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Design and ThermodynamicAnalysis of an H2O–LiBr AHP system for naval surfaceship application
Cüneyt Ezgi
PII: S0140-7007(14)00227-8
DOI: 10.1016/j.ijrefrig.2014.08.016
Reference: JIJR 2865
To appear in: International Journal of Refrigeration
Received Date: 4 June 2014
Revised Date: 15 August 2014
Accepted Date: 27 August 2014
Please cite this article as: Ezgi, C., Design and ThermodynamicAnalysis of an H2O–LiBr AHPsystem for naval surface ship application, International Journal of Refrigeration (2014), doi: 10.1016/j.ijrefrig.2014.08.016.
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Design and Thermodynamic Analysis of an H 2O–LiBr AHP System for Naval Surface Ship Application
Cüneyt Ezgi
Mechanical Engineering Department, Turkish Naval Academy, Istanbul, 34942, Turkey
Corresponding author tel.: +90 530 606 53 95
E-mail address: [email protected]
Abstract
Absorption heat pump (AHP) systems are cleaner and more efficient energy solutions than
vapour–compression heat pump systems for heating and cooling on board naval surface
ships. Thermal management is a critical requirement for naval surface ships and submarines
as well as commercial vessels and land-based industrial plants. Approximately 25% of a
ship’s thermal load is removed through the heating, ventilation and air conditioning (HVAC)
system. In this study, design and thermodynamic analysis of a water-lithium bromide (H2O-
LiBr) AHP as an HVAC system for a naval surface ship application are presented and
compared with those of a vapour–compression heat pump.
Keywords: Ship; Engine; Sea water; Water-lithium bromide; Absorption system; Heat pump
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Nomenclature Variables
C flow stream heat capacity, pcm& ,W K-1
cp specific heat capacity, J(kgK)-1
f solution circulation ratio
h specific enthalpy, Jkg-1
m& mass flow rate, kgs-1
mf mass fraction,-
M molecular weight, kgkmol-1
P pressure, Pa
P∆ difference between inlet and exit pressures, Pa
Q& heat transfer rate, W
T temperature, °C
W& power, W
v& volume rate of flow, m3s-1
X mass fraction of lithium bromide in solution
x, y molar amount
Greek letters
η efficiency,-
ρ density, kgm-3
ε effectiveness,-
Subscripts
A absorber
C condenser
cv control volume
dm driving motor
E evaporator
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e exit
exh exhaust gas
exht exhaust
f fluid
F fan coil unit
g gas
G generator
h hot
i inlet, inner
m mean value
p pump
S seawater
SHX solution heat exchanger
LiBr lithium bromide
ss strong solution
w water
ws weak solution
Abbreviations
AHP Absorption heat pump
AHU Air handling unit
ASHRAE American Society of Heating, Refrigerating and Air-Conditioning Engineers
COP Coefficient of performance
EEDI Energy Efficiency Design Index
HVAC Heating, ventilation and air conditioning
IMO International Maritime Organization
SEEMP Ship Energy Efficiency Management Plan
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1. Introduction
Thermal management is a critical requirement for naval surface ships and submarines as
well as commercial vessels and land-based industrial plants. Approximately 25% of a ship’s
thermal load is removed through the heating, ventilation and air conditioning (HVAC) system.
Projected Next Navy’s thermal loads are 2-5 times those of today’s ships. It is expected that
much of the increased load will be rejected via the HVAC system or directly to the chilled
water system Frank and Helmick (2007). Despite the great technological development of
modern marine diesel engines, only a small part of the energy contained in the fuel is
converted to power output. The maximum efficiency remains lower than 45%. The main
losses are dissipated as heat in exhaust gases and coolants and then transferred to the
environment Ouadha and El-Gotni (2013).
The International Maritime Organization has developed the first ever global CO2 reduction
index in the world known as the Energy Efficiency Design Index (EEDI) for new ships and the
Ship Energy Efficiency Management Plan (SEEMP) for all ships. The new chapter added to
MARPOL ANNEX VI Regulations for the prevention of air pollution from ships, which was
implemented on January 1, 2013, aims to reduce the emission of greenhouse gases,
specifically CO2 emissions, as CO2 is the most important greenhouse gas emitted by ships
(IMO, 2010). Implementing CO2 reduction measures will result in a significant reduction in
fuel consumption, leading to a significant saving in fuel costs to the shipping industry. If EEDI
and SEEMP are applied, the results obtained on naval ships can be evaluated.
Reduction and management of ship signatures should be taken as the major input during the
whole design and operating phase. Moreover, many classified precautions should be taken
to reduce hydrodynamic, acoustic, magnetic, infrared (IR) and radar signatures to achieve
the specified level of stealth feature.
IR-guided missiles represent a major threat to naval ships such as in military applications.
This threat will increase in the near future. Therefore, reducing or eliminating IR signature in
naval ship susceptibility to IR-guided anti-ship missiles is vital. Also, acoustic signature on a
naval ship should be reduced as well as IR.
