- Email: [email protected]
Design and thermodynamic analysis of an H2O–LiBr AHP system for naval surface ship application
Accepted Manuscript Design and ThermodynamicAnalysis of an H2O–LiBr AHP system for naval surface ship 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 AHP system for naval surface ship application, International Journal of Refrigeration (2014), doi: 10.1016/ j.ijrefrig.2014.08.016. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
ACCEPTED MANUSCRIPT
RI PT
Design and Thermodynamic Analysis of an H2O–LiBr AHP System for Naval Surface Ship Application
Cüneyt Ezgi
SC
Mechanical Engineering Department, Turkish Naval Academy, Istanbul, 34942, Turkey Corresponding author tel.: +90 530 606 53 95
M AN U
E-mail address: [email protected]
Abstract
TE D
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
EP
ship’s thermal load is removed through the heating, ventilation and air conditioning (HVAC)
AC C
system. In this study, design and thermodynamic analysis of a water-lithium bromide (H2OLiBr) 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
ACCEPTED MANUSCRIPT Nomenclature Variables
& c p ,W K-1 flow stream heat capacity, m
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
ρ
ε
efficiency,-
SC M AN U
TE D
AC C
η
EP
Greek letters
RI PT
C
density, kgm-3
effectiveness,-
Subscripts A
absorber
C
condenser
cv
control volume
dm
driving motor
E
evaporator
2
ACCEPTED MANUSCRIPT 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
SC M AN U
TE D
EP
Abbreviations
RI PT
e
Absorption heat pump
AHU
Air handling unit
AC C
AHP
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
3
ACCEPTED MANUSCRIPT 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.
RI PT
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
SC
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
M AN U
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
TE D
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
EP
(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
AC C
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.
4
ACCEPTED MANUSCRIPT 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
RI PT
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
SC
heat from a diesel engine. However, until now, the technique has been confined to landbased installations. AHP systems are particularly attractive in applications that have a
M AN U
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
TE D
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.
EP
(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
AC C
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
5
ACCEPTED MANUSCRIPT levels were over 50 dB only for frequency components equal and higher than 400Hz. Aweighted 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
RI PT
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
SC
H2O-LiBr AHP technology can effectively eliminate these disadvantages.
In the literature, although the absorption cycle is most commonly used for refrigeration in
M AN U
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 Moreover, no investigation has been conducted yet on a seawater-
TE D
for fishing vessels.
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. HVAC architecture and system selection for naval surface ships
EP
2.
The HVAC system of a naval ship is a vital part of the overall ship thermal management
AC C
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
6
ACCEPTED MANUSCRIPT 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
performance in the extreme environment of warship duty.
RI PT
by military specifications, which ensure an excellent air-conditioning system and equipment
Design conditions for naval surface ships have been established as a compromise. These
machinery, weapons, electronics and personnel.
SC
conditions consider the large cooling plants required for internal heat loads generated by
The cooling load consists of the following ASHRAE (2011):
M AN U
• Solar radiation
• Heat transmission through hull, decks and bulkheads • Heat (latent and sensible) dissipation of occupants • Heat gain from lights
TE D
• Heat (latent and sensible) gain from ventilation air
• Heat gain from motors or other electrical equipment • Heat gain from piping, machinery and equipment
EP
The heating load consists of the following ASHRAE (2011): • Heat losses through hull, decks and bulkheads
AC C
• 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
7
ACCEPTED MANUSCRIPT 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
RI PT
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 hightemperature heat input to extract heat from a low-temperature source and upgrade its
SC
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
M AN U
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
3.
System design
TE D
the pumping process involves a liquid.
The seawater-cooled H2O-LiBr AHP system is designed for cooling and heating on board
EP
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
AC C
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.
8
ACCEPTED MANUSCRIPT 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
RI PT
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.
SC
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
M AN U
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
TE D
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
EP
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
AC C
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.
9
ACCEPTED MANUSCRIPT 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
RI PT
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.
SC
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
M AN U
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.
TE D
Table 1 General system properties Type of heat pump
H2O-LiBr AHP
Energy source
Diesel engine exhaust gas
EP
Diesel engine fuel type NATO F-76 Diesel Heating Mode
AC C
Hot water temperature flows 45 °C–40 °C
through the condenser Evaporator
Seawater (-2 °C–+32 °C) Cooling Mode
Chilled water temperature 7 °C–12 °C flows through the evaporator Condenser
Seawater (-2 °C–+32 °C)
10
ACCEPTED MANUSCRIPT Table 2 Specification of a diesel engine on a naval ship 3,000 kW 750 rpm 320 mm 400 mm 6, in-line 4.0 kPa
Fuel
Exhaust gas
consumption,
temperature
Exhaust gas flow,
% load g(kWh) 191
75
182
85
181
100
185
kgs-1
after turbocharger, °C 315
3.71
345
4.43
336
4.96
380
5.40
TE D
50
-1
M AN U
Engine
SC
Table 3 Propeller demand data
RI PT
Engine maximum continuous output Engine speed Cylinder bore Stroke Cylinder configuration Backpressure, max.
EP
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
AC C
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.
11
SC
RI PT
ACCEPTED MANUSCRIPT
M AN U
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
TE D
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-
EP
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
AC C
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.
12
RI PT
ACCEPTED MANUSCRIPT
EP
TE D
M AN U
SC
Figure 2 Supply 4/3 rotary valve positions a) Heating mode, b) Closed, c) Cooling mode
4.
