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Applied Thermal Engineering 82 (2015) 18e29

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Applied Thermal Engineering

journal homepage: www.elsevier .com/locate/apthermeng

Research paper

Design and test of a single effect thermal desalination plant usingwaste heat from m-CHP units

Luca Cioccolanti a, *, Andrea Savoretti b, Massimiliano Renzi c, Flavio Caresana a,Gabriele Comodi a

a Dipartimento di Ingegneria Industriale e Scienze Matematiche, Universit�a Politecnica delle Marche, Ancona, AN, Italyb S.TRA.TE.G.I.E. srl, Ancona, AN, Italyc Libera Universit�a di Bolzano, Bolzano, BZ, Italy

h i g h l i g h t s

� A single effect thermal desalination plant has been designed and tested.� Waste heat from different m-CHP units can be used as energy input for the plant.� The fresh water production reached a maximum of about 7 L/h.� The plant average efficiency was ~ 1.3 L/kWh with a thermal power input of ~ 5.5 kWt.� The experimental analysis has been in good agreement with the predicted results.

a r t i c l e i n f o

Article history:Received 7 November 2014Accepted 20 February 2015Available online 3 March 2015

Keywords:Stirling engineHeat recoveryEnergy water nexusFresh water productionPrototype apparatus

* Corresponding author. Tel.: þ39 071 290 5176; faE-mail address: l.cioccolanti@univpm.it (L. Cioccol

http://dx.doi.org/10.1016/j.applthermaleng.2015.02.041359-4311/© 2015 Elsevier Ltd. All rights reserved.

a b s t r a c t

This work refers to an innovative integrated system for the simultaneous production of fresh water andelectricity. In particular, a 1 kWe Stirling engine coupled with a thermal desalination plant has beenconsidered for the purpose. The prototype, which refers to the distributed micro cogeneration field, hasthe final aim of building and testing a single effect distillation plant with a fresh water production ofabout 150 L/d.

Firstly, thermodynamic theories and numerical analysis have been carried out to define the finalprototype configuration. Then, an experimental test phase has been carried out to evaluate the actualplant performance.

The experimental analysis has been in good agreement with the predicted results. In particular, atnominal operating conditions (@50 �C) the maximum heat transfer rate was higher than the evaporatorheat exchanger designed condition (5.5 kWt). Despite the non-ideal plant thermal insulation, fresh waterproduction reached a maximum of about 7 L/h at best operating conditions, proving a good processefficiency. According to the behavior predicted by the model, fresh water production is stronglydependent on the temperature difference between the heating fluid and the salt water in the evaporatortank while it is weakly influenced by the salt content of the treated water. Moreover, the apparatusexhibited a very good response to varying thermal power input thus confirming the opportunity to feedthe desalination plant also with different forms of waste heat. More precisely, the plant average efficiencywas about 1.3 L/kWh of energy input with minimum and maximum values equal to 1.16 and 1.42 L/kWh.

Definitely the proposed solution, studied for a coupling with a 1 kWe Stirling engine, can be easilyapplied also to the other micro-CHP technologies.

© 2015 Elsevier Ltd. All rights reserved.

x: þ39 071 960 5109.anti).

7

1. Introduction

Among the possible solutions for a more efficient and rationaluse of energy, distributed micro Combined Heat and Power (m-CHP) generation has by far a huge potential in terms of secureenergy supply, high energy efficiency and low CO2 emissions [1,2].

Nomenclature

RomanA heat exchanger area [m2]CHP combined heat and powerCO2 carbon dioxidecp specific heat at constant pressure [J/kg K]D internal diameter [m]FC Fuel Cellg gravitational constant [m/s2]h heat transfer coefficient [W/m2 K]hlv latent heat [J/kg]HTC heat transfer coefficient kW/m2 KICE internal combustion engineṁ fluid flow rate [kg/s]m- CHP micro combined heat and powerMED multiple effect distillationMSF multi stage flashMT micro turbineMVC multi vapor compressionNu nusselt numberPr prandtl numberqII heat flux per unit surface [W/m2]R thermal resistance [m2$K/W]Re reynolds number

RO reverse osmosiss thickness [m]SE stirling engineT Temperature [�C]

GreekDT temperature difference [�C]l thermal conductibility [W/m K]m dynamic viscosity [Pa s]r density [kg/m3]s surface tension [Pa]

Subscriptsc cooling fluidfoul, int internal foulingfoul, ext external foulingext externalin heating fluid inputint internall liquidLM Log MeanM Meanout heating fluid outputsat evaporator tank saturation temperaturev vaporw wall

L. Cioccolanti et al. / Applied Thermal Engineering 82 (2015) 18e29 19

Cogeneration or Combined Heat and Power (CHP) is thesimultaneous production in one process of useful thermal energyand electrical and/or mechanical energy. As a result, CHP systemsexhibit an increase in the efficiency of energy conversion, which isaccompanied with a net reduction in fuel consumption and relatedgreenhouse gas emissions [3].

