2023 International Ocean Energy Symposium &20th Joint Young Research Forum

Date

September 15-16, 2023

Venue

Industry-Academic Hub Hall (G2, 201), KMOU
Korea Maritime & Ocean University (KMOU, Korea)
Pukyong National University (PKNU, Korea)

Participation

Mokpo National Maritime University (MMU, Korea)
Mokpo National University (MNU, Korea)
Saga University (SAGA-U, Japan)
National Fisheries University (NFU, Japan)

SponsorGraduate School & Department of Green and Smart Ship Equipment(KMOU)

CONTENTS◇ Words for invitation◇ ProgramOral Presentation (1)1

An Experimental Observation on Thermal Runawayof 18650 Lithium-Ion Battery

Ju-Won PARK

KMOU

1

2

Study on Evaluation of Heat Transfer Performanceof Heat Exchanger in Hybrid Cycle of Ocean
Thermal Energy Conversion

Tomohide
NISHIMURA

SAGA-U

2

3

Fabrication of a Heat Exchanger for Two-Phase
Flow in Organic Rankine Cycle System for Ships

Dae-Jung
HWANG

KMOU

3

4

Experimental Study of Marine Biofouling for the
Plate Type Heat Exchanger with Seawater Cooling Efficacy of Electromagnetic Processing Equipment

Naoki
KIYAMANFU

45

N2O Decomposition and Kinetic Modeling with
Varying Reaction Temperatures and Residence
Times

Su-Hyeon
KIM

KMOU

6

6

Posture Control of Counter-rotating Propeller Type
Tidal Power Unit with Winglet

HarutoFUCHIWAKI

SAGA-U

7

7

A Study on Performance Improvement of Boil-off
Gas Re-liquefaction Systems for LNG-fueled Ships

Jun-Seong
KIM

KMOU

9

Oral Presentation (2)8Flame Surfaces Variation through Ionic Wind andFlame Stretch in Mass-Diffusively Unstable Flame

Si-Yeong YU

KMOU

11

9

Research on buoy shape to improve wave energy
conversion efficiency

TakumaIMAIZUMI

SAGA-U

12

10

Correlation between Ionic Wind and ThermoAcoustic Instability in Downward-Propagating
Premixed Flame under Electric

Dae-Won IM

KMOU

16

11

Study of Power Generation System Utilizing Waste
Atsuyoshi
Heat from Marine Engine (Possibility of introducing
ETSUKAWAHFO Refrigerant)

NFU

17

12

Combustion field measurement of temperature,
concentration and velocity using TDLAS

Jeong-Woong
HONG

KMOU

19

13

Tip clearance gap effect on the hydraulicperformance of a mixed flow pump with semi-open
casingUjjwalSHRESTHA

MNU

20

14

A study of fuel cell-battery propulsion system for
four types of vessels

Jong-In Kim

KMOU

21

Poster Presentation
15

Characteristics of HFO Refrigerants in Double-Stage Tetsuya
Reverse Rankine Cycle Refrigerator
HONDA

NFU

22

16

Oceanographic Survey for Installation of OTEC
Plant in Kumejima Island (Influence from the Aguni
Basin)

ShodaiNAGATA

NFU

24

17

Research on Periodic Characteristics by
Visualization of Plate Type Evaporator Interior

Ayano
NAKAGAWA

SAGA-U

26

18

Development of Optimization Method of Jacket
Structures of Offshore Wind Turbines

Takahiro
NAGATSUSAGA-U

27

19

Heat transfer characteristics of FCU and plate heatexchanger in seawater air conditioning system

Haruki IDA

SAGA-U

29

20

Boiling heat transfer characteristics of ammonia in
micro channel and plate heat exchanger

Shoichi
SUEHIROSAGA-U

31

21

Research on optimal design of ocean thermal energy
conversion system

Takayuki
SAGA-UHARAGUCHI

35

22

Research on hot spring water power generation
using hybrid cycle

Tomoyuki
HIRAISHI38

SAGA-U2023 International Ocean Energy Symposium & 20th
Joint Young Research Forum
Words for invitation
This symposium is held every year between Korea and Japan University
every other year.
This symposium topics of interest include technology of ocean energy.
And this symposium is aimed at promoting cooperation not only amongprofessors and researchers but also between all experts in the oceanenergy technology field.
This symposium will provide various programs, such as academic
paper presentations, poster sessions, invited talks and also some otheractivities for promoting cooperative relation.
We sincerely wish to have your interest and participation that will makethis forum a success.
Sincerely yours,
Organizer:
Ph.D Cheol OH,

Ph.D Sung-Hwan YOUN,

Ph.D Jun-Seong KIM

Program
September 15 (Fri.)
Time

Event

Remark

09:30 - 10:00

Registration

Participant

10:00 – 10:15

Greetings

10:15 – 12:00

Oral Presentation (1)12:00 – 14:00

Lunch time

14:00 – 15:00

Poster Presentation

Presenters: 8

15:00 – 17:00

Oral Presentation (2)Presenters: 7

17:00 – 18:00

Go to dinner

18:00 – 20:00

Welcome dinner

Presenters: 7

September 16 (Sat.)
Time

Event

09:00 - 13:00

Meeting

Remark◎ Oral Presentation (1)Chairman: Jun-Seong KIM (KMOU)
10:15~10:30

An Experimental Observation on Thermal
Ju-Won PARK
Runaway of 18650 Lithium-Ion Battery

KMOU

10:30~10:45

Study on Evaluation of Heat Transfer
Tomohide
Performance of Heat Exchanger in Hybrid
NISHIMURA
Cycle of Ocean Thermal Energy Conversion

SAGA-U

10:45~11:00

Fabrication of a Heat Exchanger for Two-Phase
Dae-Jung
Flow in Organic Rankine Cycle System for
HWANG
Ships

KMOU

11:00~11:15

Experimental Study of Marine Biofouling for
the Plate Type Heat Exchanger with Seawater Naoki
Cooling -Efficacy of Electromagnetic KIYAMA
Processing Equipment

NFU

11:15~11:30

N2O Decomposition and Kinetic Modeling
Su-Hyeonwith Varying Reaction Temperatures and
KIM
Residence Times

KMOU

11:30~11:45

Posture Control of Counter-rotating Propeller Haruto
Type Tidal Power Unit with Winglet
FUCHIWAKI

SAGA-U

11:45~12:00

A Study on Performance Improvement of BoilJun-Seong
off Gas Re-liquefaction Systems for LNGKIM
fueled Ships

KMOU

◎ Oral Presentation (2)Chairman: Yasutaka IMAI (SAGA-U)
15:00~15:15

Flame Surfaces Variation through Ionic Windand Flame Stretch in Mass-Diffusively Si-Yeong YU
Unstable Flame

KMOU

15:15~15:30

Research on buoy shape to improve wave Takuma
energy conversion efficiency
IMAIZUMI

SAGA-U

15:30~15:45

Correlation between Ionic Wind and ThermoAcoustic Instability in Downward-Propagating Dae-Won IM
Premixed Flame under Electric

KMOU

15:45~16:00

Study of Power Generation System Utilizing
Atsuyoshi
Waste Heat from Marine Engine (Possibility of
ETSUKAWA
introducing HFO Refrigerant)

NFU

16:00~16:15

Break

16:15~16:30

Combustion field measurement of temperature, Jeong-Woong
concentration and velocity using TDLAS
HONG

KMOU

16:30~16:45

Tip clearance gap effect on the hydraulicUjjwalperformance of a mixed flow pump with semiSHRESTHA
open casing

MNU

16:45~17:00

A study of fuel cell-battery propulsion systemJong-In Kim
for four types of vessels

KMOU

◎ Poster Presentation
Chairman: Junichi OHARA (NFU)
Characteristics of HFO Refrigerants in Double-Stage Reverse Tetsuya
Rankine Cycle Refrigerator
HONDA

NFU

Oceanographic Survey for Installation of OTEC Plant in Shodai
Kumejima Island (Influence from the Aguni Basin)
NAGATA

NFU

Research on Periodic Characteristics by Visualization of Plate Ayano
Type Evaporator Interior
NAKAGAWA

SAGA-U

Development of Optimization Method of Jacket Structures of Takahiro
Offshore Wind Turbines
NAGATSUSAGA-UHeat transfer characteristics of FCU and plate heat exchanger inHaruki IDA
seawater air conditioning system

SAGA-U

Boiling heat transfer characteristics of ammonia in micro Shoichichannel and plate heat exchanger
SUEHIROSAGA-U

Research on optimal design of ocean thermal energy conversion Takayuki
system
HARAGUCHI

SAGA-U

Research on hot spring water power generation using hybrid Tomoyuki
cycle
HIRAISHISAGA-UORAL
PRESENTATION

An Experimental Observation on Thermal Runaway of 18650 Lithium-Ion Battery
Juwon Park1,2), Daegeun Park3) and Sung Hwan Yoon1,2)
Department of Marine System Engineering, Korea Maritime and Ocean University, 727 Taejong-ro, Yeongdo-gu,
Busan, 49112, Republic of Korea1)
Interdisciplinary Major of Maritime AI Convergence, Korea Maritime and Ocean University, 727 Taejong-ro,
Yeongdo-gu, Busan, 49112, Republic of Korea2)
Carbon Neutral Technology R&D Department, Korea Institute of Industrial Technology (KITECH), 89
Yangdaegiro-gil, Seobuk-gu, Cheonan, 31056, Republic of Korea3)
internal defect of the lithium-ion battery would inevitably occur. The lithium-ion battery consists ofcathode material, anode material, separator, andelectrolyte, but if the structure of the separator isdamaged due to internal or external defects, thechemical reaction caused by the short circuit between
the cathode material and anode material causes athermal runaway. However, it is observed that thermal
runaway is suppressed because of the reduction ofinternal energy if the SOC-level is below 80% in thisstudy, through this study, we will discuss how toquickly extinguish a fire caused by thermal runaway
even if thermal runaway occurs.

The need for a ‘Rechargeable battery’ is growing asa driving force in the information and communication
device and electric vehicle industry. Unlike disposable
primary cell, rechargeable battery that can be semipermanently reused through charging after
discharging is a device that store external electricalenergy after converting external electrical energy into
chemical energy form and make electric energy whennecessary.
For about 120 years, rechargeable batteries havebeen useful in everyday life in various forms, such as
lead-acid batteries, nickel-cadmium batteries, nickelhydrogen batteries, etc. Among them, a lithium-ion
battery, which can be used for a long time without
frequent charging, is currently used as a typicalrechargeable battery. Lithium-ion battery has anadvantage of high energy density and excellent power,
and have a long life due to no memory effect.
Therefore, lithium-ion battery is used in various fields
such as laptops, mobile phones, drones, and electricvehicles. In particular, in the case of electric vehicle
that uses a lithium-ion battery as the main power, the2021 Global Electric Vehicle Outlook Report releasedby the International Energy Agency (IEA) reported
that the share of electric vehicle in the global
automobile market will soar to 12% by 2030.
However, the issue of safety is constantly being
raised due to the thermal runaway that could occur due
to internal or external factors of a lithium-ion battery.
Thermal runaway has been reported to rapidly burnand explode due to mechanical, electrical, and thermal
defects1-3). If a thermal runaway occurs in the battery
pack, adjacent cells also cause a thermal runaway due
to the generation of very high heat. As a lithium-ion
battery has large internal energy, it is very difficult tosuppress the thermal runaway once generated.
Therefore, a suppression study on the thermal
runaway phenomenon is essential by experimentallyobserving the mechanism for the thermal runaway
phenomenon.
The purpose of this study is to experimentallyobserve the mechanism for the thermal runaway of alithium-ion battery and propose a technology tosuppress the thermal runaway. Thermal runaway of thelithium-ion battery was observed by adjusting the
supplied heat transfer coefficient and SOC level, andan inert gas was formed in a chamber to suppress thethermal runaway. As a result, if the internal
temperature of the lithium-ion battery increased, it wasdetermined that a thermal runaway caused by the

References(1) Wang, Q., et al., “A Review of Lithium Ion
Battery Failure Mechanisms and Fire Prevention
Strategies.” Combust. Sci., 2019, 73, pp. 95-131.
(2) Zhong, G., et al., “Experimental Analysis ofThermal Runaway Propagation Risk within
18650 Lithium-Ion Battery Modules.” J.
Electrochem. Soc., 2018, 165, pp. A1925-A1934.
(3) Baird, A.R., et al., “Explosion Hazards from
Lithium-Ion Battery Vent Gas.” J. Power
Sources, 2020, 446, pp.1-13.

1

・Title「Study on Evaluation of Heat Transfer Performance of Heat Exchanger in Hybrid Cycleof Ocean Thermal Energy Conversion」
・Authors
Yasuyuki Ikegami (Saga Univ.) Takeshi Yasunaga (Saga Univ.)
Takafumi Morisaki (Saga Univ.) Shuji Miyazono (Saga Univ.)
Yoshitaka Otakara (Saga Univ.) Tomohide Nishimura (Saga Univ.)
・Abstract
Ocean Thermal Energy Conversion (OTEC) is a type of power generation that utilizes the temperature
difference between the surface and deep layers of the ocean and is attracting attention as a form of renewableenergy. However, because OTEC directly passes seawater, which is the heat source, through an evaporator,
it is necessary to use a heat exchanger made of titanium, which is expensive due to the deterioration of theheat transfer performance of the heat exchanger in the evaporator caused by contamination from marineorganisms and the corrosion of seawater. To solve this problem, a hybrid cycle OTEC (H-OTEC) was
proposed.
H-OTEC combines a seawater desalination system with OTEC. The evaporator in H-OTEC is called anevaporator-condenser (Eva-Con) because it condenses the heat source at the same time. The Eva-Con is
also a particularly important piece of equipment because it connects the power generation system to thedesalination system. My objective was to experimentally evaluate the heat transfer performance of the EvaCon by conducting performance tests. Furthermore, I aimed to calculate experimental equations for the heattransfer coefficients of both the heat source and the working fluid (Ammonia) of the Eva-Con.
・Structure of the paper
1. Background
2. Experimental apparatus
3. Heat Exchanger Performance Methods
4. Experimental Results and Discussion
5. Conclusion

2

Fabrication of a Heat Exchanger for Two-Phase Flow in Organic Rankine Cycle System for Ships
Dae Jung Hwang 1 ㆍ Tae Woo Lim 2 ㆍ Jae Hoon Jee 3 ㆍ Eun Shin Bang 4 ㆍ Cheol Oh †Unlike conventional heat exchangers, a meticulous design approach is required for Waste Heat Recovery Units (WHRU)
and condensers utilized in Organic Rankine Cycle (ORC) power systems. This is due to incorporating phase change phenomena
and utilizing a refrigerant as a working fluid instead of water. The purpose of this study is to fabricate a waste heat recovery
unit and condenser that are capable of operating stably and efficiently for the use of ORC systems in ships.
In this study, WHRU and Condenser were designed using numerical analysis techniques, and prototypes were fabricated based onthese designs. In addition, the validity of the heat exchanger CFD modeling was confirmed through comparison of the simulation resultswith those obtained from the prototype's experimental testing. The starting position of phase change, a crucial factor for stable operationin terms of two-phase flow, was predicted for each refrigerant when an actual heat exchanger was employed in a ship.
The WHRU experiment confirmed a heat transfer rate of 403 ㎾ and a phase change, where the refrigerant changed froma liquid state at an inlet temperature 31℃ to a superheated vapor state 78℃ upon discharge. As a result of the experiment of
the condenser, it has been confirmed that there is a heat transfer rate of 113 ㎾ and a phase change occurs when a refrigerant,
in a gaseous state at a temperature of 74℃, is introduced at the inlet and discharged in a supercooled state at 38℃.

Fig. 1 Process of design and fabrication of WHRU

† Corresponding Author: Professor, Division of Marine System Engineering, Korea Maritime & Ocean University, E -mail: ohcheol@kmou.ac.kr
1 Researcher, Department of Marine Environment, Korea Maritime Cooperation Center, E-mail: hdaejung@imkmc.or.kr
2 Professor, Division of Marine System Engineering, Korea Maritime & Ocean University, E-mail: kyunlim@kmou.ac.kr
3 Professor, Division of Maritime Police, Mokpo National Maritime University, E-mail: jhjee@mmu.ac.kr
4 Ph. D., Division of Marine System Engineering, Korea Maritime & Ocean University, E-mail: esbang627@kmou.ac.kr

3

Experimental Study of Marine Biofouling for the Plate Type Heat Exchanger with
Seawater Cooling -Efficacy of Electromagnetic Processing Equipment
Naoki KIYAMA1), Junichi OHARA1), Hideyuki YOSHIMURA1), Masaki ARIYOSHI1),
Syuhei KUGIMIYA1), Minoru MATSUSHITA2) and Yasuyuki IKEGAMI,3)
National Fisheries University, 2-7-1 Nagatahonmachi, Shimonoseki-shi, Yamaguchi 759-6595, Japan 1)
Hisaka Corporation, 2-1-48 Higashikonoikechou, Higashiosaka-shi, Osaka 578-0973, Japan 2)
Institute of Ocean Thermal Energy, Saga University, 1-Honjomachi, Saga-shi, Saga 840-8502, Japan3)
Abstract
In this study, the antifouling effect of the electromagnetic processing equipment on seawater was verified byconducting long-term experiments using the electromagnetic processing equipment, which is an antifouling device, toreduce fouling of the heat transfer surface on the seawater side of a plate heat exchanger with continuous seawater
flow. The operation of the electromagnetic processing equipment in the seawater system of the plate heat exchanger
for 90 days improved the rate of decrease of the heat transfer coefficient by 7.4% and the increase of the fouling factor
by 32.5%. These results proved that the installation of an electromagnetic processing equipment in a plate heatexchanger has an antifouling effect against seawater fouling.
Key Words: Electromagnetic Processing Equipment , Plate Heat Exchanger , Fouling Factor
1. Introduction
Currently, the shipping industry is trying to reduce
operating costs by introducing various technologies in
response to soaring fuel prices. Plate heat exchangers,which are important equipment on ships, often utilize
seawater and incur large maintenance costs.
Therefore, in this study, the antifouling effect of theelectromagnetic processing equipment in seawater tube
was verified by conducting long-term experiments. The
equipment is an antifouling device to reduce fouling,which seawater continuously feeds on the seawater-side
heat transfer surface.

