Article
Design and Thermodynamic Analysis of a Steam Ejector Refrigeration/Heat Pump System for Naval Surface Ship Applications Cüneyt Ezgi * and Ibrahim Girgin Received: 8 September 2015; Accepted: 7 December 2015; Published: 11 December 2015 Academic Editor: Kevin H. Knuth Mechanical Engineering Department, Turkish Naval Academy, Istanbul 34942, Turkey;
[email protected] * Correspondence:
[email protected]; Tel.: +90-530-6065-395; Fax: +90-216-3952-658
Abstract: Naval surface ships should use thermally driven heating and cooling technologies to continue the Navy’s leadership role in protecting the marine environment. Steam ejector refrigeration (SER) or steam ejector heat pump (SEHP) systems are thermally driven heating and cooling technologies and seem to be a promising technology to reduce emissions for heating and cooling on board naval surface ships. In this study, design and thermodynamic analysis of a seawater cooled SER and SEHP as an HVAC system for a naval surface ship application are presented and compared with those of a current typical naval ship system case, an H2 O-LiBr absorption heat pump and a vapour-compression heat pump. The off-design study estimated the coefficient of performances (COPs) were 0.29–0.11 for the cooling mode and 1.29–1.11 for the heating mode, depending on the pressure of the exhaust gas boiler at off-design conditions. In the system operating at the exhaust gas boiler pressure of 0.2 MPa, the optimum area ratio obtained was 23.30. Keywords: ship; engine; sea water; ejector system; refrigeration; heat pump
1. Introduction The International Maritime Organization’s [1] Marine Environment Protection Committee published its final third IMO greenhouse gas (GHG) study report in 2014 providing updated estimates for GHG emissions from ships. In that report, for the year 2012, total shipping emissions were approximately 949 million tonnes CO2 and 972 million tonnes CO2 e for GHGs combining CO2 , CH4 and N2 O. International shipping emissions for 2012 were estimated to be 796 million tonnes CO2 and 816 million tonnes CO2 e. International shipping thus accounted for approximately 2.2% and 2.1% of global CO2 and GHG emissions on a CO2 equivalent (CO2 e) basis, respectively. In addition, refrigerant and air conditioning gas released from shipping contributed an additional 15 million tons in CO2 equivalent emissions. In the study, military forces were excluded from the total and international shipping calculations. To reduce the emission of greenhouse gases, specifically CO2 emissions, emitted by ships, the Energy Efficiency Design Index (EEDI) for new ships and the Ship Energy Efficiency Management Plan (SEEMP) for all ships entered into force on 1 January 2013. New ships are ships that enter the fleet from 2013. 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. The main and auxiliary engines are used to propel the ship, and to drive generators to produce electricity, respectively. Despite the great technological development of engines, the maximum efficiency is still less than 50%. The main and auxiliary engines onboard ships produce significant quantities of heat. The primary source of waste heat of main and auxiliary engines is the exhaust
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gas heat dissipation, which accounts for about half of the total waste heat, i.e., about 25% of the total fuel energy. For naval ships, the environment (decreased emissions), stealth demands (hydrodynamic, acoustic, magnetic, infrared, radar signatures) and efficiency (decreased fuel costs) are of importance. The required precautions to achieve the specified level of stealth features should be taken during both the design and operating phase. CO2 reduction measures as well as reduction and management of ship signatures must also be implemented by naval ships. To continue the Navy’s leadership role in protecting the marine environment, the search for new energy conservation methods that can be applied on board naval ships is necessary. Thermally driven heat pumps for heating and cooling can be promising solutions. Today, heating, cooling and refrigeration systems onboard ships are based on mechanical vapour compression. These cycles are powered by electrical energy generated by the combustion of fossil fuels, and the process contributes to an increase in greenhouse gases and the generation of air pollutants. One way to find a new solution to this problem is to apply an absorption refrigeration (AR)/heat pump (AHP) system to provide the required heating and cooling loads for the HVAC system instead of the traditional vapour-compression heat pump. The main problem with absorption systems is that they are more bulky than conventional vapour-compression systems. This is because an AR/AHP has more components and as the heat and mass transfer of absorption equipment is poor, large surface areas are required. Ezgi [2] presented design and thermodynamic analysis of a water-lithium bromide AHP as an HVAC system for a naval surface ship application and compared with those of a vapour-compression heat pump. Ezgi [2] performed calculations for diesel engine loads of 50%, 75%, 85%, 100%. The temperatures of evaporator, condenser, absorber are 4 ˝ C, 50 ˝ C, 35 ˝ C, respectively, and generator temperatures were 90 ˝ C, 95 ˝ C, 100 ˝ C, 105 ˝ C and 110 ˝ C. At the end of 1000 operating hours a year of a naval surface ship, it was stated that the AHP system could save 22,952–81,961 L of diesel fuel in the heating cycle and 21,135–75,477 L of diesel fuel in the cooling cycle and would reduce its annual CO2 emissions by 60.41–215.74 tons in the heating cycle and 55.63–198.67 tons in the cooling cycle, depending on the engine load and COP of the vapour-compression heat pump. Another type of system is known as an ejector refrigeration/heat pump. Riffat et al. [3] and Ma et al. [4] stated that ejector refrigeration is one of the most promising technologies because of its relative simplicity and low capital cost when compared to an absorption refrigerator. This is a heat-operated cycle capable of utilising solar energy, waste energy, natural gas or hybrid sources (e.g., solar/gas). The ejector heat refrigeration/heat pump has no moving parts and so is simple and reliable. In addition, it has the potential of long life and, unlike vapour compression systems, produces no noise or vibration. Chen et al. [5] provided a literature review on the recent developments in ejectors, applications of ejector refrigeration systems. Ejector refrigeration systems (ERS) are more attractive compared with traditional vapour compression refrigeration systems, with the advantage of simplicity in construction, installation and maintenance. Moreover, in an ERS, compression can be achieved without consuming mechanical energy directly. Furthermore, the utilization of low-grade thermal energy (such as solar energy and industrial waste heat) in the system can help mitigate the problems related to the environment, particularly by reduction of CO2 emissions from the combustion of fossil fuels. An ejector is a simple device in which high pressure stream (the primary fluid) is used to compress low pressure stream (the secondary fluid) to a higher pressure (discharge pressure). A number of experimental studies involving ejectors have been reported in the literature. Hsu [6] investigated by analytical methods the efficiency of an ejector heat pump with the refrigerants 11, 113, and 114. Huang et al. [7] carried out a 1-D analysis for the prediction of the ejector performance at critical mode operation. El-Dessouky et al. [8] developed semi-empirical models for the design and rating of steam jet ejectors. Alexis [9] modelled the steam-ejector refrigeration system and estimated the maximum flow entrainment ratio for constant generator pressure (6–8 bar), condenser (40–50 ˝ C) and evaporator
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temperature (4–10 ˝ C). Sanaye and Niroomand [10] presented the modelling and optimization of a ground-coupled steam ejector heat pump with a closed vertical ground heat exchanger. Steam-ejector refrigeration systems are used in food processing plants, gas plants, breweries, the rubber and vulcanizing, paper and pulp, paints and dyes, pharmaceutical and chemical industries, and edible oil refineries. In the literature, although the ejector refrigeration cycle is most commonly used for refrigeration in land-based plants, there has been no report that an ejector heat pump system has been installed on board ships. The most important difference between land and marine refrigeration is the need for a cooling tower in land-based applications and seawater in marine applications to eject heat from the condenser to ambient air. Cooling towers increase the initial system costs, and require regular maintenance and require extra space for their installation. The development of seawater-cooled steam-ejector refrigeration technology can effectively eliminate these disadvantages. Also, no investigation has been conducted yet on a seawater-cooled steam ejector refrigeration/heat pump system for a naval ship application. One possible reason why steam ejector refrigeration systems have never been applied in ships before is the lower COP—0.2~0.3—compared to vapour compression systems and other thermally driven technologies. The COP also drops significantly at operation away from the design point. Another is the lack of performance data from commercial applications to provide confidence in the ship applications of the technology and the need for uninterrupted cooling and heating on ships, especially in naval ships. Therefore, this study focuses on the dual use of steam ejector technology to produce heating and cooling onboard a naval ship. A theoretical study of the possibilities for application of steam ejector refrigeration/heat pump systems for cooling and heating of naval surface ships is presented. Waste heat from the main diesel engine exhaust gas is utilized. The technical characteristics of the system are analyzed and its economical and environmental benefits are discussed. 2. System Selection for Naval Surface Ships In general, the details of merchant ship air conditioning also apply to warships. However, all ships are governed by their specific ship specifications, and naval ships are usually governed by military specifications, which require an excellent air-conditioning system and equipment performance in the extreme environment of naval ship duty. Design conditions for naval surface ships have been established as a compromise. These conditions consider the large cooling plants required for internal heat loads generated by machinery, weapons, electronics and personnel. 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 units (AHU) which are used to regulate and circulate air as part of an HVAC system are used on board naval ships. Steam ejector refrigeration (SER) and steam ejector heat pump (SEHP) systems are a good choice for naval surface ships because of their lower energy consumption, CO2 emissions, EEDI, infrared and acoustic signature, simple, compact construction and corrosion resistance and because they can be used with water, which is the most environment friendly refrigerant. The SER/SEHP system can be fitted into the machinery space of a ship. The system serves by supplying chilled water in cooling mode and hot water in heating mode to fan coil units located in various parts of the naval ship. 3. The Case Naval Ship The case naval surface ship has two main internal combustion engines that propel the ship. The specification of the each diesel engine on the naval ship is given in Table 1 [11]. The propeller demand data are presented in Table 2 [11].