In the household sector the thermal energy output of m-CHPunit such as ICEs or SEs is usually recovered for heating or hot waterproduction purposes thus linking the system operating hours toclimate conditions [4]. However, connection of m-CHP with mul-tiple users by means of district heating micro-grid and according toa proper load sharing approach allows to obtain a higher number ofoperating hours and related energy and environmental benefits [5].

Another SEs potential application has been investigated inRef. [6] where Ulloa et al. studied the feasibility of adopting a SE m-CHP onboard a 35 ft yacht. The TRaNsient SYstem Simulation pro-gram has been used to model the yacht heat demand in 15 Euro-pean coastal location. Results show that a m-CHP SE has an annualrunning operation 3 times greater in Northern Europe locationsrather than in Southern Europe locations. Therefore, in warmerlocations m-CHP annual running hours are limited whilst theycould be sensibly extended using the thermal power output forcooling [7] by means of thermally activated technologies or forfresh water production not only in such onboard applications. Infact, fresh water production and supply, its sanitation and waste-water disposal consume high amount of energy. The energy-waternexus in urban utilities was investigated by Venkatesh et al. [8]referring to four case studies; for each case the authors suggestpossible solutions for the optimization in the efficiency of watertreatment. Ackerman et al. [9] analyzed the importance ofconsidering a correct energy planning in water use and productionin arid areas of western US. Lubega et al. [10] developed models tostudy the energy-water nexus that highlight the significant weightof energy flows inwater systems. Santhosh et al. [11] also proposed

a model that focus the attention on the optimization of the costsfigure in water production and management processes.

On large scale, dual-purpose plants for the combined productionof electric energy and fresh water are a well-known topic. Never-theless, the optimization of plants' operation is a crucial topic toachieve significant energy efficiency targets: Shakib et al. [12]studied the behavior of a combined cycle with steam extractionfor the thermal production of fresh water and proposed a man-agement strategy to reduce its energy impact; the same authors[13] also applied an optimization approach to a heat recovery steamgenerator downstream a gas turbine for thermal desalination pur-poses. The main thermal desalination processes includeMulti StageFlash (MSF), Multiple-Effect Distillation (MED) and MechanicalVapor Compression (MVC) accounting for about 50% of the entiremarket. The remaining desalination market is largely dominated byReverse Osmosis (RO) plants [14]. In general, thermal processesrequire high amount of energy, approximately 10e15 kWh/m3,while large scale RO plants less than 5 kWh/m3 [14]. In Refs. [15],Comodi et al. examined the use of distributed power plants coupledwith reverse osmosis systems in order to largely satisfy the demandfor freshwater and reduce the electrical energy demand by up to50% at Regional scale.

On the contrary, only few researches were conducted on smallscale water thermal desalination even though clean water supplyhas become necessary for households, small factories, laboratories,and emergency use in remote areas. In particular Sen et al. havereported in three successive papers [16e18] the design, the fabri-cation, the performance analysis and the heat transfer consider-ations of a multi-effect distillation plant showing the applicabilityof such a solution to small-scale plants.

In this work a single effect thermal desalination plant coupledwith a 1 kWel Stirling engine has been designed and tested for afresh water production of about 150 L/d. The electrical efficiency ofsuch cogeneration unit is in the range of 15%, but the overall

L. Cioccolanti et al. / Applied Thermal Engineering 82 (2015) 18e2920

efficiency can be as high as 90% because the exhausts still retain ahigh temperature and energy content and can be used for seawaterevaporation.

The feasibility of coupling a m-CHP technologies (namely SE)with a RO desalination plant has been investigated also experi-mentally by the authors. RO plants, indeed, have a considerablygreater water production, but at the expense of a higher electricenergy consumption. Therefore, the net electric power available forthe end user is reduced. In addition, the thermal power at thedischarge of the SE would not be recovered. In fact, one of theobjective of this study was also to better exploit the m-CHPdischarge heat that typically is overabundant [6].

The paper is organized as follows: after the Introduction, Section2 and 3 provide state of the art of m-CHP technologies and smallscale thermal desalination plants; Section 4 reports a shortdescription of prototype design; Section 5 describes the method-ology of the testing phase; Section 6 presents and discusses themain results of the work and Section 7 reports the conclusions.