Table 1 Experimental conditions
Experimental conditions
Electric water heater setting temperatureWarm water inlet temperatureWarm water flow rate
Sea water flow rate

[℃]
[℃]
[L/s]
[L/s]

40.0
40.0
0.30
0.30

seawater sides, respectively. The electromagnetic
processing equipment was installed in the No.1 test plate
heat exchanger system, while the No.2 system wasdesigned to have no effect of the electromagnetic
processing equipment, so that comparisons can be madebetween the two systems with and without theelectromagnetic processing equipment.

2. Experimental Equipment
Figure 1 shows the schematic diagram of theexperimental apparatus for seawater side contamination
using a plate heat exchanger. The experimental
apparatus consists of two plate heat exchangers, twoseawater tanks (indoor and outdoor), a hot water tank, acooling fan for the seawater tank (outdoor), a seawater
pump, a hot water pump, a hot water circulation pump,
two electromagnetic flow meters, and an electric water
heater. The experiment was conducted with anelectromagnetic processing equipment ZETAWAVE 1)[Koshin Chemical Corporation, Technolab Co, Takashin
Chemical Co.]. Installed plate heat exchangers are five[RX-015A-KNHJ-11, Hisaka Seisakusho], an effective
total heat transfer area of 0.213 m2, and three channels
on
the
hot
water
and

3. Experimental Method and Conditions
Experimental data were measured for temperature,
flow rate, and differential pressure, using a data lo
gger at 12:00 AM. Experimental conditions are sho
wn in Table 1. The experiment was conducted from April 9, 2023 to July 8, 2023, operating contin
uously for 90 days.
4. Experimental Results
In this experiment, we verified the heat transfer
coefficient, which indicates the performance of the plate
heat exchanger, and the fouling factor, which evaluates
the contamination of the plate surface, based on the
results of the experiment. As reference data, theexperimental results of the heat transfer coefficient andfouling factor for the case of an electromagnetic
processing equipment conducted under the same
conditions last year are also shown in the figure2). The
heat transfer coefficient and fouling factor for the caseof an electromagnetic processing equipment are also
shown in the figure.
4.1 Heat Transfer Coefficient
The heat transfer coefficient U is obtained from thefollowing equation.
𝑈=

Fig.1 Diagrammatical view of the device

𝑄
𝐴∆𝑇𝑚

(1)

Figure 2 shows the change over time of the heat4

transfer coefficient. Figure 2 shows that the heat transfer
coefficient decreases continuously from the beginningof the experiment. After 90 days, the heat transfer
coefficient decreased from 4418.7[W/(m2-K)] to
2199.3[W/(m2-K)] with the electromagnetic processing
equipment, the decrease of 50.2%. Without theelectromagnetic processing equipment, the heat transfer
coefficient decreased from 4016.8 [W/(m2-K)] to 1704.7[W/(m2-K)], the decrease of 57.6%. Therefore, it is clearthat the electromagnetic processing equipment installedin the seawater system of the plate heat exchanger for 90
days can reduce the reduction of the heat transfer
coefficient against seawater contamination by 7.4%
4.2 Fouling Factor
In actual heat exchangers, dirt adheres to the heattransfer surfaces during operation, degrading heattransfer performance. Although some guidelines havebeen proposed for the fouling factor, it is said to vary
greatly depending on the nature of the fluid, temperature,
flow velocity, etc. Here, we calculated the fouling factor
under the experimental conditions and verified the effectof the electromagnetic processing equipment. In thisexperiment, the fouling factor Rfre was defined by thefollowing equation
𝑅𝑓𝑟𝑒 =

1
1
−
𝑈2 𝑈1

3) The installation of an electromagnetic processing
equipment in a plate heat exchanger proved to beeffective in preventing seawater fouling.
References(1) ZETAWAVE([Koshin Chemical Corporation, Te
chno lab Co , Tak ash in Ch em ical Co.] ) ：https://www.koshin-chem.co.jp/20151211114446(2) Kaede OKUSAKO : Study on the antifoulingeffect of plate heat exchangers on seawater Verification of antifouling effect by long-term
operationofelectromagnetic
processing
equipment -. 2022 Graduation Thesis Presentationof the Department of Marine Machinery andEngineering, Fisheries University, Shimonoseki(2023).
(3) S.J.PUGH,
G.F.HEWITT,H.MULLERSTEINHAGEN, “Fouling During the Use of
Seawater as Coolant-the Development of a User
Guide”, Heat Transfer Engineering, Vol. 26, No. 1(2005), pp. 35-43.

(2)

Where U1 is the heat transfer coefficient before foulingand U2 is the heat transfer coefficient after fouling.
Figure 3 shows the change over time of the foulingfactor. Figure 3 shows that the fouling factor has
decreased continuously since the beginning of theexperiment. After 90 days, the stain coefficientincreasedto
2.28×10-4[m2·K/W]
withtheelectromagnetic processing equipment and to 3.38×104[m2·K/W] without the electromagnetic processing
equipment. Therefore, it is clear that 32.5% antifoulingeffect against seawater fouling is possible by installingand operating an electromagnetic processing equipmentin the seawater system of a plate heat exchanger for 90
days.

Fig. 2 Overall heat transfer coefficient with the day

5. Conclusions
In this study, the antifouling effect of theelectromagnetic processing equipment in seawater
system was verified by conducting long-term
experimentusing the electromagnetic processing
equipment, which is an antifouling device, to reducefouling of the heat transfer surface on the seawater side
by continuously passing seawater through a plate heatexchanger, and the following results were obtained.
1) The rate of decrease in the heat transfer coefficient
was improved by operating the plate heat exchanger
with an electromagnetic processing equipment in
the seawater system for 90 days, and it was foundthat the reduction in the heat transfer coefficient for
7.4% seawater contamination could be mitigated.
2) The increase in fouling factor was improved byoperatingtheelectromagnetic
processing
equipment in the seawater system of the plate heatexchanger for 90 days, and it was found that
antifouling effectiveness against seawater foulingof 32.5% was possible.

Fig. 3 The variation of fouling thermal
resistance Rfre with the day
5N2O Decomposition and Kinetic Modeling with Varying Reaction Temperatures andResidence Times
Suhyeon Kim1,2), Daegeun Park3), Seung-Gon Kim4) and Sung Hwan Yoon1,2)
Department of Marine System Engineering, Korea Maritime and Ocean University, 727 Taejong-ro, Yeongdo-gu,
Busan, 49112, Republic of Korea1)
Interdisciplinary Major of Maritime AI Convergence, Korea Maritime and Ocean University, 727 Taejong-ro,
Yeongdo-gu, Busan, 49112, Republic of Korea2)
Carbon Neutral Technology R&D Department, Korea Institute of Industrial Technology (KITECH), 89 Yangdaegiro-gil, Seobuk-gu, Cheonan, 31056, Republic of Korea2)
Energy Efficiency Research Division, Korea Institute of Energy Research (KIER), 152 Gajeong-ro, Yuseonggu, Daejeon, 34129, Republic of Korea2)

In July 2023, both Earth's surface temperature and sea
surface temperature reached their highest recordedlevels. If current greenhouse gas emissions persist, it's
anticipated that all Arctic glaciers will completely meltby 20301). A NASA climate scientist cautioned, "Mostpeople STILL underestimate the impending peril. This
may well be the coolest summer for the rest of your life."
Because even if global greenhouse gas reduction targets
are met, it's anticipated that by 2050, all Arctic glaciers
will vanish, even more substantial reductions ingreenhouse gas emissions are needed than initiallyplanned.
Within the maritime industry, significant efforts havebeen made to reduce greenhouse gas emissions throughGHG (Greenhouse Gas) strategies. For carbon dioxide(CO2) and methane(CH4), technologies such as Carbon
Capture, Utilization, and Storage (CCUS) and methane
oxidation catalytic systems are employed for reduction.
However, there is no significant progress on othergreenhouse gases.
Nitrous oxide (N2O), the third-largest emitted
greenhouse gases, has a global warming potential
approximately 300 times higher than CO2 and produces
NO and NO2 in the stratosphere2). Since these NOx
emissions contribute to the gradual degradation of theozone layer, it is imperative to reduce N2O emissions.
N2O is produced from sources such as fuel combustion
in boilers and fossil fuel combustion. Notably, ammonia
has recently received attention as a carbon substitutefuel, despite emitting significant amounts of NOx andN2O during decomposition, thus necessitating
technologies for their reduction.
This study focuses on decomposing high concentrationsof N2O under various reaction temperatures andresidence times to determine the most efficientconditions. It also examines the pathways andconcentrations of NOx generated during N2O reduction.
Additionally, we employed various rate constants to
conduct numerical analysis using the Plug Flow Reactor(PFR), a 1D simulation model within Chemkin-Pro3).
Further details will be presented during an oralpresentation3).

Communications, 2023, 14.1, pp. 3139.
(2) Griggs, D.J., Noguer, M., “Climate Change 2001:
The Scientific Basis. Contribution of Working
Group I to The Third Assessment Report of The
Intergovernmental Panel on Climate Change.”
Weather 2002, 57, pp. 267-269.
(3) Kee, R.J., Rupley, F.M., Miller. J.A., “Chemkin-II:
A Fortran Chemical Kinetics Package for The
Analysis of Gas-phase Chemical Kinetics." Sandia
National Lab.(SNL-CA), Livermore, CA (UnitedStates), 1989.

References(1) Kim, Yeon-Hee. et al., "Observationallyconstrained projections of an ice-free Arctic evenunder a low emission scenario." Nature
6

Posture Control of Counter-rotating Propeller Type Tidal
Power Unit with Winglet
Haruto Fuchiwaki*, Tengen MURAKAMI** and Toshiaki KANEMOTO***Faculty of Science and Engineering, Saga University
1 Honjo, Saga 840-8502, Japan
E-mail: 22728014@edu.cc.saga-u.ac.jp**Institute of Ocean Energy, Saga University
1 Honjo, Saga 840-8502, Japan

1. Introduction
The tidal current is expected as an attractive renewable energy source because the time change of flow direction
and velocity is predictable. In this research, a mooring support structure is applied to the horizontal axis counterrotating type power unit, which is not only save the initial cost due to the civil construction but also appropriate forusing at narrow strait due to set up easily. The counter-rotating type power unit can keep the posture stable withoutrolling motion in the stream when it is mooring with a cable at moment center among the buoyancy, the gravity, andthe drag force. In such an arrangement, the drag force induced from downstream type propellers works effectively tokeep the horizontal posture. However, once the unit inclines against the mainstream by unexpected change of local
flow velocity, the horizontal posture is disrupted and further aggravated by the drag force induced from the flow
separation around the casing which includes a nacelle supporting a generator and two hubs of the front and rearpropellers. Therefore, the winglet used in aircraft is set to the blade tip of the rear propeller to help the power unit
recover to horizontal posture. In this study, the effects of winglet length on posture and blade performance were
investigated by circulation type water channel tests and CFD.
2. Circulation type water channel tests
Fig. 1 shows the model power units with and without winglet at rear propeller. The length of winglets was variedin 3 cases (9 mm,19 mm,29 mm), and each ratio between length of winglet and the diameter of rear propeller are
0.038, 0.080, and 0.122. The generator is not installed in the models. The power unit consists of front propeller withthree blades and rear propeller with five blades. The unit is set in a circulation type water channel with a cross-sectional
area of 0.7 m (water depth) x 1.0 m (width) and moored with a wire at the nose spinner. In the experiments, the postureof the model was analyzed with the images by a camera. Besides, the rotational speeds of front and rear propeller were
measured using a stroboscope (SATOTECH DT-2349JP). In the following discussion, the power unit model withwinglet is called as PUW in comparison with PUO of the model without the winglet. The flow velocity Vin was changed
from 0.7 m/s to 0.8 m/s, 0.9 m/s and 1.0 m/s.

Fig. 1 Model units with and without winglet at rear propeller
Fig. 2 shows the tilt angle  between horizontal direction and the axes of propellers. Besides, Fig. 3 is the counter7

rotational speed Nt. As shown in these figures, the tilt angle  became low with increase of the winglet length. On theother hand, the counter-rotational speed Nt was higher in case PUW0.080.

Fig. 2 Changes in tilt angle  due to flow velocity Vin

Fig. 3 Changes in counter-rotational speed Nt

3. Numerical simulations
The 3-D turbulent flow was simulated at the steady state condition by the commercial CFD code of ANSYS-CFX
ver.19 with SST turbulent model to understand the propeller performances with and without winglet. The simulation
domain is composed of 1 pitch passage areas of front and rear propellers. Fig. 4 is the changes in torque due to tip
speed ratio  and Fig. 5 shows the power coefficient Cp, where the  is defined as the ratio of the circumferential
velocity at blade tip and the flow velocity. The torque and the Cp were divided by ones at  = 6.4 giving the maximum
power coefficient Cp in case PUO. As shown in Fig. 4 and Fig. 5, the torque and power coefficient Cp were higher incase PUW0.080 in comparison with the other cases with winglet. So, it seems that the ratio 0.080 between length of
winglet and the diameter of rear propeller is suitable for posture control.

Fig. 4 Changes in torque ratio due to 

Fig. 5 Changes in power coefficient ratio

4. Conclusion
In this study, the effects of winglet length on posture and blade performance were investigated by circulation typewater channel tests and CFD. As the result, it seems that the ratio 0.080 between length of winglet and the diameterof rear propeller is suitable for posture control.

8

A Study on Performance Improvement of Boil-off Gas Re-liquefaction Systems
for LNG-fueled Ships
Jun-Seong Kim1 ▪ Do-Yeop Kim 2 ▪ You-Taek Kim†
† Division of Marine System Engineering, Korea Maritime & Ocean University, 727 Taejong-ro, Yeongdo-Gu,
Busan 49112, South Korea, kimyt@kmou.ac.kr
1 Division of Marine System Engineering, Korea Maritime & Ocean University, jskim87@kmou.ac.kr
2 Division of Marine System Engineering, Korea Maritime & Ocean University, doyeop@kmou.ac.kr

Key Words: LNG-fueled ship, Re-liquefaction system, Boil-off gas, LNG-fueled ship, N2 reverse Brayton cycle, LNG cold energy

1. Introduction

Figure 2 is a system that applies LNG cold energy

Recently, regulations on environmental pollution

to the BOG re-liquefaction system based on the N2from ships have been strengthened internationally [1].

reverse Brayton cycle [4]. This system can be expected

LNG-fueled ships can reduce most pollutants, which

to improve thermodynamic performance through LNG

is advantageous in responding to regulations [2].

cold energy.

Meanwhile, LNG-fueled ships require a boil-off gas(BOG) re-liquefaction system for condensing BOG
generated from an LNG fuel tank [3]. In this study,
research was conducted to improve the performance of
the BOG re-liquefaction system for LNG-fueled ships.

2. System analysis
2.1 BOG Re-liquefaction system
Figure 1 is a BOG re-liquefaction system for LNGfueled ships proposed by Kwak et al [3]. This system

Figure 2.

is characterized by applying the N2 reverse Brayton

system for LNG-fueled ship proposed using LNG cold energy

cycle as a re-liquefaction method and using BOG[4]

Schematic diagram of the BOG re-liquefaction

compressors.
In this study, each system was simulated using
Aspen HYSYS v12.1 [5], and the thermodynamic
performance was compared and analyzed.

2.2 Energy and exergy analysis
The thermodynamic performance of the BOG reliquefaction system for LNG-fueled ships using LNG
cold energy (Figure 2) was compared and analyzed
Figure 1.

with the Kwak’s system (Figure 1). As performance

Schematic diagram of the BOG re-liquefaction

indicators of the system, specific energy consumption

system for LNG-fueled ship proposed by Kwak et al. [3](SEC) was used for energy analysis as shown in9Equation 1. In addition, for exergy analysis, exergy(𝜂𝑒𝑥 ) of the system using LNG cold energy was 88.8 %,efficiency ( 𝜂𝑒𝑥 ) expressed as Equation 2 waswhich was better than Kwak's system.

examined [6].
Table 2. Results of the exergy analysis of BOG re-liquefaction
system for LNG-fueled ship

∑ 𝑊̇𝑐
𝑆𝐸𝐶 =
𝑚̇𝐿𝐵𝑂𝐺(1)𝐸̇𝐿𝐵𝑂𝐺 − 𝐸̇𝐵𝑂𝐺
𝜂𝑒𝑥 =
∑ 𝑊̇𝑐

(2)

SystemFigure 1

Figure 2

𝜂𝑒𝑥 [%]

19.6

88.8

4. Conclusion
In this study, the performance improvement of the

2.3 BOG conditions

BOG re-liquefaction system for LNG-fueled ships was

The conditions of BOG in the simulated system arestudied. A system that applied LNG cold energy to areferred to Ref. [3], and can be summarized as shown

BOG re-liquefaction system based on N2 reverse

in Table 1.