AC C
Figure 3 Return 4/3 rotary valve positions a) Heating mode, b) Closed, c) Cooling mode
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:
13
ACCEPTED MANUSCRIPT - Steady state of the AHP - No radiation heat transfer - Water at the condenser outlet is saturated liquid - Saturated vapour in the absorber
RI PT
- 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
SC
- Pressure losses in the pipes and all heat exchangers are negligible
4. 1 Exhaust gas properties
M AN U
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
TE D
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
EP
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,
4
AC C
specific heat and density of exhaust gas can be calculated as follows Zhang et al. (2013):
hm = ∑ mf i hi i =1
(1)
4
c p , m = ∑ mf i c p ,i
(2)
i =1
4
ρm =
∑ mf M i
i =1
(3)
4
∑ mf M i =1
i
i
i
/ ρi
14
ACCEPTED MANUSCRIPT 4.2 Nominal heat balance The conservation of mass principle is expressed as
dmcv = m& i − m& e . dt
(4)
∑ m& = ∑ m& i
i
e
.
e
The energy rate balance is expressed as Moran et al. (2002):
SC
2 2 dEcv Vi Ve & & = Qcv − Wcv + ∑ m& i hi + + gzi − ∑ m& e he + + gze . dt 2 2 i e
RI PT
To obtain a control volume at steady state, the equation is reduced to (5)
(6)
M AN U
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
0 = Q& cv − W&cv + ∑ m& i hi − ∑ m& e he . i
e
(7)
4.3 Thermodynamic properties of the H2O–LiBr solution
m& ws X ws = m& ss X ss .
TE D
LiBr mass balance on absorber is
(8)
The concentrations are defined as the ratio of the mass fraction of LiBr in a solution to the total
mass LiBr . mass LiBr + mass H 2O
(9)
AC C
X =
EP
mass of LiBr and H2O in the solution.
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):
f =
m& ws . m& w
(10)
15
ACCEPTED MANUSCRIPT 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
McNeely
(1979)
developed
polynomial
correlations
RI PT
formulation of the properties of H2O–LiBr is made by McNeely (1979) valid from 15 to 165 oC. 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
SC
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
M AN U
(whichever is greater) up to 500 K in temperatures and over the full range of compositions and calculated maximum difference of 14 kJkg −1 of enthalpy according to formulations of McNeely (1979).
The solution enthalpy in kJ/kg, for range 15 p t p 165 oC and 40 p X p 70% LiBr is given as
4
4
n=0
n=0
TE D
(McNeely, 1979; Keith and Goswami, 2008; ASHRAE, 2009) 4
h = ∑ An X n + t ∑ Bn X n + t 2 ∑ Cn X n .
(11)
n =0
EP
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
AC C
and Goswami, 2008; ASHRAE, 2009)
Table 4 Coefficients and exponents of Eq. (11)
i 0 1 2 3 4
Ai -2.0243x10+3 1.6331x10+2 -4.8816x100 6.3029x10-2 -2.9137x10-4
Bi 1.8283x10+1 - 1.1692x100 3.2480x10-2 -4.0342x10-4 1.8520x10-6
16
Ci -3.7008x10-2 2.8878x10-3 -8.1313x10-5 9.9117x10-7 -4.4441x10-9
ACCEPTED MANUSCRIPT The
solution,
refrigerant
temperature
and
vapour
pressure
are
given
for
range 5 p t p 175 oC ; − 15 p t ′ p 110 oC ; 45 p X p 70 % LiBr in (McNeely, 1979; Keith and Goswami, 2008; ASHRAE,2009) The solution temperature, t, in °C is 3
3
n=0
n=0
RI PT
t = t ′∑ An X n + ∑ Bn X n . where t’ is the refrigerant temperature in °C.
(12)
Ai -2.0075x100 1.6976x10-1 -3.1334x10-3 1.9767x10-5
Vapour pressure P, in kPa is
Bi 1.2494x10+2 -7.7165x100 1.5229x10-1 -7.9509x10-4
M AN U
i 0 1 2 3
SC
Table 5 Coefficients and exponents of Eq. (12)
log P = 7.05 − 1,596.49 /(t ′ + 273.15) − 104,095.5 /(t ′ + 273.15) 2 .
(13)
TE D
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
EP
± 0.3 kJkg −1 for the solution enthalpy and ± 0.1 oC for the solution temperature. The number of significant figures given in coefficients in Tables 4 and 5 is necessary and sufficient to
AC C
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
Q& exht = Q& G = m& exht (hexh,in − hexh, out ) ,
(14)
17
ACCEPTED MANUSCRIPT & exht is the engine exhaust gas mass flow rate, hexh,in is the engine exhaust gas where m specific enthalpy at the entrance of the vapour generator and hexh,out is the engine exhaust gas specific enthalpy at the exit of the vapour generator. The rate of heat transfer to the solution is
RI PT
Q& G = m& 1h1 + m& 8h8 − m& 7 h7 . Q& G = m& w h1 + m& ss h8 − m& ws h7 . 4.5 Absorber
SC
The rate of heat transfer from the absorber is
Q& G = m& ss h10 + m& w h4 − m& ws h5 . 4.6 Solution pressure restrictor
M AN U
Q& A = m& 10 h10 + m& 4 h4 − m& 5h5 .
(15) (16)
(17) (18)
The enthalpy value at point 9 is determined from a throttling model on the solution flow restrictor which yields
TE D
h9 = h10 .
(19)
4.7 Solution heat exchanger
Assuming that heat losses to the surroundings are negligible, the rate of heat transfer
EP
between the strong and weak solutions is
Q& SHX = m& ss (h8 − h9 ) = m& ws (h7 − h6 ) .
AC C
(20)
The heat exchanger effectiveness, ε , is expressed as
ε≡
ε≡
q
qmax
.
(21)
Ch (Th, i − Th , o ) . Cmin (Th, i − Tc ,i )
(22)
If Cmin = Ch , then
ε≡
Th, i − Th , o . Th ,i − Tc ,i
(23)
18
ACCEPTED MANUSCRIPT 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:
T9 = εT6 + (1 − ε )T8 .
RI PT
m& ss (h8 − h9 ) . m& ws
4.8 Condenser The rate of heat transfer from the condenser is
4.9 Water pressure restrictor
M AN U
Q& C = m& w (h1 − h2 ) .