2. Micro combined heat and power technologies

The use of m-CHP technologies in residential applications as anoption for domestic boiler is already adopted and its diffusion isexpected to increase in the next future [19]. Use of m-CHP tech-nologies embedded together with renewable systems in domesticbuildings has been extensively investigated in Refs. [20,21] provingimportant primary energy reduction.

Although challenges exist around, the m-CHP technologies canalso positively contribute to meet the targets of the energy waternexus. In household applications the most interesting m-CHPtechnologies are: Fuel Cell (FC), Stirling Engine (SE), Micro-Turbine(MT) and Internal Combustion Engine (ICE) [19].

Table 1 summarizes the main commercial m-CHP technologies[22e26] technical data:

Compared to these technologies, SE shows a lower electricalefficiency due to the fact that it is an external combustion deviceand, therefore, it operates with a closed cycle. Anyhow, electricperformance can be optimized if the thermal energy content of thecombustion exhaust are recovered, making it competitive withsmall ICEs [2]. On the other hand, the device is simpler as there is noneed to install valves or other mechanisms; in addition, its heat rateis higher, the combustion is controlled and, as a consequence, it hasthe advantage of fuel flexibility [27], low-noise operation and lowmaintenance costs [28]; these characteristics make it suitable forhouseholds and small-scale applications.

3. Small scale thermal desalination plant

Most of the ongoing researches on desalination technologies[29e31] concentrate on large-scale plants that are suitable for massproduction of fresh water while only few refer to small scalethermal desalination plants. In Ref. [29] a 100000 t/d MED seawaterdesalination plant is modeled while in Ref. [30] a hybrid MED/

Table 1m-CHP technologies etechnical data.

m-CHP technology Size (kWe) Heat recoveryrate (kWt)

Electricalefficiencyb

Thermalefficiencyb,a

Fuel cell (SOFC) >1 >0.5 35e50% 25e40%Micro turbine 30e100 60e170 26e30% 45e55%Internal combustion

engine0.3e100 ~1e150 20e37% 50e61%

Stirling engine >1 ~5.5e110 12e15% 75e90%

a Ratio between heat supplied to the end-user and primary power input.b At maximum load.

adsorption desalination system is studied and the performanceadvantages are depicted. A reviewof themost used technologies forwater desalination and their economic aspects are reported inRef. [31]. In particular, distillation proves to be one of the best andthe most economical methods for small scale production of freshwater from high saline water like seawater [32].

Solar distillation has always been very attractive thanks to theexploitation of the free energy from the sun. Several solar stillconfigurations were studied but the simplest and most practicaldesign is the single-basin type [33]. The still consists of an air tightbasin, usually made of galvanized iron sheet which contains saltwater. It has a top cover made of a transparent material, e.g. glass,and the interior surface of its base is blackened to absorb solarradiation to the maximum extent.

In Ref. [34] Kabeel and El-Said developed a hybrid solar desali-nation system for air humidification, dehumidification and waterflashing evaporation with a maximum productivity of 41.8 kg offresh water per day. They carried out an experimental investigationin order to determine the main parameters that influence the waterproduction capacity.

Refro-distiller is another proved technology for small scale ap-plications [32]. A refro-distiller uses a normal refrigeration cyclewhere the heat rejected by the condenser is used to evaporate thesalt water; steam is then condensed using the heat absorbed by theevaporator of the refrigeration cycle. This solution allows to achievesignificant savings of electrical energy if compared with traditionaldistillers that use electrical heaters.

The adsorption cycle is a practical and inexpensive method todesalt the saline and brackish water to produce potable water forboth industrial and residential applications [35]. It has also beenproved that similar desalination plant can work with heat sourcetemperatures down to 65 �C, thus allowing the recovery of low-temperature waste heat [36].

Anyway, the simplest way to obtain fresh water from salt wateris the simple traditional distillation method [37]. In this method,the saline water in a boiling flask at atmospheric pressure is heatedto boil and evaporate by a heat source. Water vapor goes into thecondensing unit which is cooled by separate cooling water.

In a single stage distillation system, salt water in the evaporatortank is usually heated by hot steam/water by means of a properheat exchanger [38] until the boiling temperature is reached. Thesteam passes through the tubes of the heat exchanger, condensesand returns back to the boiler while the salt water outside the tubesboils and evaporates. The water vapor rises and moves towards thecondenser tank where it turns to pure liquid water. To ensure thatthe heat of the steam flows towards the salt water, the condensa-tion temperature of steam has to be higher than the boiling tem-perature of the saline water.