Brayton cycle was shown.
The thermodynamic performance of this system was

Table 1. BOG conditions of the BOG re-liquefaction system for
LNG-fueled shipcompared and analyzed with the re-liquefaction

Variables

Unit

Value

Inlet mass flow

kg/h

83.3

Inlet pressure

kPa

160.0

Outlet temperature℃

Saturated liquidcondition

Outlet pressure

kPa

150.0

system proposed in Ref. [3]. As a result, the BOG reliquefaction system using LNG cold energy showed avery significant performance improvement in energy
and exergy analysis.

References[1] Vedachalam, S., Baquerizo, N., and Dalai, A. K.,
2022, "Review on impacts of low sulfur
regulations on marine fuels and complianceoptions," Fuel, Vol. 310, pp. 122243.

3. Results of system analysis
3.1 Energy analysis[2] Son, H., and Kim, J.-K., 2019, "Operability study
on small-scale BOG (boil-off gas) re-liquefaction
processes," Energy, Vol. 185, pp. 1263–1281.

Table 2 shows the energy analysis results for thesystem using LNG cold energy (Figure 2) and Kwak's[3] Kwak, D.-H., Heo, J.-H., Park, S.-H., Seo, S.-J.,
and Kim, J.-K., 2018, "Energy-efficient design andoptimization of boil-off gas (BOG) re-liquefaction
process for liquefied natural gas (LNG)-fuelled
ship," Energy, Vol. 148, pp. 915–929.

system (Figure 1), respectively. The SEC of the systemusing LNG cold energy was 0.3 kWh/kgLBOG, which
was better than Kwak's system.

[4] Kim, J.-S., and Kim, D.-Y., 2023, "Energy, Exergy,
and Economic (3E) Analysis of Boil-Off Gas ReLiquefaction Systems Using LNG Cold Energy for
LNG-Fueled Ships," J. Mar. Sci. Eng., Vol. 11, pp.
587.

Table 2. Results of the energy analysis of BOG re-liquefaction
system for LNG-fueled ship

SystemFigure 1

Figure 2

SEC [kWh/kgLBOG]

1.4

0.3

[5] AspenTech, 2021, HYSYS User’s Guide, Release
V12.1, Bedford, MA, USA.

3.2 Exergy analysis[6] Shen, J., Yan, S., Li, Z., Xiao, Y., and Tan, N.,
2021, "Design and analysis of boil-off gas
reliquefaction processes for the LNG-fueled
ships," Applied Thermal Engineering, Vol. 199, pp.
117505.

Table 3 shows the exergy analysis results for thesystem using LNG cold energy (Figure 2) and Kwak's
system (Figure 1), respectively. The exergy efficiency
10Flame Surfaces Variation through Ionic Wind and Flame Stretch in Mass-Diffusively
Unstable Flame
Siyeong Yu1, 2), Dae Geun Park3), Sung Hwan Yoon1, 2)
Department of Marine System Engineering, Korea Maritime and Ocean University, 727 Taejong-ro, Yeongdo-gu,
Busan, 49112, Republic of Korea1)
Interdisciplinary Major of Maritime AI Convergence, Korea Maritime and Ocean University, 727 Taejong-ro,
Yeongdo-gu, Busan, 49112, Republic of Korea2)
Carbon Neutral Technology R&D Department, Korea Institute of Industrial Technology (KITECH), 89 Yangdaegiro-gil, Seobuk-gu, Cheonan, 31056, Republic of Korea3)

Electrically assisted combustion is a novel approachto combustion technology that utilizes electricity,
contributing to the reduction of soot, increase of flame
propagation speed, control of flame surface area, andflame stabilization1). In particular, the reduction ofsoot can improve incomplete combustion, aligningwith the main interest of the maritime industry, whichis actively focused on emission reduction.
Additionally, while existing combustion systems
regulated output by supplying more fuel to achieve
higher power, our team aims to propose a newly
proposed combustion system using electricallyassisted combustion.
One of the most representative phenomena inelectrically assisted combustion is ionic wind. When aflame is putted in the electric field, charges by chemiionization are accelerated to corresponding electrodeby the Lorentz force. Then, charges diffuse acollisional momentum to neutrons, which account for
90% of production by chemi-ionization, during theiracceleration. Consequently, neutrons are move toelectrodes with charges, and it is called ionic wind 2).
Using a combustor made from an acrylic tube, ourresearch team applied electric fields to downwardpropagating premixed flames in stationary mixture,
resulting in significant changes in flame displacement
velocity and flame surface area. Also, detailed
information regarding flame area variations withdifferent types of fuels, dilution gases, and applied
voltages will be presented through an oral presentation.
References(1) Park, D.G., Chung, S.H., Cha, M.S., “Dynamic
responses of counterflow nonpremixed flames to
AC electric field.” Combust. Flame, 198, pp.240248, 2018.
(2) Park, D.G., Chung, S.H., Cha, M.S.,“Visualization of ionic wind in laminar jet
flames.” Combust. Flame, 18, pp. 246-248,
2017.

11

Research on buoy shape to improve wave energy conversion efficiency
Takuma IMAIZUMI*1, Yasutaka IMAI*2 and Shuichi NAGATA*3*1

Department of Science and Engineering, Saga University*2,*3

Institute of Ocean Energy, Saga University

1. Introduction
We are currently conducting research on the practical application of a floating oscillating water column wave power
generation device called a "back bending duct buoy". This device, proposed by Masuda, consists of a buoyant body,an L-shaped duct, an air chamber, and a turbine/generator, as shown in Figure 1. Wave energy is converted into air
energy that moves in and out of the air chamber to the turbine through the oscillating water column in the duct of the
BBDB, which is agitated by the waves (primary conversion), and then converted into electrical energy by theturbine/generator (secondary conversion).

In order to design floating OWC (Oscillating Water Column)-type wave

energy converter such as Backward Bent Duct Buoy (BBDB), it is necessary to develop a numerical method to makeclear an optimal hull shape which maximizes the generating electrical energy. Therefore, this study proposes a twodimensional analysis method that can evaluate the first-order conversion performance of floating oscillating watercolumn wave power generators such as BBDBs in regular waves. In this study, the boundary element method basedon velocity potentials is used for the analysis of water, while the equation of state, mass conservation law, and energy
conservation law are used for the analysis of air in the air chamber. From these equations, motions of floating body,
air pressure in air chamber, transmission and reflection coefficient of water waves and primary conversion efficiency
are calculated.

Fig.1 BBDB overview

2. Analysis Model
The analytical model considered in this study is shown in Figure 2. The water depth is assumed to vary arbitrarily
from h~h', the floating body is moored by n* linear springs, and a nozzle of diameter D is assumed to be installed in
the ceiling of the air chamber. The origin is on the surface of the still water, the x-axis is horizontal and the z-axis isvertical, and the coordinates of the center of gravity of the floating body at rest are 𝐺(𝑥̅̅̅,
0 𝑧̅0 ) and 𝑄(𝑥𝑏 , 𝑧𝑏 ). Assume
that the center of gravity of the floating body at time t is at (𝑥0 , 𝑧0 ), and that it receives incident waves of angularfrequency σ and amplitude 𝜁0 , and translates in the x- and z-directions by ξ and η, respectively, and rotates clockwiseby an angle θ. If the complex amplitudes of the horizontal displacement, vertical displacement, and rotation angle are
X, Y, and Θ, respectively, the motion of the floating body can be expressed as follows,12𝜁 = 𝑥0 − 𝑥̅̅̅0 = 𝑅𝑒 [𝑋・𝑒 𝑖𝜎𝑡 ]𝜂 = 𝑧0 − 𝑧̅0 = 𝑅𝑒 [𝑍・𝑒 𝑖𝜎𝑡 ]𝜃 = 𝑅𝑒 [Θ・𝑒 𝑖𝜎𝑡 ]where 𝑖 = √−1 and Re denotes the real part.

Fig3 Analysis model
If we delimit the fluid domain by setting virtual boundaries AB, A'B', and FF' as shown in Figure 2, the velocity
potentials of each domain can be expressed as follows.
・Region-1∇2 𝜙 (1) = 0(Inside Fluid)𝜕𝜙 (1) 𝜎 2 ℎ (1)=𝜙𝜕𝜐𝑔𝜕𝜙 (1)=0𝜕𝜐(Free surface: external)(Bottom)𝜕𝜙 (1) 𝜎 2 ℎ 𝑋 𝑑𝑧 𝑍 𝑑𝑥 Θℎ
𝑑𝑥
𝑑𝑧
=
[
−
−
{(𝑥 − ̅̅̅)𝑥0
+ (𝑧 − 𝑧̅0 ) }]𝜕𝜐𝑔 𝜁0 𝑑𝑠 𝜁0 𝑑𝑠 𝜁0𝑑𝑠𝑑𝑠
𝜙 (1) (𝑙, 𝑧) = [𝑒 𝑖𝜆𝑙 + 𝐾𝑟 𝑒 −𝑖𝜆𝑙 ]𝐴(𝜆𝑧){𝜕𝜙 (1)𝜕𝜐(𝑙, 𝑧) = 𝑖𝜆[𝑒 𝑖𝜆𝑙 − 𝐾𝑟 𝑒 −𝑖𝜆𝑙 ]𝐴(𝜆𝑧)𝜙 (1) (−𝑙′, 𝑧) = 𝐾𝑡 𝑒 −𝑖𝜆𝑙′ 𝐴(𝜆′𝑧){𝜕𝜙 (1)(−𝑙′, 𝑧) = −𝑖𝜆𝐾𝑡 𝑒 −𝑖𝜆𝑙′ (𝜆′𝑧)𝜕𝜐𝐴(𝜆𝑧) =𝐴(𝜆′𝑧) =(On Body surface)cosh 𝜆(𝑧 + 1)
cosh 𝜆cosh 𝜆′(𝑧 + ℎ′)cosh 𝜆 ℎ′(Incident wave + Reflected wave)(𝑇𝑟𝑎𝑛𝑠𝑚𝑖𝑠𝑠𝑖𝑜𝑛 𝑤𝑎𝑣𝑒)

where Kr and Kt are complex constants for reflectance and transmittance, respectively.
・Region -2∇2 𝜙 (2) = 0(Inside Fluid)𝜕𝜙 (2) 𝜎 2 ℎ
𝜇={𝑖𝑝𝑠𝑎 − (1 − 𝑖 ) 𝜙 (2) }𝜕𝜐𝑔𝜎(On the free surface: internal)

𝜕𝜙 (1) 𝜎 2 ℎ 𝑋 𝑑𝑧 𝑍 𝑑𝑥 Θℎ
𝑑𝑥
𝑑𝑧
=
[
−
−
{(𝑥 − ̅̅̅)𝑥0
+ (𝑧 − 𝑧̅0 ) }]𝜕𝜐𝑔 𝜁0 𝑑𝑠 𝜁0 𝑑𝑠 𝜁0
𝑑𝑠
𝑑𝑠

13

(On the floating surface)・Connection condition between Region-1 and Region-2
𝜙 (1) = 𝜙 (2) ,𝜕𝜙 (1) 𝜕𝜙 (2)=𝜕𝜐𝜕𝜐

3. Calculation
The calculations are performed by coupling the boundary integral equations for the fluid section, the equations forthe air in the air chamber, and the equations of motion for the floating body.
・Boundary integral equations
The boundary line surrounding the fluid is decomposed into N micro line segments, the center coordinates of eachelement are denoted by (𝜉𝑗 , 𝜂𝑗 ) (𝑗 = 1~𝑁), and the length of the line segment is Δ𝑠𝑗 .
Assuming that the potential function and its normal derivative at each element are constant, the following
relationship between the potential function value 𝜙(𝑗) and its normal derivative 𝜙̅(𝑗) at each point is establishedfrom Green's formula.
𝑁∑[𝐹𝑚𝑗 ∙ 𝜙(𝑗) − 𝐸𝑚𝑗 ∙ 𝜙̅(𝑗)] = 0(m = 1~N)

(1)

𝑗=1

where,
𝐹𝑚𝑗 = −𝛿𝑚𝑗 + 𝐸̅𝑚𝑗
𝐸𝑚𝑗 =(Region − 1),

𝐹𝑚𝑗 = 𝛿𝑚𝑗 + 𝐸̅𝑚𝑗

(Region − 2)

1
1
𝜕
2
2 1⁄2
∫ log 𝑅𝑚𝑗 𝑑𝑠 , 𝐸̅𝑚𝑗 = ∫log 𝑅𝑚𝑗 𝑑𝑠 , 𝑅𝑚𝑗 = [(𝜉𝑗 − 𝜉𝑚 ) + (𝜂𝑗 − 𝜂𝑚 ) ]𝜋 Δ𝑆𝑗
𝜋 Δ𝑆𝑗 𝜕𝜈𝛿𝑚𝑗 is the Kronecker delta.
Substituting the boundary condition equations obtained in Chapter 2 into Equation 2, we obtain the algebraic
equations for Regions -1 and 2.
・Equations for the air in the air chamber
From the equation of state, the law of conservation of mass, and the law of conservation of energy, establish thebasic equation for the motion of air in an air chamber.
𝐓𝐡𝐞 𝐞𝐪𝐮𝐚𝐭𝐢𝐨𝐧 𝐨𝐟 𝐬𝐭𝐚𝐭𝐞：𝑝𝑎 = 𝜌𝑎 𝑅𝑇𝑎(2)

𝐓𝐡𝐞 𝐥𝐚𝐰 𝐨𝐟 𝐜𝐨𝐧𝐬𝐞𝐫𝐯𝐚𝐭𝐢𝐨𝐧 𝐨𝐟 𝐦𝐚𝐬𝐬：𝑑 𝑝𝑎 𝑉𝑎𝑑𝑚𝑎()+=0𝑑𝑡 𝑅𝑇𝑎𝑑𝑡

𝐓𝐡𝐞 𝐥𝐚𝐰 𝐨𝐟 𝐜𝐨𝐧𝐬𝐞𝐫𝐯𝐚𝐭𝐢𝐨𝐧 𝐨𝐟 𝐞𝐧𝐞𝐫𝐠𝐲：𝒑𝒂

(3)

𝑑𝑉𝑎
𝑑 𝑝𝑎 𝑉𝑎𝑑𝑚𝑎+ 𝐶𝑣 () + 𝐶𝑝 𝑇𝑒
=0
𝑑𝑡
𝑑𝑡 𝑅
𝑑𝑡

(4)

where,𝑑𝑚𝑎= ±𝜌𝑒 𝜀𝐶𝑑 𝐶𝑠 𝐴𝑊 √2𝐶𝑝 |𝑇𝑎 − 𝑇0 |𝑑𝑡
𝑝𝑎 ：Air chamber pressure, 𝑇𝑎 ：Chamber temp, 𝑉𝑎 ：Chamber Volume, 𝑡：time, 𝑑𝑚𝑎 ⁄𝑑𝑡：Mass outflow rate of air, 𝑅：Gas
constant, 𝐶𝑣 ：specific heat of formation, 𝐶𝑝 ：specific heat at constant pressure, 𝜀：Nozzle ratio, 𝐶𝑑 ：contraction coefficient,
𝐶𝑠 ：velocity coefficientLinearizing and rearranging Equations 2~4 yields the basic Equation 5.
(2)

𝑁1(1 + 𝑖𝐶𝐸 )𝑙𝐶 𝑎𝜇 (2)Θℎ∑ (1 − 𝑖 ) 𝜙1 (𝑗) ∙ ∆𝑥𝑗 −𝑝𝑠 − 𝑖𝑙𝐶 (𝑥0𝑟 − ̅̅̅)
𝑥0
=0
𝜌
𝐶𝐸
𝜁0
𝑗=1

14

(5)

・The equations of motion for the floating body
The equation of motion of a floating body is formulated separately for surge, heave, and pitch.
𝑀

𝑑2 (𝑋𝑒 𝑖𝜎𝑡 )𝑑2 (𝑍𝑒 𝑖𝜎𝑡 )𝑑2 (𝑍𝑒 𝑖𝜎𝑡 )
(𝑤)
(𝑎)
(𝑤)
(𝑠)
(𝑎)
(𝑤)
(𝑠)
(𝑎)
= 𝑃𝑋 + 𝐹𝑋 + 𝑃𝑋 , 𝑀= 𝑃𝑍 + 𝑃𝑍 + 𝐹𝑍 + 𝑃𝑋 , 𝑀= 𝑃𝑍 + 𝑃𝑍 + 𝐹𝑍 + 𝑃𝑋
2
2
𝑑𝑡
𝑑𝑡
𝑑𝑡 2

where,(𝑤)(𝑤)

𝑃𝑋 , 𝑃𝑍(𝑤), 𝑇𝜃 ：Fluid power by water,(𝑎)

𝐹𝑋 , 𝐹𝑍 , 𝑀𝜃 ：Mooring force,

(𝑎)

(𝑠)

(𝑠)

𝑃𝑍 , 𝑇𝜃 ：Static Righting Moment,(𝑎)

𝑃𝑋 , 𝑃𝑍 , 𝑇𝜃 ：Air pressure acting on air chamber wall

Equation 6 is obtained from the boundary integral equation and the equation of motion of the BBDB.
(2)(1) (𝑎)𝜙1 (𝑗) = 𝐶𝑗 𝑃𝑆
(𝑎)

(1)

+ 𝑑𝑗

(6)

(2)

𝐾𝑡 = 𝐶 (2) 𝑃𝑆

+ 𝑑𝑗
(𝑎)

𝑋⁄𝜁0 = 𝐶 (3) 𝑃𝑆

(3)

+ 𝑑𝑗

𝑒𝑡𝑐.

