − 3 B 2 ) Z − ( AB − B 2 − B 3 ) = 0

pv RT

(5)

ap ( RT ) 2

(6)

bp RT

(7)

Z=

A=

(4)

B=

where Z is a compression factor, A is a dimensionless attractive parameter, B is dimensionless effective molecular volume, A and B are the coefficients relating to the gas state parameters.

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ACCEPTED MANUSCRIPT 2.4. Process optimization In the proposed process as shown in Fig. 3, some key parameters play dominant roles in affecting the process performance. As a result, these key parameters need to be optimized, such as

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the pressures of NG1 and point 2, the temperature of point 6 after precooling in the multi-stream heat exchanger, the flow rate of NG.

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The steady-state optimizer with the original method in Aspen HYSYS is employed to conduct the optimization of this natural gas liquefaction process. The tolerance of the optimizer is set to

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1.0e-5 [42]. The optimizer includes three parts: objective functions, constraint functions, adjusted (primary) variables. Several parameters need to be defined before optimization as follows: (1) NG liquefaction ratio

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The NG liquefaction ratio is used to evaluate the liquefaction capability of a process and can be defined as:

qm, LNG qm, NG

(8)

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ϕ=

where φ is the NG liquefaction ratio, qm,LNG is molar mass flow of LNG, kmol·h-1, and qm,NG is molar

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mass flow of NG, kmol·h-1.

(2) Specific power consumption The evaluation index for a natural gas liquefaction process is the specific power consumption ωcon and it denotes the amount of power per liquified 1 kmol natural gas. It can be calculated as follows:

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Wnet qm , LNG

(9)

Wnet = ∑Wcompressors − ∑Wexp anders

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(10)

where ωcon is the specific power consumption of the natural gas liquefaction process, kWh·kmol-1,

∑Wexpanders is the output power of all expanders, kW.

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Wnet is the net input power of the process, kW, ∑Wcompressors is the input power of all compressors, kW,

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The specific power consumption is adopted as an objective function as shown in Equation (11) and the optimization problem is finding out the optimum parameters values to make the specific power consumption lowest under the constraint conditions as follows: (A) The flow rate of the refrigerant is set to 16.06 kmol·h-1.

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(B) The minimum approach temperature of the multi-stream heat exchanger is set to 4

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(C) The NG liquefaction ratio of the NG liquefaction process is specified as 100%.

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(D) The highest pressure of the refrigerant is set to 2.916 MPa. 11

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f ( x ) = min (ωcon )

Adjusted (primary) variables include the molar flow rate of feed gas, the pressure of NG1, the temperature of point 6 and the pressure of point 2. The low bound and high bound of them are listed as shown in Table 2.

Based on the optimization results as shown in Table 3. the molar flow rate of the feed gas is 4.376 kmol·h-1. The pressure of NG1 is 5.739 MPa, it means that the power consumption will be lowest if the initial pressure of the feed gas is close to 5.739 MPa. The temperature of point 6 is

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. The pressure of point 2 is 0.54 MPa (compression ratio is 5.4). Finally, the specific power

consumption is 15.43 kWh·kmol-1.

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3. Design and performance test for a single-screw compressor Compressors are key component of the natural gas liquefaction process, so based on the optimization results for the above NG liquefaction process, a single-screw compressor (the low

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pressure compressor shown in Fig. 3) was designed and a test system was set up to obtain the

3.1. Single-screw compressor prototype

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performance of the single-screw compressor.

The key components of a single-screw compressor are one screw and two star-wheels, which create a working chamber with the shell, as shown in Fig. 4. The two star-wheels are made of

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engineering plastics. In order to prevent the star-wheels from being damaged, a metal support is matched with them. The spindle is connected with the screw in the form of hot charging. The

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star-wheel and metal support rotate following with the screw synchronously. The rotation direction

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of the screw complies with the right hand screw rule. When the single-screw compressor works, the working chamber becomes smaller and smaller, and thus the working fluid is compressed. Fig. 5 shows a single-screw compressor prototype. There is an oil injection hole and its diameter is 7 mm. When the prototype starts working, the lubricating oil will be injected into the working chamber to cool the air and seal the leakage passage. So the lubricating oil is not only having the effect of cooling, but also has the effect of sealing. Table 4 gives the physical properties of the lubricating oil and this lubricating oil is named as 46# Shell synthetic lubricating oil and

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ACCEPTED MANUSCRIPT provided by Shell Co., Ltd. Furthermore, in order to ensure the air inflow, this working chamber is surrounded by the screw, rotor and shell completely before the lubricating oil is injected into the working chamber. So the oil injection hole is not designed casually. Table 5 lists the parameters of

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the prototype. The diameter of the screw and star-wheel is designed to be 200 mm and the flow rate of low pressure compressor is 6.5 Nm3·min-1. In general, groove numbers of the screw and tooth numbers of the star-wheel are set to 6 and 11, respectively. Based on the optimization results, the

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3.2. Experimental system description

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compression ratio of the low pressure compressor is designed as 5.4.

