Sensitivity Analysis of Evaporative Condensers Performance Using an Experimental Approach

Sensitivity Analysis of Evaporative Condensers Performance Using an Experimental Approach

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Energy Procedia 00 (2017) 000–000 Energy Procedia 00 (2017) 000–000 Energy Procedia 126 345–352 Energy Procedia 00(201709) (2017) 000–000

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nd 72 72nd Conference Conference of of the the Italian Italian Thermal Thermal Machines Machines Engineering Engineering Association, Association, ATI2017, ATI2017, 6–8 6–8 September 2017, Lecce, Italy September 2017, Lecce, Italy

Sensitivity Analysis of Condensers Performance Using Sensitivity The Analysis of Evaporative Evaporative Performance 15th International SymposiumCondensers on District Heating and Cooling Using an an Experimental Experimental Approach Approach Assessing the Maria feasibility ofaa,using the heat b,∗ demand-outdoor Maria Fiorentino Fiorentino , Giuseppe Giuseppe Starace Staraceb,∗ Cappelletta, 27058district Voghera, Italy heat demand forecast temperature functionDecsa forS.r.l., a Via long-term Decsa S.r.l., Via Cappelletta, 27058 Voghera, Italy b Department b Department

a,b,c

I. Andrić

a a

of Engineering for Innovation, University of Salento, Via per Monteroni, 73100 Lecce, Italy of Engineering for Innovation, University of Salento, Via per Monteroni, 73100 Lecce, Italy

*, A. Pinaa, P. Ferrãoa, J. Fournierb., B. Lacarrièrec, O. Le Correc

a

IN+ Center for Innovation, Technology and Policy Research - Instituto Superior Técnico, Av. Rovisco Pais 1, 1049-001 Lisbon, Portugal Abstract b Abstract Veolia Recherche & Innovation, 291 Avenue Dreyfous Daniel, 78520 Limay, France c Département Systèmes et Environnement - IMT Atlantique, 4 rueto Alfred 44300units Nantes, The evaporative condensers operateÉnergétiques at lower condensing temperatures with respect dry Kastler, condensing andFrance involve reduced The evaporative condensers operate at lower condensing temperatures with respect to dry condensing units and involve reduced water consumption if compared with water cooled condensers. A test rig to investigate the evaporative condenser at small scale has water consumption if compared with water cooled condensers. A test rig to investigate the evaporative condenser at small scale has been designed and built up. The condensing refrigerant has been simulated by electrical heaters and an air handling unit provides air been designed and built up. The condensing refrigerant has been simulated by electrical heaters and an air handling unit provides air with dry bulb temperature and relative humidity set by the user. All the parameters affecting the evaporative condenser performance with dry bulb temperature and relative humidity set by the user. All the parameters affecting the evaporative condenser performance Abstract can be monitored and adjusted, in order to carry out either an extensive experimental campaign or a sensitivity analysis. The results, can be monitored and adjusted, in order to carry out either an extensive experimental campaign or a sensitivity analysis. The results, as expected, clearly show that the heat released to air increases with the outer surface temperature of electrical heaters and decreases asDistrict expected, clearlynetworks show thatare thecommonly heat released to air increases with the outer surface temperature of electrical heaters and decreases heating therate literature as one of the most leads effective decreasing the with relative humidity. An increase of 37.5%addressed of the air in flow (at constant sprayed water) to a solutions maximumfor reduction of the with relative humidity. An increase of 37.5% of the air flow rate (at constant sprayed water) leads to a maximum reduction of the greenhouse gas of emissions from the building sector. These systems requireshowing high investments which aretransversal returned through the heat heat transfer rate 50%. Different tubes arrangements have been compared, as a decrease of the pitch involves heat transfer rate of 50%. Different tubes arrangements have been compared, showing as a decrease of the transversal pitch involves sales.performance. Due to the changed climate conditions and building renovation policies, heat demand in the future could decrease, worse worse performance. prolonging the investment return period. main of this paper isby to Elsevier assess the feasibility of using the heat demand – outdoor temperature function for heat demand ©The 2017 Thescope Authors. Published Ltd. ©forecast. 2017 The Authors. Published by Elsevier Ltd. nd nd Conference The district of Alvalade, in Lisbon (Portugal), used as aof study. Thermal The district is consisted of 665 Peer-review under responsibility the Machines Engineering Conference ofcase theItalian Italian Thermal Machines Engineering Peer-review under responsibility of thelocated scientific committee of the 72was Peer-review under responsibility of the scientific committee of the 72nd Conference of the Italian Thermal Machines Engineering Association buildings that vary in both construction period and typology. Three weather scenarios (low, medium, high) and three district Association. Association. renovation scenarios were developed (shallow, intermediate, deep). To estimate the error, obtained heat demand values were Keywords: Evaporative condensers, Experimental tests, Heat rejection, Test bench, Sensitivity analysis compared Evaporative with resultscondensers, from a dynamic heat demand model, previously developed and validated by the authors. Experimental tests, Heat rejection, Test bench, Sensitivity analysis Keywords: The results showed that when only weather change is considered, the margin of error could be acceptable for some applications (the error in annual demand was lower than 20% for all weather scenarios considered). However, after introducing renovation scenarios, the error value increased up to 59.5% (depending on the weather and renovation scenarios combination considered). The value of slope coefficient increased on average within the range of 3.8% up to 8% per decade, that corresponds to the decrease in the number of heating hours of 22-139h during the heating season (depending on the combination of weather and 1. Introduction 1.renovation Introduction scenarios considered). On the other hand, function intercept increased for 7.8-12.7% per decade (depending on the coupled scenarios). The values suggested could be used to modify the function parameters for the scenarios considered, and The is saving improve the accuracycooling of heat demand estimations. The evaporative evaporative cooling is an an energy energy saving technology technology widely widely used used both both in in the the industrial industrial refrigeration refrigeration and and in in