Substituting Equation 6 into Equation 5 yields Equation 7 for the air chamber pressure amplitude.
(𝑎)

(𝑎)

𝐶1 𝑃𝑆

+ 𝐵2

𝑃𝑆

√|𝑃𝑆(𝑎) |+ 𝐶3 = 0(7)(𝑎)Since 𝐶1 , 𝐵2 , and 𝐶3 are constants in the above equation, 𝑃𝑆

is obtained by iterative calculation. Equation 6 andother equations are then used to obtain the potential on the fluid boundary and its normal direction derivative, thewave transmission coefficient, and the motion of the floating body.

4. Conclusion
Using these methods, we obtained the primary conversion efficiency and motion of the BBDB.

15

Correlation between Ionic Wind and Thermo-Acoustic Instability in DownwardPropagating Premixed Flame under Electric
Dae won Im1), Dae Geun Park3), Sung Hwan Yoon1, 2)
Department of Marine System Engineering, Korea Maritime and Ocean University, 727 Taejong-ro, Yeongdo-gu,
Busan, 49112, Republic of Korea1)
Interdisciplinary Major of Maritime AI Convergence, Korea Maritime and Ocean University, 727 Taejong-ro,
Yeongdo-gu, Busan, 49112, Republic of Korea2)
Carbon Neutral Technology R&D Department, Korea Institute of Industrial Technology (KITECH), 89 Yangdaegiro-gil, Seobuk-gu, Cheonan, 31056, Republic of Korea3)
The recent incidents involving the explosion of
spacecraft developed by NASA and SpaceX have once
again sparked global interest in the stability of rocket
engines. Among the significant challenges in rocket
engine development, thermoacoustic instability stands
out, rooted in the fundamental coupling of rapid heatrelease from combustion and inherent resonances
within the combustion chamber1). This phenomenon,
which can lead to compromised engine lifespan and,in extreme cases, catastrophic failure, necessitates theanalysis of the combustion chamber's inherent natural
frequencies to induce a phase mismatch between
thermal energy and pressure waves, thus ensuring
combustion system stability2).
To address these challenges, various studies havebeen undertaken to mitigate combustion instability
within the combustion chamber, encompassing thedevelopment of injector array patterns, combustion
stabilization mechanisms such as baffles, acoustic
cavities, and Helmholtz resonators3, 4).
This research team aims to propose a novel ap
proach utilizing electric fields to mitigate combus
tion instability. Employing electric fields enables
active control, effectively suppressing thermoacous
tic instability. Furthermore, it facilitates simplifica
tion of internal engine components. Detailed expe
rimental conditions and results are intended to bepresented through the publication.
References(1) J. Rayleigh, “The explanation of certain
acoustical phenomena.” Nature, 18 ,1878, pp.
319-321.
(2) C. Lieuwen, V. Yang, “Combustion instabilities ingas turbine engines.” AIAA, 2005.
(3) D.T Harrje, F.H. Reardon, “Liquid Propellant
Rocket Combustion Instability.” NASA Special
Publication, 1972. pp.194.
(4) D. Zhao and A.S. Morgans, “Tuned passive
control of combustion instability using multiple
Helmholtz resonators.” J. Sound and Vibration,
320, 2009, pp.744-757.

16

Study of Power Generation System Utilizing Waste Heat from Marine Engine(Possibility of introducing HFO Refrigerant)
Atsuyoshi ETSUKAWA*1, Junichi OHARA*1, Hideyuki YOSHIMURA*1, Masaki ARIYOSHI *1,
Yoshihiro SHIMAOKA*2 and Yasuyuki IKEGAMI*3
National Fisheries University, 2-7-1 Nagatahonmachi, Shimonoseki-shi, Yamaguchi 759-6595, Japan*1
National Institute of Technology, Toba College1-1 Ikegami, Toba-shi, Mie 517-8501, Japan*2
Institute of Ocean Thermal Energy, Saga University, 1-Honjomachi, Saga-shi, Saga 840-8502, Japan*3
Abstract
In the present study, the characteristics of cycle thermal efficiency and net power output were clarified for a marine
waste heat power generation system to verify if HFO refrigerant (HFO-1234yf, HFO-1233zd(E)), which is a green
refrigerant, can be a substitute for HFC refrigerant when HFC refrigerant (HFC-134a, HFC-245fa), which will nolonger be used, is employed. When the heat source temperature, flow rate, and heat transfer performance of the heatexchanger are defined, HFO-1233zd(E) and HFO-1234yf show almost the same turbine output values obtained
compared to conventionally used HFC-245fa and HFC-134a. Therefore, HFO-1233zd(E) should be adopted in thispower generation system as an alternative refrigerant to HFC-245fa and HFC-134a, which are conventionally used.
Key Words: Waste Heat Energy, Marine Engine, Turbine Output, Rankine Cycle, Green Working Fluid
1. Introduction
Recent years in particular, the regulation ofgreenhouse gas emissions for the control of global
warming is being tightened for the operation of ships1).
As one of the countermeasures, Exhaust heattemperature difference power generation, which
recovers and effectively utilises the exhaust heat frommarine engines, is attracting attention. There are few
research results on power generation systems using
unregulated green refrigerants. Therefore, this study
clarifies the characteristics of cycle thermal efficiency
and net power output for marine waste heat power
generation systems to verify whether HFO refrigerant
can be a substitute for HFC refrigerant when HFC
refrigerant (HFC-134a, HFC-245fa) and green
refrigerant (HFO-1234yf, HFO-1233zd(E)), which
will no longer be available, are employed.
2. Binary power generation system of Rankine
Cycle using waste heat from marine engine
Figure 1 shows the outline of a binary power
generation system of organic rankine cycle, utilizing
waste heat from marine engine. Binary power
generation generally utilizes steam or hot water at 80
to 150°C as a heat source (primary system) to heat andevaporate a low-boiling working fluid (secondary
system) and utilize the steam to turn the turbine togenerate electricity. Hot water from marine engine isutilized as the high heat source and seawater is utilized

Table 1 Thermophysical properties of working fluids
Boiling
point

ODP

GWP

［℃］

Heat ofvaporization［kJ/kg］

C3H2F4

−29.0

145.2

0

<1

C3H2F3Cl

19.0

194.0

0

<1

CH2FCF3

−26.1

217.0

0

1,430

C3H3F5

15.3

190.9

0

950

Working
fluid

ChemicalformulaHFO-1234yf
HFO-1233zd(E)
HFC-134a
HFC-245fa

Table 2

Calculated conditions

Input Data
Warm water inlet temperature

TWWI

90.0 [℃]

Warm water outlet temperature

TWWO

40.0～80.0 [℃]

Cold water inlet temperature

TCWI

20.0 [℃]

Cold water outlet temperature

TCWO

30.0 [℃]

Flow rate of warm water

mWW

20 [kg/s]

Flow rate of cold watermCW

Calculated through convergence

Heat conductance of evaporator(UA)E

200,400 [kW/K]

Heat conductance of condenser(UA)C

200,400 [kW/K]

as the low heat source. The numbers 1 to 4 in the
diagram indicate the state points between eachcomponent.
3. Relational expression used in cycle calculations
The equations used in the cycle calculations in thisreport are given below. The cycle thermal efficiency
and turbine power are defined by the following
equations.
𝜂𝑡ℎ =𝑊𝑇 − 𝑊𝑃 (ℎ4 − ℎ1 ) − (ℎ3 − ℎ2 )=𝑄𝐸
ℎ4 − ℎ3

𝑊𝑇 = 𝑚𝑊𝐹 (ℎ4 − ℎ1 )

(1)

(2)

𝑊𝑇 indicates turbine output, 𝑊𝑃 indicates pump
output and 𝑄𝐸 indicates evaporator exchange heatrate. ℎ1 , ℎ2 , ℎ3 and ℎ4 are the specific enthalpies at
the turbine outlet, condenser outlet, evaporator inlet
and turbine inlet in Figure 1, respectively, and 𝑚𝑊𝐹 is
the refrigerant flow rate.

Fig. 1 Power Generation System Utilizing Waste Heatfrom Marine Engine
17

4. Calculation Conditions and Methods
Table 1 shows the refrigerants evaluated and thethermodynamic properties of each working fluid
compared. REFPROP2) was used to calculate these
physical properties. The calculation conditions for thecycle calculations in this study are shown in Table 2.
The cold water flow rate is obtained by convergencecalculations in the heat balance on the condenser side.
The cycle thermal efficiency and turbine power are
calculated by changing the heat transfer performanceof the evaporator and condenser to 200 W/K and 400
W/K, respective.

3)

References
1) Yasunobu Araki, "Energy Efficiency Design
Index (EEDI) Regulations and Corresponding
Technology Trends", Journal of the Japan
Institute of Marine Engineering. Vol. 52, No. 4(2017), pp. 480-484.
2) NIST：REFPROP Version 10.0，(2018)．

5.
Calculation Results and Considerations
5・1
Evaluation of Thermophysical Properties
Figure 2 shows the P-h diagram for each refrigerant.
Figure 2 shows that HFO-1233zd(E) and HFC-245fa
have similar thermophysical properties. In addition,
HFO-1234yf
andHFC-134a
have
similarthermophysical properties, although there is adifference in latent heat content. Focusing on adoptinga refrigerant in this power generation system, HFO1233zd(E) and HFC-245fa, which have low pressureand high latent heat, should be used.
Figure 3 shows the T-s diagram for each refrigerant.
Figure 3 shows that HFO-1233zd(E) and HFC-245fa
have similar thermophysical properties, as in Figure 2.
In addition, HFO-1234yf and HFC-134a have similar
thermophysical properties, although there is adifference in entropy. Focusing on adopting a
refrigerant in this power generation system, HFO1233zd(E) and HFC-245fa, which have large entropy
difference, should be used.
5・2
Assessment of turbine power versus
refrigerant flow rate
Figure 4 shows that the turbine power value changes
convexly upward as the refrigerant flow rate changes,
and that the turbine power values obtained with HFO1233zd(E) and HFO-1234yf are almost equivalent tothose obtained with HFC-245fa and HFC-134a, which
are conventionally used. There is an optimum value ofrefrigerant flow rate at which the turbine output value
is the highest, and it is smaller for HFO-1233zd(E),HFC-245fa, HFC-134a, and HFO-1234yf, in thatorder.
The above results indicate that the lower the
refrigerant flow rate, the lower the refrigerant
circulation pump power, which may affect the net
power output of the entire system, depending on thescale of the power generation system. Therefore,
HFO-1233zd(E) and HFO-1234yf can be adopted asalternative refrigerants to HFC refrigerants.

2)

Pressure [MPa]

101

100

10-1
：HFC-134a：HFC-245fa：HFO-1234yf：HFO-1233zd(E)

10-2
100

Temperature [℃]

200
150

200

300
400
Enthalpy [kJ/kg]
Fig. 2 P - h diagram

500：HFC-134a：HFC-245fa：HFO-1234yf：HFO-1233zd(E)

100
50
0
-50
0.5

1
1.5
Entropy [kJ/(kg・K)]
Fig. 3 T - s diagram

2

200
Turbine output [kW]6.
1)1234yf or HFC-134a. The lower refrigerant flowrate results in lower refrigerant circulation pump
power, which may affect the net power output
obtained by the entire system, depending on thescale of the power generation system.
The ratio of refrigerant circulation pump power
to turbine power was small for each refrigerant,
and had no effect on net power output. Therefore,
HFO-1233zd(E) and HFO-1234yf have thepotential to be adopted as alternative refrigerants
to HFC refrigerants.

Conclusions
When the heat source temperature, flow rate, andheat transfer performance of the heat exchanger
are defined, HFO-1233zd(E) and HFO-1234yf
show almost the same turbine output values
obtained compared to conventionally used HFC245fa and HFC-134a.
Similar to HFC-245fa, HFO-1233zd(E) has
operating conditions where the turbine output
peaks at a lower refrigerant flow rate than HFO-

mWW=20[kg/s](UA)E=200[kW/K](UA)C=200[kW/K]
TWWI =90.0[℃]
TCWI=20.0[℃]
TCWO=30.0[℃]

150
100
50 〇：HFC-134a△：HFC-245fa□：HFO-1234yf▽：HFO-1233zd(E)

0

10
20
30
Flow rate of working fuild [kg/s]
Fig. 4 Characteristics of Turbine output forvariable Flow rate of working fluid
18

Combustion field measurement of temperature, concentration and velocity using TDLAS
Jeong Woong Hong1, 2), Min Gyu Jeon 3) Sung Hwan Yoon1, 2)
Department of Marine System Engineering, Korea Maritime and Ocean University, 727 Taejong-ro, Yeongdo-gu,
Busan, 49112, Republic of Korea1)
Interdisciplinary Major of Maritime AI Convergence, Korea Maritime and Ocean University, 727 Taejong-ro,
Yeongdo-gu, Busan, 49112, Republic of Korea2)
Department of Mechanical Engineering, Republic of Korea Naval Academy, Changwon 51704, Republic of Korea 3
Tracking and monitoring greenhouse gases emitted
from industries to comply with environmental
regulations is becoming increasingly important. In
response, research and development of power units that
increase combustion efficiency, and the use of lowcarbon fuels such as LNG, hydrogen, and ammonia arebeing conducted. To accurately diagnose the
combustion state and combustion efficiency, it isnecessary to track the combustion products generated by
the combustion reaction. Temperature, concentration,and speed are important factors to check the combustion
efficiency of the flame.
However, current measurement methods have
limitations. Such as, temperature measurement using
thermocouple and velocity measurement using hot-wireanemometer are contact to measuring area, it is cuased
error by thermal radiation. And non-contact opticaltechniques such as CARS(Coherent anti-Stokes Raman
Spectroscopy) and LIF(Laser Induced Fluorescence) are
expensive and complex optical systems.
In this study, we aimed to confirm the possibility of
non-contact measurement technology in places where
direct measurement is impossible. We used TDLAS(Tunable Diode Laser Absorption Spectroscopy) (1-3), a
wavelength-variable absorption spectroscopy system, tomeasure temperature, concentration, and velocity
simultaneously in combustion field. And we arranged
various types of cells for gas supply flow conditions,temperature, concentration, and velocity measurements,
and compared the measurement results to confirm thepossibility of simultaneous measurement. Details on theexperimental method and results will be intend to
through the presentation.
References(1) Y. Zaatar, J. Bechara, A. Khoury, D. Zaouk and J. P.
Charles, Diode laser sensor for process control andenvironmental mon-itoring, Applied Energy, 65 (1)(2000) 107-113(2) M.G. Jeon, J. W. Hong, K. M. Kim and D. H. Doh,Development of measurement technique of
explosive flame using an optical measurement
method, Journal of Mechanical Science andTechnology 34(21) (2020) 5109-5115(3) M. G. Jeon, D. H. Doh and Y. deguchi, Opticaltemperature measurement method of premixed
flames using a multi-laser system, Journal of
Mechanical Science and Technology 35(6) (2021)
2535-2542

19

Tip clearance gap effect on the hydraulic performance of a mixed flow pumpwith semi-open casing
Ujjwal Shrestha
Department of Mechanical Engineering, Mokpo National University
Muan, 58554, Republic of Korea, ujjwalshrestha10@yahoo.com
Young-Do Choi
Department of Mechanical Engineering, Institute of New and Renewable Energy, Mokpo National University
Muan, 58554, Republic of Korea, ydchoi@mnu.ac.kr

Abstract
Mixed-flow pumps are widely used in agriculture, water conservancy, shipping, power, and other industries dueto the advantages of high flow rate, high efficiency, and cavitation resistance. The tip clearance gap between theblade tip and shroud is inevitable in fluid machinery. The tip clearance gap can induce leakage flow and interact
with the mainstream flow, which causes efficiency drop and cavitation in hydraulic machinery. The pressure
fluctuation is associated with the stable operation of hydraulic machinery, and leakage flow induces the unstableflow pattern and increases the pressure fluctuation amplitude in the hydraulic machinery. The concept ofclearance gap is unique in a mixed flow pump with a semi-open casing. The systematic investigation of thehydraulic and suction performance of a mixed flow pump with a semi-open casing is rarely performed. In thepresent work, a detailed analysis is conducted to evaluate the influence of the tip clearance gap on theperformance curves, pressure fluctuations, and vortex intensity. Figure 1 shows the performance curves of amixed-flow pump with various clearance gaps. Figure 2 shows the leakage flow in the mixed-flow pump. The
reverse flow from the high-pressure region to the low-pressure zone is observed in the clearance gap. The
magnitude of reverse flow increases with an increase in clearance gap width. The clearance gap shows asignificant influence on the hydraulic and suction performance in the mixed flow pump with a semi-open casing.
The leakage flow is dependent on the width of the clearance gap between the blade tip and the shroud wall. The
efficiency and head of the mixed flow pump decreased significantly with an increase in the clearance gap.