An experimental system was built as shown in Fig. 6. The prototype is driven by a motor. The speed of the motor is adjusted by a frequency converter. When the prototype starts, the air in

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environment is sucked into the compressor through an air filter. The compressed air flows into an oil-gas separator. Under the action of the minimum pressure valve, the pressure is built up slowly. The lubricating oil in the oil-gas separator begins to flow. Then the lubricating oil flows into a

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plate-fin heat exchanger. Its temperature is then reduced by the air cooling. After that the lubricating

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oil flows into the compressor to absorb heat. In order to prevent dirty substances from flowing into the compressor, the lubricating oil must be filtered. After that, the compressed air and the lubricating oil flows into an oil-gas separator together. The lubricating oil continues to flow circularly. When the pressure of the oil-gas separator was reached a certain value (0.45 MPa), the compressed air will flow from the oil-gas separator to the plate-fin heat exchanger. The cooled and compressed air is discharged into the atmosphere through a buffer tank. The effect of the ball valve is to adjust the magnitude of the outlet pressure.

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ACCEPTED MANUSCRIPT During the experiment, the motor of MN2-200L2-2 is provided by Hengshui motor Co., Ltd. (Hengshui, China). Its rated power, rated current and rated voltage are 37 kW, 67.9 A and 380V, respectively. The oil-gas separator includes an oil drum, a minimum pressure valve, a relief valve

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and an oil mirror. The oil drum is used to hold the oil. The minimum pressure valve is used to form certain pressure to promote lubricating oil flowing. The relief valve is applied to prevent the excessive pressure in the oil drum. The oil mirror is used to determine the amount of lubricating oil.

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In order to ensure the accuracy of the measurement data, the buffer tank is used to reduce the

tank are 0.84 MPa, 150

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instability of the compressed gas. The design pressure, design temperature, volume of the buffer and 300 L, respectively. The power of fan in the plate-fin heat exchanger

is 0.75 kW and its speed is 930 r·min-1. The air filter, ball valve and fine oil separator are provided

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by Shaanxi di kai electrical technology Co., Ltd.

In the process of experimental tests, the parameters such as rotational speed and active power of the motor are measured by a speed-meter and a power-meter, respectively. The measurement

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range of the speed-meter is 2.5~99999 r·min-1 and its accuracy is 1 r·min-1. The power-meter of

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PF9830C type is provided and adjusted by EVERFINE Co., Ltd. (Hangzhou, China). The measurement range of the power-meter is 0~12 kW and its accuracy is ±0.5%. In order to broaden the scope of application, three mutual inductors were used. The temperature of the mixture flowing from the prototype is measured by a temperature transmitter of WAP021. The measurement range is -50~150

and its accuracy is ±0.5%. The temperature, pressure and volumetric flow rate of the

cooled air are measured by another temperature transmitter of SBWZP231, a pressure transmitter and a flow meter, respectively. The measurement range of the temperature transmitter of

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ACCEPTED MANUSCRIPT SBWZP231 is -25~80

and its accuracy is ±0.5%. The measurement range of the pressure

transmitter is 0~1.6 MPa and its accuracy is ±0.3%. The temperature transmitters and pressure transmitter are provided and adjusted by Xinmin instruments Co., Ltd. (Xi'an, China). All data were

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collected by data collectors including Agilent 34972A provided by Agilent technologies Co., Ltd (USA) and Toprie TP700 provided by Toprie electronics Co., Ltd (Shenzhen, China). The details of the measurement instruments are shown in Table 6. A photo of the experimental system is presented

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in Fig. 7. In addition, in order to arrange the space layout more expediently, the two plate-fin heat

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exchangers were manufactured together as shown in the Fig. 7.

4. Parameter definition

4.1. Evaluation index of a single-screw compressor

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The main indices for evaluating the performance of a single-screw compressor are the volumetric efficiency, specific power rate, shaft efficiency and irreversible loss rate. These are

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conventional indices to evaluate the performance of compressors. In addition, the cooling efficiency

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of the lubricating oil is also defined.

(1) Volumetric efficiency

The volumetric efficiency is used to measure the deviation between the actual volumetric flow and theoretical volumetric flow. Due to the leakage and inlet pressure loss, its value is usually less than 100%. The volumetric efficiency is defined as follows:

ηV =

qV ,a qV ,th

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(12)

ACCEPTED MANUSCRIPT qV ,th = 2 ⋅ n ⋅ Z1 ⋅ V0

(13)

where ηV is volumetric efficiency, qV, a is actual inlet volume flow through the compressor, m3·min-1, qV, th is theoretical inlet volume flow, m3·min-1, n is the rotational speed of the compressor, r·min-1,

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Z1 is the groove number of the screw and its value is 6, V0 is the primitive volume of one groove and its value is 2.2×10-4 m³.

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(2) Specific power rate

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The specific power rate reflects the situation of the actual power consumption. The smaller the value is, the less energy will be used. According to the regulation of the compressor industry, the specific power rate is defined as follows:

Pel qV ,a

(14)

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ε =

where ε is specific power rate, kW·(m3·min-1)-1. In ac circuits, the power includes apparent power,

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active power and reactive power and their units are kVA, kW and kVAr, respectively. Here, Pel is the

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active power of the motor, kW.