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Table 1. Specification of a diesel engine on a naval ship. Engine Maximum Continuous Output
3000 kW
Engine speed Cylinder bore Stroke Cylinder configuration Backpressure, max.
750 rpm 320 mm 400 mm 6, in-line 4.0 kPa
Table 2. Propeller demand data. Engine % Load
Fuel Consumption, g¨ (kW¨ h)´1
Exhaust Gas Temperature after Turbocharger, ˝ C
Exhaust Gas Flow kg¨ s´1
50 75 85 100
191 182 181 185
315 345 336 380
3.71 4.43 4.96 5.40
The total heating and cooling loads of the case naval surface ship are 144 kW and 116 kW, respectively. The case naval surface ship has two marine diesel generator sets for auxiliary power production, each with a generator power of 240 kVA. Heating load (144 kW) is met by electric duct heaters. Cooling load (116 kW) is met by a seawater cooled chiller unit. 4. System Design This study focuses on the dual use of ejector technology to produce heating and cooling on board a naval ship. The design criteria are given in Table 3, which is determined according to Loydu’s Rules for the Classification of Naval Ships and based on the naval distillate fuel (NATO symbol F-76) used on power systems by the navies of NATO countries. Table 3. The design criteria. Type of Heat Pump
Steam Ejector
Energy source Diesel engine fuel type
Diesel engine exhaust gas NATO F-76 Diesel
Heating Mode Hot water temperature flows through the condenser Evaporator
45 ˝ C–40 ˝ C Seawater (´2 ˝ C–+32 ˝ C)
Cooling Mode Chilled water temperature flows through the evaporator Condenser
7 ˝ C–12 ˝ C Seawater (´2 ˝ C–+32 ˝ C)
The working fluid is water in SER/SEHP system. Water is a natural working fluid. It is an excellent working fluid for high-temperature industrial heat pumps because of its favourable thermodynamic properties and the fact that it is neither flammable nor toxic. Water is inexpensive and has no environment impact (zero ozone depletion and global warming potential). Water has an extremely high heat of vaporization that causes a low circulation rate for given heating and cooling capacity. At low temperatures the saturation pressures are low (0.008129 bar at 4 ˝ C) and the specific volumes are high (157.3 m3 ¨ kg´1 at 4 ˝ C) at typical evaporator conditions. Therefore, low mechanical power is required for the pump. The system is not considered under the diesel engine load of 50% as long term operation at lower loads, typically below 50% of its maximum rated load, reduces engine service life. If the engine of the 8155
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naval ship is in standby or under very low engine loads (below 50%), the heating and cooling load forEntropy the naval surface 2015, 17, 1–22 ship will be met from a vapour compression heat pump. In addition, vapour compression heat pumps onboard naval ship are operated as reserves in emergency situations, at naval compression heat pumps onboard navalinship areshipyards. operated as reserves in emergency situations, at base or during periodic engine overhauls naval naval or during periodic engine overhauls in navalfor shipyards. Thebase expansion tanks are designed to compensate the changing volume of the water in the The expansion tanks are designed to compensate for the of the water in water the SER/SEHP system to maintain the static pressure created by thechanging pump atvolume the utilisation level in SER/SEHP system to maintain the static pressure created by the pump at the utilisation level in water production and to compensate for any changes in the water flow rate. production and to compensate for any changes in the water flow rate. A SER/SEHP system design for naval ship application is presented in Figure 1. The SER/SEHP is A SER/SEHP system design for naval ship application is presented in Figure 1. The SER/SEHP driven by the thermal energy of a diesel engine and consists of an ejector, a boiler, an evaporator, a is driven by the thermal energy of a diesel engine and consists of an ejector, a boiler, an evaporator, condenser, an expansion valve and a circulation pump. In the shown system, high-pressure steam a condenser, an expansion valve and a circulation pump. In the shown system, high-pressure steam expands while flowing through the nozzle. The expansion causes a drop in pressure and an enormous expands while flowing through the nozzle. The expansion causes a drop in pressure and an enormous increase in in velocity. Due the evaporator evaporatorisisdrawn drawninto intothe the swiftly increase velocity. Duetotothe thehigh highvelocity, velocity,vapour vapour from from the swiftly moving steam and the mixture enters the diffuser. The velocity is gradually reduced in the diffuser but moving steam and the mixture enters the diffuser. The velocity is gradually reduced in the diffuser thebut pressure of the steam at the at condenser is increased 5–10 times moremore than than that at theatentrance of the the pressure of the steam the condenser is increased 5–10 times that the entrance ˝ C. The latent heat of diffuser. This pressure value corresponds to a condensation temperature of 50 of the diffuser. This pressure value corresponds to a condensation temperature of 50 °C. The latent condensation is transferred to the condenser water. Thewater. condensate is pumped to the boiler. The heat of condensation is transferred to the condenser The condensate is back pumped back to the cooled water pumped as is the refrigeration to the carrier fan-coiltounit. The heated is pumped boiler. The is cooled water pumped as the carrier refrigeration the fan-coil unit. water The heated water as pumped as theto heating to the fan coil unit. theisheating carrier the fancarrier coil unit.
Figure1.1.SER/SEHP SER/SEHP system system design Figure designfor foraanaval navalship. ship.
The evaporator temperature has to be designed at 4 °C to cool water of 12 °C for the air The evaporator temperature has to be designed at 4 ˝ C to cool water of 12 ˝ C for the air conditioning of the space to be maintained between 24 ˝°C and 27 °C. conditioning of the space to be maintained between 24 C and 27 ˝ C. The condenser temperature depends on the seawater temperature. According to Loydu’s [12] The condenser temperature depends on the seawater temperature. According to Loydu’s [12] Rules for the Classification of Naval Ships, the selection, layout and arrangement of all shipboard 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 machinery, appliances ensure faultless continuous operation under the seawater seawater equipment temperatureand of −2 °C to +32 should °C and the air temperature outside the ship of −25 °C to +45 °C. ˝ C to +32 ˝ C and the air temperature outside the ship of ´25 ˝ C to +45 ˝ C. temperature of ´2 The exhaust gas boiler is used to recover exhaust waste heat from the diesel engine. Each engine The exhaustexhaust gas boiler used The to recover exhaust waste heat from theexhaust diesel engine. Each engine has a separate gasisboiler. pressure drop of the gas side of the gas boiler is set to has3 akPa separate exhaust gasback boiler. The pressure drop gas of the gas side of the exhaustFresh gas boiler maximum and the pressure of the exhaust system is 4 kPa maximum. water is setconservation to 3 kPa maximum the back of the exhaust gas is coolant 4 kPa maximum. Fresh onboardand a naval ship ispressure very important. Seawater is system the main on board naval water conservation onboard naval ship veryand important. Seawater main coolant board ships, like in other ships. aSeawater is ais free renewable source is forthe cooling onboardon ships. naval ships, like in other ships. Seawater is a free and system renewable source for cooling ships. Therefore, the condenser and evaporator used in this are seawater-cooled heatonboard exchangers. However, seawater-cooled exchangers are in prone fouling, causes the thermohydraulic Therefore, the condenser andheat evaporator used this to system arewhich seawater-cooled heat exchangers. performance of heat transfer to decrease time [13]. Design fouling must However, seawater-cooled heatequipment exchangers are pronewith to fouling, which causes theresistances thermohydraulic also be taken into account dimensioning the exhaust gas[13]. boiler and heat exchangers because performance of heat transferwhen equipment to decrease with time Design fouling resistances must the heat exchangers subject to fouling and mustgas operate long periods withoutbecause being also beboiler takenand into account whenare dimensioning the exhaust boilerforand heat exchangers Thus, theexchangers design fouling resistances of the engine exhaust gas for andlong seawater are without set to thecleaned. boiler and heat are subject to fouling and must operate periods Rf,exh = 1.761 m2·K·(kW)−1, Rf,s = 0.088 m2·K·(kW)−1, respectively [14].