Figure 1 Effect of clearance gap on semi-open mixed flow pump performance curves

Figure 2 Velocity vectors on the leakage gap
20

A study of fuel cell-battery propulsion system for four types of vessels
Jong In Kim 1 ㆍ San Kim 2 ㆍ Eun Shin Bang 3 ㆍ Kyoung Kuk Yoon 4 ㆍ Sang Kyun Park †The shipping industry accounts for a large proportion of global carbon emissions, and various methods are being tried to reduce theseemissions. One of the methods is to replace the existing internal combustion engine, which is the power source of the ship, with a fuel cell.
Among many fuel cells, Polymer Electrolyte Membrane Fuel Cell (PEMFC) produces power using a chemical reaction between hydrogenand oxygen, does not cause environmental pollution, and is used for small cars and public transportation on land because of its high efficiency
and difficult operating conditions. Considering this usability, it would not be a long way to apply PEMFC to ships.
In theory, PEMFC has a high efficiency of about 60%, and it is easy to maintain/manage at an operating temperature of 70 to 80℃. In
addition, it has the advantage of being able to respond quickly to the rapidly changing operating load of the ship due to its high reactivity.
There have been cases in which small cruise ships with PEMFC have been operated overseas, but in Korea, ships using fuel cells as powersources have not yet been operated. In order for ships powered by fuel cells to be generalized and built in the future, related research needs tobe sufficiently preceded.
In this study, four types of ships were identified, with PEMFC as the main power source and Li-ion batteries as auxiliary power sources
for mutual supplementation in order to establish a ship system to which fuel cells are applied. The required output size and hull size were
calculated according to the size and characteristics of each ship, the sizes of PEMFC and Li-ion batteries were designated accordingly, and apower system including a Balance of plant (BOP) was proposed. Fuel cells and batteries mainly play the role of base load and peak shaving,and propulsion motors and BOPs were configured depending on the size of the ship's power. The fuel cell was modularized with 50kW
PEMFC and arranged according to the size of the ship, and redundancy of the power system was applied.

Fig.1 Piping and electric diagram of Leisure vessel

Fig.2 Electric propulsion system of Leisure vessel† Corresponding Author: Professor, Division of Maritime AI & Cyber Security, Korea Maritime & Ocean University, E-mail: skpark@kmou.ac.kr
1 Researcher, Division of Marine Information Technology, Korea Maritime & Ocean University, E-mail: darkkrow@kmou.ac.kr
2 Researcher, Division of Marine Information Technology, Korea Maritime & Ocean University, E-mail: ks9524@g.kmou.ac.kr
3 Researcher, Korea Maritime & Ocean University, E-mail: esbang627@ kmou.ac.kr
4 Professor, Division of Maritime AI & Cyber Security, Korea Maritime & Ocean University, E-mail: kkyoon@kmou.ac.kr

21

POSTERPRESENTATION

Characteristics of HFO Refrigerants in Double-Stage Reverse Rankine Cycle Refrigerator
Tetsuya HONDA1), Takafumi KOJIMA1) Junichi OHARA1), Hideyuki YOSHIMURA1),
Masaki ARIYOSHI1), Takafumi MORISAKI2) and Yasuyuki IKEGAMI2)
National Fisheries University, 2-7-1 Nagatahonmachi, Shimonoseki-shi, Yamaguchi 759-6595, Japan 1)
Institute of Ocean Thermal Energy, Saga University, 1-Honjomachi, Saga-shi, Saga 840-8502, Japan2)

Abstract
The Kigali Amendment to the Montreal Protocol (effective 2019) regulates HFC refrigerants and other CFC substitutes
with high global warming potential, and there is a need to convert to green refrigerants with low global warming
potential. Therefore, in this study, cycle simulations of a double-stage reverse Rankine cycle using HFO1233zd(E),
HFO1234yf, and HFO1234ze(E), Ammonia, and HFC32 were conducted to evaluate refrigerant performance undercertain heat source conditions. As a result, in descending order of 𝐶𝑂𝑃, they were HFO1233zd(E), Ammonia, HFC32,
HFO1234ze(E), and HFO1234yf, and HFO1233zd(E) could be operated most efficiently.
Key Words: Reverse Rankine Cycle, Refrigerator, Double-Stage, HFOs, 𝐶𝑂𝑃
1. Introduction
The Kigali Amendment led to the international
regulation of CFC substitutes with high global warming
potential (GWP). Therefore, it is necessary to convert torefrigerants with low GWP and to research refrigeration
and air-conditioning equipment using low GWP
refrigerants1)．
Energy-saving and high-efficiency systems forrefrigeration cycles have been proposed, and this study
focuses on the multi-stage reverse Rankine cycle. A
multistage reverse Rankine cycle is configured with two
or more reverse Rankine cycles working between one
low-temperature and one high-temperature heat source.
Saga University group researches on this multistage
reverse Rankine cycle and clarify these performances
better than other typical refrigeration cycles 2)．
The system characteristics and refrigerant properties of
the new HFOs-as green refrigerants and naturalrefrigerants have not yet been fully clarified.
This study investigates the refrigerant performance ofthe double-Stage reverse Rankine cycle using cycle
simulations applied under certain heat source conditions
for refrigerants: HFO1233zd(E), HFO1234yf, andHFO1234ze(E), Ammonia, HFC32.

compressor.

Fig.1 T-s diagram

Fig.2 Double-stage Reverse Rankine cycle

2. Double-Stage reverse Rankine cycle
Figure 1 shows a conceptual T-s diagram of a Doublestage Reverse Rankine cycle．To increase 𝐶𝑂𝑃 in abasic reverse Rankine system, it is important to keep
compression power lower relative to heat absorption anddissipation. For the attainment of this purpose, it isnecessary to increase the heat transfer performance ofthe heat exchanger and reduce the temperature
difference between the heat source and the working fluid.
As shown in Figure 1, the temperature difference
between the heat source and the working fluid during theheat exchange process is reduced by using a doublestage inverted Rankine cycle, which is expected to
improve the system performance.
Figure 2 shows a system schematic of the double-stage
Reverse Rankine cycle. Two reverse Rankine cycles are
installed in series for the high and low temperature heatsources, each cycle has its own evaporator, condenser,

3. Calculation Formula
The equations used in this study are shown ． The
overall coefficient of performance 𝐶𝑂𝑃 andcompression power 𝑊𝑃 are defined by (1) and (2).
𝐶𝑂𝑃 =(𝑚𝑊𝐹1 × (ℎ1 − ℎ4 ) + 𝑚𝑊𝐹2 × (ℎ5 − ℎ8 ))(𝑚𝑊𝐹1 × (ℎ2 − ℎ1 ) + 𝑚𝑊𝐹2 × (ℎ6 − ℎ5 ))𝑊𝑃 = 𝑊𝑃1 + 𝑊𝑃2 = (𝑚𝑊𝐹1 × (ℎ2 − ℎ1 )+ 𝑚𝑊𝐹2 × (ℎ6 − ℎ5 )

(1)

(2)

𝐶𝑂𝑃 represents the performance factor, 𝑚𝑊𝐹1 and𝑚𝑊𝐹2 represent the working fluid flow rate for eachcycle. ℎ1 ，ℎ2, and ℎ4 represent the enthalpies at eachpoint in the cycle No.1, and ℎ5 ，ℎ6 , and ℎ8 represent
the enthalpies at each point in the cycle No.2 ; 𝑊𝑃
represents the overall compression power; 𝑊𝑃1 and
𝑊𝑃2 are the compression powers for each cycle.

22

40

ammonia
HFC32
HFO1234zeE
HFO1234yf
HFO1233zdE

Table 1 Test medium
Classification

Toxicity・
flammability

HFC32

Alternative freon

A1

Ammonia

Natural refrigerant

B2L

HFO1233zd(E)Green refrigerant

A1

HFO1234yf

Green refrigerant

A1

HFO1234ze(E)Green refrigerant

A2L

30

mWF [t/h]

Name

20

10

0

0.2

Table 2 Calculated conditions

35.0[℃]

101.325[kPa]

Thermal conductance of Evaporator（1.8～100）×10４[W/K]

Thermal conductance of Condenser（1.8～100）×10４[W/K]

Refrigeration Capacity

ammonia
HFC32
HFO1234zeE
HFO1234yf
HFO1233zdE

170

25.0[℃]

Pressure of High temperature fluids

1[106]

180-30.0[℃]101.325[kPa]

0.8

Fig.3 Working fluid flow rate

0.0[℃]

Pressure of Low temperature fluids

0.6

Wp [kW]

Inlet temperature of Low
temperature fluids
Outlet temperature of Low
temperature fluids
Inlet temperature of Hightemperature fluids
Outlet temperature of Hightemperature fluids

0.4

UA(E) [W/k]

160

150

140

130
0

0.2

500[kW]

0.4

0.6

UA(E) [W/K]0.81[106]

Fig.4 Compression power
4. Calculation conditions
In this study, air and water were used as low and hightemperature heat source, respectively．Table 1 shows
the classification of the five refrigerants used in thesimulation in terms of composition and the index of
toxicity and flammability of the refrigerants defined by
ASHRAE．Table 2 shows the simulation conditions:
inlet and outlet temperatures of the low and hightemperature fluids, thermal conductance andrefrigeration capacity of the condenser and evaporator.

3.8

COP [-]

3.6

3.4

3.2

ammonia
HFC32
HFO1234zeE
HFO1234yf
HFO1233zdE

3

2.8
0

5. Calculation Results and Considerations
Figure 3 shows the overall mass flow rate 𝑚𝑊𝐹 withsolid color-coded lines for each refrigerant．In an ove
rall charcteristics, it can be seen that the mass flow rateis decreasing with increase of 𝑈𝐴(𝐸) in the order ofAmmonia, HFC32, and the three HFO refrigerants. In
other words, the flow rate required for heat exchange isconsidered to decrease as the heat dissipation andrefrigerating effect increases.
Figure 4 shows the overall compression power 𝑊𝑃
with solid color-coded lines for each refrigerant. Whenthe value of HFO1233zd(E), which has the lowest
compression power among all refrigerants, is set to
100%, it can be confirmed that the value of each 𝑊𝑃 is
increased by 1% for Ammonia, by 7% for both HFC32
and HFO1234ze(E), and by 13% for HFO1234yf.
Figure 5 shows the overall coefficient of performance(𝐶𝑂𝑃) as a solid color-coded line for each refrigerant.
If the value of HFO1233zd(E), which has the highest
𝐶𝑂𝑃 among all refrigerants, is set to 100%, the 𝐶𝑂𝑃
values are respectively Ammonia decreased by 2%, both
HFO1234ze(E) and HFC32 decreased by 7%, andHFO1234yf decreased by 12%, respectively. It can also
be confirmed that the ratio difference and order of

0.2

0.4

0.6

UA(E) [W/K]0.81[106]

Fig.5 Coefficient of performance
performance is almost the same as that of thecompression power 𝑊𝑃 .
6. Conclusion
The performance of five refrigerants; three green
refrigerants HFOs (HFO1233zd(E), HFO1234yf, andHFO1234ze(E)), Ammonia as a natural refrigerant, and
HFC32 as a conventional CFC substitute, were
investigated under certain heat source conditions using
a double-stage reverse Rankine cycle. HFO1233zd(E)could be operated more efficiently than the other four
refrigerants under the present simulation conditions,while in the case of HFO1234yf, it shows lower
efficiency overall.
References(1) Mitsuyo Uno, Approaches to next-generation
refrigerant compatibility, Marine Engineering 53(6), 793-799, 2018-11-01, (2018)
Y. Ikegami, T. Morisaki, Study on heat pump andrefrigeration cycle using multi-stage inverted
Rankine cycle, Academic paper, Saga Univ. (2013)(2)23

Oceanographic Survey for Installation of OTEC Plant in Kumejima Island（Influence from the Aguni Basin）
Shodai NAGATA1), Junichi OHARA1), Hideyuki YOSHIMURA1),
Masaki ARIYOSHI1) and Yasuyuki IKEGAMI2)
National Fisheries University, 2-7-1 Nagatahonmachi, Shimonoseki-shi, Yamaguchi 759-6595, Japan 1)
Institute of Ocean Thermal Energy, Saga University, 1-Honjomachi, Saga-shi, Saga 840-8502, Japan2)

Abstract
In this study, oceanographic surveys were conducted in the sea area of Kumejima Island, Okinawa Prefecture, whichis the best candidate in Japan for the installation of Ocean Thermal Energy Conversion (OTEC), and the surrounding
Aguni Basin, and an evaluation was made to clarify whether the seawater in the Kumejima Island sea area is derived
from the Aguni Basin. Seawater temperature, salinity, and dissolved oxygen relative to water depth were comparable
between the Kumejima Sea area and the Aguni Basin, indicating that the seawater in the Kumejima Sea area is likely
to be deep ocean water originating from the Aguni Basin.
Key Words: Oceanographic Survey, Kumejima , Aguni Basin, OTEC, Seawater Temperature
1. Introduction

said to satisfy these conditions.

In October 2020, the Japanese government sat the goal
which Japan become carbon neutral country by 2050,
and the spread of renewable energy is necessary toachieve this goal. Therefore, this study focuses on ocean
thermal energy conversion (OTEC) as a renewableenergy source. 70% of the Earth is composed of ocean,and Japan is an island country surrounded by ocean,making it easy to make effective use of ocean energy.
OTEC is used not only for power generation, but also
for various secondary uses such as air-conditioning,aquaculture, and agriculture1).
In this study, oceanographic surveys were conducted in
the waters of Kumejima, Okinawa Prefecture, the mostpromising candidate in Japan for the installation of
OTEC, and the surrounding Aguni Basin, and anassessment was made to clarify whether the seawater inKumejima sea area is derived from the Aguni Basin ornot.

3. Period for investigation, survey item, method ofwater sampling and investigation site
Figure 1 shows the oceanographic survey area. A totalof seven sites were surveyed: four in the Kumejima sea
area and three in the Aguni Basin.
The electrical conductivity of seawater was measured
using a conductivity temperature depth profiler (CTD)aboard the Koyo Maru to investigate water depth,
seawater temperature, dissolved oxygen, and salinity.
The survey was conducted in August 2021.
4. Result and Discussion
4.1 Seawater temperature
Figure 2 shows the results of the vertical distribution ofseawater temperature. The results for the Kumejima sea
area are shown by the red line, and those for the Aguni
Basin are shown by the blue line. Figure 2 shows that
there is no difference between the Kumejima sea area
and the Aguni Basin, with a rapid decrease from thesurface to a depth of around 600 m, followed by agradual decrease thereafter. The temperature of theseawater at the depth of 900m or deeper is 4 to 5°C.
Therefore, the Kumejima sea area and the Aguni Basinare comparable in terms of seawater temperature,

2. Ocean Basin
An ocean basin is a flat basin at the bottom of the deepsea, where the seawater temperature is extremely lower
as one moves toward the bottom. Deep-sea basins arecharacterized by the lack of light, the absence ofphotosynthesis by phytoplankton and other organisms,
and the lack of currents, which results in stable levels ofnutrients and dissolved oxygen. In addition, when theseawater flows from one ocean basin to other, fixedflows are easily formed due to the characteristics of the
topography. If we assume that the seawater which flowfrom the Aguni Basin to the Kumejima sea area is also
fixed in this study, there is a high possibility that the two
types of seawater mix and form a unique seawater withsimilar characteristics2).
In order to establish OTEC, it is desirable to have astable temperature difference of 20°C or more between
surface seawater and deep seawater, as well as stabledeep seawater with salinity for secondary use of deep
seawater, and the endemic seawater in the basin can be

Fig.1 Investigation site
24and it is highly likely that the seawater in the Kumejima
sea area is fed by stable, low-temperature deep seawateroriginating from the Aguni Basin. The temperature
difference between surface seawater and deep seawater,which is the most important factor in OTEC, can bestably maintained.

00

Seawater temperature (℃)
5 10 15 20 25 30 35

Depth (m)500

4.2 Salinity
Figure 3 shows the results of the vertical distribution ofsalinity. Salinity is one of the secondary uses of deep
seawater used for OTEC and also for aquaculture. The
results of the survey in the Kumejima sea area are shownby the red line, and the Aguni Basin is shown by the blueline. Figure 3 shows that there is no difference between
the Kumejima sea area and the Aguni Basin, with a rapid
increase from the surface to a depth of about 200 m,
followed by a rapid decrease to a depth of about 700 m.
The salinity increases slowly at depths deeper than 700
m, and then remains stable.
Therefore, as with seawater temperature, theKumejima sea area and the Aguni Basin are comparable
in terms of salinity, indicating that the seawater in the
Kumejima sea area is likely to be fed by deep seawaterwith stable salinity originating from the Aguni Basin,
and that stable secondary use of deep seawater ispossible.

1000
1500

Kumejima 2021/08
St.1～4
Aguni kaibon 2021/08
St. 5～7

2000

Fig.2 Water temperature vertical distribution

34
0

Salinity (PSU)
34.5

35

Depth (m)500

4.3 Dissolved oxygen
Figure 4 shows the results of the vertical distribution ofdissolved oxygen content. Dissolved oxygen is an
important parameter for aquaculture. The results of thesurvey in the Kumejima sea area are shown by the redline, and the Aguni Basin is shown by the blue line.
Figure 3 shows that there is almost no difference
between the Kumejima sea area and the Aguni Basin,with a gradual decrease from the surface to a depth ofaround 700 m, followed by a stable transition.
Therefore, the Kumejima sea area and the Aguni Basinare comparable in terms of dissolved oxygen as well as
seawater temperature and salinity, and it is highly likelythat the seawater in the Kumejima sea area is fed by deep
seawater with dissolved oxygen from the Aguni Basin.

1000
1500
2000

kumejima 2021/8
St.1～4
Aguni kaibon 2021/08
St. 5～7

Fig.3 Salinity vertical distribution

5. Conclusions
4) The seawater temperatures in the Kumejima sea
area and the Aguni Basin show similar verticalprofiles, indicating that the Kumejima sea area islikely fed by cold and stable deep seawateroriginating from the Aguni Basin.
5) The salinity distribution in the Kumejima sea area
and the Aguni Basin is comparable, indicating thatthe seawater in the Kumejima sea area is likely to
be deep seawater with a stable salinity originatingfrom the Aguni Basin.
6) The dissolved oxygen in Kumejima sea area and theAguni Basin show similar vertical profiles,
indicating that the seawater in Kumejima sea area islikely to be deep seawater with dissolved oxygenfrom the Aguni Basin.