residential residential air air conditioning conditioning systems. systems. Savings Savings are are based based on on aa heat heat transfer transfer increase increase occurring occurring when when water water is is sprayed sprayed on heat transfer surfaces of a condensing unit and a reduced temperature difference between air and refrigerant © 2017 The Authors. Published by Elsevier Ltd. on heat transfer surfaces of a condensing unit and a reduced temperature difference between air and refrigerant is is reached; this leads to cycles at condensing temperatures and COP values. Peer-review of the Scientific Committee of The 15th International Symposium District Heating and The reached; thisunder leadsresponsibility to operate operate refrigeration refrigeration cycles at lower lower condensing temperatures and at atonhigher higher COP values. The evaporative Cooling. cooling is a well-established technology; nonetheless, the involved physical phenomena are not easy to evaporative cooling is a well-established technology; nonetheless, the involved physical phenomena are not easy to be be modelled modelled so so that that the the topic topic represents represents an an important important field field of of research research widely widely explored. explored. Parker Parker and and Treybal Treybal [1] [1] carried carried Keywords: Heat demand; Forecast; Climate change ∗ ∗

Giuseppe Starace. Tel.: +39-0809645195; fax: +39-1782745616. Giuseppe Starace. Tel.: +39-0809645195; fax: +39-1782745616. E-mail address: [email protected] E-mail address: [email protected] 1876-6102©©2017 2017The TheAuthors. Authors.Published Publishedby byElsevier ElsevierLtd. Ltd. 1876-6102 1876-6102 © under 2017 The Authors. Published by Elsevier Ltd. of Thend15th International Symposium on District Heating and Cooling. Peer-review responsibility thescientific Scientific Committee Peer-review under responsibility ofofthe committee of the 72 Conference of the Italian Thermal Machines Engineering Association. Peer-review under responsibility of the scientific committee of the 72nd Conference of the Italian Thermal Machines Engineering Association. 1876-6102 © 2017 The Authors. Published by Elsevier Ltd. Peer-review under responsibility of the scientific committee of the 72nd Conference of the Italian Thermal Machines Engineering Association 10.1016/j.egypro.2017.08.232

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out an experimental campaign on evaporative coolers and derived empirical relationships for heat and mass transfer coefficients.