In Refs. [39,40] Armenta-Deu designed and built a cylindricalevaporator-condenser unit which operated at 0.1 atm. Two heatexchangers inside the element play the role of the condenser andevaporator sensibly reducing the overall dimensions of the plant.However, results showed scarce performance and fresh waterproduction limited to less than 0.2 kg/kWh.

In Ref. [41] Tanaka and Park examined the use of waste heatfrom a portable electric generator for thermal desalination pur-poses. The tested distiller is a vertical single-effect diffusion still,whose structure is equivalent to the first distilling cell of a verticalmultiple-effect diffusion still. Experimental results show that40e50% of thermal energy of the exhausts can be transported bythe heat pipe and about 35% can be used for the distillation of salinewater.

Low grade heat resulted from industrial processes waste can beusefully recovered also for large scale desalination plants. InRef. [42] Ammar et all considered two potential scenarios: (i) low

L. Cioccolanti et al. / Applied Thermal Engineering 82 (2015) 18e29 21

grade heat is used directly to feed the desalination process, (ii) lowgrade heat is enhanced using a heat pump coupled with a MEDsystem. In the former, a Humidification Dehumidification processwas identified as a suitable technology due to its low operatingtemperature. In the second scenario, the low grade heat tempera-ture was upgraded using a hybrid absorption heat pump and sub-sequently used to feed a Multiple Effect Distillation desalinationsystem.

In Ref. [43] performance of a submerged vertical tube evapo-rator were extensively investigated. In particular, 175 single flutedaluminum tubes have been used to evaporate the salt water makinguse of waste heat. A vapor production rate of 3.3 t/d and an averageoverall heat transfer coefficient of about 1000 W/m2$K were ach-ieved. Nevertheless, the overall plant dimensions sensibly limitedits potential application to few extent.

4. Prototype design

A part from target specifications, prototype design has beendesigned to suite the exploitation of the heat produced by theWhisperGen EU1 Stirling engine burner. Table 2 reports the EU1Stirling engine specification (main burner only):

Anyway, in the prototype configuration Stirling engine wasreplaced by a 6 kWe electric boiler for a better control of theevaporation process and for a better comprehension of the thermaldesalination plant performance.

Because of the limited thermal power output, a simple singlestage distillation system has been considered for the purpose.Despite its lower efficiency compared tomultiple effects distillationplants, the system presented reduced overall dimensions and easyoperation that are fundamental for its application. For the experi-mental purposes, fresh water has been used both as heating andcooling fluid.

In the proposed plant, a commercial titanium helical tube heatexchanger was designed and placed co-axially to the evaporatortank to reduce the height of the salt water column. The internalheat exchange coefficient hint was calculated using the relation:

hint ¼ Nu$l

D(1)

where l is the thermal conductibility of the water from the boiler, Dis the internal diameter of the pipe and Nu is the Nusselt number.This latter parameter has been calculated according to the Colburncorrelation [44] considering the mean value of the viscosity of thefluid:

Nu ¼ 0;027$Re45$Pr

13 (2)

As regards the external side of the pipe, where the evaporationprocess takes place, the evaporator heat exchanger was designedaccording to the Nukiyama's boiling curve [45]; according to thistheory, nucleate boiling occurs for temperature difference DT be-tween the hot surface and the salt water saturation temperature inthe range 5 �C<DT < 30 �C. In particular, heat flow in boiling

Table 2EU1 stirling engine specifications.

Engine 4 cylinders double-acting Stirling cycleMain burner premix surface burnerGenerator 4 pole single phase induction motorElectrical supply 230 Ve50 HzHeat input 7 kW nominalElectrical output up to 1 kWeThermal output 5.5 kWt minCooling water flow rate 5 to 8.5 L/minMaximum flow temperature 85 �C

condition has been calculated according to the Rosehnow's formula[46]:

q00 ¼ mlhlv

�gðrl � rvÞ

s

�12

"cplðTw � TsatÞCs;lhlvPrnl

#3(3)

where qII is the heat flux per unit surface, ml is the liquid dynamicviscosity, hlv the fluid latent heat, g the gravitational constant, rland rv liquid and vapor density, s the surface tension, cpl the liquidspecific heat at constant pressure, Prl the liquid Prandtl number andCs,l and n coefficients depending on the solideliquid combination.