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being cleaned. Thus, the design fouling resistances of the engine exhaust gas and seawater are set to 2 ¨ K¨ (kW)´1 , R 2 ´1 R f ,exh = 1.761 m1–22 Entropy 2015, 17, f ,s = 0.088 m ¨ K¨ (kW) , respectively [14]. Seawater flows through the evaporator during the heating mode, through the condenser during Seawater flows through during heating mode, through the condenser during the cooling mode all the time.the Toevaporator achieve this flow,the two 4-way, 3-position (4/3) rotary valves (also the cooling mode are all the time. To achieve flow,intwo (4/3) called CEFS valves) used for heating andthis cooling the 4-way, system.3-position One of the 4/3rotary valvesvalves is the (also supply called CEFS is valves) are used for heating and cooling in thethe 4/3 heating valves ismode, the supply and the other the return valve. The three positions ofthe thesystem. rotary One valveofare closed and the other is the return valve. The three positions of the rotary valve are the heating mode, closed in and cooling mode. A fan coil system is integrated with CEFS valves in a SEHP system as shown and cooling mode. A fan coil system is integrated with CEFS valves in a SEHP system as shown in Figure 2. Heat produced with the SER or SEHP system is transferred to the fan-coil units. The heat is Figure 2. Heat produced with the SER or SEHP system is transferred to the fan-coil units. The heat is mainly consumed in fan coil units. A heat meter is used to measure the heat returning from fan coil mainly consumed in fan coil units. A heat meter is used to measure the heat returning from fan coil units. The residual heat is sent to hot water production in heating mode or to machine room cooling units. The residual heat is sent to hot water production in heating mode or to machine room cooling in cooling mode. An analog-to-digital converter (ADC) is used that converts a continuous physical in cooling mode. An analog-to-digital converter (ADC) is used that converts a continuous physical quantity (usually voltage) to a digital number that represents the quantity's amplitude. A servo valve quantity (usually voltage) to a digital number that represents the quantity's amplitude. A servo valve is used to control flow is used to control flowthrough throughthe theheat heatmeter. meter.
Figure 2. Integration of fan coil and CEFS valves design for a naval ship. Figure 2. Integration of fan coil and CEFS valves design for a naval ship.
5. Thermodynamic Analysis 5. Thermodynamic Analysis The thermodynamic design of the ejector heat pump system by the first law only is usually based thermodynamic design of operating the ejectorconditions. heat pumpThe system by theand firstevaporator law only istemperatures usually based onThe given or assumed steady-state condenser onare given or The assumed steady-state conditions. The condenser and evaporator temperatures fixed. system boiler heat operating transfer rates are also known. are fixed. The system boiler heat transfer rates are also known. The fundamental simplifications assumed for the model are as follows: The fundamental simplifications assumed for the model are as follows: Steady state of the SER/SEHP No radiation heat transfer ‚ Steady state of the SER/SEHP The primary and secondary fluids are saturated and have the same molecular weight and ‚ No radiation heat transfer ratio of specific heats. ‚ The primary and secondary fluids are saturated and have the same molecular weight and ratio of Water at the condenser outlet is saturated liquid specific heats. Water at the evaporator outlet is saturated vapour ‚ Water at the condenser outlet is saturated liquid Pressure losses in the pipes and all heat exchangers are negligible ‚ Water theprimary evaporator is saturated vapour at The fluidoutlet expands through the nozzle from the boiler pressure to the evaporator ‚ Pressure losses in the pipes and all heat exchangers are negligible pressure. ‚ The fluid expands through the nozzle the boiler pressure to the evaporator pressure. primary The pressure drop and momentum of thefrom secondary flow are negligible. There is no wall friction. ‚ The pressure drop and momentum of the secondary flow are negligible. is All ‚ There nofluid wallproperties friction. are uniform over the cross section after complete mixing at section 2. Potential energy negligibly small the section energy equations. ‚ All fluid properties areisuniform over the in cross after complete mixing at Section 2. The exit velocity at the ejector outlet is ignored. ‚ Potential energy is negligibly small in the energy equations. The T-s diagram steamoutlet ejector pump cycle is shown in Figure 3. Saturated motive ‚ The exit velocity at of thethe ejector is heat ignored. steam enters the ejector at a high pressure P0, temperature T0 and zero velocity corresponding to state
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The T-s diagram of the steam ejector heat pump cycle is shown in Figure 3. Saturated motive 1–22 steamEntropy enters2015, the17,ejector at a high pressure P0 , temperature T0 and zero velocity corresponding to state (0) and to a pressure at state secondary vapour enters the ejector at pressure (0)expands and expands to a pressure at (1). stateThe (1).saturated The saturated secondary vapour enters the ejector at P4 and zero velocity corresponding to state (4). pressure P4 and zero velocity corresponding to state (4).
Figure 3. T-s diagram of the steam ejector refrigeration and heat pump cycle.
Figure 3. T-s diagram of the steam ejector refrigeration and heat pump cycle.
5.1. Nominal Heat Balance
5.1. Nominal Heat Balance
The conservation of mass principle is expressed as:
The conservation of mass principle is expressed as: dmcv
m i m e
. dmcvdt . “ mi ´ m e dt the equation is reduced to: To obtain a control volume at steady state,
To obtain a control volume at steady state, the m i equation e m e is reduced to: ÿ i. ÿ . mi “ me The energy rate balance is expressed as: i
(1)
(1) (2)
(2)
e
dEcvis expressed V V The energy rate balance as: Q cv Wcv m i hi i gzi m e he e gze (3) 2 2 dt ¸ i˜ ¸ e ˜ 2 2 ÿ . ÿ . . . V Ve dEcv Qcv ´volume W cv ` at m ` i and ` gz ´ me thehechanges ` ` gze (3) i hi state To obtain a “ control steady toi disregard dt 2 2 in the kinetic and e 2
2
i potential energies of the flowing streams from inlet to exit, the equation is reduced to:
To obtain a control volume at steady and to 0 state Q cv W m i hi m ethe he changes in the kinetic and potential i disregard cv (4) e energies of the flowing streams from inlet to exit, the equation is reduced to: ÿ . ÿ . . . Momentum equation is: 0 “ Qcv ´ W cv ` mi hi ´ me he (4) Pi Ai m iVi i Pe Ae meeVe
Momentum equation is: 5.2. Main Engine
Pi Ai `
ÿ
.
mi Vi “ Pe Ae `
ÿ
(5)
.
me Ve
(5)
The heat transfer rate from the engine exhaust gas to the exhaust gas boiler is expressed as:
5.2. Main Engine
Q exh m exh hexh ,in hexh ,out
(6)
The heatm transfer rate from the engine exhaust gas to the exhaust gas boiler is expressed as: where exh is the engine exhaust gas mass flow rate, hexh ,in is the engine exhaust gas specific . ` . enthalpy at the entrance of the Q exhaust gas boiler and hexh ,out ˘is the engine exhaust gas specific (6) exh “ mexh hexh,in ´ hexh,out enthalpy at the exit of the exhaust gas boiler. .
where mexh is the engine exhaust gas mass flow rate, hexh,in is the engine exhaust gas specific enthalpy at the entrance of the exhaust gas boiler and hexh,out is the engine exhaust gas specific enthalpy at the exit of the exhaust gas boiler. 7 8158
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5.3. Exhaust Gas Properties The exhaust gas properties, which include specific heat at constant pressure, dynamic viscosity and thermal conductivity, should be determined for the heat transfer analysis. The main components of the exhaust gas of a diesel engine are CO2 , H2 O, N2 and O2 . The mass fractions of these components vary with the operating conditions of the engine. When the engine operates at steady state, the injected fuel quantity and the intake air amount can be measured on the engine test bench. Except for very low engine loads, the exhaust temperature of a marine engine is between 300 ˝ C and 380 ˝ C, and the exhaust pressure is slightly higher than the atmospheric pressure. Therefore, exhaust gas can be treated as a mixture of ideal gases. The specific enthalpy, specific heat and density of exhaust gas can be calculated as follows [15]: hm “
4 ÿ
m f i hi
(7)
m f i c p,i
(8)
i“1
c p,m “
4 ÿ i“1 4 ř
ρm “
i“1 4 ř i“1
m f i Mi (9)
m f i Mi {ρi
5.4. Exhaust Gas Boiler The heat transfer rate of the exhaust gas boiler is expressed as: .
.
m0 “
QB h0 ´ h9
Qexh “ Q B
(10)
The mass flow rate of the steam is: .
.