Fig.4 Dissolved Oxygen vertical distribution
References(1) Yukio NAKAMURA(2017)：Kumejima Island

Model Challenging the Cascaded Use of Deep
Seawater, Deep Ocean Water Research, Vol.18
No.3 ,210-211(2) Koshiro
TAYA
etal.(2022)：Horizontal
distribution of the benthic front formed at theboundary between Japan Basin Bottom Water and
Yamato Basin Bottom Water in the Japan Sea andmodification processes due to mixing of both
Bottom Waters, Oceanography in Japan, Vol.31
No.6 ,99-110
25

・Title「Research on Periodic Characteristics by Visualization of Plate Type Evaporator Interior」

・Authors
Ayano Nakagawa (Saga Univ.)

Ryosuke Nikushi (Saga Univ.)

Takeshi Yasunaga (Saga Univ.)

Takafumi Morisaki (Saga Univ.)

Yasuyuki Ikegami (Saga Univ.)
・Abstract
Plate-type evaporators have been applied to low-source temperature difference power generation becauseof their high heat transfer performance and compact size. The working fluid in a plate evaporator is acomplex gas-liquid two-phase flow with evaporation phenomena, and it is important to understand theinternal flow phenomena. However, the unsteady-state void fraction measurement in two-phase flow has
not been sufficiently investigated for the observation time. In this research, a herringbone plate evaporator
made of transparent resin is used to visualize the evaporation phenomena inside the evaporator to measure
the void fraction and to clarify the observation time suitable for the flow conditions.
・Structure of the paper
1. Background
2. Experimental apparatus and method
3. Image processing method
4. Experimental Results and Discussion
5. Conclusion

26

Development of an Optimization Method for Offshore Jacket Foundation Hosting a Wind
Turbine under Static Loads
Takahiro NAGATSU
Institute of Ocean Thermal Energy, Saga University, 1-Honjo machi, Saga shi, Saga 840-8502, Japan
E-mail :23728017@edu.cc.saga-u.ac.jp
Abstract
Wind energy is abundantly available in nature, but due to its low energy density, the size of wind turbines needs tobecome larger and require vast sites where a large number of wind turbines can be installed in clusters. Therefore,
there are high hopes for offshore wind power, where wind resources are abundant and vast sites are available. Offshore
wind turbines have so far been developed in Europe and China, where low-cost monopile foundations have been used
for the most part. However, as wind turbines become larger and the water depth of the installation area increases,
jacket foundations will become increasingly important, as they are easier to ensure strength and rigidity and havefewer restrictions on construction equipment. In particular, the cost of foundations accounts for a large proportion of
the cost of offshore wind turbines, so rationalizing the design of jacket foundations is an important issue for reducing
the cost of power generation from offshore wind turbines. A jacket foundation basically consists of main structure(four legs) forming the periphery of the platform, and braces connected between the legs.
In this study, an optimization software “JacketDes.m” is developed in Matlab to optimize for the lightest weight ofthe entire jacket foundation under static wind loads. Overall, five parameters are considered in this optimization study,
the diameters and thicknesses of both the main structure and braces of the jacket foundation, and the base width of thefour legs. The optimization software “JacketDes.m” controls the calculations of the freeware structural analysis
software Frame3DD and uses the results to obtain a lightweight jacket foundation with sufficient resistance tocompressive/tensile stresses and buckling. As there are only five parameters in this study, the optimal values are
obtained by round-robin method within the pre-defined optimization range for each parameter.
As a demonstration of the optimization method, a case study was conducted, to design the foundation of a 15 MW
wind turbine. A four-legged jacket with a deck width 20 m and height 45 m off the coast of northern Kyushu(installation site water depth 20 m) was selected and optimized. As a result, jacket foundations with a base width of
26 m, main structure and brace diameters of 1.2 m and 0.2 m, and respective plate thicknesses of 32 mm and 20 mm
were obtained, as the the lightest weighing jacket foundation (1128 t).
The future work includes extending the optimization method to include hydrodynamic forces of waves and currents,
seismic loads, terminal and fatigue loads during power generation and welding effects, which are not considered inthis study. In the optimization, the development of optimization algorithms for an increased number of parameters,such as higher resolution, individual diameter and thickness settings for each member, topology optimization, etc.,
should also be considered.
terms of compressive/tensile stresses and buckling,
and the optimization parameters are the diameters andthicknesses of the main structure and braces, and thebase widths. As there are five parameters listed above,
the optimum value is obtained by round-robin method
within the pre-defined optimization range for eachparameter.
The validity of the analysis was verified by the
response under known loads.
The ratio of the stress to the allowable values ofcompressive stress and buckling stress (0.605Et/r; E:
Young's modulus, t: plate thickness, r: radius [5]) for
each member was taken as effort (ηs, ηb). If these
efforts exceed 0.5, the member can be considered tohave failed. However, in reality, more loads areapplied and there are many conditions that are notassumed in this study. Therefore, in this study, thelightest geometry among those whose effort does notexceed 0.5 is selected as the optimum value.

1. INTRODUCTION
This study develops an optimization method for jacket
foundations (Fig. 1), which are promising as
foundations for very large or deep-water offshore windturbines.

Fig 1 Example of Jacket Foundation [3]2. DESIGN METHOD OVERVIEW

3. DESIGNIn this study, the freeware for 3D frame analysis
Frame3DD [1], the optimization software
JacketDes.m (newly developed in Matlab) were used
to perform the analysis and optimization of the jacket
foundations. Here, the bearing capacity is evaluated in

3.1 DESIGN CONDITION
Optimization of a four-leg jacket (deck width 20 m)
foundation consisting of SS400 (strength: 400 N/mm2[4]) cylinders was carried out for an IEA Wind 15 MW
Reference Wind Turbine [1] to be installed offshore in27

northern Kyushu (base wind speed 34 m/s, roughness
category I, water depth 20 m). The design load (at the
top of the jacket) was calculated as follows. The
design loads (at the top of the jacket) were 16
combinations of 6.1 MN shear force, 634 MNmbending moment, ±39 MNm torque and 18.4 MN
axial force from 8 directions.
3.2 DESIGN RESULTThe distribution of effort and mass for a base width of
26 m according to the above conditions and design
method is shown in Fig. 2. The optimum shape for thisbase width is the lightest one with an effort notexceeding 0.5. A similar analysis was carried out forbase widths between 20 and 30 m (2 m each) and theresults are shown in Fig 3. This determined thefollowing optimum (lightest weight) shapes. The
values in brackets are the resolution of each variable.
・Base width：26 m（2 m）
・Main structure diameter ：1200 mm（200 mm）
Thickness：32 mm（2mm）・Brace diameter ：200 mm（100 mm）
Thickness：20 mm（5 mm）
・Mass：1128 tFig 3 Mass by Base Width

4. CONCLUSION
1) An optimization (lightest weight) method for thejacket foundation structure of a land-based offshore
wind turbine has been developed.
2) The same design method was applied to a 15 MW
Reference Wind Turbine for the IEA Wind Turbine to
be installed at a water depth of 20 m off the coast ofnorthern Kyushu, with a base width of 26 m and an
optimum geometry of 1128 t. The wind turbine wasdesigned to be installed in the same location as the IEAWind Turbine.
3) Loads such as during waves and currents,
earthquakes and power generation, consideration of
fatigue and optimization algorithms are future issues.

Fig 4 Appearance of The Optimum Jacket
Foundation

References[1] Frame3DD， http://frame3dd.sourceforge.net/[2] Gaertner, E, et al, Definition of the IEA 15Megawatt
Offshore Reference Wind, IEA Wind, Technical
Report,2020.
[3] NEDO，着床式洋上風力発電導入ガイドブッ
ク（最終版）2018.
[4] JFE ，
材 料 の 許 容 応 力 度 ，binran_chapter04.pdf(jfe-steel.co.jp)，Accessed 2023/02/13.
[5] Wind Energy Handbook

Fig 2 Base Width 22 m Jacket Effort and Mass

28

Heat transfer characteristics of FCU and plate heat exchanger
in seawater air conditioning system
Haruki Ida*, Hirofumi ARIMA***

Graduate school of Sci. Eng., Saga Univ., 1-Honjo, Saga-shi, Saga 840-8502, Japan
Email: 23728001@edu.cc.saga-u.ac.jp**
Institute of Ocean Energy, Saga Univ., 1-48, Hirao, Kubara-aza, Yamashiro-cho, Imari-shi, Saga, 849-4256, Japan

Abstract
In this study, the air conditioning using seawater was investigated, focusing on the cascade utilization of deep-sea water. The aimof this research that a developing a heat exchanger has optimized performance of a cooling system. Then, the heat transfer
performance of the heat exchanger and the seawater cooling system were evaluated. In this experiment, a plate heat exchanger withaluminum plate was employed. In order to evaluate the heat transfer performance at high flow rate conditions, the number of plates
was reduced against previous study. In order to evaluate the heat transfer performance of the heat exchanger, a water-water
experiment was performed with hot and cold water. Next, in order to evaluate the heat transfer performance of the seawater coolingsystem with FCU (Fan Coil Unit). In the FCU experiment, cold water and tap water flowed through the heat exchanger and the tap
water cooled by the cold water. The cooled water flowed into FCU and heat exchange the cooled water and hot air. The result of
these experiments was as follows. As the water-water experiments, the smaller the number of plates in the heat exchanger indicatesthe larger heat transfer coefficient. As the FCU experiments, it was found that the heat transfer coefficient of the FCU increases
with an increase in the mass flow rate of tap water and the flow rate of the air. It was also found that the cooling performance ofthe cooling system improved at increasing the heat transfer rate of the heat exchanger.
1. Introduction
In this study, the application to the cooling system using deepsea water was paid attention. This cooling system is thatindirectly transfers cold heat from deep ocean water using aheat exchanger and an FCU. It is one of the sea water air
conditioning system (SWAC). In this study, the heat transfer
performance of the heat exchanger and the heat transfer
performance of the seawater cooling system were evaluated.
In this presentation, the result of heat transfer characteristics
at the higher water flow velocity was reported.

,

,

,
,

,

,

2. Experiment
To evaluate the heat transfer characteristics of the PHE, awater-water experiment and the heat transfer performance ofthe entire cooling system using the FCU and heat exchanger
were performed.
Fig. 2.1 and 2.2 show the schematic diagrams of the waterwater experiment apparatus and the FCU experimentapparatus. The water-water experimental apparatus consists ofa PHE, a refrigerator, and hot water tank. CU experimentapparatus consists of PHE, a refrigerator, tap water tank, andFCU (Fan Coil Unit). Fig. 2.3 shows that the FCU unit. The
equipment of FCU experiment apparatus except of therefrigerator were installed in a storeroom. The inlet
temperature of air at the FCU was controlled by an air
conditioner.
The material of the PHE test plate is pure aluminum A1050,
and the chevron angle is 60°. The experiments were performed
using PHE with 5 and 13 plates, in order to compare the heat
result of previous research (1) which plate heat exchanger was
installed 21 plates.
,

Fig. 2.2 Schematic diagram of FCU experimental apparatus.

Fig. 2.3 FCU unit
At the water-water experiment, while counter flowing thecold and hot water in the PHE, the investigation the change inheat transfer characteristics due to the difference flow velocity
and number of plates were performed.
The experimental conditions were 10°C for cold water and
30°C for hot water.
In the FCU experiment, the hot air and cold tap water washeat exchanged in the FCU., Also, the heat obtained by the
FCU is heat exchanged with cold water in the PHE. The air
inlet temperature of the FCU was 30°C. Cold water was o 5°C.
The flow rate of cold water, tap water, and FCU air were given
as experimental conditions. The temperature and flow rate at
the inlet and outlet of PHE and FCU were measured.

,

,

,

3. Data reduction
In the water-water experiment and the FCU experiment, theheat transfer rate was calculated by the following equation.
Fig. 2.1 Schematic diagram of water-water experimental
apparatus.

𝑈 = 𝑄/(𝐴𝑠 Δ𝑇𝑙𝑚 )

29

(1)

flow velocity is high. Also, in Fig. 4.3, it was found that thecooling performance improves with an increases in𝑈𝑡𝑤,𝑃𝐻𝐸 .

(2)

𝑄 = 𝑚𝐶𝑝 (𝑇𝑖𝑛 − 𝑇𝑜𝑢𝑡 )Δ𝑇𝑙𝑚 =

UFCU,tw [W/m2K]

𝑄 is the heat transfer rate of the heat exchanger, Δ𝑇𝑙𝑚 is thelogarithmic mean temperature difference, and 𝐴𝑠 is the heattransfer area. Each value was obtained by the following
equations (2) and (3).

(𝑇ℎ,𝑖𝑛 − 𝑇𝑐,𝑜𝑢𝑡 ) − (𝑇ℎ,𝑜𝑢𝑡 − 𝑇𝑐,𝑖𝑛 )ln{(𝑇ℎ,𝑖𝑛 − 𝑇𝑐,𝑜𝑢𝑡 )/(𝑇ℎ,𝑜𝑢𝑡 − 𝑇𝑐,𝑖𝑛 )}

(3)

15
10
5

FCU-H
FCU-M
FCU-L

0
0

where 𝑚 [kg/s] is the mass flow rate of cold water,
𝐶𝑝 [J/(kg ∙ K)] is the specific heat of cold water at constantpressure, 𝑇ℎ,𝑖𝑛 , 𝑇ℎ,𝑜𝑢𝑡 is the inlet and outlet temperature of
the heat exchanger at high temperature side , 𝑇𝑐,𝑖𝑛 , 𝑇𝑐,𝑜𝑢𝑡 are
the inlet and outlet temperatures at the cold temperature side.
The heat transfer coefficient ℎ of the heat exchanger was
defined by using the Wilson plot method.

0.1

0.2

0.3

Gtw [kg/s]

UFCU,tw [W/m2K](a) 5 plates

4. Experimental result
4.1 Water-water experimentFig. 4.1 shows the correlation between 𝑁𝑢ℎ and 𝑅𝑒ℎobtained from the heat transfer coefficient on the hot water sideat the different number of plates heat exchanger. Table 4.1
shows the Nusselt number at different number of plates. As
shown in Fig. 4.1, heat exchanger has smaller the number ofplates indicates high 𝑁𝑢ℎ .
The cause is considered that the influence of the flow
distribution in each channel. As the number of plates increases,
the water flow velocity in the channel which is furthest fromthe entrance becomes slower than near the entrance due to thepressure loss.

15
10
5
0
0

FCU-H
FCU-M
FCU-L

0.1

0.2

0.3

Gtw [kg/s](b) 13 plates
Fig. 4.2 Correlation between mass flow rate 𝐺𝑡𝑤 and overall
heat transfer coefficient 𝑈𝐹𝐶𝑈,𝑡𝑤 .

[-]

102

Nuh/Prh

1/3

Number ofplates [-]

101

5
13
21

100 2
10

103

Fig. 4.3 Cooling performance against heat transfer rate ofPHE.

Equation drived by
Nishida et al
Equation drived by
Nakaoka et al

5. Conclusion
At the plate heat exchangers with different numbers of plates,
the heat transfer performance of the heat exchanger was
evaluated and the heat transfer performance of the seawatercooling system installed the heat exchanger. At the reducing
the number of plates in the plate heat exchanger against theprevious research, the heat transfer characteristics of the fluidin the heat exchanger were investigated under higher flow
velocity conditions.
In water-water experiments, it was found that the smaller thenumber of plates, the higher the heat transfer coefficient. The
cause is that the influence of distribution of flow velocity ineach channel in plate heat exchanger.
In the FCU experiment, it was found that the heat transfer
coefficient on the FCU side increased with an increasing in the
tap water flow rate under each number of plates condition.
However, the heat transfer coefficient decreased under thecondition of the highest flow velocity., It is considered due tothe improvement in heat transfer due to the increase in flow
velocity exceeded the peak point. Regarding the cooling
performance, it was found that the higher the overall heattransfer coefficient of the heat exchanger, the better the cooling
performance. By comparing experiments with two different
number of plates, it has a wider distribution of the correlation
between the overall heat transfer coefficient and the airtemperature difference in the FCU.

104

Reh [-]

Fig. 4.1 Correlation between 𝑁𝑢ℎ and 𝑅𝑒ℎ
5 plates
13 plates
21 plates

Table 4.1 𝑁𝑢ℎ number
𝑁𝑢ℎ = 0.349𝑅𝑒ℎ0.592 𝑃𝑟ℎ1/3
𝑁𝑢ℎ = 0.349𝑅𝑒ℎ0.584 𝑃𝑟ℎ1/3
𝑁𝑢ℎ = 0.302𝑅𝑒ℎ0.583 𝑃𝑟ℎ1/3

4.2 FCU experimentFig. 4.2 shows that the relationship between the mass flowrate 𝐺𝑡𝑤 of tap water and the heat transfer coefficient
𝑈𝐹𝐶𝑈,𝑡𝑤 of the FCU, and Fig. 4.3 shows that the comparing
the cooling performance at the difference in the heat transfer
coefficient of the PHE. In addition, the temperature difference𝑑𝑇𝑎𝑖𝑟 [°C] was calculated by Eq. (4) using the inlet and outlet
temperatures on the air side of the FCU.
𝑑𝑇𝑎𝑖𝑟 = 𝑇𝑎𝑖𝑟−𝐹𝐶𝑈,𝑖𝑛 − 𝑇𝑎𝑖𝑟−𝐹𝐶𝑈,𝑜𝑢𝑡(4)

Fig. 4.2 shows that the 𝑈𝐹𝐶𝑈,𝑡𝑤 increases with an increasingin 𝐺𝑡𝑤 . Although, the heat transfer coefficient decreased at thehighest flow velocity. It is considerd that one possible reason
is that the value of the heat transfer coefficient and overall heattransfer coefficient become constant in the region where theReferences(1) Nishiguchi M., Master's thesis, Graduate School of Scienceand Engineering, Saga University, 2022.