SC
h7 = h6 +
(24) (25)
(26)
The throttling model yields the result that
h2 = h3 . 4.10 Evaporator
(27)
Q& E = m& w (h4 − h3 ) . 4.11 Solution pump
TE D
The rate of heat transfer to the evaporator is
(28)
(29)
AC C
W& p = m& ws (h6 − h5 ) .
EP
The power input to the pump is
The following is an alternative to Eq. 29 for evaluating the pump work:
W& p m& ws
6 = vdP . ∫ 5
(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,
19
ACCEPTED MANUSCRIPT h6 = h5 + (P6 − P5 )v5 .
(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
v&∆P W&dm = .
ηp
4.12 Overall mass and energy balance of the system The energy rate balance at steady state is
SC
Q& cv − W& cv = 0 .
RI PT
to size the driving motor.
M AN U
Q& G + Q& E − Q& A − Q& C + W& p = 0 . The mass flow rates are
m& 1 = m& 2 = m& 3 = m& 4 = m& w . m& 5 = m& 6 = m& 7 = m& ws .
TE D
m& 8 = m& 9 = m& 10 = m& ss .
(32)
(33) (34)
(35) (36) (37)
The mass flow rates of water and the weak and strong solutions are expressed as
Q& G
X ws X ss h1 + h8 − h7 X ss − X ws X ss − X ws
.
EP
m& w =
(38)
.
(39)
X ws m& ss = m& w X ss − X ws
.
(40)
AC C
X ss m& ws = m& w X ss − X ws
4.13 COP The COP is the most common measurement used to rate heat pump efficiency. The COP of the cooling cycle is
COPc =
cooling capacity obtained at evaporator . heat input for the generator + work input for the pump
20
(41)
ACCEPTED MANUSCRIPT Q& COPc = & E . QG + W& p
(42)
The COP of the heating cycle is
heating capacity obtained at condenser . heat input for the generator + work input for the pump
Q& COPh = & C . QG + W& p
(44)
SC
4.3 Operating temperatures and pressures
(43)
RI PT
COPh =
As lower condenser and absorber temperatures increase cycle efficiency, they should be
M AN U
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
TE D
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
EP
°C to satisfy both practical requirements and needs of the higher second law efficiency. The condenser and absorber temperatures depend on the seawater temperature.
AC C
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
21
ACCEPTED MANUSCRIPT 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. Results and comparison
RI PT
5
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.
SC
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
M AN U
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-
TE D
2,000 1,500 1,000 500 0
EP
The rates of heat transfer (kW)
Gotni (2013).
AC C
50
75
Generator Evaporator Absorber Condenser
85
100
Engine load (%)
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.
22
ACCEPTED MANUSCRIPT 0.9 COP
0.8 0.7
cooling heating
0.6 0.5 90
95
100
105
110
o
RI PT
Generator temperature ( C)
Figure 5 COPs versus generator temperature
SC
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
M AN U
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
30 25 20 15 10 5 0 90
TE D
Mass flow rate ratio
effectiveness.
95
100
105
110
o
Generator temperature ( C)
AC C
EP
Figure 6 Mass flow rate ratio versus generator temperature
0.9
COP
0.85 0.8
Cooling Heating
0.75 0.7
0.65 0.6
0.65
0.7
0.75
0.8
Heat exchanger effectiveness
Figure 7 COPs versus heat exchanger effectiveness
23
ACCEPTED MANUSCRIPT 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
RI PT
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%.
SC
5.1 Comparison between an AHP system and a vapour–compression heat pump system
M AN U
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
TE D
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
EP
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
AC C
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 (η DG ) for the diesel generator is 95%. Power systems for the navies of
24
ACCEPTED MANUSCRIPT 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 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
M AN U
SC
RI PT
Molecular formula (avg)
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.
SFC =
W&c − W& p H u .η DG
.
TE D
The saved F-76 diesel fuel consumption (SFC) for heating and cooling can be calculated as (45)
EP
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
AC C
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.
25
80 60 40 20 0 906.32
1206.78
1303.21
1618.12
RI PT
The saved fuel consumption -1 (kgh )
ACCEPTED MANUSCRIPT
Heating capacity (kW) Vapour-compression heat pump of COP=3
Vapour-compression heat pump of COP=4
Vapour-compression heat pump of COP=2
M AN U
70 60 50 40 30 20 10 0 834.62
TE D
The saved fuel consumption -1 (kgh )
SC
Figure 8 SFC versus heating capacity
1111.31
1200.11
1490.11
EP
Cooling capacity (kW)
AC C
Vapour-compression heat pump of COP=3 Vapour-compression heat pump of COP=2
Vapour-compression heat pump of COP=4
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.
26
250 200 150 100 50 0 906.32
1206.78
1303.21
Heating capacity (kW)
1618.12
Vapour-compression heat pump of COP=4
SC
Vapour-compression heat pump of COP=3 Vapour-compression heat pump of COP=2
RI PT
(kgh -1)
The reduced CO 2 emission
ACCEPTED MANUSCRIPT
M AN U
Figure 10 Reduced CO2 emission versus heating capacity
150
TE D
(kgh )
200
-1
100 50 0
EP
The reduced CO 2 emission
250
834.62
1111.31
1200.11
1490.11
Cooling capacity (kW)
AC C
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.
27
ACCEPTED MANUSCRIPT 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
RI PT
heating cycle and US$20,622–US$73,641 in the cooling cycle and has a simple payback period of about between 2.5 and 9.5 years depending on the engine load and COP of the vapour–compression heat pump.
SC
80,000 70,000 60,000 50,000
M AN U
Profitability (US$year-1)
90,000
40,000 30,000 20,000 10,000 0 906.32
1206.78
1303.21
1618.12
Heating capacity (kW)
TE D
Vapour-compression heat pump of COP=3
Vapour-compression heat pump of COP=4
Vapour-compression heat pump of COP=2
70,000
AC C
Profitability (US$year-1)
80,000
EP
Figure 12 Profitability of the AHP system as a function of heating capacity
60,000 50,000 40,000 30,000 20,000 10,000
0 834.62
1111.31
1200.11
1490.11
Cooling capacity (kW) 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
28
ACCEPTED MANUSCRIPT 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
6
RI PT
difference to the human ear.