(3) Shaft efficiency

The shaft efficiency reflects the effective utilization degree of a shaft power and is defined as follows:

ηs =

Pad Ps

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(15)

ACCEPTED MANUSCRIPT ηel =

Pad Pad Ps = ⋅ = η s ⋅η mo P el Ps Pel



κ Rg T1  p2  Pad = ∆h ⋅ qm =   κ − 1  p1 

 − 1 ⋅ qm  

(17)

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

κ −1 κ

(16)

where ηs is the shaft efficiency, Pad is the power consumption of an adiabatic compression process, kW, and Ps is the shaft power or input power of a compressor, kW, ηel is the electrical efficiency, ηmo

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is the efficiency of the motor and can be assumed as 0.9 according to a manufacturer, ∆h is the

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adiabatic specific enthalpy difference, kJ·kg-1, qm is the mass flow of the air, kg·s-1, T1 and p1 are the inlet temperature and inlet pressure of the air in the atmosphere, respectively, K, MPa, p2 is the outlet pressure, MPa, Rg is the gas constant, kJ·kg-1·K-1, κ is the isentropic exponent and set to be 1.4.

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(4) Irreversible loss rate

According to different external conditions, there are many typical compression processes. For example, adiabatic compression, isothermal compression, polytropic compression and so on. Fig.8

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shows different compression processes in a T-S diagram. A-B1 represents a polytropic compression

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without any external cooling process (AB1FEA represents heat absorption capacity). A-Bad represents an adiabatic compression. A-B2 represents a polytropic compression process along with an external cooling process (AB2DEA represents heat release capacity). A-B3 represents an isothermal compression. AB1FEA can also be considered as the internal power loss caused by friction, and it can be assumed that it is the same in all non-isentropic processes. ABadB3CDEA, AB2B3CDEA and AB3CDEA represent the compression power. Due to the external cooling process, a part of the energy can be saved compared with the polytropic processes such as AB1B2A and

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ACCEPTED MANUSCRIPT AB1B3A. In processing the experiments, the air is compressed along with the lubricating oil cooling like the process A-B2. Thus the actual power consumption of the compression process is represented by

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AB2B3CDEA and AB1FEA. AB2B3CDEA is the power consumption of the reversible polytropic process and AB1EDA is the power loss of the irreversible process. Equations (18) and (19) can be

p  T2 =  2   p1 

n2 −1 n2

T1

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used to calculate the polytropic process index and power consumption of the reversible process.

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n2 −1   n2   n2 p  2 Pn2 = RgT1   − 1 ⋅ qm  p1   n2 − 1  

(18)

(19)

where n2 is polytropic process index of the process A-B2, Pn2 is reversible compression power of

calculated by:

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the process A-B2, kW. Then the irreversible loss can be obtained and the irreversible loss rate is

ηl =

Pl Ps

(20)

(21)

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Pl = Ps − Pn2

where Pl is the irreversible loss, kW, ηl is the irreversible loss rate. (5) Cooling efficiency of lubricating oil

In order to describe the role of the lubricating oil in the process of compression, the process A-B1 without any external cooling process (as shown in Fig. 8) is necessary to be studied. A-B1 represents the polytropic compression without any external cooling process and its internal power loss can be assumed to be the same as for the process A-B2. Its irreversible compression power can

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ACCEPTED MANUSCRIPT be determined by:

n1 −1   n1   p  2 Pn1′ = Pn1 + Pl = ⋅ RgT1   − 1 ⋅ qm  p  κ −1  1  

κ

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(23)

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n1 −1    p2  n1 n1  Pn1 = RgT1   − 1 ⋅ qm  p1   n1 − 1  

(22)

where Pn1′ is the irreversible compression power of the process A-B1 without any external cooling

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process, kW, Pn1 is the reversible compression power of the process A-B1, kW, and n1 is the polytropic process index of the process A-B1. Equation (22) can be used to calculate n1 of the process A-B1. Finally, the compression power of the process A-B1 without the lubricating oil

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cooling can be obtained. The cooling efficiency of the lubricating oil is defined by Equation (24)

α=

Pn1 ′ − Ps Pn1′

(24)

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and it is used to characterize the size of reducing the compression power for the lubricating oil.

where α is the cooling efficiency of the lubricating oil.

4.2. Evaluation index of a natural gas liquefaction process

In the section 2.4, NG liquefaction ratio and specific power consumption are defined. Furthermore, the cold and hot composite curves can be used to evaluate the energy utilization of a system. The larger the temperature difference of the cold and hot fluid is, the greater the exergy loss is. The meaning of their slopes can be expressed by the following equations:

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ACCEPTED MANUSCRIPT T2

Q = ∫ CPdT = CP(T2 − T1 )

(25)

CP = qm × cp

(26)

dT 1 = dQ CP

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T1

(27)

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where Q is the heat flow of the cold and hot fluids, kW, CP is the heat capacity flow rate and it can be expressed by Equation (26), kW·K-1, T2 and T1 are the outlet and inlet temperatures of the cold and

heat of the cold and hot fluids, kJ·kg-1·K-1.