The external heat exchange coefficient is:

hext ¼ q00

DT(4)

The properties of the salt water were calculated according to therelations found in Ref. [47] while the fouling factors affecting theinternal and the external surface of the heat exchanger have beenfound in Ref. [48]. It is worth underlining that the fouling factorresistance have a value that is one order of magnitude higher thanthe external heat exchange coefficient one (on the boiling side) andit is sensibly higher than the internal heat exchange resistance.These high values found in literature refer to heat exchangershaving several operating hours for which the fouling process haspartially penalized the heat exchange process. Therefore, theoverall heat exchange coefficient is much more influenced by thefouling effect than by the direct heat exchange:

HTC ¼ 1PR

(5)

XR ¼ 1

hintþ Rfoul;int þ

slþ 1hext

þ Rfoul;ext (6)

The external and the internal exchange resistance are 5.0$10�5

(m2$K)/Wand 1.2$10�4 (m2 K)/W respectively and are calculated onthe basis of the aforementioned relations; the fouling resistance is9.0$10�5 (m2 K)/W for the internal side and 3.5$10�4 (m2 K)/W forthe external side.

Table 3 reports the dimension of the evaporator tank and of theheat exchanger:

Table 4 summarizes the nominal design conditions and theresulting operating conditions of the prototype unit:

Since lower coolant temperatures entail higher electrical effi-ciencies, the nominal coolant temperature has been set to 65 �C. Atthe same time, lower coolant temperatures mean lower saturationpressures in the evaporator tank to assure an adequate temperaturedifference between the coolant and the seawater thus operating atmore severe conditions.

A vacuum pump initially allows to obtain the necessary pressurein the evaporator tank to enable boiling of the seawater at tem-peratures of about 50e55 �C.Water vapor from the evaporator tankrises up and flows to an external cold heat exchanger. A proper gridat the exit of the evaporator tank limits the salt content transport tothe cold heat exchanger and prevent the risk of its corrosion.

Table 3Evaporator dimensions.

Helical tube heat exchanger Evaporator tank

Helical radius 165 mm Evaporator diameter 500 mmHelical pitch 16 mm Evaporator height 550 mmNumber of helical

turns9 Water volume to submerge

the coil30 L

Radial dimension 343 mm Water volume over the coil 60 LAxial dimension 160 mm Fresh water production ratio 2

Table 4Nominal design and operating conditions of the prototype unit.

Nominal design conditionsMaximum input thermal power 5.5 kWtSalt water saturation temperature 50 �COutput engine coolant temperature 65 �CInput engine coolant temperature 55 �CResulting operating conditionsMean temperature difference at the

heat exchanger DTM10 �C

Engine coolant flow rate 0.13 kg/sLog mean temperature difference DTLM 9.1 �CEvaporator tank pressure (@40/50/60 �C) 0.074/0.123/0.199 bar

L. Cioccolanti et al. / Applied Thermal Engineering 82 (2015) 18e2922

Thanks to this device, condenser is less prone to corrosion and ithas been made with low cost material. In particular, a commercialAISI316 stainless steel brazed plate condenser has been selectedand its overall dimensions sensibly reduced. Condensate is thencollected in a depressurized stainless steel tank and periodicallyextracted.

Since tap water is used as cooling fluid in the condenser, inlettemperature changes from one day to another with a mean value ofabout 21 �C. On the contrary, the pressure of the tank in which thecondensate is collected is correlated to the saturation temperatureof the seawater in the evaporator tank. In any case, cooling fluidflow rate has beenmanually adjusted in order to limit the influenceof the changing inlet temperature of the tap water on the plantperformance and operating conditions.

Measurements and controls have been embedded in the plant tomake the operation of the system semi-automatic. In particular,depending on the process parameters settings, heating input andpumps operation are automatically controlled whilst coolant flowrate is manually adjusted acting on a ball valve. Schneider input/output devices have been adopted and the prototype is remotelycontrolled using a Labview [49] interface.

Fig. 1 shows the thermal desalination prototype:The plant has been designed to operate in batch conditionwhich

means that the evaporator tank is initially filled with the salt waterup to its maximum. The evaporation process ends when the freesurface of the salt water reaches the top of the coil. In this way,brine salt content is sensibly higher at the end of the process.Anyway, the batch operation mode of the desalination plant iscompatible with the continuous operation of a m-CHP. Indeed, thetwo systems can be decoupled when fresh water production is notrequired or during the refilling of the evaporator tank which re-quires less than half hour. Therefore, the uninterrupted electricitysupply can be granted.

Fig. 1. The thermal desalination prototype.