Conclusions
An H2O-LiBr AHP system for a naval surface ship was designed and thermodynamically
SC
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
M AN U
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
TE D
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
EP
results show that the generator temperature is an important factor to consider the optimum temperature at which an AHP cycle operates.
AC C
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.
29
ACCEPTED MANUSCRIPT •
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.
RI PT
•
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.
SC
References
ASHRAE, 2009. Handbook Fundamentals, SI Edition, ASHRAE Inc., Atlanta, GA.
ASHRAE Inc., Atlanta, GA.
M AN U
ASHRAE, 2011 Handbook Heating, Ventilating, and Air-conditioning Applications, SI Edition,
Bakhtiari B, Fradette L, Legros R, Paris J., 2011. A model for analysis and design of H2O– LiBr absorption heat pumps, Energy Convers Manage;52:1439–48.
TE D
Cotana, F., Nicolini, A.,2003. Noise Impact of Absorption Machines for Civil Applications, The 32nd International Congress and Exposition on Noise Control Engineering, Jeju International Convention Center, Seogwipo, Korea,
EP
Ezgi, C., Turhan Çoban, M., and Selvi, Ö., 2013. Design and thermodynamic analysis of an SOFC system for naval surface ship application. ASME J.Fuel Cell Sci. Technol.
AC C
10(3), 031006.
Fernandez-Seara, J.,Vales, A., Vazquez, M.,1998. Heat recovery system to power an onboard NH3-H2O absorption refrigeration plant in trawler chiller fishing vessels, Appl. Therm. Eng.18 1189-1205
Frank, M.V. and Helmick D.,2007. 21st Century HVAC System for Future Naval Surface Combatants-Concept Development Report, NAVSEA Philadelphia Naval Surface Warfare Center, Philadelphia. IMO, 2010. Prevention of Air Pollution from Ships: Assessment of IMO Efficiency Measures for the Control of GHG Emissions from Ships, London.
30
ACCEPTED MANUSCRIPT Keith, F., Goswami, D.Y., 2008. Energy Management and Conservation Handbook. CRC Press, Boca Raton. Liang, Y., Shu, G., Tian, H., Liang, X., Wei, H., Liu, L., 2013. Analysis of an electricity– cooling cogeneration system based on RC–ARS combined cycle aboard ship, Energy
RI PT
Convers Manage, 76, 1053–1060. Loydu, T., 2007. Rules for the Classification of Naval Ships Ship Operation Installations and Auxiliary Systems, Istanbul, Turkey.
McNeely, L.A. 1979. Thermodynamic properties of aqueous solution of lithium bromide,
SC
ASHRAE Transactions 85(1):413.
Moran, M. J., Shapiro, H.N., Munson, B.R., DeWitt, D.P., 2002. Introduction to Thermal
John & Sons, USA.
M AN U
Systems Engineering: Thermodynamics, Fluid Mechanics, and Heat Transfer. Wiley,
Patek, J., Klomfar, J., 2006. A computationally effective formulation of the thermodynamic properties of LiBr–H2O solutions from 273 to 500 K over full composition range, Int. J.
TE D
Refrigeration. 29, 566–578
Ruiz, V.,2012. Analysis of existing refrigeration plants onboard fishing vessels and improvement possibilities, Second International Symposium on Fishing Vessel
EP
Energy Efficiency, E-Fishing, Vigo, Spain. Ouadha A., El-Gotni, Y., 2013. Integration of an ammonia–water absorption refrigeration
AC C
system with a marine diesel engine: A thermodynamic study. Procedia Comput Sci. 19,754 –761.
Seddiek, I.S., Mosleh, M., Banawan, A. A., 2012. Thermo-economic approach for absorption air condition onboard high-speed crafts. Inter. J. Nav. Archit. Oc. Engng. 4 460–476.
Singh, Y., Kumar, D., Kumar, A., Singla, A., et al., 2013. Mathematical modeling and analysis of absorption refrigeration system using waste heat of diesel genset. Int. J. Emerg. Manag. Res. Manag & Tech. 46–50. Srikhirin P, Aphornratana S, Chungpaibulpatana S.,2001. A review of absorption refrigeration technologies, Renew Sustain Energy Rev, 5:343-72.
31
ACCEPTED MANUSCRIPT Steinfeld, G., Sanderson, R., Ghezel-Ayagh, H., Abens, S., and Cervi, M. C., 2000. Distillate Fuel Processing for Marine Fuel Cell Applications. Proceedings of the AICHE Spring 2000 Meeting, Atlanta, GA, March 5–9. Sun, J., Fu, L., Zhang, S., Hou, W., 2010. A mathematical model with experiments of single
RI PT
effect absorption heat pump using LiBr-H2O. Appl. Therm. Eng. 30 2753–2762. Táboas, F., Bourouis, M., Vallès, M., 2014. Analysis of ammonia/water and ammonia/salt mixture absorption cycles for refrigeration purposes in fishing ships, Appl. Therm. Eng. 66,603-611
SC
Wang, K., Abdelaziz, O., Kisari, P., Vineyard, E.A., 2011. State-of-the-art review on
Refrigeration. 34,1325–1337.
M AN U
crystallization control technologies for water/LiBr absorption heat pumps. Int. J.
Wang, S.G.,Wang R.Z., 2005. Recent developments of refrigeration technology in fishing vessels, Renew Energ 30,589–600 Wärtsılä 32 Product Guide, 2013, Vaasa.
TE D
Zhang, H.G., Wang, E.H., Fan, B.Y., 2013. Heat transfer analysis of a finned-tube evaporator for engine exhaust heat recovery. Energy Convers Manage, 65, 438–447.
EP
Acknowledgments
The views and conclusions contained herein are those of the author and should not be
AC C
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.