(11)
5.5. Steam Ejector The thermodynamic properties at initial states for steam are calculated from the saturation property. At 50 ˝ C, the saturation pressure of water is 12.350 kPa. At pressures below this value, water vapour can be treated as an ideal gas with negligible error (under 0.2 percent), even when it is a saturated vapour [16]. The performance of an ejector is measured by its entrainment ratio w which is the mass flow rate ratio of the secondary flow to that of the primary flow. It is given as [6]: .
m w“ .4 m0
(12)
The efficiency of an ejector can be defined by [6]: ηE “
wa wi
(13)
The ejector structure is normally characterised by the area ratio ξ which is defined as the cross-section area of the constant area section divided by that of the primary nozzle throat. It is given as [17]: A2 ξ“ (14) At For inlet and outlet conditions, the conservation of energy is given as: `. . . . ˘ m0 h0 ` m4 h4 “ m0 ` m4 h3
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(15)
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5.5.1. Nozzle Section Nozzle efficiency is given as [16]: h0 ´ h1 h0 ´ h1s
ηn “
(16)
Assuming zero entrance velocity, the velocity of the primary fluid at the nozzle exit is: V1 “ r2ηn ph0 ´ h1s qs0.5
(17)
The Mach number, M, is the ratio of the actual velocity of the fluid to the velocity of sound at the same state. The properties of a fluid at a location where the Mach number is unity (the throat) are called critical properties, and the above ratios are called critical ratios. When steam enters the nozzle as a saturated vapour, the critical-pressure ratio becomes [16]: P˚ “ Po
ˆ
2 k`1
˙k{pk´1q “ 0.576
(18)
Since the ratio of the exit-to-inlet pressure (P4 /P0 ) is less than the critical pressure ratio, the flow is choked at the nozzle. Then the velocity at the throat, Vt , is the sonic velocity, and the throat pressure, Pt , is the critical pressure. 5.5.2. Mixing Section Conservation of mass is:
.
.
A2 V2 v2
.
.
m0 ` m4 “
(19)
Conservation of energy is: ˜ .
.
`
m0 h0 ` m4 h4 “ m0 ` m4
˘
V2 2 h2 ` 2
¸ (20)
Conservation of momentum is [6]: `. . . ˘ m0 V1 ` P4 A2 “ m0 ` m4 V2 ` P2 A2 Including mixing efficiency for the mixing chamber, the momentum equation reduces to: `. ˘ `. . ˘ ηm m0 V1 ` P4 A2 “ m0 ` m4 V2 ` P2 A2
(21)
(22)
5.5.3. Constant Area Section A normal shock wave occurs if the velocity of the mixing fluid entering the constant area section is supersonic. In this case, a sudden reaction in the mixture velocity and a rise in the pressure take place [18]. Also, that the primary and secondary streams mix in the mixing chamber and enter a normally choked (sonic flow conditions at the throat) secondary converging-diverging nozzle was stated by [19]: A2i “ A2e “ A2 (23) Conservation of mass is:
V2i V2e “ v2i v2e
(24)
V2i2 V2 “ h2e ` 2e 2 2
(25)
Conservation of energy is: h2i ` Conservation of momentum is: `. . ˘ A2 pP2i ´ P2e q “ mo ` m4 pV2e ´ V2i q 8160
(26)
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The Mach number of the mixed flow at Section 2 after the shock wave is M2 “ ?
V2 kRT2
(27)
V2
M2 “
(28)
pkP2 v2 q0.5
From Equations (19), (22) and (28), the cross-sectional area, A2 , is solved by: .
A2 “
mo V1 ηm ˘ P2 kM22 ` 1 ´ P4 ηm
(29)
`
ASHRAE [20] proposes that the constant area throat section be typically 3–5 throat diameters long to accommodate the shock pattern and its axial movement under load. 5.5.4. Diffuser Conservation of energy is: h2 `
V2 2 “ h3 2
(30)
The performance of a diffuser is usually expressed in terms of the diffuser and is given as Çengel and Boles [16]: h3s ´ h2 (31) ηd “ h3 ´ h2 5.5.5. The Optimum Mixing Constant Area .
The optimum mixing constant section area, A2, can be found by maximizing m4 . Hsu [6] calculated six simultaneous equations containing six unknown differentials, ds2 , dV2 , dh2 , dv2 , dA2 and dP2 . The optimum criterion to operate an ejector for given conditions is given as: ˙, » ¨$ ˆ ˛fi T2 / ’ / ’ d ln & — ffi T2 ˚ k ‹ v2 . ˚ ‹ffi “ 1 M22 — 1 ` p1 ´ η q ` (32) d – ‚ fl P / q T3s ˝’ d plnv 2 / ’ % 1´ 4 P2 s Finally, M2 and P2 values for the optimum mixing constant section area are calculated as: » fi k´1 ˆ ˙ k M22 –1 ` p1 ´ ηd q pP2 {P3 q k ´ k fl “ 1 1 ´ P4 {P2 k´1 pP3 {P2 q k ´ ηd M22 pk ´ 1q {2 “ 1
(33)
(34)
5.6. Condenser The rate of heat transfer from the condenser is: .
.
QC “ m3 ph3 ´ h5 q
(35)
5.7. Expansion Valve The throttling model yields the result that: h6 “ h7 5.8. Evaporator The rate of heat transfer to the evaporator is: 8161
(36)
Entropy 2015, 17, 8152–8173 .
.
.
.
Q E “ m4 ph4 ´ h7 q
(37)
5.9. Feed Water Pump The power input to the pump is: W p “ m0 ph9 ´ h8 q
(38)
The following is an alternative to Equation (38) for evaluating the pump work: ˜
.
Wp
¸
ż9 vdP
“
.
m0
(39)
8
Then: h9 “ h8 ` pP9 ´ P8 q v8
(40)
The work input to the system in the pump is small and neglected in the calculation of the coefficient of performance (COP) and efficiency. However; in practice, it is usually estimated to size the driving motor: . . v∆P W dm “ (41) ηp 5.10. Overall Mass and Energy Balance of the System The energy rate balance at steady state is: .
.
Qcv ´ W cv “ 0 .
.
.
(42) .
QC “ Q B ` Q E ` W p The mass flow rates are:
.
.
.
.
(43) .
m2i “ m2e “ m2 “ m3 “ m5 .
.
.
.
.
.
m8 “ m9 “ m0
(44) (45)
m4 “ m6 “ m7
(46)
5.11. COP The performance of refrigerators and heat pumps is expressed in terms of the COP, which is defined for a vapour-compression refrigeration/heat pump as: .
cooling e f f ect Q COPR “ “ . E work input W net;in
(47)
.
COPHP
heating e f f ect Q “ “ . C work input W net;in
(48)
The COP of the SER is: COPR “
cooling capacity obtained at evaporator heat input f or the boiler ` work input f or the pump
(49)
.
COPR “
.
QE
.
QB ` W p
8162
(50)
Entropy 2015, 17, 8152–8173
The COP of the SEHP is: COPHP “
heating capacity obtained at condenser heat input f or the boiler ` work input f or the pump
(51)
.
COPHP “
.
QC
(52)
.
QB ` W p
5.12. Solution Procedure .
‚ ‚ ‚ ‚ ‚ ‚ ‚ ‚ ‚ ‚ ‚ ‚ ‚ ‚ ‚
Define the design parameters, which include the heat input for the boiler (Q B ), pressures of the primary motive steam (P0 ), secondary suction vapour (P4 ), condenser (P3 ) Define the efficiencies of the nozzle, mixing and diffuser (ηn , ηm , ηd ) . Calculate the mass flow rate of steam (m0 ) Calculate the primary flow velocity (V1 ) Calculate the Mach number (M2 ) and pressure (P2 ) at Section 2 for optimum mixing constant area (A2 ) Calculate the optimum mixing constant cross-sectional area (A2 ) Assume value of the entropy (s2 ) with pressure (P2 ) Calculate the enthalpy (h2 ) and specific volume (v2 ) Calculate the enthalpy (h3s ) at state 3s from entropy (s2 ) and pressure (P3 ) Calculate the enthalpy (h3 ) Calculate the velocity (V2 ) . Calculate the secondary mass flow rateˇ (m4 ) ˇ ˇ` . . ˘ A2 V2 ˇˇ Check convergence ˇˇ m0 ` m4 ´ ďε v2 ˇ Assume a new value of the entropy (s2 ) and carry out the above calculations again, if tolerance of convergence is no. Calculate cooling and heating capacities and COP for SER/SEHP system, if tolerance of convergence is yes. For off-design study:
‚ ‚ ‚
Calculate the Mach number (M2 ) and pressure (P2 ) with known values of A2 , k, ηd , P3 and P4 , if the boiler pressure is lower than boiler operating pressure Calculate the pressure (P2 ) with the Mach number (M2 = 1), if the boiler pressure is higher than boiler operating pressure Calculate COP for cooling and heating of SER/SEHP.