30

Boiling heat transfer characteristics of ammonia
in micro channel and plate heat exchanger
Shoichi Suehiro*, Hirofumi ARIMA***

Graduate school of Sci. Eng., Saga Univ., 3-hara, Fukuoka-shi, Fukuoka 814-0022, Japan
Email: 23728014@edu.cc.saga-u.ac.jp**
Institute of Ocean Energy, Saga Univ., 1-48, Hirao, Kubara-aza, Yamashiro-cho, Imari-shi, Saga, 849-4256, Japan

Abstract
To improve the performance of the heat exchanger for OTEC, it is necessary to clarify the boiling heat transfer
characteristics of ammonia of the heat exchanger. In this study, the effect of the channel shape of heat exchanger andshape of heat transfer surface of PHE were focused, and the overall and the boiling heat transfer coefficient were
measured using ammonia as the working fluid on micro channel (MCHE) and plate heat exchanger (PHE). The MCHE
has 35 stages with 51 channels per stage. The heat transfer surfaces of PHE has a herringbone surface which wereanodized aluminum plates. The anodic oxidation was performed to increase the corrosion resistance for ammonia. As
a result, it was clarified that the effect of the mass flow rate and temperature of hot water and on the overall heattransfer coefficient and boiling heat transfer coefficient. Further, an empirical correlation for the boiling heat transfer
coefficient based on the Lockhart-Martinelli parameter was derived.
Key Words: Overall heat transfer coefficient, Boiling heat transfer coefficient, MCHE, PHE,
Lockhart-Martinelli parameter

1. Introduction
In order to improve the performance of heat exchanger for OTEC, it is essential to clarify the boiling heat transfer
characteristics of heat exchanger. The overall heat transfer coefficient and boiling heat transfer coefficient were
measured to clarify the boiling heat transfer characteristics on MCHE and PHE. To improve the performance of heatexchangers used in OTEC, various studies have been conducted, such as heat transfer characteristics of ammonia,
channel shape of heat exchanger, shape of heat transfer surface of PHE, enhancement boiling heat transfer coefficient,
Pressure drop and material of heat transfer surface. Therefore, in this study, MCHE was investigated the effect ofchannel shape of heat exchanger. In addition, PHE was investigated the effect of shape of heat transfer surface.

2. Experiment
2.1 Experimental apparatus
Fig. 1 shows that the schematic diagram of experimental apparatus. The experimental apparatus consists of a heatexchanger, a refrigerator 1 and 2, a subcooler, a hot water tank, a hot water pump and a working fluid tank. The
ammonia is used as the working fluid. The MCHE and PHE has been adopted as a heat exchange. The quantity ofstate which measured by each sensor are collected and recorded in a data logger. In addition, the experimental
apparatus has three circulation cycles: a working fluid, a cold water and a hot water circulation cycle.
Fig. 2 shows that an external view of the MCHE which is made of SUS316. The size of MCHE is width of 140 mm
× length of 200 mm × thickness of 40 mm. The MCHE has 35 stages with 51 channels per stage.
Fig. 3 shows that an external view of test plate which is made of A1050 for PHE. The size of PHE is width of 100
mm × length of 350 mm × thickness of 1 mm. The surface shape of the plate is herringbone and the chevron angle
is 60 degrees. This test plate is surface-treated by anodic oxidation specified in JIS H8601 in order to provide corrosion
resistance.
Table 1 and 2 show that the experimental conditions of this experiment in MCHE and PHE.

31

Fig. 1 Schematic diagram of experimental apparatus.

Fig. 2 MCHE

Fig. 3 PHE

Table 1 Experimental conditions (MCHE).

Pressure [kPa]
Mass flow flux of working [kg/m2s]
Flow rate of hot water [L/min]
Inlet temperature of hot water [℃]

7008.2, 10.2, 12.3, 14.3
1, 2, 3, 4, 5
25,30,35,40

Table 2 Experimental conditions (PHE).

Pressure [kPa]
Mass flow flux of working [kg/m2s]
Flow rate of hot water [L/min]
Inlet temperature of hot water [℃]

7002.7, 3.9, 5.8, 7.7, 11.6
1, 2, 3, 4, 5
25, 30, 40

2.2 Data reduction
The overall heat transfer coefficient 𝑈and boiling heat transfer coefficient hwf were obtained by the experimental
data. These values were calculated by following equations:
𝑈 = 𝑄ℎ /(𝐴𝑆 ∆𝑇𝑙𝑚 )

(1)

1

(2)

ℎ𝑤𝑓

=

1
𝑈

−(

1
ℎℎ

𝑡

+ )
𝑘

Further, heat transfer of hot water hh was calculated by a following equation proposed by Nishiguchi (1).
𝑁𝑢 = ℎℎ 𝐷𝑒𝑞 /𝑘ℎ = 𝐶1 𝑅𝑒 𝑛 𝑃𝑟 1⁄3(3)

where 𝐶1 = 0.4, 𝑛 = 0.23 at MCHE and 𝐶1 = 0.31, 𝑛 = 0.58 at PHE (Nishiguchi(1)).

32

3. Result and discussion
3.1 Overall heat transfer coefficientFig 4 (a) and (b) show that the overall heat transfer coefficient 𝑈 (at 𝑇ℎ,𝑖𝑛 = 30℃) of in MCHE and PHE. At theresult of the MCHE in Fig. 4 (a), the overall heat transfer coefficient tends to increase with an increase in hot watervelocity at hot water velocity 𝑉ℎ < 0.1. However, it remains almost constant at 𝑉ℎ > 0.1.
At the result of the PHE in Fig. 4 (b), the overall heat transfer coefficient tends to increase at 𝑉ℎ < 0.2. However, incase of 𝑉ℎ > 0.2, it remains almost constant at the large mass flow flux G of working fluid.
4000

4000
Gwf [kg/(m2s)]
14.3
12.3
10.2
8.2

3000
U [W/m2K]

U [W/m2K]

3000

Gwf [kg/(m 2s)]
11.6
7.7
5.8
3.9
2.7

2000

2000

1000

1000
0
0

0.1

0.2

0
0

0.3

0.1

0.2

0.3

Vh [m/s]Vh [m/s](a) MCHE, 𝑇ℎ,𝑖𝑛 = 30℃(b) PHE, 𝑇ℎ,𝑖𝑛 = 30℃
Fig. 4 Overall heat transfer coefficient against hot water velocity.

3.2 Boiling heat transfer coefficientFig 5 (a) and (b) show that the boiling heat transfer coefficient hwf (at 𝑇ℎ,𝑖𝑛 = 30℃) in MCHE and PHE. According
to Fig. 5 (a) and (b), the boiling heat transfer coefficient tends to increase at vapor quality x < 0.4. However, it decreases
at x > 0.4.
20000

6000

hwf [W/m2K]

4000

hwf [W/m2K]

Gwf [kg/(m2s)]
14.3
12.3
10.2
8.2

5000

10000

3000
2000

Gwf [kg/(m 2s)]
11.6
7.7
5.8
3.9
2.7

1000
0
0

0.2

0.4

0.6

0.8

0
0

1

0.2

0.4

0.6

0.8

1

x [-]

x [-]

(a) MCHE, 𝑇ℎ,𝑖𝑛 = 30℃(b) PHE, 𝑇ℎ,𝑖𝑛 = 30℃
Fig. 5 Boiling heat transfer coefficient against quality.

3.3 𝒉𝒘𝒇 ⁄𝒉𝑳𝒁as a function of 1/𝜲𝒗𝒗
Fig 6 (a) and (b) show that the relationship between the ratio ℎ𝑤𝑓 ⁄ℎ𝐿𝑍 and 1/𝛸𝑣𝑣 at MCHE and PHE. LockhartMartinelli parameter 𝛸𝑣𝑣 and ℎ𝐿𝑍 was calculated by following equations:
𝑋𝑣𝑣 = {(1 − 𝑥)/𝑥}0.5 (𝜌𝑔 /𝜌𝑙 )ℎ𝐿𝑍 = 0.23𝑘𝑙
𝐷𝑒𝑞

𝐺(1−𝑥)𝐷𝑒𝑞 0.8

[

𝜇𝑙

]

0.5

(𝜇𝑙 /𝜇𝑔 )

0.5

(4)

𝑃𝑟𝑙0.4(5)

The derived correlation of MCHE and PHE with Lockhart-Martinelli parameter are as follows:[MCHE] ℎ𝑤𝑓 ⁄ℎ𝐿𝑍 = 3.3(1⁄Χ 𝑣𝑣 )0.80(6)ℎ𝑤𝑓 ⁄ℎ𝐿𝑍 = 22.6(1⁄Χ𝑣𝑣 )0.97

(7)

[PHE]

On the other hand, the correlation equation of Arima et al. (2) which is correlation for flat plate are as follows:ℎ𝑤𝑓 ⁄ℎ𝐿𝑍 = 16.4(1⁄Χ𝑣𝑣 )1.08(8)33

According to Fig. 6 (a), the correlation of Lockhart-Martinelli parameter for MCHE is lower than the correlation of
Arima et al. (2) However, in case of PHE, the value of present study is larger than the correlation.

104
hwf /hLZ=22.6(1/Xvv)0.97

hwf /hLZ [-]

103
102
+25%

101

Present date
Correlation date
Arima et al.

-25%

100 -110(a) MCHE

Arima et al.

100
101
1/Xvv [-]

102(b) PHE
Fig. 6 ℎ𝑤𝑓 ⁄ℎ𝐿𝑍 as a function of 1/𝛸𝑣𝑣

4. Conclusion(1) At the MCHE, the overall heat transfer coefficient tends to increase with an increase in hot water velocity at hot
water velocity 𝑉ℎ < 0.1. However, it remains almost constant at 𝑉ℎ > 0.1.
At PHE, the overall heat transfer coefficient tends to increase at 𝑉ℎ < 0.2. However, in case of 𝑉ℎ > 0.2, it
remains almost constant at the large mass flow flux G of working fluid.
(2) At the MCHE and PHE, the boiling heat transfer coefficient tends to increase at vapor quality x < 0.4. However,it decreases at x > 0.4.
(3) The correlation of Lockhart-Martinelli parameter for MCHE is lower than the that of Arima et al. However, thatof PHE is large than the that of Arima et al. (2).

5. References(1) Nishiguchi M., Master's thesis, Graduate School of Science and Engineering, Saga University, 2022.
(2) H. Arima, J. H. Kim, A. Okamoto, Y. Ikegami, “Local boiling heat transfer characteristics of ammonia in a verticalplate evaporator”, International Journal of Refrigeration, Vol. 33, No. 2 (2010), pp. 359-370.

34

Research on optimal design of ocean thermal energy conversion system
Takayuki HARAGUCHI, Takafumi MORISAKI, Takeshi YASUNAGA and Yasuyuki IKEGAMI
Institute of Ocean Thermal Energy, Saga University, 1-Honjo machi, Saga shi, Saga 840-8502, Japan
E-mail :23728018@edu.cc.saga-u.ac.jp
Abstract
Since ocean thermal energy conversion (OTEC) uses the temperature difference of about 20°C in seawater to generate electricity,
the available temperature difference is very small compared to thermal power generation method. In order to promote practical use,
it is important to design with consideration for higher performance and installation requirements. In this paper, we focused on the
optimization of the OTEC system under the conditions of the set Deep ocean water (DOW) discharge, targeting the combined useof heat and seawater resources such as the Kumejima model. We clarified the characteristics of each objective function of
conventional optimization, along with the optimum Surface ocean water (SOW) flow rate at a set DOW flow rate. As a result, inthe multi-stage use of DOW, the optimization of OTEC when the DOW flow rate is set was examined, and the relationship between
the SOW flow rate and the conventional objective function was clarified. Furthermore, as a theoretical study when the DOW flowrate is set, an estimation formula for 𝐶𝐻 that maximizes the net power for 𝐶𝐿 was derived by numerical analysis.
1．Introduction

expected as a Kumejima model. If the main purpose is to use

Since ocean thermal energy conversion (OTEC) uses the

DOW like the Kumejima model, it is important to optimize thetemperature difference of about 20°C in seawater to generate

OTEC system under the set DOW flow rate. There are notelectricity, the available temperature difference is very small

many studies of OTEC optimization in such cases.

compared to other power generation methods. In other words,

Therefore, in this research, with the aim of optimizing thetheoretically, the amount of heat required for the work of the

OTEC system with a constant DOW flow rate, we clarified theheat engine is large, and the heat transfer area of the heatcharacteristics of each objective function along with theexchanger is large. Therefore, in order to promote practical use,

optimal SOW flow rate under this condition. For this optimal

it is important to design with consideration for further

design, we conducted experiments in a wider range than theperformance and installation requirements. Heat exchanger

conventional applicable range for the heat transfer equation ofoccupies the largest volume in the total volume of the floating

the Cross flow plate heat exchanger to be used, and derived abody, and since titanium is used as the material, the heatnew estimation equation. Furthermore, as a theoretical study

exchanger is the component that has the greatest impact on thewhen the DOW flow rate is set, the optimum SOW flow rate

construction cost. Therefore, in the design of the OTEC power

estimation formula is examined by numerical analysis.

plant, minimization of the construction cost is simply
calculated by assigning it to the optimization with the ratio of

2．Evaporator performance test

the total heat transfer area of the heat exchanger per net power

In the experiment, a performance test was conducted on a cross-as the objective function [1].

flow plate heat exchanger using a 15kW OTEC experimental

On the other hand, it is estimated that water intake facilities

apparatus. Schematic diagram of the OTEC system using Rankine

account for about 70% of the total installation cost for a small-cycle is shown in Fig.1. This system consists of an evaporator, ascale (1~5MW) land-based OTEC. Above all, the cost ofworking fluid pump, piping, a turbine, a condenser, a separator, andintake equipment for DOW accounts for 80% of thea generator.

installation cost [2]. For the purpose of reducing the installation

The experimental conditions were a hot water temperature

cost of water intake equipment, attempts have been made toof 29°C, a cold water temperature of 9°C, a hot water flow rate

design single-stage Rankine cycle (S-R) and two-stageof 120-390 m3/h, and a vapor quality of the working fluid at

Rankine cycle (D-R) to minimize the total amount of seawater

the evaporator outlet of about 0.8. Pure ammonia was used as

intake. Furthermore, in order to reduce the DOW water intake,the working fluid. In this study, in order to extend thean optimization method was constructed to minimize theapplication range, the current study was conducted at 1400 to

DOW water intake [3]. However, results have been obtained by

4500, compared to Re = 1400 to 2500 in conventional

optimization, but the Reynolds number is high, and an

experimental condition.

empirical formula for the heat transfer coefficient at a highFrom the experimental results, 𝑁𝑢𝑊𝑆 on the hot water side

Reynolds number is required in order to improve the accuracy

obtained the following estimation formula.
1/3of the optimization.

0.8
𝑁𝑢𝑊𝑆 = 0.0858𝑅𝑒𝑊𝑆
𝑃𝑟𝑊𝑆

In addition, combined use of DOW including OTEC is35

6

1 [m2/kW]

lDOW=1500 [m]lDOW=1000 [m]lDOW= 800 [m]5QV,L=5000 QV,L=20000
QV,L=10000
QV,L=30000 [m3/h]

4

0

10000

20000

30000

QV,H [m3/h]

Fig.1 Rankine Cycle

Fig. 3 Relationship between 𝑄𝑉,𝐻 and 𝛾1
3．Evaluation of the objective function when the amountof DOW intake is constantIn this study, the OTEC system was optimized for the DOW
flow rate, assuming that the power generation output of the
OTEC is determined from the set DOW flow rate when the
OTEC is installed mainly for the purpose of multi-stage use of
the DOW. In this optimal design, the following objectivefunction 𝛾1 , 𝛾2 , 𝛾3 , which has been used conventionally, wasused.
𝛾1 = (𝐴𝐸𝑠𝑢𝑚 + 𝐴𝐶𝑠𝑢𝑚 )/𝑊𝑛𝑒𝑡
𝛾2 = (𝐴𝐸𝑠𝑢𝑚 + 𝐴𝐶𝑠𝑢𝑚 )𝑊𝑚𝑎𝑥 /𝑊𝑛𝑒𝑡
𝛾3 = (𝐴𝐸𝑠𝑢𝑚 + 𝐴𝐶𝑠𝑢𝑚 )𝑚𝑐𝑠 /𝑊𝑛𝑒𝑡

Fig. 4 Relationship between 𝑄𝑉,𝐿 and 𝑄𝑉,𝐿

where 𝛾1 is the total heat transfer area per net power, 𝛾2

at the optimum point (𝑙𝐷𝑂𝑊 = 1000m)

is the total heat transfer area for the exergy efficiency, and 𝛾3is the objective function for reducing DOW water intake.