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Therefore, searching for new energy conservation methods that can be applied on board
naval surface ships is necessary. One way to find a new solution to this problem is to apply
an absorption heat pump (AHP) system to provide the required heating and cooling loads for
the HVAC system instead of the traditional vapour–compression heat pump. Compared with
automobile engine, marine engine onboard ship has some advantages: more stable
operation, larger spacing for installing, and larger quantity of exhaust gas and engine
coolant, using sea water as cooling source directly Liang et al.(2013).
Unlike traditional heat pump units, which are powered by electricity, AHP works on surplus
heat from a diesel engine. However, until now, the technique has been confined to land-
based installations. AHP systems are particularly attractive in applications that have a
cooling demand and at the same time a source of heat, which if not used will be ejected to
the environment. For instance, Wärtsılä has produced 4977 kW chilled water (7/12 °C) using
a direct exhaust gas-driven absorption chiller through a diesel engine generator, which has
an electric power of 9730 kWe for district cooling. A number of research options, such as
various types of absorption refrigeration systems, on working fluids and improvement of
absorption processes, were discussed in Srikhirin et al. (2001). A single-stage H2O-LiBr
absorption chiller of 14 kW was experimentally characterised and modelled by Bakhtiari et al.
(2011). It was reported that the heat pump cooling capacity was more sensitive to cooling
stream and generator inlet temperature than it was to chilled stream temperature and the
COP is primarily influenced by the cooling stream temperature and flow rate. A mathematical
model of a single-effect H2O–LiBr AHP operated at steady conditions was presented by Sun
et al. (2010). They found that the mass flux of vapour increased with the increase of absorber
pressure, coolant flow rate, spray density of LiBr solution and decrease of coolant and input
temperature of solution and the vapour mass flux increased almost linearly with the increase
of absorber pressure.
Noise measurements were carried out on a single-effect H2O–LiBr AHP by Cotana and
Nicolini (2003) and their measurement results shown that noise 1/3 octave-band spectrum
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levels were over 50 dB only for frequency components equal and higher than 400Hz. A-
weighted power level of AHP was 65.9 dBA.
The main market barrier to the application of H2O–LiBr absorption chiller technology in
combined heat and power systems is the need for a cooling tower to eject heat from the
condenser and absorber to ambient air. The use of cooling towers in light commercial
absorption chiller systems is unpopular because cooling towers 1) provide a breeding ground
for bacteria, 2) increase initial system costs, 3) require regular maintenance and 4) require
extra space for their installation Wang et al. (2011). The development of seawater-cooled
H2O-LiBr AHP technology can effectively eliminate these disadvantages.
In the literature, although the absorption cycle is most commonly used for refrigeration in
land-based plants, there has been no report that an AHP system has been installed on board
ships. In particular, many researchers (Fernandez-Seara et al.1998; Wang and Wang 2005;
Ruiz, 2012; Táboas et al. 2014) concentrated on designing, modelling and analyzing of
absorption refrigerant which is needed for food preservation, air-conditioning and icemaker
for fishing vessels. Moreover, no investigation has been conducted yet on a seawater-
cooled H2O-LiBr AHP system for a naval ship application. Therefore, this study focuses on
the dual use of absorption technology to produce heating and cooling on board a naval ship.
2. HVAC architecture and system selection for naval surface ships
The HVAC system of a naval ship is a vital part of the overall ship thermal management
system. Shipboard HVAC is a large, complex, vital system which impacts every ship
compartment. The HVAC system is divided into zones and integrated with the ship chilled
water system. There are three types of HVAC systems on a ship: supply, exhaust and
recirculation. Compartments are either air conditioned or ventilated. In compartments that are
ventilated, there is a supply system which brings air to the compartment, and an exhaust
system which returns the air to the weather. In air conditioned compartments, the air is
recirculated, a portion of the compartment air is exhausted to the weather and a makeup
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portion of replenishment (weather) air added. In general, air enters the ship via fan rooms
where heaters, cooling coils and fans may also be located Frank and Helmick (2007).
In general, the details of merchant ship air conditioning also apply to warships. However, all
ships are governed by their specific ship specifications, and warships are usually governed
by military specifications, which ensure an excellent air-conditioning system and equipment
performance in the extreme environment of warship duty.
Design conditions for naval surface ships have been established as a compromise. These
conditions consider the large cooling plants required for internal heat loads generated by
machinery, weapons, electronics and personnel.
The cooling load consists of the following ASHRAE (2011):
• Solar radiation
• Heat transmission through hull, decks and bulkheads
• Heat (latent and sensible) dissipation of occupants
• Heat gain from lights
• Heat (latent and sensible) gain from ventilation air
• Heat gain from motors or other electrical equipment
• Heat gain from piping, machinery and equipment
The heating load consists of the following ASHRAE (2011):
• Heat losses through hull, decks and bulkheads
• Ventilation air
• Infiltration
Some electronic spaces require adding 15% to the calculated cooling load for future growth
and using one-third of the cooling season’s equipment heat dissipation (less the 15% added
for growth) as heat gain in the heating season.
Today, heat pumps for heating and cooling on board naval ships are mechanically driven.