5. Test bench set up

Fig. 2 reports the pipe and instrumentation diagram of thethermal desalination prototype:

During the start-up phase, the electric boiler heats by means ofan intermediate water circuit the salt water in the evaporator tankup to the saturation temperature while a vacuum pump extracts airuntil the set point pressure is reached. At regime, evaporation oc-curs and vapor from the evaporator tank flows to an external coldheat exchanger where it condenses. Fresh water is then collected ina depressurized collection tank and periodically extracted. Duringplant operation a 0.25 kWe vacuum pump runs intermittently towithdraw the non-condensable gases formed in the evaporationprocess. Since the vacuum pump operation is limited to few mi-nutes per test, its electrical consumption has been neglected.

Considering the prototype configuration, four key process pa-rameters mostly affect the salt water evaporation process:

1. Heating fluid input temperature, Tin;2. Heating fluid flow rate, ṁin;3. Evaporator tank saturation temperature, Tsat;4. Cooling fluid flow rate ṁc;

The heating fluid input temperature Tin strictly affects the heattransfer rate according to [45]. Heating fluid flow and return tem-peratures are monitored by two Pt100 resistance thermometers.The control system receives an analog input and controls theelectrical heaters according to the temperatures set-point with anon-off logic. The heating fluid flow rate ṁin affects plant perfor-mance in terms of the inner convective heat transfer coefficient ofthe heat exchanger in the evaporator tank. This flow rate iscontrolled by regulating the pump speed and it is monitored by alow accuracy pulse flow meter (0.25 L/pulse).

The saturation temperature Tsat is controlled adjusting the cor-responding saturation pressure by means of a vacuum pump. Thecooling fluid flow rate ṁc affects pressure equilibrium between theevaporator and the storage tank. In the prototype configuration, tapwater is used as cooling fluid and its mean temperature is manuallycontrolled by regulating its flow rate in order to allow thecondensation of all the vapor and keep the saturation temperatureconstant.

Two pairs of pressure transducers placed both in the evaporatorand in the storage tank allow the measuring of the water level in-side and control the plant start and stop. Whereas fresh waterproduction can be deduced from the measured pressure difference,it is weighted bymeans of an high accuracy scale for a further checkat the end of each experiment.

Table 5 reports the characteristics of the measuring and controlinstrumentations embedded in the prototype plant:

A design of experiments analysis [50] has been carried out inorder to evaluate plant performance at different operating condi-tions. In particular, three independent parameters have beenconsidered: (i) the temperature difference between the heatingfluid and the salt water saturation temperature; (ii) the heatingfluid flow rate; (iii) the evaporator tank saturation temperature;using this technique a three level full factorial plan performed for atotal of 27 experiments.

In particular, (i), (ii), (iii) have been varied in the range 13e17 �C,4.15e7.30 L/min and 40e60 �C respectively.

6. Results and discussions

Each experiment required about 5e8 h to achieve the targetfresh water production of 40 L. In particular, tests were per-formed according to the following procedure: at the beginning,

Fig. 2. P&ID of the prototype thermal desalination plant.

Table 5Measuring and control instrumentations characteristics.

Parameter controlled Sensor Number Resolution Accuracy Signal

Temperature Pt 100 resistance thermometer 8 0.1 �C ±(0.15 þ 0.002*T) �C 4e20 mAPressure Capacitive pressure transducer 4 2 mbar 10 mbar 4e20 mAFlow rate Pulse flow meter 2 0.01 L 4 pulse/L digitalElectrical conductivity Conductivity meter 1 1 mS 10 mS 4e20 mA

L. Cioccolanti et al. / Applied Thermal Engineering 82 (2015) 18e29 23

the electric heaters in the boiler warm the primary water circuitup while the vacuum pump starts and extracts air in the evap-oration tank, till the set-point pressure is reached. The electroniccontrol unit sets and controls also the temperature differencebetween the saturation temperature in the evaporation tank andthe water temperature entering the evaporator coil. As soon asthis temperature approaches the set-point, the thermal powerinput is reduced and continuously controlled thus keeping thetemperature difference close to the set point values. The testsautomatically end as the target fresh water production isobtained.

However, initial operating conditions were different for eachtest. Therefore, transient state effects have been neglected andplant performance was evaluated and compared only at regime.

In particular, plant performance is expressed in terms of: (i)thermal power input Pev; (ii) heat transfer coefficient of the evap-orator heat exchanger HTC; (iii) fresh water production _mfw; (iv)overall plant efficiency E in liters of fresh water per kWh of heattransfer in the evaporator tank.

The thermal power input is estimated according to equation (7):

Pev ¼ _mincplðTin � ToutÞ (7)

where Tin and Tout are the inlet and the outlet temperature of theheating fluid.

HTC is calculated as in equation (8):

HTC ¼_min$cpl$ðTin � ToutÞ

DTLM$A(8)

where DTLM is the Log Mean Temperature Difference and A the heatexchanger area.