5. Results and discussions

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5.1. Experimental results and discussions

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hot fluids, respectively, K, qm is the mass flow rate of the cold or hot fluid, kg·s-1, cp is the specific

Experimental tests were conducted at different rotational frequencies and outlet pressures,

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ranging from 25 to 50 Hz, and from 0.4 to 0.7 MPa, respectively. Among them, the design rotation

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frequency is 50 Hz and the design outlet pressure is 0.54 MPa according to the optimization results as shown in section 2.4.

5.1.1. Volumetric efficiency

Fig. 9 shows the influence of different rotational frequency on the volumetric efficiency. It can be seen that, except for the case of the outlet pressure of 0.54 MPa, the volumetric efficiency increases with the decrease of the outlet pressure at the same rotational frequency. This is because the high pressure can easily break the sealing layer formed by the lubricating oil and lead to serious

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ACCEPTED MANUSCRIPT leakage. There are some perturbations for the situation of the outlet pressure of 0.4 MPa, 0.5 MPa and 0.54 MPa. The experimental errors of the perturbations are 2.54%, 2.55% and 2.53%, respectively. If the experimental errors are considered, the law is very obvious and the volumetric efficiency

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increases slowly with the increase of rotational frequency at the same outlet pressure. The reason is that the compressor rotating at a high frequency can reduce the chance for leakage. It is worth mentioning that when the outlet pressure is 0.54 MPa and the rotational frequency is in the range of

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25~45 Hz, the volumetric efficiency is larger than for the situation with the outlet pressure of 0.5

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MPa. However, when the rotational frequency is 50 Hz, the above phenomenon is reversed, because the design outlet pressure of the compressor is 0.54 MPa and the corresponding effect of an eddy current is marginal. Furthermore, a high rotational frequency can overcome the faults caused by the eddy current and increase the volumetric efficiency.

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Another feature can be analyzed from this figure. When the outlet pressure deviates from the design outlet pressure by 7.4%, the maximum deviation for the volumetric efficiency is 3.9% at 30

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Hz and the minimum one is 0.71% at 45 Hz. However, when the outlet pressure deviates from the

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design outlet pressure by 26%, the maximum deviation for volumetric efficiency is 19.9% at 45 Hz and the minimum one is 2.9% at 25 Hz. This feature shows that the magnitude of the outlet pressure deviation from the design value has a great influence on the volumetric efficiency. When the outlet pressure is less than or more than the design outlet pressure, the volumetric efficiency would be improved or reduced, accordingly. The maximum volumetric efficiency is 0.91 at the outlet pressure of 0.4 MPa and the rotational frequency of 45 Hz (actual rotational speed is 2641 rpm). On the other hand, at the design outlet

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ACCEPTED MANUSCRIPT pressure of 0.54 MPa and the design rotational frequency of 50 Hz, the volumetric efficiency is only 0.78 and the error is 2.21%.

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5.1.2. Specific power rate

Fig. 10 presents the influence of different rotational frequency on the specific power rate. As shown in this figure, the specific power rate increases with the increase of the outlet pressure at the

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same rotational frequency (the situation of 0.5 MPa is the same as that of 0.54 MPa). The reason is that more power is needed to overcome the high pressure, and the leakage is more serious as the

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pressure becomes higher. At the same outlet pressure, there are some perturbations for the situation of the outlet pressure of 0.4 MPa and 0.5 MPa. The experimental errors of the perturbations are 2.12% and 2.13%, respectively. If the experimental errors are considered, the law is very obvious and the

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specific power rate is not sensitive to the influence of the rotational frequency. This is because the growth rate of the outlet volumetric flow is close to the growth rate of the compression work. In other words, in order to obtain 1 m3 of gas, the same cost is spent at the same rotational frequency.

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At the design outlet pressure of 0.54 MPa and design rotational frequency of 50 Hz, the specific

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power rate is 7.8 kW·(m3·min-1)-1 and the error is 2.2%. The minimum specific power rate is 6.5 kW·(m3·min-1)-1 and the error is 2.1%.

5.1.3. Shaft efficiency

Fig. 11 shows the influence of different rotational frequency on the shaft efficiency. It can be seen that the shaft efficiency is larger when the outlet pressure is close to the design outlet pressure of 0.54 MPa. This is because the eddy current loss increases gradually as the degree of deviation

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ACCEPTED MANUSCRIPT from the design outlet pressure increases. However, the situation of the outlet pressure of 0.4 MPa is an exception, as the lubricating oil is insufficient when the outlet pressure is low. At the same outlet pressure, there is a perturbation for the situation of the outlet pressure of 0.4 MPa. The experimental

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error of the perturbation is 2.43%. If the experimental error is considered, the law is very obvious: the shaft efficiency increases firstly and then decreases with the increase of the rotational frequency in general. There is an optimum rotation frequency to make the shaft efficiency maximum. At the

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design outlet pressure of 0.54 MPa and the design rotational frequency of 50 Hz, the shaft

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efficiency is 0.6 and the error is 2.38%.