Seawater is used for condenser cooling. The equipment described for merchant ships also
applies to naval surface ships. Fans, cooling coils, heating coils with steam or electric duct
heaters and air handling unit (AHU) are used on board naval ships. An AHU is a device used
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to regulate and circulate air as part of an HVAC system. An air handler is usually a large
metal box containing a blower, heating or cooling elements filter racks or chambers, sound
attenuators and dampers. Air handlers usually connect to a ductwork ventilation system that
distributes conditioned air through the building and returns it to the AHU. Occasionally, AHUs
discharge (supply) and admit (return) air directly to and from the space served without
ductwork.
AHP is a type of heat-driven heat pump that utilises the thermodynamic availability of a high-
temperature heat input to extract heat from a low-temperature source and upgrade its
temperature to a useful level. AHPs supplied with waste energy are attractive options but
only if they are correctly implemented. These devices are environmentally friendly as they
use working fluids that do not cause ozone depletion. For the majority of AHPs used in
industrial applications, H2O-LiBr is the working fluid pair of choice because it is not toxic, has
a high enthalpy of vaporisation and does not require a rectification step Srikhirin et al. (2001).
The distinctive feature of the absorption system is that little work input is required because
the pumping process involves a liquid.
3. System design
The seawater-cooled H2O-LiBr AHP system is designed for cooling and heating on board
naval ships. General system specifications are given in Table 1.
The system is not considered under the diesel engine load of 50% as running an engine
under low loads causes low cylinder pressures and consequent poor piston ring sealing,
which relies on the gas pressure to force it against the oil film on the bores to form the seal.
Low cylinder pressures cause poor combustion and low combustion pressures and
temperatures. Ideally, diesel engines should be run at least 50% of their maximum rated
load. When the ship engine is in stand by or very low engine loads (under 50%), the heating
and cooling load for naval surface ship will be met from vapour compression heat pump. In
addition, vapour compression heat pump on board naval ship will be operated as reserve in
emergency situations, in naval base or during periodic engine overhaul in naval shipyard.
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The case naval surface ship has two main internal combustion engines that propel the ship.
The specification of the diesel engine on the naval ship is given in Table 2 Wärtsılä (2013).
The main internal combustion engines emit exhaust gas. The purpose of the exhaust system
is to transport the burned exhaust gases of combustion from the cylinders to the atmosphere
as silently as possible. The system includes exhaust valves and ports, headers and pipes,
main inboard and outboard exhaust valves and engine mufflers. The propeller demand data
are presented in Table 3 Wärtsılä (2013). The total heating and cooling loads of the case
naval surface ship are 144 kW and 116 kW, respectively.
AHP uses water as the refrigerant and a solution of LiBr in water as the absorbent. Solid salt,
such as LiBr, that is dissolved in water becomes a solution. If aqueous solutions of LiBr are
boiled, the vapour produced will become pure water vapour as LiBr is virtually volatile. The
AHP system consists of a generator, absorber, solution heat exchanger, condenser,
evaporator, expansion valves, solution pump, exhaust heat exchanger and supply and return
4/3 rotary valves. The exhaust heat exchanger converts waste heat from engine exhaust into
useable heat for space heating. The solution heat exchanger is used for internal heat
recovery to preheat the solution leaving the absorber. The hot concentrated LiBr solution
leaves the generator to improve system efficiency. The solution is circulated by the solution
pump. The use of a pump prevents crystallisation and reduces submergence in pool boiling
generators. Crystallisation is the solidification of LiBr from the solution, and it can block the
flow of fluid within the unit. When the concentration of salt exceeds the solubility limit,
crystallisation causes the precipitation of the salt component. Solubility is a strong function of
mass fraction and temperature and a weak function of pressure. Once the salt starts to
precipitate, it forms crystals that enhance the possibility of the creation of more crystals. The
formation of crystals will accelerate and continue even when satiety has declined. The
crystals can clog the system and create blockages in the flow. The highest risk for
crystallisation is when the strong solution has been cooled by the solution heat exchanger.
This point is where the concentration is the highest and the temperature is the lowest.
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Superheated refrigerant vapour, which separates from the solution, is generated in the steam
generator by means of the available heating source. The exiting vapour stream is sent to the
condenser. The absorber cooled by seawater acts as the condenser and is the component in
the solution cycle that sucks the vapour that condenses in the solution at low pressure and
low temperature. The final absorption temperature for a given solution concentration
determines the equilibrium exit condition of the solution, which is pumped and preheated
before entering the steam generator. Mass and energy balances are solved for each
component.
The expansion tanks are designed to compensate for the changing volume of the water in
the AHP system to maintain the static pressure created by the pump at the utilisation level in
water production and to compensate for the changes in the water flow rate. A right choice of
expansion tank prevents sudden changes in pressure and provides longer life for the pump
and other elements of the system.