32
ACCEPTED MANUSCRIPT HIGHLIGHTS •
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
RI PT
•
AC C
EP
TE D
M AN U
SC
calculated.
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 AHP system for naval surface ship application, International Journal of Refrigeration (2014), doi: 10.1016/ j.ijrefrig.2014.08.016. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
ACCEPTED MANUSCRIPT
RI PT
Design and Thermodynamic Analysis of an H2O–LiBr AHP System for Naval Surface Ship Application
Cüneyt Ezgi
SC
Mechanical Engineering Department, Turkish Naval Academy, Istanbul, 34942, Turkey Corresponding author tel.: +90 530 606 53 95
M AN U
E-mail address: [email protected]
Abstract
TE D
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
EP
ship’s thermal load is removed through the heating, ventilation and air conditioning (HVAC)
AC C
system. In this study, design and thermodynamic analysis of a water-lithium bromide (H2OLiBr) 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
ACCEPTED MANUSCRIPT Nomenclature Variables
& c p ,W K-1 flow stream heat capacity, m
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
ρ
ε
efficiency,-
SC M AN U
TE D
AC C
η
EP
Greek letters
RI PT
C
density, kgm-3
effectiveness,-
Subscripts A
absorber
C
condenser
cv
control volume
dm
driving motor
E
evaporator
2
ACCEPTED MANUSCRIPT 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
SC M AN U
TE D
EP
Abbreviations
RI PT
e
Absorption heat pump
AHU
Air handling unit
AC C
AHP
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
3
ACCEPTED MANUSCRIPT 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.
RI PT
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
SC
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
M AN U
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
TE D
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
EP
(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
AC C
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.
4
ACCEPTED MANUSCRIPT 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
RI PT
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
SC
heat from a diesel engine. However, until now, the technique has been confined to landbased installations. AHP systems are particularly attractive in applications that have a
M AN U
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
TE D
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.
EP
(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
AC C
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
5
ACCEPTED MANUSCRIPT levels were over 50 dB only for frequency components equal and higher than 400Hz. Aweighted 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
RI PT
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
SC
H2O-LiBr AHP technology can effectively eliminate these disadvantages.
In the literature, although the absorption cycle is most commonly used for refrigeration in
M AN U
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 Moreover, no investigation has been conducted yet on a seawater-
TE D
for fishing vessels.
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. HVAC architecture and system selection for naval surface ships
EP
2.
The HVAC system of a naval ship is a vital part of the overall ship thermal management
AC C
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
6
ACCEPTED MANUSCRIPT 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
performance in the extreme environment of warship duty.
RI PT
by military specifications, which ensure an excellent air-conditioning system and equipment
Design conditions for naval surface ships have been established as a compromise. These
machinery, weapons, electronics and personnel.
SC
conditions consider the large cooling plants required for internal heat loads generated by
The cooling load consists of the following ASHRAE (2011):
M AN U
• Solar radiation
• Heat transmission through hull, decks and bulkheads • Heat (latent and sensible) dissipation of occupants • Heat gain from lights
TE D
• Heat (latent and sensible) gain from ventilation air
• Heat gain from motors or other electrical equipment • Heat gain from piping, machinery and equipment
EP
The heating load consists of the following ASHRAE (2011): • Heat losses through hull, decks and bulkheads
AC C
• 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
7
ACCEPTED MANUSCRIPT 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
RI PT
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 hightemperature heat input to extract heat from a low-temperature source and upgrade its
SC
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
M AN U
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
3.
System design
TE D
the pumping process involves a liquid.
The seawater-cooled H2O-LiBr AHP system is designed for cooling and heating on board
EP
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
AC C
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.
8
ACCEPTED MANUSCRIPT 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
RI PT
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.
SC
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
M AN U
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
TE D
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
EP
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
AC C
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.
9
ACCEPTED MANUSCRIPT 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
RI PT
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.
SC
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
M AN U
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.
TE D
Table 1 General system properties Type of heat pump
H2O-LiBr AHP
Energy source
Diesel engine exhaust gas
EP
Diesel engine fuel type NATO F-76 Diesel Heating Mode
AC C
Hot water temperature flows 45 °C–40 °C
through the condenser Evaporator
Seawater (-2 °C–+32 °C) Cooling Mode
Chilled water temperature 7 °C–12 °C flows through the evaporator Condenser
Seawater (-2 °C–+32 °C)
10
ACCEPTED MANUSCRIPT Table 2 Specification of a diesel engine on a naval ship 3,000 kW 750 rpm 320 mm 400 mm 6, in-line 4.0 kPa
Fuel
Exhaust gas
consumption,
temperature
Exhaust gas flow,
% load g(kWh) 191
75
182
85
181
100
185
kgs-1
after turbocharger, °C 315
3.71
345
4.43
336
4.96
380
5.40
TE D
50
-1
M AN U
Engine
SC
Table 3 Propeller demand data
RI PT
Engine maximum continuous output Engine speed Cylinder bore Stroke Cylinder configuration Backpressure, max.
EP
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
AC C
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.
11
SC
RI PT
ACCEPTED MANUSCRIPT
M AN U
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
TE D
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-
EP
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
AC C
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.
12
RI PT
ACCEPTED MANUSCRIPT
EP
TE D
M AN U
SC
Figure 2 Supply 4/3 rotary valve positions a) Heating mode, b) Closed, c) Cooling mode
4.