5.1.4. Irreversible loss rate

Fig. 12 provides the influence of different rotational frequency on the irreversible loss rate.

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According to the degree of deviation from the design outlet pressure, the design pressure is regarded as the symmetry axis. The situations of the outlet pressures of 0.4 MPa, 0.7 MPa and 0.5 MPa, 0.6 MPa are compared, respectively. Firstly, the situations of the outlet pressures of 0.4 MPa and 0.7

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MPa are compared. When the rotational frequency is less than 40Hz, the irreversible loss rate is

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close. However, when the rotational frequency is more than 40 Hz, the difference between them becomes larger. Next, the situations of the outlet pressures of 0.5 MPa and 0.6 MPa are compared. When the rotational frequency is less than 40 Hz, the difference of the irreversible loss rate is small, while when the rotational frequency is more than 40Hz, the difference between them changes greatly. The above indicates that if the degree of deviation from the design pressure is the same, the difference of the irreversible loss rate is a constant value at low rotational frequency, but for high rotational frequency, the difference changes greatly. It can be also seen from the figure, at low

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ACCEPTED MANUSCRIPT rotational frequency, the larger the degree of deviation from the design pressure is, the more is the irreversible loss rate. Except for the case with the outlet pressure of 0.4 MPa, the reason is that the lubricating oil is insufficient when the outlet pressure is low.

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5.1.5. Cooling efficiency of lubricating oil

Fig. 13 plots the influence of different rotational frequency on the cooling efficiency. As shown

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in this figure, at the same rotational frequency, the cooling efficiency increases with the increase of the outlet pressure. This is the reason why the mass flow of the lubricating oil is larger when the

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outlet pressure becomes higher. At the same outlet pressure, the cooling efficiency decreases with the increase of the rotational frequency. This result is caused by the fact that the volume ratios of oil and gas are decreased as the rotational frequency is increased, and thus insufficiently cooling

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appears.

5.1.6. The influence of lubricating oil on the volumetric efficiency

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Lubricating oil has not only the effect of cooling, but also the effect of sealing. The above test data was obtained with a 46# Shell synthetic lubricating oil (S4 R 46). In order to explore the

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influence of the lubricating oil on the volumetric efficiency, another lubricating oil, 46# BP mineral lubricating oil, was used to obtain experimental data. Its kinematic viscosity is 6.8×10-6 m2·s-1 while the kinematic viscosity of 46# Shell synthetic lubricating oil is 7.5×10-6 m2·s-1 as listed in Table 4. The lubricating oil injection quantity is consistent in both experimental conditions and is equal to 12.5 L. Fig. 14 gives the influence of the lubricating oils on the volumetric efficiency. The characteristics show clearly that, the volumetric efficiency is larger with the 46# Shell synthetic

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ACCEPTED MANUSCRIPT lubricating oil, as the kinematic viscosity of 46# Shell synthetic lubricating oil is higher than that of the 46# BP mineral lubricating oil. This feature indicates that the effect of sealing is much better at a higher kinematic viscosity of the lubricating oil. However, individual condition does not follow the

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above conclusions because of experimental deviations. Furthermore, a larger kinematic viscosity may lead to an increase of the frictional power consumption.

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5.2. The influence of correction results on natural gas liquefaction process

In the natural gas liquefaction process as shown in Fig. 3, only the low pressure compressor

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was manufactured and its performance was tested, so only the parameters of the low pressure compressor were corrected. The design volumetric flow, design outlet pressure and design rotational frequency of the low pressure compressor are 6.5 Nm3·min-1, 0.54 MPa and 50 Hz, respectively.

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The isentropic efficiency of the low pressure compressor is assumed to be 0.8. According to the experimental data, when the actual rotational frequency and actual outlet pressure are 50 Hz and 0.54 MPa, respectively, the actual volumetric flow and the actual isentropic efficiency are 6.08

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Nm3·min-1 and 0.6, respectively.

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In order to describe the influential mechanism, Fig. 15 shows the thermodynamic cycle of the NG liquefaction process. Simulation

is the original simulation process, and simulation

is

another simulation process where the parameters were corrected according to the experimental data. These simulations were all optimized. From this figure, it is obvious that the refrigeration cycle (1-7) of the simulation

consumes more power than the simulation

compressing the NG in simulation

. Meanwhile, the power for

is larger than that in simulation

. This is because that the

initial pressure of the NG needs to be increased and should be matched well with another fluid in

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ACCEPTED MANUSCRIPT the multi-stream heat exchanger. Fig. 16 plots the cold and hot composite curves in the multi-stream heat exchanger. It is seen that the heat transfer temperature difference in the simulation

changes

significantly and as well does the exergy loss. The specific power consumption is changed from

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15.43 kWh·kmol-1 to 22.12 kWh·kmol-1.