Table 1 General system properties
Type of heat pump H2O-LiBr AHP
Energy source Diesel engine exhaust gas
Diesel engine fuel type NATO F-76 Diesel Heating Mode
Hot water temperature flows
through the condenser 45 °C–40 °C
Evaporator Seawater (-2 °C–+32 °C)
Cooling Mode
Chilled water temperature
flows through the evaporator 7 °C–12 °C
Condenser Seawater (-2 °C–+32 °C)
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Table 2 Specification of a diesel engine on a naval ship
Engine maximum continuous output
3,000 kW
Engine speed 750 rpm Cylinder bore 320 mm Stroke 400 mm Cylinder configuration 6, in-line Backpressure, max. 4.0 kPa
Table 3 Propeller demand data
Engine
% load
Fuel
consumption,
g(kWh)-1
Exhaust gas
temperature
after turbocharger, °C
Exhaust gas flow,
kgs-1
50 191 315 3.71
75 182 345 4.43
85 181 336 4.96
100 185 380 5.40
The AHP system design for naval ship application is presented in Figure 1. Fresh water
conservation on board a naval ship is very important. Seawater is the main coolant on board
naval ships similar to other ships. Seawater is a free and renewable source for cooling on
board ships. Therefore, the condenser, evaporator and absorber used in an AHP system are
seawater-cooled heat exchangers.
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Figure 1 AHP system design for a naval ship
The exhaust heat exchanger is used to recover exhaust waste heat from the diesel engine. A
finned-tube evaporator is selected to enhance the heat transfer rate between exhaust gas
and H2O. The pressure drop of the gas side of the exhaust heat exchanger is set to 3 kPa
maximum and the backpressure of the exhaust gas system is set to 4 kPa maximum.
Seawater flows through the evaporator during the heating mode, through the condenser
during the cooling mode, and through the absorber all the time. To achieve this flow, two 4-
way, 3-position (4/3) rotary valves are used for heating and cooling in the system. One of the
4/3 valves is the supply and the other is the return valve. The three positions of the rotary
valve are the heating mode, closed and cooling mode. The supply and return 4/3 rotary valve
positions are presented in Figures 2 and 3, respectively.
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Figure 2 Supply 4/3 rotary valve positions a) Heating mode, b) Closed, c) Cooling mode
Figure 3 Return 4/3 rotary valve positions a) Heating mode, b) Closed, c) Cooling mode
4. Thermodynamic analysis
The thermodynamic design of the H2O–LiBr AHP system by the first law only is usually based
on given or assumed steady-state operating conditions. The absorber, condenser and
evaporator temperatures are fixed as well as the temperature approach in the solution heat
exchanger. The system generator heat transfer rates are also known.
The data demanded by a fixed pitch propeller used in a displacement hull are given in Table 1.
The fundamental simplifications assumed for the model are as follows:
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- Steady state of the AHP
- No radiation heat transfer
- Water at the condenser outlet is saturated liquid
- Saturated vapour in the absorber
- Water at the evaporator outlet is saturated vapour
- The generator and condenser are assumed to have the same pressure at equilibrium
- The absorber and evaporator are assumed to have the same pressure at equilibrium
- Pressure losses in the pipes and all heat exchangers are negligible
4. 1 Exhaust gas properties
The exhaust gas properties, which include specific heat at constant pressure, dynamic
viscosity and thermal conductivity, should be determined for the heat transfer analysis. The
main components of the exhaust gas of a diesel engine are CO2, H2O, N2 and O2. The mass
fractions of these components vary with the operating condition of the engine. When the
engine operates at steady state, the injected fuel quantity and the intake air amount can be
measured on the engine test bench.
Except for very low engine loads, the exhaust temperature of a marine engine is between 300
and 380 °C, and the exhaust pressure is slightly hi gher than the atmospheric pressure.
Therefore, exhaust gas can be treated as a mixture of ideal gases. The specific enthalpy,
specific heat and density of exhaust gas can be calculated as follows Zhang et al. (2013):
ii
im hmfh ∑=
=4
1
(1)
ipi
imp cmfc ,
4
1, ∑
=
= (2)
∑
∑
=
== 4
1
4
1
/i
iii
iii
m
Mmf
Mmf
ρρ (3)
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4.2 Nominal heat balance The conservation of mass principle is expressed as
eicv mm
dt
dm&& −= . (4)
To obtain a control volume at steady state, the equation is reduced to
∑∑ =e
ei
i mm && . (5)
The energy rate balance is expressed as Moran et al. (2002):
++−
+++−= ∑∑ e
ee
eei
ii
iicvcv
cv gzV
hmgzV
hmWQdt
dE
22
22
&&&& . (6)
To obtain a control volume at steady state and to disregard the changes in the kinetic and
potential energies of the flowing streams from inlet to exit, the equation is reduced to
ee
eii
icvcv hmhmWQ ∑∑ −+−= &&&&0 . (7)
4.3 Thermodynamic properties of the H 2O–LiBr solution LiBr mass balance on absorber is
sssswsws XmXm && = . (8) The concentrations are defined as the ratio of the mass fraction of LiBr in a solution to the total
mass of LiBr and H2O in the solution.
OHmassLiBrmass
LiBrmassX
2+= . (9)
Another characteristic value is the solution circulation ratio (pump rate). The circulation ratio is
defined as the ratio of the mass flow rate of the solution through the pump to the mass flow
rate of the working fluid. Note that f represents the required pumping energy. It can be
expressed in terms of concentrations as follows (Singh et al., 2013; Táboas et al. 2014):
w
ws
m
mf
&
&= . (10)
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A high circulation ratio involves high pump consumption which implies a high electrical
consumption, eliminating the advantage of absorption cycles compared to vapour
compression cycles Táboas et al. (2014).