AC C
Figure 3 Return 4/3 rotary valve positions a) Heating mode, b) Closed, c) Cooling mode
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:
13
ACCEPTED MANUSCRIPT - Steady state of the AHP - No radiation heat transfer - Water at the condenser outlet is saturated liquid - Saturated vapour in the absorber
RI PT
- 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
SC
- Pressure losses in the pipes and all heat exchangers are negligible
4. 1 Exhaust gas properties
M AN U
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
TE D
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
EP
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,
4
AC C
specific heat and density of exhaust gas can be calculated as follows Zhang et al. (2013):
hm = ∑ mf i hi i =1
(1)
4
c p , m = ∑ mf i c p ,i
(2)
i =1
4
ρm =
∑ mf M i
i =1
(3)
4
∑ mf M i =1
i
i
i
/ ρi
14
ACCEPTED MANUSCRIPT 4.2 Nominal heat balance The conservation of mass principle is expressed as
dmcv = m& i − m& e . dt
(4)
∑ m& = ∑ m& i
i
e
.
e
The energy rate balance is expressed as Moran et al. (2002):
SC
2 2 dEcv Vi Ve & & = Qcv − Wcv + ∑ m& i hi + + gzi − ∑ m& e he + + gze . dt 2 2 i e
RI PT
To obtain a control volume at steady state, the equation is reduced to (5)
(6)
M AN U
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
0 = Q& cv − W&cv + ∑ m& i hi − ∑ m& e he . i
e
(7)
4.3 Thermodynamic properties of the H2O–LiBr solution
m& ws X ws = m& ss X ss .
TE D
LiBr mass balance on absorber is
(8)
The concentrations are defined as the ratio of the mass fraction of LiBr in a solution to the total
mass LiBr . mass LiBr + mass H 2O
(9)
AC C
X =
EP
mass of LiBr and H2O in the solution.
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):
f =
m& ws . m& w
(10)
15
ACCEPTED MANUSCRIPT 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
McNeely
(1979)
developed
polynomial
correlations
RI PT
formulation of the properties of H2O–LiBr is made by McNeely (1979) valid from 15 to 165 oC. 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
SC
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
M AN U
(whichever is greater) up to 500 K in temperatures and over the full range of compositions and calculated maximum difference of 14 kJkg −1 of enthalpy according to formulations of McNeely (1979).
The solution enthalpy in kJ/kg, for range 15 p t p 165 oC and 40 p X p 70% LiBr is given as
4
4
n=0
n=0
TE D
(McNeely, 1979; Keith and Goswami, 2008; ASHRAE, 2009) 4
h = ∑ An X n + t ∑ Bn X n + t 2 ∑ Cn X n .
(11)
n =0
EP
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
AC C
and Goswami, 2008; ASHRAE, 2009)
Table 4 Coefficients and exponents of Eq. (11)
i 0 1 2 3 4
Ai -2.0243x10+3 1.6331x10+2 -4.8816x100 6.3029x10-2 -2.9137x10-4
Bi 1.8283x10+1 - 1.1692x100 3.2480x10-2 -4.0342x10-4 1.8520x10-6
16
Ci -3.7008x10-2 2.8878x10-3 -8.1313x10-5 9.9117x10-7 -4.4441x10-9
ACCEPTED MANUSCRIPT The
solution,
refrigerant
temperature
and
vapour
pressure
are
given
for
range 5 p t p 175 oC ; − 15 p t ′ p 110 oC ; 45 p X p 70 % LiBr in (McNeely, 1979; Keith and Goswami, 2008; ASHRAE,2009) The solution temperature, t, in °C is 3
3
n=0
n=0
RI PT
t = t ′∑ An X n + ∑ Bn X n . where t’ is the refrigerant temperature in °C.
(12)
Ai -2.0075x100 1.6976x10-1 -3.1334x10-3 1.9767x10-5
Vapour pressure P, in kPa is
Bi 1.2494x10+2 -7.7165x100 1.5229x10-1 -7.9509x10-4
M AN U
i 0 1 2 3
SC
Table 5 Coefficients and exponents of Eq. (12)
log P = 7.05 − 1,596.49 /(t ′ + 273.15) − 104,095.5 /(t ′ + 273.15) 2 .
(13)
TE D
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
EP
± 0.3 kJkg −1 for the solution enthalpy and ± 0.1 oC for the solution temperature. The number of significant figures given in coefficients in Tables 4 and 5 is necessary and sufficient to
AC C
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
Q& exht = Q& G = m& exht (hexh,in − hexh, out ) ,
(14)
17
ACCEPTED MANUSCRIPT & exht is the engine exhaust gas mass flow rate, hexh,in is the engine exhaust gas where m specific enthalpy at the entrance of the vapour generator and hexh,out is the engine exhaust gas specific enthalpy at the exit of the vapour generator. The rate of heat transfer to the solution is
RI PT
Q& G = m& 1h1 + m& 8h8 − m& 7 h7 . Q& G = m& w h1 + m& ss h8 − m& ws h7 . 4.5 Absorber
SC
The rate of heat transfer from the absorber is
Q& G = m& ss h10 + m& w h4 − m& ws h5 . 4.6 Solution pressure restrictor
M AN U
Q& A = m& 10 h10 + m& 4 h4 − m& 5h5 .
(15) (16)
(17) (18)
The enthalpy value at point 9 is determined from a throttling model on the solution flow restrictor which yields
TE D
h9 = h10 .
(19)
4.7 Solution heat exchanger
Assuming that heat losses to the surroundings are negligible, the rate of heat transfer
EP
between the strong and weak solutions is
Q& SHX = m& ss (h8 − h9 ) = m& ws (h7 − h6 ) .
AC C
(20)
The heat exchanger effectiveness, ε , is expressed as
ε≡
ε≡
q
qmax
.
(21)
Ch (Th, i − Th , o ) . Cmin (Th, i − Tc ,i )
(22)
If Cmin = Ch , then
ε≡
Th, i − Th , o . Th ,i − Tc ,i
(23)
18
ACCEPTED MANUSCRIPT 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:
T9 = εT6 + (1 − ε )T8 .
RI PT
m& ss (h8 − h9 ) . m& ws
4.8 Condenser The rate of heat transfer from the condenser is
4.9 Water pressure restrictor
M AN U
Q& C = m& w (h1 − h2 ) .
SC
h7 = h6 +
(24) (25)
(26)
The throttling model yields the result that
h2 = h3 . 4.10 Evaporator
(27)
Q& E = m& w (h4 − h3 ) . 4.11 Solution pump
TE D
The rate of heat transfer to the evaporator is
(28)
(29)
AC C
W& p = m& ws (h6 − h5 ) .