6. Results and Comparison In this study, calculations were performed for diesel engine loads of 50%, 75%, 85% and 100%. The parameters were taken as evaporator temperature TE = 4 ˝ C, condenser temperature TC = 50 ˝ C, and boiler pressures PB = 0.2, 0.3, 0.4, 0.5, 0.6, 0.7, 0.8 and 0.9 MPa. Efficiencies are assumed to be 0.90, 0.85 and 0.90 for the nozzle, mixing and diffuser processes, respectively. Tolerance of convergence is set to 1%. Specific heat ratio, k, for superheated steam is assumed to be 1.3 in the superheated steam region. Figure 4 shows that the variation of Mach number and pressure versus diffuser efficiency for the optimum mixing section area at Section 2. The results indicate that the pressure decreases as the diffuser efficiency increases at Section 2. For diffuser efficiency, ηd “ 0.90, solving Equations (33) and (34) simultaneously yields M2 = 0.7887 and P2 = 8.697 kPa.
8163
set to 1%. Specific ratio, k, for superheated steam isrespectively. assumed toTolerance be 1.3 in of the superheated 0.85 and 0.90 for theheat nozzle, mixing and diffuser processes, convergence is steam set to region. 1%. Specific heat ratio, k, for superheated steam is assumed to be 1.3 in the superheated 4 shows that the variation of Mach number and pressure versus diffuser efficiency for the steamFigure region. optimum section area at Section 2. The resultsand indicate thatversus the pressure decreases for as the Figuremixing 4 shows that the variation of Mach number pressure diffuser efficiency 0.90 diffuser efficiency increases at Section 2. For diffuser efficiency, , solving Equations (33) and optimum mixing section area at Section 2. The results indicate that the pressure decreases as the d Entropy 2015, 17, 8152–8173 0.90 diffuser efficiency increases at Section 2. For diffuser efficiency, , solving Equations (33) and (34) simultaneously yields M2 = 0.7887 and P2 = 8.697 kPa. d (34) simultaneously yields M2 = 0.7887 and P2 = 8.697 kPa.
Mach number and Pressure (kPa) Mach number and Pressure (kPa)
13 12 13 11 12 10 11 9 10 8 9 7 8 6 7 5 6 4 5 3 4 2 3 1 2 0 1 00.45
0.50
0.55
0.60
0.65
0.70
0.80
0.85
0.90
0.95
1.00
1.05
0.45
0.50
0.55
0.60
0.65
0.70 0.75 0.80 Diffuser efficiency
0.85
0.90
0.95
1.00
1.05
Pressure (kPa) Mach number Pressure (kPa) Mach number
0.75
Diffuser efficiency
Figure 4. Mach numberand andpressure pressure versus versus diffuser at Section 2. 2. Figure 4. Mach number diffuserefficiency efficiency at Section Figure 4. Mach number and pressure versus diffuser efficiency at Section 2.
The boiler heat transfer rates depend on the engine load. As the engine load increases, the boiler The heatThe transfer rates depend the load. Asthe the engine load the heat boiler increases. boiler, condenser andon evaporator heat transfer rates increase as increases, the engine loadboiler The boiler heat transfer rates depend on the engine engine load. As engine load increases, the boiler heat increases. The boiler, condenser and evaporator heat transfer rates increase as the engine increases. Figure 5 shows COPs versus boiler pressure for heating cooling given the boiler heat increases. The boiler,the condenser and the evaporator heat transfer rates and increase as the engine load load heat transfer rate. increases. Figure 5 shows thethe COPs pressurefor forheating heating and cooling given the boiler increases. Figure 5 shows COPsversus versusthe the boiler boiler pressure and cooling given the boiler heat transfer heat transfer rate.rate. 1.6 1.4 1.6 1.2 1.4
COP COP
1 1.2 Heating
0.8 1
Cooling Heating
0.6 0.8
Cooling
0.4 0.6 0.2 0.4 0 0.2 0
0.1
0.2
0.3
0.4
0.1
0.2
0.3
0.4
0.5
0.6
0.5 0.6 Boiler Pressure (MPa)
0.7
0.8
0.9
1
0.7
0.8
0.9
1
Boiler Pressure (MPa)
Figure 5. COPs versus boiler pressure. Figure5.5.COPs COPs versus versus boiler Figure boilerpressure. pressure. Figure 6 depicts the effect of boiler pressure on entrainment ratio for fixed evaporator and condenser entrainment ratio of the system increases the increasing Figuretemperatures. 6 depicts the The effect of boiler pressure on entrainment ratio with for fixed evaporatorboiler and Figure the effect of condenser boiler pressure on system entrainment for evaporator pressure6atdepicts given evaporator temperatures since increases there isratio more motive energy. condenser temperatures. Theand entrainment ratio of the with thefixed increasing boiler and condenser temperatures. The entrainment of the system increases theenergy. increasing boiler pressure at given evaporator and condenserratio temperatures since there is morewith motive
pressure at 2015, given evaporator and condenser temperatures since there is more motive energy. Entropy 17, 1–22 13 13 0.6
Entrainment Ratio
0.5 0.4 0.3 0.2 0.1 0 0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1
Boiler Pressure (MPa)
Figure Effectofofboiler boiler pressure pressure on ratio. Figure 6. 6. Effect onentrainment entrainment ratio.
Figure 7 depicts the effect of boiler pressure on area ratio for fixed evaporator and condenser temperatures. The area ratio of the system increases with the increasing boiler pressure at given 8164 evaporator and condenser temperatures. 120
Entr
0.2 0.1 0 0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1
Boiler Pressure (MPa)
Entropy 2015, 17, 8152–8173
Figure 6. Effect of boiler pressure on entrainment ratio.
FigureFigure 7 depicts the the effect of of boiler on area arearatio ratioforfor fixed evaporator and condenser 7 depicts effect boilerpressure pressure on fixed evaporator and condenser temperatures. The area ratio of the system increases with the increasing boiler pressure at given temperatures. The area ratio of the system increases with the increasing boiler pressure at given evaporator and condenser temperatures. evaporator and condenser temperatures. 120
Area Ratio
100 80 60 40 20 0 0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1
Boiler Pressure (MPa)
Figure Effectof ofboiler boiler pressure area ratio. Figure 7. 7. Effect pressureononthe the area ratio.
Figures 8 and 9 show the variation of cooling and heating capacity with change in entrainment
Figures 9 show theload. variation of cooling heating capacity with change in entrainment ratio at8 aand given engine The cooling and and heating capacities of the system increase with the ratio increase in entrainment because latter iscapacities directly proportional to theincrease mass flow ratethe given to at a given engine load. Theratio cooling andthe heating of the system with increase in the evaporator. entrainment ratio because the latter is directly proportional to the mass flow rate given to the evaporator. . , decreases with the decreasing boiler heat transfer m Of course the mass flow rate thesteam, steam, m Of course the mass flow rate ofofthe 00 , decreases with the decreasing boiler heat transfer rate at a constant boiler pressure, condenser and evaporator temperatures. Therefore, COP, heating rate at a constant boiler pressure, condenser and evaporator temperatures. Therefore, COP, heating and cooling capacity decrease but the entrainment ratio and optimum mixing area ratio do not and cooling capacity decrease but thegas entrainment ratio pressure and optimum mixing ratio not change. change. In this case, the exhaust boiler operating is increased to area achieve thedo needed In this heating case, the exhaust gas cooling boiler operating pressure is increased to achieve thegas needed heating and and especially capacity. The minimum recommended exhaust temperature can be kept ascapacity. high as 285 dependingrecommended on the sulphur content of gas the fuel oil in ordercan to avoid anyas high especially cooling The°Cminimum exhaust temperature be kept risk of hydrocarbon and ammonium sulphate condensation. Therefore, the boiler heat transfer ˝ as 285 C depending on the sulphur content of the fuel oil in order to avoid any risk of hydrocarbon rate decreases. and ammonium sulphate condensation. Therefore, the boiler heat transfer rate decreases. For example, Figures 10 and 11 show the cooling and heating capacity and COP versus boiler For example, 10 and 11 show cooling heating capacity andand COP pressures for aFigures boiler heat transfer rate of the 488 kW,. For and a boiler pressure of 0.2 MPa theversus boiler boiler pressures a boiler heat transfer rate of 488 kW,. For a boiler pressurethe of heating 0.2 MPa the boiler heat for transfer rate of 488 kW at given evaporator and condenser temperatures, andand cooling heat transfer of 488 kWtoatbegiven andkW, condenser temperatures, the heating and capacityrate are calculated 630.51evaporator kW and 142.34 respectively. COP is calculated to be 1.29 andcooling 0.29 forcalculated the heatingto and respectively. The kW, heating and coolingCOP capacities can meetto heating capacity are becooling, 630.51 kW and 142.34 respectively. is calculated be 1.29 and and cooling loads of the case naval surface ship. The residual heat, 486.51 kW in heating mode and 0.29 for the heating and cooling, respectively. The heating and cooling capacities can meet heating Entropy 2015, 17, 1–22 and cooling loads of the case naval surface ship. The residual heat, 486.51 kW in heating mode and 14 26.34 kW in cooling mode, is sent to hot water production in heating mode or to machine room cooling 26.34 kW in cooling mode, is sent to hot water production in heating mode or to machine room in cooling mode, respectively. cooling in cooling mode, respectively. 900
Cooling capacity (kW)
800 700 600
100%
500
85%
400
75% 50%
300 200 100 0 0.3
0.4
0.4
0.5
0.5
0.6
Entrainment ratio
Figure Effectofofentrainment entrainment ratio capacity. Figure 8. 8.Effect ratioon oncooling cooling capacity. 3,000
8165 acity (kW)
2,500 2,000
100% 85%
100 0.3
0.4
0.4
0 0.3
0.5
0.5
Entrainment ratio0.5 0.4
0.4
0.6
0.5
0.6
Entrainment ratio
Figure 8. Effect of entrainment ratio on cooling capacity.