5.3. Economic analysis

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Capital and operating costs represent the final figure of merit investigated [6]. Here, the operating profit is considered only because other facilities have not been manufactured. The specific

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power consumption is 22.12 kWh·kmol-1 when the actual performance of the single-screw compressor is considered. The amount of the liquefied natural gas is 4.665 kmol·h-1 (92 kg·h-1). In China, the price of the liquefied natural gas is about 1.0 $·kg-1. The price of the electricity is 0.15

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$·kWh-1 in the rush hour or 0.062 $·kWh-1 in the slack period. The average price is 0.106 $·kWh-1. So the cost of electricity and liquefied natural gas is 131.3 $·d-1 and 1104 $·d-1 if this equipment runs 12

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6. Conclusions

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hours a day. The operating profit is 972.7 $·d-1.

In this article, a new NG liquefaction process was simulated and optimized. A single-screw compressor prototype was designed for this liquefaction process according to the optimization results. To verify the performance of the prototype, an experimental system was set up. Based on the actual performance of the prototype, the efficiency of the NG liquefaction process was also observed. The main findings of this study are summarized as follows: (1) The degree of the outlet pressure deviation from the design outlet pressure has a great

29

ACCEPTED MANUSCRIPT influence on the volumetric efficiency. The volumetric efficiency increases with decreasing outlet pressure at the same rotational frequency and increases slowly with increasing rotational frequency at the same outlet pressure. The maximum volumetric efficiency is 0.91, which is superior to that of

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0.78 at the design point conditions. (2) The specific power rate increases with the increase of the outlet pressure at the same rotational frequency while is not very sensitive to the rotational frequency. The shaft efficiency is

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larger when the outlet pressure approaches the design outlet pressure. The difference of the

the difference changes greatly.

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irreversible loss rate is constant at low rotational frequencies, while at high rotational frequencies,

(3) The cooling of the lubricating oil can reduce the power consumption of the compressor effectively. The cooling efficiency increases by increasing the outlet pressure at a fixed rotational

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frequency, and decreases by increasing the rotational frequency at the same outlet pressure. The lubricating oil also has an effect on the sealing. The volumetric efficiency can be improved by 10.5%

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when lubricating oil was used.

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(4) Considering the actual performance of the studied compressor, the refrigeration cycle (1-7) consumes more power, and the consumption for compressing the NG is larger. The additional specific power consumption is 43.3%. The operating profit of the portable natural gas liquefaction process is 972.7 $·d-1.

Acknowledgment This work was sponsored by National Natural Science Foundation of China (51676163), by the National Key Research and Development Program of China (Grant Number 2017YFB0903603), by

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ACCEPTED MANUSCRIPT the Fundamental Research Funds for the Shaanxi province (2015KJXX-12), by the Fundamental

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Research Funds of Shenzhen City (JCYJ20170306155153048).

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ACCEPTED MANUSCRIPT

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pre-cooled mixed refrigerant LNG plant. Appl Therm Eng 2011;31(6-7):1091-8. [41] Mortazavi A, Alabdulkarem A, Hwang Y, Radermacher R. Novel combined cycle configurations for propane pre-cooled mixed refrigerant (APCI) natural gas liquefaction cycle. Appl Energy 2014;117(3):76-86.

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small-scale LNG plant. Cryogenics 2014;61(5):111-9.

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Table and figure captions list

Table 3. The optimization results of adjusted variables. Table 4. The physical property of lubricating oil. Table 5. Parameters of the prototype.

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Table 6. The details of the measuring instruments.

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Table 2. The adjusted variables in the process.

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Table 1. Mole fraction of the feed gas and other parameters in the process.

Fig. 1 The feed gas T-S diagram at different pressure. Fig. 2 The actual natural gas liquefaction process.

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Fig. 3 The natural gas liquefaction process simulated by Aspen HYSYS. Fig. 4 The main structure of the single-screw compressor. Fig. 5 A single-screw compressor prototype.

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Fig. 6 A schematic diagram of the experimental system.

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Fig. 7 A photo of the experimental system. Fig. 8 Compression processes in T-S diagram. Fig. 9 The influence of different rotational frequencies on the volumetric efficiency. Fig. 10 The influence of different rotational frequencies on the specific power rate. Fig. 11 The influence of different rotational frequencies on the shaft efficiency. Fig. 12 The influence of different rotational frequencies on the irreversible loss rate. Fig. 13 The influence of different rotational frequencies on the cooling efficiency.

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ACCEPTED MANUSCRIPT Fig. 14 The influence of lubricating oils on the volumetric efficiency. Fig. 15 The thermodynamic cycle of the NG liquefaction process

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Fig. 16 The cold and hot composite curves of the multi-stream heat exchanger

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Table 1. Mole fraction of the feed gas and other parameters in the process Parameters

Value

Notes

Feed gas pressure

––

Need to be optimized

Feed gas mole fraction components

CH4 C2H6 C3H8 n-C4H10 i- C4H10 N2

Refrigerant initial temperature

0.82 0.112 0.040 0.009 0.012 0.007

0.1 MPa 20

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16.06 kmol·h-1

Refrigerant flow rate

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Refrigerant mole fraction components LNG storage pressure

(6.5 Nm3·min-1) N2

0.79

O2

0.21

2.0 MPa 30

Ambient temperature

20

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Temperature after air cooler