Figs 3e6 report the thermal power input, the fresh water pro-duction, the heat transfer coefficient and the plant efficiency for

each experiment. For the tests carried out at the maximum hotwater flow rate (7.3 L/min), it was not possible to obtain a tem-perature difference in the hot water higher than 15.5 �C because theelectric heater could not provide the required thermal power. As aconsequence, in this working condition, plant equilibriummoves tolower hot water temperatures than the maximum set-point (about17 �C).

Curves at constant heating fluid flow rate set point are groupedtogether showing that variations of the flow rate have a highereffect than the saturation temperature on the thermal power inputand fresh water production.

In particular, the plant average efficiency was about 1.3 L/kWh ofenergy input withminimum andmaximumvalues equal to 1.16 and1.42 L/kWh.

The energy required in the transient phase has been estimatedusing the following approximate analysis. Considering a startingtemperature of the salt water in the evaporator tank of 20 �C, thesensible heat required for the start-up with the evaporation tem-perature of 40 �C and 60 �C, was calculated to be respectively 5%and 10% of the overall thermal energy demand.

Experimental data are then elaborated using a second orderanalysis with a specific Matlab tool [51] to predict plant perfor-mance at different operating conditions. This methodology allowsto have a better understanding of the behavior of the system undervarying operating conditions and it allows to extrapolate additionaloperating data points that could not be measured during the re-ported tests.

Figs 7e10 show themain plant performance data obtained usingthe above mentioned methodology. The values reported in thesquare near the x axis are the reference values of the parametersused to describe the trends. The green lines refer to the linear fit ofthe experimental data while the red dotted lines indicate the 95%global confidence intervals of variation of the performance ac-cording to the second order analysis. From the reported trend, it is

Fig. 3. Thermal power input with temperature difference.

Fig. 4. Fresh water production with temperature difference.

L. Cioccolanti et al. / Applied Thermal Engineering 82 (2015) 18e2924

possible to conclude that the performance of the system has alinear dependence with the key process parameters; for thisreason, quadratic terms are neglected while the interaction termsare considered.

In particular, it is possible to affirm that the thermal power inputslightly depends on saturation temperature while the temperaturedifference and the heating fluid flow rate have an important effect.More precisely, increase in temperature difference and heatingfluid flow rate entail 0.31 kW/K and 0.53 kW/(kg/s) increases inthermal power input.

Fig. 8 reports the trend of the fresh water production as afunction of the key process parameters. In particular, it has a lineardependence both with the temperature difference and the heatingfluid flow rate. Saturation temperature has no effect on fresh waterproduction although it affects thermal power input. This meansthat the process is less efficient at higher temperatures. The higherthe global process temperatures, the higher the heat thermal lossesin the evaporator tank.

Depending on the operating conditions, the prototype unit canreach a maximum fresh water production >7 L/h. Taking into ac-count the transient state effects (tank refilling and other minorsystem losses), the target daily production of 150 L/d can be ach-ieved at best operating conditions.

Fig. 9 shows the heat transfer coefficient with the key processparameters. More precisely, the heat transfer coefficient strictlydepends on saturation temperature due to its influence on both theinner and outer convective heat transfer coefficient. In particular,the higher the saturation temperature, the lower the kinematicviscosity and the related inner convective thermal resistance. Inaddition, according to (3) a higher saturation temperature means ahigher external heat transfer coefficient. The temperature differ-ence influences the outer convective heat transfer coefficient ac-cording to [46]. On the contrary, the heating fluid flow rateinfluences the inner convective coefficient in terms of fluid Rey-nolds number and the outer convective coefficient in terms ofMeanLog Temperature Difference (DTLM).

Fig. 5. Heat transfer coefficient with saturation temperature.

Fig. 6. Plant efficiency with saturation temperature.

L. Cioccolanti et al. / Applied Thermal Engineering 82 (2015) 18e29 25

Fig. 10 shows the trend of the plant efficiency with the keyprocess parameters. In particular, the higher the saturation tem-perature the lower the overall plant efficiency due to higher ther-mal losses in the evaporator tank. On the contrary, highertemperature difference and heating fluid flow rate have a positiveeffect on plant efficiency.

Table 6 reports the coefficient of determination (R-squared) ofthe considered response variables explained by the independentvariables, namely the saturation temperature, the temperaturedifference and the heating fluid flow rate. R-squared values showthat the regression model is in good agreement with the measuredtests apart from the plant efficiency which is weakly influenced bychanges of the independent variable.