Entropy 2015, 17, 8152–8173
Figure 8. Effect of entrainment ratio on cooling capacity. 3,000 3,000 2,500
2,000 Heating capacity (kW)
Heating capacity (kW)
2,500
100%
2,000
1,500 1,500
1,000 500
100%
85%
85%
75%
75% 50%
1,000
50%
500
0
0
0.3
0.35 0.35
0.3
0.4 0.4
0.45 0.45
0.5 0.5
0.55 0.55
Entrainment ratio Entrainment ratio
Figure Effect of entrainment ratio capacity. Figure 9. onheating heating capacity. Figure 9. 9. Effect of entrainment ratioon capacity.
Cooling and heating capacity (kW)
Cooling and heating capacity (kW)
800 700 600
800 700 600 500 Heating
500 400
Cooling
Heating
400 300
Cooling
300 200 200 100 0
100
0.1
0.2
0.3
0.4
0 0.1
0.2
0.3
0.4
0.5
0.6
0.7
Boiler pressure (MPa)
0.5
0.6
0.7
0.8
0.8
0.9
1
0.9
1
Boiler pressure (MPa)
Figure 10. Cooling and heating capacity versus boiler pressure.
Figure 10. 10. Cooling and heating heating capacity capacity versus versus boiler boiler pressure. pressure. Figure Cooling and
Entropy 2015, 17, 1–22
1.6
15
1.4 1.2
15
COP
1
Heating
0.8
Cooling
0.6 0.4 0.2 0 0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1
Boiler pressure (MPa)
Figure 11. COP Figure 11. COP versus versus boiler boiler pressure. pressure.
6.1. Verification of the the Results Results 6.1. Verification of The entrainment entrainment ratios ratios computed computed in in the the SER/SEHP SER/SEHP system the theoretical theoretical The system are are compared compared with with the results presented presentedin inreferences references[9,10]. [9,10].The Thenumerical numericalresults resultsfor forthe theentrainment entrainment ratios compared results ratios areare compared in in Table 4. The difference in this comparison is acceptable. There is no experimental study based on the same working conditions of this study or similar ones. 8166 Table 4. Comparison between the entrainment ratios of presented system and the references.
Boiler Pressure (P0), Temperatures of
This Study
Ref. [10]
Ref. [9]
Figure 11. COP versus boiler pressure.
6.1. Verification of the Results The entrainment ratios computed in the SER/SEHP system are compared with the theoretical results presented in references [9,10]. The numerical results for the entrainment ratios are compared in Table 4. The difference in this comparison is acceptable. There is no experimental study based on the same working conditions this studyisoracceptable. similar ones. Table 4. The difference in this of comparison There is no experimental study based on the same working conditions of this study or similar ones. Entropy 2015, 17, 8152–8173
Table 4. Comparison between the entrainment ratios of presented system and the references. Table 4. Comparison between the entrainment ratios of presented system and the references.
This Study Ref. [10] Ref. [9] Boiler Pressure (P0), Temperatures of This Study Ref. [10] Ref. [9] Boiler Pressure (P ), Temperatures of Condenser (TC) and Evaporator (TE) 0 Entrainment Ratio Entrainment Ratio Condenser (T ) and Evaporator (TE ) C
˝ C, T = 4 ˝ C P0 = 0.6 MPa,P0T=C 0.6 = 50MPa, °C, TTCE == 450°C 0.430 E ˝ C, T = 4 ˝ C P = 0.8 MPa, T = 50 0 C E P0 = 0.8 MPa, TC = 50 °C, TE = 4 °C 0.482
0.430 0.482
0.43 0.43 -
0.45 0.48
0.45 0.48
6.2. Off-Design Study There is an optimum mixing constant section area area A22 for each operating condition, yet an ejector heat pump doesn’t always operate at the design conditions. Therefore, an off-design study should be performed forfor thethe SER/SEHP system. Heating andand cooling capacity are performed to todetermine determinethe theeffect effectononCOP COP SER/SEHP system. Heating cooling capacity calculated to be 1332.58 kW and 300.83 kW, respectively, for a boiler pressure of 0.2 MPa and engine are calculated to be 1332.58 kW and 300.83 kW, respectively, for a boiler pressure of 0.2 MPa and load of load 50% at and condenser temperatures. engine of given 50% atevaporator given evaporator and condenser temperatures. 1.4 1.2
COP
1 0.8
Heating Cooling
0.6 0.4 0.2 0 0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1
Boiler pressure (MPa)
Figure for off-design. off-design. Figure 12. 12. COP COP versus versus boiler boiler pressure pressure for
16 In the system operating at the exhaust gas boiler pressure of 0.2 MPa, the optimum area ratio is determined to be ξ “ 23.30. With the designed the area ratio (ξ “ 23.30), the COP values for heating and cooling are calculated and plotted in Figure 12 for different boiler pressures. 6.3. Comparison Between an SER/SEHP System and Current System of Naval Ship The electric power inputs for heating and cooling of the case naval ship are 37.5% and 15.1% of the naval ship’s installed diesel generator power, respectively. The saved F-76 diesel fuel consumption (SFC) for heating and cooling can be calculated as: .
W net SFC “ Hu .ηDG
(53)
The electrical efficiency (ηDG ) for the diesel generator is 95%. Some values of the logistic fuel NATO F-76 are presented in Table 5 [2,21,22].
8167
6.3. Comparison Between an SER/SEHP System and Current System of Naval Ship The electric power inputs for heating and cooling of the case naval ship are 37.5% and 15.1% of the naval ship’s installed diesel generator power, respectively. The saved F-76 diesel fuel consumption (SFC) for heating and cooling can be calculated as: Entropy 2015, 17, 8152–8173
SFC
Wnet H u .DG
(53)
Table the logistic fuelisNATO F-76. values of the logistic fuel for theof diesel generator 95%. Some The electrical efficiency ( DG5.)Values NATO F-76 are presented in Table 5 [2,21,22]. Molecular Formula (Average)
C14.8 H26.9
Table Molecular 5. Values ofweight the logistic fuel NATO F-76. 205 Sulphur content, wt.% (max) 0.1 C14.8876 H26.9 Molecular Formula (Average) ˝ C, kg¨ m ´3 (max) Density, at 15 ´1 , (2015) 205 Molecular weight 3.69 Fuel price, US$¨ gallon 42,700 Lower heating value,wt.% Hu , kJ¨ kg´1 0.1 Sulphur content, (max)
876 Density, at 15 °C, kg·m−3 (max) Fuel price, US$·gallon−1, (2015) 3.69 If an SEHP system is usedLower instead of electric duct heaters in 42,700 heating mode, 14,550 L of diesel fuel heating value, Hu, kJ·kg−1
(US$14,511) can be saved and 38.30 tons of CO2 emissions reduced at the end of 1000 operating hours a year of a naval ship.isIfused an SER system is used of heating seawater cooled chiller unit fuel in cooling If an surface SEHP system instead of electric ductinstead heaters in mode, 14,550 L of diesel can befuel saved and 38.30can tonsbe of CO 2 emissions reduced of 1000 operating hours mode, (US$14,511) 5837 L of diesel (US$5822) saved and 15.36 tons at ofthe COend reduced at the end 2 emissions year of a naval surface an SER systemship. is used instead of seawater cooled chiller unit in of 1000aoperating hours a yearship. of aIfnaval surface cooling mode, 5837 L of diesel fuel (US$5822) can be saved and 15.36 tons of CO2 emissions reduced
at the end Between of 1000 operating hours aSystem year of aand naval ship.AHP System 6.4. Comparison an SER/SEHP an surface H2 O-LiBr
The of an H2 O-LiBr AHP system driven exhaust heat of a 3000 kW 6.4.reference Comparisonsystem Between consists an SER/SEHP System and an H 2O-LiBr AHP System diesel engine [2]. Figures 13 and 14 show cooling and heating capacities versus engine the SER The reference system consists of an H2O-LiBr AHP system driven exhaust heat of load a 3000for kW and SEHP an H2 O-LiBr AHP system. and heating capacities of for thethe SER and dieselsystem engine and [2]. Figures 13 and 14 show coolingThe and cooling heating capacities versus engine load SEHP and H2 O-LiBr AHP and system with the increasing load capacities at given of evaporator SER and SEHP system an Hincreases 2O-LiBr AHP system. The coolingengine and heating the SER and and SEHP and H2O-LiBr system increasesofwith the increasing engine load giventhan evaporator condenser temperatures. TheAHP cooling capacities H2 O-LiBr AHP system are at higher SER but the and condenser temperatures. The cooling capacities of H 2O-LiBr AHP system are higher than SER heating capacities are lower than SEHP for the same engine with the same load. but the heating capacities are lower than SEHP for the same engine with the same load.