The adiabatic efficiency of compressor

0.8

The adiabatic efficiency of expander

0.6

The highest pressure of refrigerant

2.916 MPa

The minimum approach temperature of multi-stream heat exchanger Pressure drop in multi-stream heat exchanger and air coolers

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––

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Feed gas flow rate

Refrigerant initial pressure

Based on the ambient temperature Depends on the results of optimization

20

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Feed gas temperature

Case study [15, 24]

Based on the ambient pressure Based on the ambient temperature Based on the primitive volume of the compressor we designed Simplify the process Case study [21, 22]

Based on the design

4 0 MPa

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Simplify the process

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Molar flow rate of feed gas qm,NG The pressure of NG1 pNG1 (MPa)

Low bound

High bound

2

10

2

20

-10

10

TE D

The temperature of point 6 T6 (K)

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Table 2. The adjusted variables in the process

0.1

AC C

EP

The pressure of point 2 p2 (MPa)

40

2.916

SC

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ACCEPTED MANUSCRIPT

Molar flow rate of

The pressure of

The temperature

The pressure of

feed gas qm,NG

NG1 pNG1

of point 6 T6

point 2 p2

(kmol·h-1) 4.376

(MPa)

(

5.739

7.9

AC C

EP

TE D

Optimization values

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Table 3. The optimization results of adjusted variables

41

)

(MPa) 0.54

SC

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ACCEPTED MANUSCRIPT

M AN U

Table 4. The physical property of lubricating oil Kinematic viscosity

Density

(m2·s-1)

(kg·m-3)

Type

7.5×10-6

843

AC C

EP

TE D

S4 R 46

42

Flash point (

)

230

Pour point (

)

-48

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Table 5. Parameters of the prototype

number of

of the

(mm)

the screw

star-wheel

200

6

11

star-wheel

(mm) 200

TE D

the screw

EP

values

Tooth number

Diameter of the

AC C

Parameter

Groove

Diameter of

43

Volume flow 3

-1

Compression

(Nm ·min )

ratio

6.5

5.4

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Type Measurement ranges

Flow meter

Power-meter

Speed-meter

CYB13

LZ series

PF9830C

0~1.6 MPa

50~500 Nm3·h-1

0~12 kW

DT6234B 2.5~99999

±0.5%

±1.5%

±0.4%

±0.3%

EP

±0.5%

-25~80

Pressure transmitter

AC C

Accuracy

-50~150

Temperature transmitter two SBWZP231

TE D

Temperature transmitter one WAP021

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Table 6. The details of the measuring instruments

44

r·min-1 ±1 r·min-1

300

0.1 MPa 0.5 MPa 1.0 MPa 2.0 MPa 4.0 MPa 6.0 MPa 8.0 MPa 10.0 MPa Two phase line

275

225 200 175 150 125

M Pa 6.0 Pa 4.0 M a MP 2.0 Pa 1.0 M a MP 0.5

M 0.1

TE D

100

a MP 0 . 10

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Temperature (K)

250

8.0 MPa

SC

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ACCEPTED MANUSCRIPT

80

90

Pa

100 110 120 130 140 150 160 Specific entropy (kJ/kmol·K)

AC C

EP

Fig. 1 The feed gas T-S diagram at different pressure

45

ACCEPTED MANUSCRIPT

6-2 6-1

6-4 6-5

6

2-4

2-2 2-1

2-6

2-7

2-10

8

4 4-5

3

M AN U

2-5

2

2-9

SC

2-8

2-3

1

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6-3

5

7

4-2

Natural gas Refrigerant - air

4-4

4-3

4-1

TE D

4-6

The mixture of natural gas and oil The mixture of air and oil

Lubricating oil Lubricating oil

1—Air pretreatment device; 2—The compressor and the lubricating oil circuit system (Air); 3—Multi -stream heat exchanger;

EP

4—The expanders and the lubricating oil circulation system; 5—NG pretreatment device; 6—The compressor and the lubricating oil circulation system (NG); 7—Throttle valve; 8—LNG storage

2-1, 2-6, 6-1—Single screw compressors; 4-1, 4-4—Single screw expanders; 2-2, 2-7, 4-2, 4-5, 6-2—Oil-gas separators; 2-3, 2-4 2-8,

AC C

2-9, 4-3, 4-6, 6-3,6-4—Plate-fin heat exchangers; 2-5, 6-5, 2-10—Motors

Fig. 2 The actual natural gas liquefaction process

46

TE D

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AC C

EP

Fig. 3 The natural gas liquefaction process simulated by Aspen HYSYS

47

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SC

Seal groove Screw groove Metal support

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Screw

TE D

Spindle

Star-wheel

EP

Flat key

AC C

Fig. 4 The main structure of the single-screw compressor

48

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TE D

Oil injection hole

AC C

EP

Fig. 5 A single-screw compressor prototype

49

Compressed air

Compressed air

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Atmospheric air

SC

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Lubricating oil

TE D

Mixture of oil and air

Compressed air

1—Motor 2—Air filter 3—Single screw compressor 4—Oil-gas separator 5—Minimum pressure valve exchanger (for air) 7—Buffer tank