Many articles have been made over the properties of H2O–LiBr solutions. An often used
formulation of the properties of H2O–LiBr is made by McNeely (1979) valid from 15 to 165 oC.
McNeely (1979) developed polynomial correlations relating solution temperature,
concentration, and vapour pressure and presented a consistent set for inclusion in the
ASHRAE Handbook of Fundamentals. Patek and Klomfar (2006) provided formulation of the
thermodynamic properties of H2O–LiBr solutions in vapour–liquid equilibrium states valid
over a broader range of parameters, from 273 K or from the crystallization temperature
(whichever is greater) up to 500 K in temperatures and over the full range of compositions
and calculated maximum difference of 114 −kJkg of enthalpy according to formulations of
McNeely (1979).
The solution enthalpy in kJ/kg, for range Ct o16515 pp and LiBrX %7040 pp is given as
(McNeely, 1979; Keith and Goswami, 2008; ASHRAE, 2009)
n
nn
n
nn
n
nn XCtXBtXAh ∑∑∑
===
++=4
0
24
0
4
0
. (11)
where t is the solution temperature in °C and X is the solution concentration in %LiBr. The
coefficients A, B and C for solution enthalpy are presented in Table 4 (McNeely, 1979; Keith
and Goswami, 2008; ASHRAE, 2009)
Table 4 Coefficients and exponents of Eq. (11)
i Ai Bi Ci
0 -2.0243x10+3 1.8283x10+1 -3.7008x10-2 1 1.6331x10+2 - 1.1692x100 2.8878x10-3 2 -4.8816x100 3.2480x10-2 -8.1313x10-5 3 6.3029x10-2 -4.0342x10-4 9.9117x10-7 4 -2.9137x10-4 1.8520x10-6 -4.4441x10-9
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The solution, refrigerant temperature and vapour pressure are given for
range Ct o1755 pp ; Ct o11015 pp ′− ; LiBrX %7045 pp in (McNeely, 1979; Keith and
Goswami, 2008; ASHRAE,2009)
The solution temperature, t, in °C is
n
nn
n
nn XBXAtt ∑∑
==
+′=3
0
3
0
. (12)
where t’ is the refrigerant temperature in °C.
Table 5 Coefficients and exponents of Eq. (12)
i Ai Bi
0 -2.0075x100 1.2494x10+2 1 1.6976x10-1 -7.7165x100 2 -3.1334x10-3 1.5229x10-1 3 1.9767x10-5 -7.9509x10-4
Vapour pressure P, in kPa is
2)15.273/(5.095,104)15.273/(49.596,105.7log +′−+′−= ttP . (13)
Tables 4 and 5 give the coefficients and exponents for the Eqs. (11)–(12). For given
temperature and mass fraction range, the maximum deviations of the values calculated from
Eqs. (11)–(12) from the (McNeely, 1979; Keith and Goswami, 2008; ASHRAE,2009) are
13.0 −± kJkg for the solution enthalpy and Co1.0± for the solution temperature. The number
of significant figures given in coefficients in Tables 4 and 5 is necessary and sufficient to
obtain the stated accuracy.
4.4 Generator
The heat transfer rate from the engine exhaust gas to the AHP system vapour generator is
expressed as
( )outexhinexhexhtGexht hhmQQ ,, −== &&& , (14)
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where exhtm& is the engine exhaust gas mass flow rate, inexhh , is the engine exhaust gas
specific enthalpy at the entrance of the vapour generator and outexhh , is the engine exhaust
gas specific enthalpy at the exit of the vapour generator.
The rate of heat transfer to the solution is
778811 hmhmhmQG &&&& −+= . (15)
781 hmhmhmQ wssswG &&&& −+= . (16)
4.5 Absorber
The rate of heat transfer from the absorber is
55441010 hmhmhmQA &&&& −+= . (17)
5410 hmhmhmQ wswssG &&&& −+= . (18)
4.6 Solution pressure restrictor
The enthalpy value at point 9 is determined from a throttling model on the solution flow
restrictor which yields
109 hh = . (19)
4.7 Solution heat exchanger
Assuming that heat losses to the surroundings are negligible, the rate of heat transfer
between the strong and weak solutions is
( ) ( )6798 hhmhhmQ wsssSHX −=−= &&& . (20)
The heat exchanger effectiveness,ε , is expressed as
maxq
q≡ε . (21)
( )( )icih
ohihh
TTC
TTC
,,min
,,
−−
≡ε . (22)
If hCC =min , then
icih
ohih
TT
TT
,,
,,
−−
≡ε . (23)
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The heat exchanger effectiveness,ε , is defined as the ratio of the temperature drop of the
strong solution to the temperature difference between the strong and weak solutions entering
the heat exchanger. The temperature of the strong solution exiting the heat exchanger is
expressed as follows:
( ) 869 1 TTT εε −+= . (24)
( )9867 hhm
mhh
ws
ss −+=&
&. (25)
4.8 Condenser
The rate of heat transfer from the condenser is
( )21 hhmQ wC −= && . (26)
4.9 Water pressure restrictor
The throttling model yields the result that
32 hh = . (27)
4.10 Evaporator
The rate of heat transfer to the evaporator is
( )34 hhmQ wE −= && . (28)
4.11 Solution pump
The power input to the pump is
( )56 hhmW wsp −= && . (29)
The following is an alternative to Eq. 29 for evaluating the pump work:
∫=
6
5vdP
m
W
ws
p
&
&
. (30)
As the specific volume of the liquid normally varies only slightly as liquid flows from the inlet
to the exit of the pump, a plausible approximation to the value of the integral can be obtained
by taking the specific volume at the pump inlet, v5, as constant for the process Moran et
al.(2002).