EP
The power input to the pump is
The following is an alternative to Eq. 29 for evaluating the pump work:
W& p m& ws
6 = vdP . ∫ 5
(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,
19
ACCEPTED MANUSCRIPT h6 = h5 + (P6 − P5 )v5 .
(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
v&∆P W&dm = .
ηp
4.12 Overall mass and energy balance of the system The energy rate balance at steady state is
SC
Q& cv − W& cv = 0 .
RI PT
to size the driving motor.
M AN U
Q& G + Q& E − Q& A − Q& C + W& p = 0 . The mass flow rates are
m& 1 = m& 2 = m& 3 = m& 4 = m& w . m& 5 = m& 6 = m& 7 = m& ws .
TE D
m& 8 = m& 9 = m& 10 = m& ss .
(32)
(33) (34)
(35) (36) (37)
The mass flow rates of water and the weak and strong solutions are expressed as
Q& G
X ws X ss h1 + h8 − h7 X ss − X ws X ss − X ws
.
EP
m& w =
(38)
.
(39)
X ws m& ss = m& w X ss − X ws
.
(40)
AC C
X ss m& ws = m& w X ss − X ws
4.13 COP The COP is the most common measurement used to rate heat pump efficiency. The COP of the cooling cycle is
COPc =
cooling capacity obtained at evaporator . heat input for the generator + work input for the pump
20
(41)
ACCEPTED MANUSCRIPT Q& COPc = & E . QG + W& p
(42)
The COP of the heating cycle is
heating capacity obtained at condenser . heat input for the generator + work input for the pump
Q& COPh = & C . QG + W& p
(44)
SC
4.3 Operating temperatures and pressures
(43)
RI PT
COPh =
As lower condenser and absorber temperatures increase cycle efficiency, they should be
M AN U
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
TE D
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
EP
°C to satisfy both practical requirements and needs of the higher second law efficiency. The condenser and absorber temperatures depend on the seawater temperature.
AC C
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
21
ACCEPTED MANUSCRIPT 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. Results and comparison
RI PT
5
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.
SC
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
M AN U
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-
TE D
2,000 1,500 1,000 500 0
EP
The rates of heat transfer (kW)
Gotni (2013).
AC C
50
75
Generator Evaporator Absorber Condenser
85
100
Engine load (%)
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.
22
ACCEPTED MANUSCRIPT 0.9 COP
0.8 0.7
cooling heating
0.6 0.5 90
95
100
105
110
o
RI PT
Generator temperature ( C)
Figure 5 COPs versus generator temperature
SC
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
M AN U
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
30 25 20 15 10 5 0 90
TE D
Mass flow rate ratio
effectiveness.
95
100
105
110
o
Generator temperature ( C)
AC C
EP
Figure 6 Mass flow rate ratio versus generator temperature
0.9
COP
0.85 0.8
Cooling Heating
0.75 0.7
0.65 0.6
0.65
0.7
0.75
0.8
Heat exchanger effectiveness
Figure 7 COPs versus heat exchanger effectiveness
23
ACCEPTED MANUSCRIPT 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
RI PT
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%.
SC
5.1 Comparison between an AHP system and a vapour–compression heat pump system
M AN U
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
TE D
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
EP
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
AC C
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 (η DG ) for the diesel generator is 95%. Power systems for the navies of
24
ACCEPTED MANUSCRIPT 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 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
M AN U
SC
RI PT
Molecular formula (avg)
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.
SFC =
W&c − W& p H u .η DG
.
TE D
The saved F-76 diesel fuel consumption (SFC) for heating and cooling can be calculated as (45)
EP
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
AC C
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.
25
80 60 40 20 0 906.32
1206.78
1303.21
1618.12
RI PT
The saved fuel consumption -1 (kgh )
ACCEPTED MANUSCRIPT
Heating capacity (kW) Vapour-compression heat pump of COP=3
Vapour-compression heat pump of COP=4
Vapour-compression heat pump of COP=2
M AN U
70 60 50 40 30 20 10 0 834.62
TE D
The saved fuel consumption -1 (kgh )
SC
Figure 8 SFC versus heating capacity
1111.31
1200.11
1490.11
EP
Cooling capacity (kW)
AC C
Vapour-compression heat pump of COP=3 Vapour-compression heat pump of COP=2
Vapour-compression heat pump of COP=4
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.
26
250 200 150 100 50 0 906.32
1206.78
1303.21
Heating capacity (kW)
1618.12
Vapour-compression heat pump of COP=4
SC
Vapour-compression heat pump of COP=3 Vapour-compression heat pump of COP=2
RI PT
(kgh -1)
The reduced CO 2 emission
ACCEPTED MANUSCRIPT
M AN U
Figure 10 Reduced CO2 emission versus heating capacity
150
TE D
(kgh )
200
-1
100 50 0
EP
The reduced CO 2 emission
250
834.62
1111.31
1200.11
1490.11
Cooling capacity (kW)
AC C
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.
27
ACCEPTED MANUSCRIPT 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
RI PT
heating cycle and US$20,622–US$73,641 in the cooling cycle and has a simple payback period of about between 2.5 and 9.5 years depending on the engine load and COP of the vapour–compression heat pump.
SC
80,000 70,000 60,000 50,000
M AN U
Profitability (US$year-1)
90,000
40,000 30,000 20,000 10,000 0 906.32
1206.78
1303.21
1618.12
Heating capacity (kW)
TE D
Vapour-compression heat pump of COP=3
Vapour-compression heat pump of COP=4
Vapour-compression heat pump of COP=2
70,000
AC C
Profitability (US$year-1)
80,000
EP
Figure 12 Profitability of the AHP system as a function of heating capacity
60,000 50,000 40,000 30,000 20,000 10,000
0 834.62
1111.31
1200.11
1490.11
Cooling capacity (kW) 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
28
ACCEPTED MANUSCRIPT 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
6
RI PT
difference to the human ear.