Cooling capacity (kW)
1,600
AHP (90 oC) AHP (95 oC) AHP (100 oC) AHP (105 oC) AHP (110 oC) SER (0.2 MPa) SER (0.3 MPa) SER (0.4 MPa) SER (0.5 MPa) SER (0.6 MPa) SER (0.7 MPa) SER (0.8 MPa) SER (0.9 MPa)
1,400 1,200 1,000 800 600 400 200 0 50
75
85
100
Engine load (%)
Figure 13. Cooling capacity versus engine load.
Figure 13. Cooling capacity versus engine load.
Entropy 2015, 17, 1–22
17 3,000
SEHP (0.9 MPa) SEHP (0.8 MPa)
Heating capacity (kW)
2,500
SEHP (0.7 MPa) SEHP (0.6 MPa) SEHP (0.5 MPa)
2,000
SEHP (0.4 MPa)
1,500
SEHP (0.3 MPa) SEHP (0.2 MPa)
1,000
AHP (110 oC)
500
AHP (105 oC) AHP (100 oC) AHP (95 oC) AHP (90 oC)
0 50
75
85
100
Engine load (%)
Figure Heating capacity capacity versus load. Figure 14.14.Heating versusengine engine load.
6.5. Comparison between an SER/SEHP System and a VCHP System The reference system in this case consists8168 of a conventional diesel generator, which provides electricity and heat, and a vapour-compression heat pump (VCHP), which produces heating and cooling. VCHPs generally have COPs of 2–4 and deliver 2–4 times more energy than they consume. The electric motor of the compressor in a VCHP is fed by a diesel generator set on board the ship. As
Heati
1,000
AHP (110 oC)
500
AHP (105 oC) AHP (100 oC) AHP (95 oC) AHP (90 oC)
0 50
75
Entropy 2015, 17, 8152–8173
85
100
Engine load (%)
6.5. Comparison between an SER/SEHP System and a VCHP System Figure 14. Heating capacity versus engine load. The reference between systemaninSER/SEHP this case System consists a conventional diesel generator, which provides 6.5. Comparison andof a VCHP System electricity and heat, and a vapour-compression heat pump (VCHP), which produces heating and The reference system in this case consists of a conventional diesel generator, which provides cooling. VCHPs generally have COPs of 2–4 and deliver 2–4 times more energy than they consume. electricity and heat, and a vapour-compression heat pump (VCHP), which produces heating and The electric motor of the compressor in a VCHP is fed by a diesel generator set on board the ship. cooling. VCHPs generally have COPs of 2–4 and deliver 2–4 times more energy than they consume. As the energy the entire naval ship’s load fromset diesel generator sets, The electricfor motor of the compressor in aelectrical VCHP is fed byisa produced diesel generator on board the ship. As each electrical load directly the overall fuel economy emissions. the energy for the affects entire naval ship’s electrical load isand produced from diesel generator sets, each Fuel consumption and CO2the emissions decrease to produce the same amount of power given by a electrical load directly affects overall fuel economy and emissions. VCHP system when an SER/SEHP system is used. As the SER/SEHP to by replace Fuel consumption and CO2 emissions decrease to produce the same system amount is of assumed power given a VCHP systemthe when an SER/SEHP system is used.will As the SER/SEHP to replace the VCHP system, results of the system studies depend on thesystem COP is ofassumed the VCHP. the VCHP the results of the system demand studies will on to thebe COP of the VCHP. Based on system, the engine load, the heating is depend assumed 1332.58, 1774.36, 1916.14 and Based the engine load,isthe heating to demand is assumed be 1332.58, 1774.36,kW, 1916.14 and COP 2379.16 kW, theon cooling demand assumed be 300.83, 400.56,to432.57 and 537.10 and the 2379.16 kW, the cooling demand is assumed to be 300.83, 400.56, 432.57 and 537.10 kW, and the COP is assumed to be 2, 3 and 4 for the VCHP. The amounts of cooling and heating produced by the is assumed to be 2, 3 and 4 for the VCHP. The amounts of cooling and heating produced by the SER/SEHP in each system can be regarded to correspond to electricity saving. The saving is calculated SER/SEHP in each system can be regarded to correspond to electricity saving. The saving is calculated by estimating fuel and input needed to produce thethe same heating and cooling by estimating fuel the andelectricity the electricity input needed to produce same heating and coolingeffect effectusing the VCHP. The calculations are performed with COPs of 2, 3 and 4 for the VCHP. Figures 15 15 and 16 using the VCHP. The calculations are performed with COPs of 2, 3 and 4 for the VCHP. Figures present the SFC versus heating and cooling capacities for a boiler pressure of 0.2 MPa. and 16 present the SFC versus heating and cooling capacities for a boiler pressure of 0.2 MPa.
The saved fuel consumption (kgh-1)
120 100 80 VCHP of COP=2 60
VCHP of COP=3 VCHP of COP=4
40 20 0 1,200
1,400
1,600
1,800
2,000
2,200
2,400
2,600
Heating capacity (kW)
Figure15. 15.SFC SFC versus versus heating Figure heatingcapacity. capacity.
Entropy 2015, 17, 1–22
18
The saved fuel consumption (kgh-1)
25
20
15
VCHP of COP=2 VCHP of COP=3 VCHP of COP=4
10
5
0 270
320
370
420
470
520
570
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Figure 16. SFC versus cooling capacity. Figure 16. SFC versus cooling capacity.
At the end of 1000 operating hours a year of a naval surface ship, the SER/SEHP system can use more than 99.5% less electricity compared with the vapour-compression heat pump for HVAC. It can save 33,712–120,447 L of diesel fuel in the heating cycle and 7581–27,139 L of diesel fuel in the cooling cycle depending on the engine load and COP of8169 the vapour-compression heat pump. Figures 17 and 18 show the reduced CO2 emission versus heating and cooling capacities for a boiler pressure of 0.2 MPa. At the end of 1000 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
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Figure 16. SFC versus cooling capacity.
At the end of 1000 operating hours a year of a naval surface ship, the SER/SEHP system can use At the end of 1000 operating hours a year of a naval surface ship, the SER/SEHP system can use more than 99.5% less electricity compared with the vapour-compression heat pump for HVAC. It can more than 99.5% less electricity compared with the vapour-compression heat pump for HVAC. It can save 33,712–120,447 L of diesel fuel in the heating cycle and 7581–27,139 L of diesel fuel in the cooling save 33,712–120,447 L of diesel fuel in the heating cycle and 7581–27,139 L of diesel fuel in the cooling cycle depending on the engine load and COP of the vapour-compression heat pump. cycle depending on the engine load and COP of the vapour-compression heat pump. Figures 17 and 18 show the reduced CO emission versus heating and cooling capacities for a Figures 17 and 18 show the reduced CO22 emission versus heating and cooling capacities for a boiler pressure of 0.2 MPa. At the end of 1000 operating hours a year of a naval surface ship, the boiler pressure of 0.2 MPa. At the end of 1000 operating hours a year of a naval surface ship, the AHP AHP system can improve the ship’s green profile because it will reduce its annual CO2 emissions system can improve the ship’s green profile because it will reduce its annual CO2 emissions by by 88.74–317.05 tons in the heating cycle and 19.95–71.43 tons in the cooling cycle depending on the 88.74–317.05 tons in the heating cycle and 19.95–71.43 tons in the cooling cycle depending on the engine load and COP of the vapour-compression heat pump. engine load and COP of the vapour-compression heat pump. Figures 19 and 20 show the profitability plotted against the heating and cooling capacities, Figures 19 and 20 show the profitability plotted against the heating and cooling capacities, respectively. The fuel price is set to US$3.69 gallon´1 and the operating hours are 1000 h a year. respectively. The fuel price is set to US$3.69 gallon−1 and the operating hours are 1000 h a year. In this In this case, the COPs used for the VCHP are 2, 3 and 4. The SEHP system can provide an annual case, the COPs used for the VCHP are 2, 3 and 4. The SEHP system can provide an annual energy energy savings of US$33,621–US$120,122 and the SER system can provide an annual energy savings of savings of US$33,621–US$120,122 and the SER system can provide an annual energy savings of US$7561–US$27,065. US$7561–US$27,065.