Then,
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( ) 55656 vPPhh −+= . (31)
The work input to the system in the pump is small and neglected in the calculation of the
coefficient of performance (COP) and efficiency. However; in practice, it is usually estimated
to size the driving motor.
pdm
PvW
η∆=&
& . (32)
4.12 Overall mass and energy balance of the syste m
The energy rate balance at steady state is
0=− cvcv WQ && . (33)
0=+−−+ pCAEG WQQQQ &&&&& . (34)
The mass flow rates are
wmmmmm &&&&& ==== 4321 . (35)
wsmmmm &&&& === 765 . (36)
ssmmmm &&&& === 1098 . (37)
The mass flow rates of water and the weak and strong solutions are expressed as
wsss
ss
wsss
ws
Gw
XX
Xh
XX
Xhh
Qm
−−
−+
=781
&
& . (38)
−=
wsss
sswws XX
Xmm && . (39)
−=
wsss
wswss XX
Xmm && . (40)
4.13 COP
The COP is the most common measurement used to rate heat pump efficiency. The COP of
the cooling cycle is
pumptheforinputworkgeneratortheforinputheat
evaporatoratobtainedcapacitycoolingCOPc +
= . (41)
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pG
Ec WQ
QCOP
&&
&
+= . (42)
The COP of the heating cycle is
pumptheforinputworkgeneratortheforinputheat
condenseratobtainedcapacityheatingCOPh +
= . (43)
pG
Ch WQ
QCOP
&&
&
+= . (44)
4.3 Operating temperatures and pressures
As lower condenser and absorber temperatures increase cycle efficiency, they should be
selected as low as possible. However, in practice, they are more or less fixed by the cooling
water available.
In most applications, the evaporator temperature is usually between 4 °C and 12 °C for the
air conditioning of the space maintained between 24 °C and 27 °C. An evaporator
temperature of 12 °C is sufficient to cool the air, but the evaporator temperature of an actual
absorption cycle has to be designed at 4 °C or 5 °C to absorb excess humidity in the air.
Reduced evaporator temperatures give higher second law efficiency of H2O–LiBr AHP
cycles. Therefore, refrigerant temperature in the evaporator should be designed at or below 4
°C to satisfy both practical requirements and needs of the higher second law efficiency. The
condenser and absorber temperatures depend on the seawater temperature.
To reduce the risk of crystallisation in an H2O–LiBr AHP, the temperature should be
sufficiently high and the concentration sufficiently low.
According to Loydu’s (2007) Rules for the Classification of Naval Ships, the selection, layout
and arrangement of all shipboard machinery, equipment and appliances should ensure
faultless continuous operation under the seawater temperature of -2 °C to +32 °C and the air
temperature outside the ship of -25 °C to +45 °C.
The absorber and evaporator are assumed to have the same pressure at equilibrium,
although a small pressure is necessary in actual processes. Therefore, the evaporator
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pressure or temperature and the absorber temperature define the concentration, Xws, of the
weak solution in the absorber. The condenser and generator are assumed to have the same
pressure at equilibrium. The condenser temperature and the concentration of the strong
solution in the generator, Xss, can determine the generator temperature.
5 Results and comparison
In this study, calculations were performed for diesel engine loads of 50%, 75%, 85% and
100%. The parameters were taken as TE=4 °C, T C= 50 °C, T A=35 °C and TG=90, 95, 100,
105 and 110 °C.
The generator heat transfer rates depend on the engine load. As the engine load increases,
the generator heat increases. The generator, absorber, condenser and evaporator heat
transfer rates increase as the engine load increases. Figure 4 presents the heat transfer
rates versus engine load. Figure 5 shows the COPs versus the generator temperature.
These results are in accordance with those of Seddiek et al. (2012) and Ouadha and El-
Gotni (2013).
0
500
1,000
1,500
2,000
50 75 85 100
Engine load (%)
The
rate
s of
hea
t tra
nsfe
r (k
W)
GeneratorEvaporatorAbsorberCondenser
Figure 4 Rates of heat transfer versus engine load
Figure 4 shows 1841.40 kW of recovered heat can produce 1618.12 kW of heating and
1490.11 kW of cooling for full-load operation.