The analysis of the variable interaction terms shows that theheating fluid flow rate and the temperature difference are moreinfluential at low saturation temperatures on the performance ofthe system. Indeed, in these conditions, inner and outer convective

coefficient values have the same order of magnitude, thus an in-crease of one of these process parameters entails an increase of theoverall HTC. On the contrary, at higher saturation temperatures, theouter convective coefficient is significantly higher than the innerone, thus reducing the potential increase of the overall HTC. Thisresult is in agreement with both the Nukiyama's boiling curve andthe Rosehnow's formula of the outer heat transfer coefficient. Theexperimental trend of the HTC is compared with the value of theHTC calculated according to the equation (8) Fig. 11.

It is possible to notice a similar trend but a sensible shift of theexperimental values to higher HTC. In the authors opinion, this ismainly due to the overestimation of the fouling heat exchangeresistance: in fact, the tubes of the heat exchanger used in theexperimental plant are new and the operating hours of the systemare few; therefore, the effect of fouling is probably much lower thanthe values suggested in literature. This implies a much better heatexchange process in the real experimental plant. Nevertheless, it is

Fig. 7. Thermal power input with key process parameters.

Fig. 8. Fresh water production with key process parameters.

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possible to conclude that the experimental and the theoreticaltrends of the HTC are in agreement and that the dimensioningprocedure proposed in this paper is correct for a small sized ther-mal desalination plant.

7. Conclusions

Most of the present combined power and desalination systemsare large-scale plants which are suitable for mass production offresh water. Only few studies were carried out on small scalethermal desalination systems and their performance is usually verylow. However, several works forecast a significant increase ofelectric energy and water demand on the medium to long term,especially in rural areas.

In this work, an innovative integrated system for the simulta-neous production of fresh water and electricity has been designedand tested. The design procedure has been reported and the cor-relations for the dimensioning of the main components have beendescribed The prototype, which refers to the distributed microcogeneration field, has the final aim of achieving a fresh waterproduction of about 150 L/d.

In particular, the single effect thermal desalination plant hasbeen designed in order to recover the thermal power output from a1 kWe Stirling engine. However, the thermal power input for thedesalination purpose could derive from any other source (wasteheat, etc.).

In order to assess the performance of the desalination plant, atest bed has been designed with the aim to evaluate the influence

Fig. 9. Heat transfer coefficient with key process parameters.

Fig. 10. Plant efficiency with key process parameters.

L. Cioccolanti et al. / Applied Thermal Engineering 82 (2015) 18e29 27

on the performance of the plant of some parameters, namelythermal power input, saturation pressure and operating tempera-tures. According to the scheduled plan of experiments, a three levelfull factorial plan has been performed and plant performance wasevaluated in batch operation mode.

Table 6Coefficient of determination of the response variables.

Response variable R-squared (coefficient of determination)

Thermal power input 0.9698Fresh water production 0.9584Heat transfer coefficient 0.7356Plant efficiency 0.2964

The experimental analysis is in good agreement with the pre-dicted results. At nominal operating conditions (50 �C of satura-tion temperature) the maximum heat transfer rate was >5.5 kWtcorresponding to the evaporator heat exchanger designed condi-tion. Despite the non-ideal plant thermal insulation, the maximumfresh water production reached about 7 L/h, proving a good pro-cess efficiency. The performance of the plant is in good accordancewith the design values, but the fouling effect on the heat exchangeis still worth of further experimental evaluations. Moreover, theapparatus exhibited a very good response to varying thermal po-wer input thus confirming the opportunity to feed it also withdifferent forms of waste heat. In conclusion, it is possible to affirmthat the response of the system can be well predicted undervarying working conditions in batch operation mode. More

Fig. 11. Heat transfer coefficient trends.

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precisely, the plant average efficiency was about 1.3 L/kWh ofenergy input with minimum and maximum values equal to 1.16and 1.42 L/kWh.

In conclusion, the test campaign has provided useful informa-tion even if additional experiments are necessary to better evaluateplant performance at different operating conditions. Indeed severalprocess parameters contribute to modify plant operating condi-tions and its overall efficiency such as the cooling fluid flow rateand inlet temperature which were manually adjusted only duringthe tests campaign.

Acknowledgements

The authors wish to thank S.TRA.TE.G.I.E. srl, PROTECNO srl andIDEA scarl for their operation and support. A special thanks also toMarche Region for partially funding this research project (BandoDDS n. 50/S11_2010 e s.m.i. DGR 1495/2010).

Appendix A. Supplementary data

Supplementary data related to this article can be found at http://dx.doi.org/10.1016/j.applthermaleng.2015.02.047.

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