Conclusions
An H2O-LiBr AHP system for a naval surface ship was designed and thermodynamically
SC
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
M AN U
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
TE D
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
EP
results show that the generator temperature is an important factor to consider the optimum temperature at which an AHP cycle operates.
AC C
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.
29
ACCEPTED MANUSCRIPT •
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.
RI PT
•
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.
SC
References
ASHRAE, 2009. Handbook Fundamentals, SI Edition, ASHRAE Inc., Atlanta, GA.
ASHRAE Inc., Atlanta, GA.
M AN U
ASHRAE, 2011 Handbook Heating, Ventilating, and Air-conditioning Applications, SI Edition,
Bakhtiari B, Fradette L, Legros R, Paris J., 2011. A model for analysis and design of H2O– LiBr absorption heat pumps, Energy Convers Manage;52:1439–48.
TE D
Cotana, F., Nicolini, A.,2003. Noise Impact of Absorption Machines for Civil Applications, The 32nd International Congress and Exposition on Noise Control Engineering, Jeju International Convention Center, Seogwipo, Korea,
EP
Ezgi, C., Turhan Çoban, M., and Selvi, Ö., 2013. Design and thermodynamic analysis of an SOFC system for naval surface ship application. ASME J.Fuel Cell Sci. Technol.
AC C
10(3), 031006.
Fernandez-Seara, J.,Vales, A., Vazquez, M.,1998. Heat recovery system to power an onboard NH3-H2O absorption refrigeration plant in trawler chiller fishing vessels, Appl. Therm. Eng.18 1189-1205
Frank, M.V. and Helmick D.,2007. 21st Century HVAC System for Future Naval Surface Combatants-Concept Development Report, NAVSEA Philadelphia Naval Surface Warfare Center, Philadelphia. IMO, 2010. Prevention of Air Pollution from Ships: Assessment of IMO Efficiency Measures for the Control of GHG Emissions from Ships, London.
30
ACCEPTED MANUSCRIPT Keith, F., Goswami, D.Y., 2008. Energy Management and Conservation Handbook. CRC Press, Boca Raton. Liang, Y., Shu, G., Tian, H., Liang, X., Wei, H., Liu, L., 2013. Analysis of an electricity– cooling cogeneration system based on RC–ARS combined cycle aboard ship, Energy
RI PT
Convers Manage, 76, 1053–1060. Loydu, T., 2007. Rules for the Classification of Naval Ships Ship Operation Installations and Auxiliary Systems, Istanbul, Turkey.
McNeely, L.A. 1979. Thermodynamic properties of aqueous solution of lithium bromide,
SC
ASHRAE Transactions 85(1):413.
Moran, M. J., Shapiro, H.N., Munson, B.R., DeWitt, D.P., 2002. Introduction to Thermal
John & Sons, USA.
M AN U
Systems Engineering: Thermodynamics, Fluid Mechanics, and Heat Transfer. Wiley,
Patek, J., Klomfar, J., 2006. A computationally effective formulation of the thermodynamic properties of LiBr–H2O solutions from 273 to 500 K over full composition range, Int. J.
TE D
Refrigeration. 29, 566–578
Ruiz, V.,2012. Analysis of existing refrigeration plants onboard fishing vessels and improvement possibilities, Second International Symposium on Fishing Vessel
EP
Energy Efficiency, E-Fishing, Vigo, Spain. Ouadha A., El-Gotni, Y., 2013. Integration of an ammonia–water absorption refrigeration
AC C
system with a marine diesel engine: A thermodynamic study. Procedia Comput Sci. 19,754 –761.
Seddiek, I.S., Mosleh, M., Banawan, A. A., 2012. Thermo-economic approach for absorption air condition onboard high-speed crafts. Inter. J. Nav. Archit. Oc. Engng. 4 460–476.
Singh, Y., Kumar, D., Kumar, A., Singla, A., et al., 2013. Mathematical modeling and analysis of absorption refrigeration system using waste heat of diesel genset. Int. J. Emerg. Manag. Res. Manag & Tech. 46–50. Srikhirin P, Aphornratana S, Chungpaibulpatana S.,2001. A review of absorption refrigeration technologies, Renew Sustain Energy Rev, 5:343-72.
31
ACCEPTED MANUSCRIPT Steinfeld, G., Sanderson, R., Ghezel-Ayagh, H., Abens, S., and Cervi, M. C., 2000. Distillate Fuel Processing for Marine Fuel Cell Applications. Proceedings of the AICHE Spring 2000 Meeting, Atlanta, GA, March 5–9. Sun, J., Fu, L., Zhang, S., Hou, W., 2010. A mathematical model with experiments of single
RI PT
effect absorption heat pump using LiBr-H2O. Appl. Therm. Eng. 30 2753–2762. Táboas, F., Bourouis, M., Vallès, M., 2014. Analysis of ammonia/water and ammonia/salt mixture absorption cycles for refrigeration purposes in fishing ships, Appl. Therm. Eng. 66,603-611
SC
Wang, K., Abdelaziz, O., Kisari, P., Vineyard, E.A., 2011. State-of-the-art review on
Refrigeration. 34,1325–1337.
M AN U
crystallization control technologies for water/LiBr absorption heat pumps. Int. J.
Wang, S.G.,Wang R.Z., 2005. Recent developments of refrigeration technology in fishing vessels, Renew Energ 30,589–600 Wärtsılä 32 Product Guide, 2013, Vaasa.
TE D
Zhang, H.G., Wang, E.H., Fan, B.Y., 2013. Heat transfer analysis of a finned-tube evaporator for engine exhaust heat recovery. Energy Convers Manage, 65, 438–447.
EP
Acknowledgments
The views and conclusions contained herein are those of the author and should not be
AC C
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.
32
ACCEPTED MANUSCRIPT HIGHLIGHTS •
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
RI PT
•
AC C
EP
TE D
M AN U
SC
calculated.