6—Plate-fin heat

8—Ball valve 9—Plate-fin heat exchanger (for lubricating oil) 10—Fine oil separator

AC C

EP

Fig. 6 A schematic diagram of the experimental system

50

EP

TE D

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AC C

Fig. 7 A photo of the experimental system

51

TE D

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AC C

EP

Fig. 8 Compression processes in T-S diagram

52

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SC M AN U

0.9 0.8 0.7

0.4 MPa 0.5 MPa 0.54 MPa

0.6

TE D

Volumetric efficiency

1.0

0.5

25

0.6 MPa 0.7 MPa

30 35 40 45 Rotational frequency (Hz)

50

AC C

EP

Fig. 9 The influence of different rotational frequencies on the volumetric efficiency

53

RI PT 0.4 MPa 0.5 MPa 0.54 MPa

11

9 8 7 6 5

0.6 MPa 0.7 MPa

M AN U

10

SC

12

TE D

Specific power rate (kW/m3·min-1)

ACCEPTED MANUSCRIPT

25

30 35 40 45 Rotational frequency (Hz)

50

AC C

EP

Fig. 10 The influence of different rotational frequencies on the specific power rate

54

0.80 0.4 MPa 0.5 MPa 0.54 MPa

0.75

0.65 0.60 0.55 0.50

TE D

0.45 0.40

0.6 MPa 0.7 MPa

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Shaft effciency

0.70

SC

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ACCEPTED MANUSCRIPT

25

30 35 40 45 Rotational frequency (Hz)

50

AC C

EP

Fig. 11 The influence of different rotational frequencies on the shaft efficiency

55

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ACCEPTED MANUSCRIPT

SC

0.65

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0.55 0.50 0.45

0.4 MPa 0.5 MPa 0.54 MPa

0.40

TE D

Irreversible loss rate

0.60

0.35

25

0.6 MPa 0.7 MPa

30 35 40 45 Rotation frequency (Hz)

50

AC C

EP

Fig. 12 The influence of different rotational frequencies on the irreversible loss rate

56

0.26

0.4 MPa 0.5 MPa 0.54 MPa

0.22 0.20 0.18 0.16

TE D

0.14 0.12

0.6 MPa 0.7 MPa

M AN U

Cooling effciency

0.24

SC

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ACCEPTED MANUSCRIPT

25

30 35 40 45 Rotational frequency (Hz)

50

AC C

EP

Fig. 13 The influence of different rotational frequencies on the cooling efficiency

57

ACCEPTED MANUSCRIPT

0.4 MPa

RI PT

0.9 0.8

0.5 MPa

0.7 0.6 0.5

25

SC

0.4 MPa 0.5 MPa 46# Shell synthetic lubricating oil 46# BP mineral lubricating oil

30 35 40 45 Rotational frequency (Hz)

M AN U

Volumetric efficiency

1.0

50

(a) outlet pressures are 0.4 MPa, 0.5 MPa

1.0

0.54 MPa 0.7 MPa

0.6 MPa

TE D

0.54 MPa

0.8 0.7 0.6

0.7 MPa

EP

Volumetric efficiency

0.9

0.6 MPa

0.5

AC C

0.4

25

46# Shell synthetic lubricating oil 46# BP mineral lubricating oil

30 35 40 45 Rotational frequency (Hz)

50

(b) outlet pressures are 0.5.4 MPa, 0.6 MPa, 0.7 MPa

Fig. 14 The influence of lubricating oils on the volumetric efficiency

58

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Simulation ℃ Simulation ℃ Simulation ℃ Simulation ℃

500

300 200 100 0

(Air) (Air) (NG) 4 (NG)

NG1

NG1

M AN U

Temperature (℃ )

400

SC

600

4

2

2

5 5 3 3 1 NG2 NG NG 6 6 NG2 1 NG3 7 NG3 NG4 7 8 NG4 8 NG (two phase line）

Air（ two phase line）

-100

TE D

-200

20

40

60 80 100 120 140 160 180 200 220 Specific entropy (kJ/kmol·K)

AC C

EP

Fig. 15 The thermodynamic cycle of the NG liquefaction process

59

RI PT

ACCEPTED MANUSCRIPT

SC

40 0

M AN U

Temperature (℃ )

20 -20 -40 -60

Simulation℃ (cold fluid) Simulation℃ (hot fluid) Simulation℃ (cold fluid) Simulation℃ (hot fluid)

-80 -100

TE D

-120 -2

0

2

4

6

8

10 12 14 16 18 20

Heat flow (kW)

AC C

EP

Fig. 16 The cold and hot composite curves of the multi-stream heat exchanger

60

ACCEPTED MANUSCRIPT

Highlights
The performance of a single-screw compressor applied in the optimized NG liquefaction process is tested.
The actual volumetric efficiency and shaft efficiency are 0.78 and 0.6, respectively.

RI PT


The additional power of 43.3% are consumed when the simulated parameters are

AC C

EP

TE D

M AN U

SC

corrected.