From Fig.2, we confirmed that 𝛾1 and 𝛾3 are minimized,

The optimum design was performed using the heat balance,and 𝛾2 decreases as the SOW flow rate decreases. From thethe performance of each component, and the empiricalfigure, it was clarified that there exists an optimum SOW flow rate

formulae for convective heat transfer coefficient. Fig.2 shows

for the set DOW flow rate at 𝛾1 and 𝛾3 . It was also confirmed

the relationship between the SOW flow rate 𝑄𝑉,𝐿 and eachthat the optimum conditions are equivalent under the constantobjective function when the DOW flow rate 𝑄𝑉,𝐻 =DOW flow rate. From Fig.3, it was confirmed that the minimum

10000m3/h and the DOW intake pipe length is 1000m. Fig. 3

value of the objective function takes the same value for each intake

shows the relationship between the change in 𝛾1 and the net

pipe length.

power when the intake pipe length and DOW flow rate are

Also, looking at the flow rate at this optimum point, it waschanged.

newly found that the SOW flow rate relative to the DOW flowrate is determined for each OTEC design condition. Fig. 4
shows the relationship between the DOW flow rate and the20000

3 [m2-kg / kW-s]

108

SOW flow rate when the length of the DOW intake pipe is
1000m. Clearly, the optimum SOW flow rate is determined by

6

the DOW flow rate.

10000

4．Theoretical study for constant DOW flow rate

4

1
2
3

2

0

0

5000

10000

QV,H [m3/h]

15000

5000

In this study, numerical analysis was conducted to find the2 [m2]

1 [m2/kW]

15000

theoretical formula for the SOW flow rate that maximizes thenet power under the condition that the DOW flow rate is set.
Equation (2) is the relational expression of the net power by

0
20000

Ikegami et al. [4]. The derivative of Eq.(2) is Eq.(3), and therelational expressions of the heat capacity flows 𝐶𝐻 and 𝐶𝐻

Fig. 2 Relationship between 𝑄𝑉,𝐻 and 𝛾1 , 𝛾2 , 𝛾3that satisfy Eq.(3) are derived.

36

𝑊𝑛𝑒𝑡 = 𝛥𝑇(𝐶𝐻−1 + 𝐶𝐿−1 )−1 − 𝑅𝐻 𝐶𝐻𝑛+1 − 𝑅𝐿 𝐶𝐿𝑛+1

(2)

10000

+10%

2

𝜕𝑊𝑛𝑒𝑡

=

𝜕𝐶𝐻

𝛥𝑇
−1 2
(𝐶−1
𝐻 + 𝐶𝐿 )− (𝑛 + 1)𝑅𝐻 𝐶𝐻𝑛+2 = 0

CH,CAL [MW/K](Δ𝑇 = (√𝑇𝐻 − √𝑇𝐿 ) )

(3)

1000

Here, DOW temperature 𝑇𝐿 = 4 ℃, SOW temperature 𝑇𝐻 =28 ℃ , 𝑛 = 1.8 , flow resistance variable 𝑅𝐿 = 𝑅𝐻 =10−7 K 2.8 /MW1.8 ,-10%Fig. 5 shows the relationship between

𝐶𝐻 and 𝑊𝑛𝑒𝑡 when𝐶𝐿 = 1000 MW/K , and Fig. 6 shows
the relationship between 𝐶𝐿 and 𝐶𝐻,𝑂𝑃𝑇 .

100
100

1000

10000

CH,EXP [MW/K]

From Fig. 5, it can be confirmed that the optimum value of 𝐶𝐻 thatmaximizes the net power (𝑊𝑛𝑒𝑡,𝑀𝐴𝑋 ) with respect to 𝐶𝐿 can be

Fig. 7 Relationship between 𝐶𝐻,𝐸𝑋𝑃 and 𝐶𝐻,𝐶𝐴𝐿

obtained. Fig. 6 shows the relationship of 𝐶𝐻,𝑂𝑃𝑇 to changes in 𝐶𝐿
under the same conditions of 𝑛, Δ𝑇and 𝑅𝐻 . From this relationship,

Table 1 Range in equation (4)𝑛, Δ𝑇 and 𝑅𝐻 are variables, and the optimal relational expression

1.5 ≤ 𝑛 ≤ 2.0

of 𝐶𝐻,𝑂𝑃𝑇 for 𝐶𝐿 is derived as shown in Equation (4). Table 1

0.3 ≤ ∆𝑇 ≤ 1.0

shows the conditions under which Eq. (4) holds. In addition, Fig. 7

1 × 10−7 ≤ 𝑅𝐻 ≤ 4 × 10−7shows the theoretical formula for obtaining the optimum value of 𝐶𝐻 ,When 𝑊𝑛𝑒𝑡 > 0，𝐶𝐿 < 44.2𝑛−8.06 𝑅𝐻−0.576 ∆𝑇 𝑛

1

and the relationship between the optimum values of 𝐶𝐻 in Eqs. (3)and (4).

5．Conclusions−0.250 0.682𝑛
𝐶𝐻,𝑂𝑃𝑇 = 4.15𝑛−1.68 ∆𝑇 0.249 𝑅𝐻
𝐶𝐿−0.928(4)

In the multi-stage use of DOW, the optimization of OTEC

From Fig.7, we confirmed that the error of Equation (4)when the DOW flow rate was investigated. Regarding thederived by numerical analysis is within ±10%.

relationship between the SOW flow rate and the conventional
objective function, it was clarified that the optimum SOW flow

400

1rate exists for the set DOW flow rate in 𝛾1 and 𝛾3 , and that300

clarified that the optimum SOW flow rate was determined when
the DOW flow rate was determined for each OTEC design

Wnet,MAXEX [-]

Wnet, Wm [MW]these objective functions can be optimized. In addition, it was
Wm

EX

200

condition. Furthermore, an estimation formula for 𝐶𝐻 that

0.5

maximizes the net power for 𝐶𝐿 was derived by numericalanalysis, and it was confirmed that the error of this estimation

100

Wnet

formula was within ±10%.

CH,OPT
0

0

500

1000

1500

2000

0
2500

References

CH [MW/K][1]上原，中岡，機論(B 編)，50-456(1984)，pp.1955-1962

Fig. 5 Relationship between 𝐶𝐻 and 𝑊𝑛𝑒𝑡

[2]沖縄県久米島町，久米島海洋深層水複合利用調査報告書，(2011)

2500

300

2000

築に関する研究”，令和 3 年度

net,OPT

200
1500

1000

500

100

Wnet,MAX [MW][4]

CH=1374 MW/K

CH,OPT [MW/K][3]中村泰誠,”陸上型海洋温度差発電の最適設計法の構CH,OPT
Wnet,MAX W

139-144

CL=1374 MW/K
0

0

500

1000

1500

2000

0
2500

CL [MW/K]

Fig. 6 Relationship between 𝐶𝐿 , and 𝐶𝐻,𝑂𝑃𝑇

37

修士論文,pp.1-141

Y. Ikegami, et al., J. Sol. Energy Eng, 120(1998),pp.

Experimental study of a hybrid hot spring water temperature difference power generation
systemTomoyuki HIRAISHI, Takahumi MORISAKI, Naoki NAKAZIMA Yasuyuki IKEGAMI
Institute of Ocean Thermal Energy, Saga University, 1-Honjo machi, Saga shi, Saga 840-8502, Japan
Tel :+81-0952-28-8624, FAX : +81-0952-28-8595, E-mail :23728019@edu.cc.saga-u.ac.jp
Abstract
Japan has about 10% of the world's active volcanoes (1) and possesses a large geothermal potential. Among these, smallscale binary power generation is expected to be one of the most promising, but adhesion to pipes and heat exchangers
is an issue. As a countermeasure, hybrid hot spring water temperature difference power generation (H-STEC), whichuses flash evaporation of hot spring water, has been proposed. This study aims to clarify the characteristics of the HSTEC power generation system and its effect on the working fluid flow rate. It is found that the upper limit of theworking fluid flow rate exists under the operating conditions of this H-STEC system because the superheat at the turbine
inlet is determined by the working fluid flow rate due to the constant evaporation temperature of the hot spring water.
In the future, we will experimentally clarify the characteristics of this system when it is applied using R1224yd, a lowGWP refrigerant.
Key Words: Flash evaporation, H-STEC, Hot spring, Maximum power, Self-supplied water system
1. INTRODUCTION
Japan has approximately 10% of the world's active
volcanoes (1), and as one of the world's largest volcanic
countries, it possesses a large amount of geothermal
potential. Japan has the world's third largest geothermal
potential, with a geothermal resource of approximately
23.47 million kW (2). In fact, Japan's geothermal power
plants have an approved output of 603 MW (in 2021) (3).
In addition, geothermal power generation is notaffected by weather conditions, and stable power
generation is possible day and night because power isgenerated by pumping up steam and hot water fromgeothermal reservoirs deep underground. Therefore,
geothermal power generation has a high utilization rate
among renewable energies, with a facility utilization rateof 56% compared to 13% for solar power and 20% forwind power (4). In Japan in particular, geothermal waterwith low temperatures is scattered throughout the country.
Figure 1 shows a schematic diagram of a binary power
generation system. Binary power generation uses anorganic Rankine cycle (ORC), in which a geothermal
fluid with a relatively low temperature of 150°C or lower
is used as a heat source to exchange heat with a lowboiling-point medium that has a lower boiling point thanwater. Because the low-boiling medium is used as theworking fluid, binary power generation can generate
power using heat at medium to low temperatures, which
is not possible with the flash method. It also has theadvantage of a large pressure difference before and after
the turbine, which allows the turbine to be downsized.
However, the heat transfer performance is degraded whensilica, calcium carbonate, and other components in thegeothermal fluid adhere to the heat exchanger as scale,causing a reduction in power generation output. In
addition, if an air-cooled condenser is used, it is larger
than a water-cooled condenser, and if a water-cooled
condenser is used, securing water resources is an issue.
This study focuses on a hybrid hot spring watertemperature difference power generation system: HSTEC, which combines a flash evaporation system and a
hot spring water temperature difference power generation
system.

Fig.1 General STEC system (hot spring binary power
generation)
Conventional studies on H-STEC have notexperimentally clarified the operational and essential
cycle characteristics to obtain the maximum power
output of H-STEC, especially the cycle characteristics
including working fluid and turbine inlet superheat. This
study experimentally clarifies the optimum operating
conditions for H-STEC in terms of working fluid flowrate and turbine inlet superheat.
Silica and calcium carbonate components dissolved in
the hot spring water precipitate and adhere to the heatexchanger, resulting in a decrease in the flow rate of the
hot spring water and a decrease in the amount of heat thatcan be recovered by the heat exchanger. As a result, it has
been reported that the power output decreases with
continued operation (5). Although many studies have been
conducted on anti-scale measures, they have yet to reacha satisfactory solution.
In the case of an air-cooled condenser, air is used as
the cold heat source, eliminating the need for water
resources. However, since the heat transfer coefficient of
the air side is lower than that of the working fluid side, alarge heat transfer area must be secured by installing finson the air side, which may make it the largest structure
among the components. On the other hand, when using awater-cooled condenser, securing the cooling waterrequired to cool the working fluid is an issue.
This study focuses on a hybrid hot spring watertemperature difference power generation system: HSTEC, which combines a flash evaporation system and a38

hot spring water temperature difference power generation
system.
Conventional studies on H-STEC have notexperimentally clarified the operational and essential
cycle characteristics to obtain the maximum power
output of H-STEC, especially the cycle characteristics
including working fluid and turbine inlet superheat.
In addition, the effectiveness of H-STEC in preventing
scale deposition in the evaporator by flash evaporation
has not been quantitatively evaluated.
In addition, the recent addition of R245fa, aconventionally used working fluid, to the regulatory list
has prompted ORC to investigate alternative working
fluids, but no evaluation of the applicability of alternative
working fluids to H-STEC has been conducted. In thisstudy, the following three points will be investigated.

3.

PE

To experimentally clarify the optimum operating
conditions for H-STEC in terms of working fluid
flow rate and turbine inlet superheat.

PC
H-STEC
STEC

600

400

200

0
0.9

1

1.1

1.2

1.3

1.4

Working fluid flow mWF [kg/s]

1.5

Fig.3 Evaporation and condensation pressures versus
working fluid flow rate

To quantitatively clarify the scale deposition
prevention effect of H-STEC in the evaporator bychemical analysis of hot spring water andcondensate.

Fig. 4 shows the relationship between the working
fluid flow rate and the power generation end output withthe refrigerant R245fa. The increase in the differential
pressure between the turbine inlet and outlet with theincrease in the working fluid flow rate is a factor. On theother hand, the turbine inlet superheat, △TTIsh, showed adecreasing trend. This is thought to be due to the lower
heat source available for H-STEC compared to STEC.

Apply R1224yd, a low-GWP refrigerant proposed as
an alternative to R245fa, as well as R245fa, to HSTEC to theoretically clarify the characteristics
inside the cycle and to clarify the applicability andvalidity of the alternative working fluid.

Generating end output WG [kW]2. ANALYSISThe specifications of the flash chamber used in theexperiment and photographs of its exterior are shown in
Fig. 2, respectively. The flash chamber has one nozzle
connected to the hot spring water inlet, and flash
evaporation occurs when the water leaves the nozzle. The
generated vapor is sent to the evaporator through ademister. The evaporator has a connection to a vacuum
pump, and the internal pressure can be changed by
opening a valve.

20

40

15

30

10

20

5

Turbine inlet superheat TTIsh [C]

2.

800

Evaporation pressure PE [kPaA]
Condensing pressure PC [kPaA]

1.

3. Result and Consideration of Calculation at cycle
Fig. 3 shows the relationship between the working
fluid flow rate and the evaporation and condensationpressures. The evaporation pressure is increasing in
relation to the working fluid flow rate, and the differential
pressure at the turbine inlet and outlet is increasing.

10

R245faWG TTIsh
H-STEC
STEC

0
0.9

1

1.1

1.2

1.3

1.4

Working fluid flow rate mWF [kg/s]

0
1.5

Fig.4 Generating end power and turbine inlet superheat
versus working fluid flow rate.
Fig. 5 shows the relationship between the working
fluid flow rate and the cycle thermal efficiency. The cycle
thermal efficiency and the maximum power ratio areincreasing for both STEC and H-STEC. In this case, thecycle thermal efficiency is increasing because the output
is increasing and the rate of increase of the heatexchanged in the evaporator is smaller than the output.

Fig.2 Hot spring steam and condensate flow
39

30

5

25

4

20

3

15

2

10
th

1
0
0.9

m
H-STEC
STEC

1

1.1

1.2

1.3

1.4

5

Maximum output efficiency m [%]thermal efficiency of cycle th [%]

6

𝐵𝑃𝑅
𝐷𝑀𝑆
𝑇
𝑃
∆𝑇𝑚
∆𝑇𝑠
∆𝑇
𝑈
𝑊
𝑟𝑓
𝑅𝑊
𝜂
𝜌
𝑣
𝜆
𝛿
𝑀

0
1.5

Working fluid flow rate mWF [kg/s]

Fig.5 Cycle thermal efficiency and maximum power
factor versus working fluid flow rate.
4. CONCLUSIONSParametric analysis was conducted to theoreticallyclarify the applicability of the green refrigerant R1224yd
to replace R245fa and its effectiveness in comparison
with STEC. As a result, it was theoretically clarified thatthe basic characteristics of R245fa and R1224yd in HSTEC are equivalent in performance for both working
fluids. The applicability of R1224yd should be
experimentally evaluated in the future.
The output of STEC and H-STEC were compared, andit was quantitatively clarified that the output of STEC
was higher than that of H-STEC under the same
analytical conditions. However, considering that the heattransfer performance of the STEC evaporator is reducedby the deposition of the hot spring water scale, there
exists an inversion condition in which the output of HSTEC is higher than that of STEC. The chemical analysis
was conducted to clarify the change of the hot springwater after the theory that the higher the heat passage
coefficient, the more the stain causes the STEC output
decrease and the more effective the H-STEC is, and thecondensate showed that it was quantitatively clarified
that the fouling components that cause scale were
removed. In addition, a 30 kW H-STEC experiment was
conducted in this study to investigate optimization
parameters. As a result, it was experimentally clarified
that the working fluid flow rate determines the turbine
inlet superheat due to the constant hot spring steam
temperature, and that there is an upper limit to theworking fluid flow rate under the operating conditions of
this H-STEC system. Therefore, the balance between theworking fluid turbine inlet superheat and the hot spring
steam temperature is important for the stabilization of thissystem.

rise in boiling point [K]demister loss [K]Temperature [oC], [K]pressure [kPa]
Logarithmic mean temperature
difference [K]simple average temperature
difference [K]Temperature drop [K]Heat Passage Coefficient[kW/(m2∙K)]
output (e.g., of dynamo) [kW]Feed water ratio [%]dirt factor [(m2∙K)/W]
efficiency [%]density [kg/m3]
speed of moving fluid [m2/s]
Thermal conductivity [W/(m∙K)]
thickness [m]Scale adhesion[g/m2]

Reference[1] Cabinet Office, Disaster Prevention Information
Page，https://www.bousai.go.jp/[2] Agency for Natural Resources and Energy, Ministryof Economy, Trade, and Industry, "Energy White
Paper 2021", (2021)[3] New Energy Foundation, "Proposal for the
Promotion of Development and Utilization of
Geothermal Energy”. （2022）[4] Organization for Promotion of Wide-Area
Electricity Utilization, "Compilation of Supply Plan
for FY2021”. （2021）[5] Motoaki Morita, Scale Problem in Hot Spring
Binary Power Plants and Development of Inhibition
Materials, Materia, Vol. 57, No. 10. （2018）

NOMENCLATURE
Heat Transfer Area [m2]𝐴specific heat at constant pressure𝑐𝑝[kJ/(kg∙K)]
specific entropy [kJ/(kg∙K)]𝑠specific enthalpy [kJ/kg]ℎlatent heat [kJ/kg]
𝐿𝑚
Mass, Volume Flow [kg/s] ，[m3/h]
Heat exchange rate [kW]𝑄
Non-equilibrium temperature
𝑁𝐸𝑇𝐷
difference [K]40