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0.5
0.6
0.7
0.8
0.9
90 95 100 105 110
Generator temperature ( oC)
CO
P
coolingheating
Figure 5 COPs versus generator temperature
The mass flow rate ratio decreases as the generator temperature increases. Figure 6 shows
the mass flow rate ratio versus the generator temperature. In practice, heat exchanger
effectiveness values typically range from 60% to 80%. The COP increases as the heat
exchanger effectiveness increases. Figure 7 presents the COPs versus heat exchanger
effectiveness.
05
1015202530
90 95 100 105 110
Generator temperature ( oC)
Mas
s flo
w ra
te ra
tio
Figure 6 Mass flow rate ratio versus generator temperature
0.65
0.7
0.75
0.8
0.85
0.9
0.6 0.65 0.7 0.75 0.8
Heat exchanger effectiveness
CO
P CoolingHeating
Figure 7 COPs versus heat exchanger effectiveness
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For a generator temperature of 90 °C, heat exchange r effectiveness of 0.8 and diesel engine
load of 50%, the heat transfer rates of the condenser and evaporator are 778.68 kW and
728.22 kW, respectively.
For a generator temperature of 110 °C, heat exchang er effectiveness of 0.8 and diesel
engine load of 50%, the heat transfer rates of the condenser and evaporator are 906.32 kW
and 834.62 kW, respectively.
The seawater-cooled H2O-LiBr AHP system can meet the entire HVAC load of the case
naval surface ship even when the engine is operating at a load of 50%.
5.1 Comparison between an AHP system and a vapour–c ompression heat pump
system
The case naval surface ship uses a vapour–compression heat pump system for heating and
cooling. Vapour–compression heat pumps generally have COPs of 2–4 and deliver 2–4
times more energy than they consume. The electric motor of the compressor in a vapour–
compression heat pump is fed by a diesel generator set on board the ship. As the energy for
the entire naval ship’s electrical load is produced from diesel generator sets, each electrical
load directly affects the overall fuel economy and emissions.
Fuel consumption and CO2 emissions decrease to produce the same amount of power given
by a vapour–compression heat pump system when an AHP system is used. As the AHP
system is assumed to replace the vapour-compression heat pump system, the results of the
system studies will depend on the COP of the vapour-compression heat pump system.
A reference system consists of a conventional diesel generator, which provides electricity
and heat, and a vapour-compression heat pump, which produces heating and cooling. Based
on the engine load, the heating demand is assumed to be 906.32, 1206.78, 1303.21 and
1618.12 kW, the cooling demand is assumed to be 834.62, 1111.31, 1200.11 and 1490.11
kW, and the COP is assumed to be 2, 3 and 4 for the vapour-compression heat pump. The
electrical efficiency ( DGη ) for the diesel generator is 95%. Power systems for the navies of
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NATO countries operate on naval distillate fuel (NATO symbol F-76). Some values of the
logistic fuel NATO F-76 are presented in Table 6 (Steinfeld et al., 2000; Ezgi et al., 2013).
Table 6 Values of the logistic fuel NATO F-76
Molecular formula (avg) C14.8H26.9
Molecular weight 205
Density, at 15 °C, kgm -3 876
Fuel price, US$gallon-1, (2014) 3.61
Net heating value, Hu, kJkg-1 42,700
The amounts of heating and cooling produced by the AHP in each system can be regarded
to correspond to electricity saving. The saving is calculated by estimating fuel and the
electricity input needed to produce the same heating and cooling effect using the vapour–
compression heat pump.
The saved F-76 diesel fuel consumption (SFC) for heating and cooling can be calculated as
DGu
pc
H
WWSFC
η.
&& −= . (45)
The calculations are made with COPs of 2, 3 and 4 for the vapour–compression heat pump.
Figures 8 and 9 present the SFC versus heating and cooling capacities for a generator
temperature of 110 °C and heat exchanger effectiven ess of 0.8, respectively. Figures 10 and
11 show the reduced CO2 emission versus heating and cooling capacities for a generator
temperature of 110 °C and heat exchanger effectiven ess of 0.8, respectively.
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0
20
40
60
80
906.32 1206.78 1303.21 1618.12
Heating capacity (kW)
The
sav
ed fu
el
cons
umpt
ion
(k
gh-1
)
Vapour-compression heat pump of COP=3 Vapour-compression heat pump of COP=4
Vapour-compression heat pump of COP=2
Figure 8 SFC versus heating capacity
0
10
20
30
40
50
60
70
834.62 1111.31 1200.11 1490.11
Cooling capacity (kW)
The
sav
ed fu
el c
onsu
mpt
ion
(kgh
-1)
Vapour-compression heat pump of COP=3 Vapour-compression heat pump of COP=4
Vapour-compression heat pump of COP=2
Figure 9 SFC versus cooling capacity
At the end of 1,000 operating hours a year of a naval surface ship, the AHP system can use
more than 99.5% less electricity compared with the vapour–compression heat pump for
HVAC. It can save 22,952 L–81,961 L of diesel fuel in the heating cycle and 21,135 L–
75,477 L of diesel fuel in the cooling cycle depending on the engine load and COP of the
vapour–compression heat pump.