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Figure 17. 17. Reduced Reduced CO CO22 emissions emissions versus versus heating heating capacity. capacity. Figure
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Figure Figure 18. 18. Reduced Reduced CO CO22 emissions emissionsversus versuscooling cooling capacity. capacity. 140,000
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Figure 18. Reduced CO2 emissions versus cooling capacity. Figure 18. Reduced CO2 emissions versus cooling capacity.
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Figure 19. Profitability of the SEHP system as a function of heating capacity. Figure Figure 19. 19. Profitability Profitability of of the the SEHP SEHP system system as as aa function function of of heating heating capacity. capacity.
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Figure 20. 20. Profitability Profitability of of the the SER SER system system as as aa function function of of cooling cooling capacity. capacity. Figure Figure 20. Profitability of the SER system as a function of cooling capacity.
7. Conclusions A seawater cooled SER/SEHP system 20 for a naval surface ship was designed and 20 thermodynamically analysed. The SER/SEHP system was compared with those of a current system in a case naval ship, an H2 O–LiBr AHP system and a VCHP system in a case naval surface ship. The dual use of ejector technology to produce heating and cooling on board a naval ship was investigated. The off-design study estimated the COPs were 0.29–0.11 for the cooling mode and 1.29–1.11 for the heating mode, depending on the pressure of the exhaust gas boiler at off-design conditions. In the system operating at the exhaust gas boiler pressure of 0.2 MPa, the optimum area ratio obtained was 23.30. The results show that the seawater-cooled SER/SEHP system not only meets the actual cooling and heating and loads of the case naval surface ship, but also provides more. The SER system is particularly attractive in applications that have a cooling demand and a source of heat, such as naval surface ships. The SEHP 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 results show that SER/SEHP 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 8171
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CO2 emissions. Moreover, as water can be used as the refrigerant in SER/SEHP, the problem of environmentally harmful refrigerants used in VCHP is avoided. Compared with current system of a case naval ship, SER and SEHP system have the advantages of lower energy consumption, CO2 emissions and EEDI. Despite the low COP values obtained in this study, the SEHP operating in the heating mode offers better performance than electrical resistance type heating. Compared with an AHP system, SER and SEHP system have the advantages of lower cost, simplicity, and minimal maintenance, even though its efficiency is still relatively low. More importantly, an absorption heat pump is prone to corrosion and occupies a large installation space, while an ejector heat pump is convenient and compact. Compared with a VCHP system, SER and SEHP system have the advantages of lower of energy consumption, CO2 emissions and EEDI. More importantly, the naval ship's susceptibility can be dramatically reduced by lowering infrared and acoustic signature. In future studies, variable area ejectors for seawater-cooled SER/SEHP system application should be investigated. Acknowledgments: The views and conclusions contained herein are those of the authors and should not be interpreted as necessarily representing official policies or endorsements, either expressed or implied, of any affiliated organisation or government. We wish to thank Mech. Eng. Azize Ezgi for helpful suggestions and critical comments. Author Contributions: Cüneyt Ezgi conceived and designed the research, analysed the data. Cüneyt Ezgi and Ibrahim Girgin worked out the theory, and wrote the manuscript. Both authors have read and approved the final manuscript. Conflicts of Interest: The authors declare no conflict of interest.
References 1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11. 12. 13. 14.
International Maritime Organization (IMO). Report of the Marine Environment Protection Committee on Its Sixty-Seventh Session; MEPC 67/6; IMO: London, UK, 2014. Ezgi, C. Design and thermodynamic analysis of an H2 O-LiBr AHP system for naval surface ship application. Int. J. Refrig. 2014, 48, 153–165. [CrossRef] Riffat, S.B.; Gan, G.; Smith, S. Computational Fluid Dynamics Applied to Ejector Heat Pumps. Appl. Therm. Eng. 1996, 16, 291–297. [CrossRef] Ma, X.; Zhang, W.; Omer, S.A.; Riffat, S.B. Experimental investigation of a novel steam ejector refrigerator suitable for solar energy applications. Appl. Therm. Eng. 2010, 30, 1320–1325. [CrossRef] Chen, X.; Omer, S.; Worall, M.; Riffat, S. Recent developments in ejector refrigeration Technologies. Renew. Sust. Energ. Rev. 2013, 19, 629–651. [CrossRef] Hsu, C.T. Investigation of an Ejector Heat Pump by Analytical Methods; Oak Ridge National Laboratuary: Oak Ridge, TN, USA, 1984. Huang, B.J.; Chang, J.M.; Wang, C.P.; Petrenko, V.A. A 1-D analysis of ejector performance. Int. J. Refrig. 1999, 22, 354–364. [CrossRef] El-Dessouky, H.; Ettouney, H.; Alatiqi, I.; Al-Nuwaibit, G. Evaluation of steam jet ejectors. Chem. Eng. Process. 2002, 41, 551–561. [CrossRef] Alexis, G.K. Estimation of ejector’s main cross sections in steam-ejector refrigeration system. Appl. Therm. Eng. 2004, 24, 2657–2663. [CrossRef] Sanaye, S.; Niroomand, B. Vertical ground coupled steam ejector heat pump; thermal-economic modeling and optimization. Int. J. Refrig. 2011, 34, 1562–1576. [CrossRef] Wärtsılä, Ship Power Technology. Wärtsılä 32 Product Guide; Wärtsılä: Vaasa, Finland, 2013. Loydu, T. Rules for the Classification of Naval Ships, Ship Operation Installations and Auxiliary Systems; Türk Loydu: Istanbul, Turkey, 2015. Ezgi, C.; Özbalta, N.; Girgin, I. Thermohydraulic and thermoeconomic performance of a marine heat exchanger on a naval surface ship. Appl. Therm. Eng. 2014, 64, 413–421. [CrossRef] Kakaç, S.; Liu, H. Heat Exchangers: Selection, Rating and Thermal Design, 2nd ed.; CRC Press: Boca Raton, FL, USA, 2002.
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15. 16. 17. 18. 19. 20. 21. 22.
Zhang, H.G.; Wang, E.H.; Fan, B.Y. Heat transfer analysis of a finned-tube evaporator for engine exhaust heat recovery. Energy Convers. Manag. 2013, 65, 438–447. [CrossRef] Cengel, Y.A.; Boles, M.A. Thermodynamics: An Engineering Approach; McGraw-Hill: Boston, MA, USA, 2001. Sun, D.W. Comparative study of the performance of an ejector refrigeration cycle operating with various refrigerants. Energy Convers. Manag. 1999, 40, 873–884. [CrossRef] Aly, N.H.; Karameldin, A.; Shamloul, M.M. Modelling and simulation of steam jet ejectors. Desalination 1999, 123, 1–8. [CrossRef] Meyer, A.J.; Harms, T.M.; Dobson, R.T. Steam jet ejector cooling powered by waste or solar heat. Renew. Energy 2009, 34, 297–306. [CrossRef] ASHRAE. Steam-Jet Refrigeration Equipment, Equipment Handbook; ASHRAE: Atlanta, GA, USA, 1979. Steinfeld, G.; Sanderson, R.; Ghezel-Ayagh, H.; Abens, S.; Cervi, M.C. Distillate Fuel Processing for Marine Fuel Cell Applications. In Proceedings of the AICHE Spring Meeting, Atlanta, GA, USA, 5–9 March 2000. Ezgi, C.; Çoban, M.T.; Selvi, Ö. Design and thermodynamic analysis of an SOFC system for naval surface ship application. ASME J. Fuel Cell Sci. Technol. 2013, 10, 031006. [CrossRef] © 2015 by the authors; licensee MDPI, Basel, Switzerland. This article is an open access article distributed under the terms and conditions of the Creative Commons by Attribution (CC-BY) license (http://creativecommons.org/licenses/by/4.0/).
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