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0
50
100
150
200
250
906.32 1206.78 1303.21 1618.12
Heating capacity (kW)
The
redu
ced
CO
2 em
issi
on
(kgh
-1)
Vapour-compression heat pump of COP=3 Vapour-compression heat pump of COP=4Vapour-compression heat pump of COP=2
Figure 10 Reduced CO2 emission versus heating capacity
0
50
100
150
200
250
834.62 1111.31 1200.11 1490.11
Cooling capacity (kW)
The
redu
ced
CO
2 em
issi
on
(kgh
-1)
Vapour-compression heat pump of COP=3 Vapour-compression heat pump of COP=4
Vapour-compression heat pump of COP=2
Figure 11 Reduced CO2 emission versus cooling capacity
At the end of 1,000 operating hours a year of a naval surface ship, the AHP system can
improve the ship’s green profile because it will reduce its annual CO2 emissions by 60.41
tons–215.74 tons in the heating cycle and 55.63 tons–198.67 tons in the cooling cycle
depending on the engine load and COP of the vapour–compression heat pump.
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Figures 12 and 13 show the profitability plotted against the heating and cooling capacities,
respectively. The fuel price is set to US$3.61gallon-1 and the operating hours are 1,000 hours
a year. In this case, the COPs used for the vapour–compression heat pump are 2, 3 and 4.
The AHP system can provide an annual energy savings of US$22,394–US$79,967 in the
heating cycle and US$20,622–US$73,641 in the cooling cycle and has a simple payback
period of about between 2.5 and 9.5 years depending on the engine load and COP of the
vapour–compression heat pump.
0
10,000
20,000
30,000
40,000
50,000
60,000
70,000
80,000
90,000
906.32 1206.78 1303.21 1618.12
Heating capacity (kW)
Pro
fitab
ility
(US
$yea
r-1
)
Vapour-compression heat pump of COP=3 Vapour-compression heat pump of COP=4
Vapour-compression heat pump of COP=2
Figure 12 Profitability of the AHP system as a function of heating capacity
0
10,000
20,000
30,000
40,000
50,000
60,000
70,000
80,000
834.62 1111.31 1200.11 1490.11
Cooling capacity (kW)
Pro
fitab
ility
(US
$yea
r-1
)
Vapour-compression heat pump of COP=3 Vapour-compression heat pump of COP=4
Vapour-compression heat pump of COP=2
Figure 13 Profitability of the AHP system as a function of cooling capacity
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The sound power levels according to product guides of miscellaneous manufacturers’ of
vapour–compression heat pumps are about between 99 and 102 dBA. The noise level of
vapour–compression heat pumps is too high compared with measurements of Cotana and
Nicolini (2003). Noise scale is logarithmic, so a difference of 10 dB (A) make a significant
difference to the human ear.
6 Conclusions
An H2O-LiBr AHP system for a naval surface ship was designed and thermodynamically
analysed. The AHP system was compared with the vapour–compression heat pump system
in a case naval surface ship. The dual use of absorption technology to produce heating and
cooling on board a naval ship was investigated. The results show that the seawater-cooled
H2O-LiBr AHP system not only meets the actual heating and cooling loads of the case naval
surface ship but also provides more. The AHP system is particularly attractive in applications
that have a cooling demand and a source of heat, such as naval surface ships. The AHP
system is needed for the HVAC as waste heat from a running ship engine may be sufficient
to provide enough heat to meet the heating load.
The COP of this cycle versus the generator temperature and engine load was analysed. The
results show that the generator temperature is an important factor to consider the optimum
temperature at which an AHP cycle operates.
The results show that AHP is an environmentally friendly way to produce heating and cooling
as it reduces the use of an electrically driven heat pump in the energy system and thus
global CO2 emissions. Moreover, as water can be used as the refrigerant in AHP, the
problem of environmentally harmful refrigerants used in vapour–compression heat pumps is
avoided. When using the AHP system instead of the vapour–compression heat pump system
on naval surface ships,
• the consumption of electricity can be reduced.
• the CO2 emissions will be lowered by the reduced electricity consumption.
• the AHP system can become an economical alternative.
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• the problem of environmentally harmful refrigerants used in vapour–compression heat
pump is avoided as water can be used as the refrigerant in AHP.
• the AHP system with exhaust gas can suppress the IR signature. Through IR
signature suppression, the ship's susceptibility can be dramatically reduced.
• acoustic signature can be reduced.
Although exhaust heat-driven AHP systems are the perfect choice for naval ships, they are
prone to corrosion. Therefore, in future studies, material selection and corrosion inhibitors for
seawater-cooled AHP system application should be investigated.
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Acknowledgments
The views and conclusions contained herein are those of the author and should not be
interpreted as necessarily representing official policies or endorsements, either expressed or
implied, of any affiliated organisation or government. I wish to thank Mech. Eng. Azize Ezgi
for helpful suggestions and critical comments.
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• An exhaust gas-driven seawater cooled H2O–LiBr AHP system is investigated.
• The analysis of the diesel engine exhaust recovery on board a naval ship is
presented.
• The dual use of absorption technology to produce heating and cooling is investigated.
• An AHP system is compared with a vapour–compression heat pump.
• The saved diesel fuel, reduced CO2 emissions and profitability of AHP system are
calculated.