Nomenclature d outer diameter [m] G˙ volumetric flow rate [m3 h−1 ] for air, [l·min−1 ] for water h specific enthalpy [kJ·kgda −1 ] m ˙ mass flow rate [kg·s−1 ] Pl longitudinal pitch [m] transversal pitch [m] Pt Q˙ heat transfer rate [kW] RH relative humidity [%] T temperature [°C] v velocity [m·s−1 ] x specific humidity of moist air [kg·kgda −1 ] Greek symbols ρ density [kg·m−3 ] Subscripts air moist air da dry air db dry bulb in air conditions before the interaction with electrical heaters max maximum mean average out air conditions after the interaction with electrical heaters setpoint conditions at the outlet of the air handling unit wall outer surface of electric heaters Mizushina et al. [2] obtained empirical relationships for heat and mass transfer coefficients corresponding to different tube diameters. Kreid et al. [3] and Korenic [4] studied the influence of fins on evaporative coolers and condensers performance. Bykov et al. [5] focused on the water temperature and air enthalpy variations with the elevation above the water basin. Webb [6] proposed a mathematical model for cooling towers, evaporative coolers and condensers, while Erens and Dreyer [7] carried out numerical analyses on both the devices. Zalewsky and Gryglaszewski [8] developed a mathematical model for counter-current evaporative condensers validated by experimental data. Ettouney et al. [9] experimentally compared the performance of evaporative condensers and dry condensing units. Qureshi and Zubair [10] modeled evaporative coolers and condensers and focused on the effect of fouling by introducing [11] a characteristic factor. Qureshi and Zubair [12] carried out an experimental campaign on evaporative coolers and condensers showing that the process fluid flow rate is the most affecting parameter for the former, while the condensing temperature and relative humidity are the relevant factors for the latter. Hajidavaloo and Eghtedari [13] found that the substitution of a dry condensing unit with an evaporative condenser in a conditioning system leads to a decrease of 20% of the power consumption and to an increase of 50% of the COP. Jahangeer et al. [14] used finite differences technique to model a single straight tube of evaporative condenser and explored the heat transfer coefficients. Tissot et al. [15] investigated on the advantages of evaporative cooling and noticed that the maximum increase of the COP is of 29%. Islam et al. [16] proposed a validated model of an air conditioning system operating with an evaporative condenser. Fiorentino and Starace [17] carried out 2D numerical activities of the falling film evaporation on horizontal tubes and studied different types of flows depending on tubes arrangement and water-to-air flow ratio. Junior and Smith-Schneider [18] carried out an experimental campaign on a small scale evaporative condenser. They collected 40 samples and obtained a predicting function of the heat transfer rate with the condensing temperature, air dry and wet bulb temperatures, water and air flow rates, water temperature. Liu et al. [19] analyzed the performance of an



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air conditioning system operating with a dual evaporative condenser. The analysis of literature showed that analytical models of the heat and mass transfer phenomena in evaporative condensers are too complex and numerical analyses are not always suitable tools to investigate condensers performance due to the associated high computational costs. A promising way to investigate evaporative condensers is so to perform experimental test at small scale, aimed at overcoming problems arising when considering condensing units as a whole. Fiorentino and Starace [20] designed and built up a test rig capable of monitoring and adjusting all the parameters affecting the evaporative condenser performance. A sensitivity analysis on the air velocity, condensing temperature, relative humidity and heat transfer geometry characteristics is presented in this work. 2. Experimental campaign 2.1. Setup The bench consists of a heat and mass transfer section, an air handling unit and a water chiller, as illustrated in Fig.1. The heat and mass transfer section is a rectangular transparent channel with dimensions 450x200x1600 mm3 , where staggered tubes are mounted on removable plates. The condensing refrigerant is simulated by three electrical heaters equipped with RTD Pt100 and PID controller for monitoring and adjusting the outer surface temperature, while the remaining tubes establish real fluid dynamics conditions. Water is distributed over the electrical heaters by three purposely designed hoses and its temperature, measured by a Pt100 sensor, is regulated by an electrical heater placed in the basin. The amount of deluge water is measured by an electromagnetic flow meter and adjusted by selecting the velocity level of the circulating pump and through a 3-way valve. Counter-current air at specified dry bulb temperature and relative humidity is supplied by an air handling unit. The heating section is made up of electric heaters, while water is used as process fluid in the cooling section. The air flow rate entering the heat and mass transfer section is measured by a differential pressure transducer and adjusted through the inverter of an induced draft fan. Cooling water is sent to a chiller condensing with air. Air is totally recirculated from the air handling unit to the heat and mass transfer section.

Fig. 1. (a) Experimental setup scheme; (b) Experimental test rig.

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2.2. Experiments The parameters set by the user on the board panel are: -

air flow rate entering the heat and mass transfer section; air dry bulb temperature and relative humidity entering the heat and mass transfer section; flow rate and temperature of water distributed over the electrical heaters; outer surface temperature of electrical heaters.

When the steady operating conditions are reached, the heat transfer geometry is characterized by measuring: - air dry bulb temperature and relative humidity before the interaction with the electrical heaters; - air dry bulb temperature and relative humidity after the interaction with the electrical heaters. Each test representative of the same operating conditions is repeated several times to ensure reproducibility of experimental results. The mass flow rate of dry air is obtained by the conditions at the outlet of the air handling unit: m ˙ da =

G˙ air · ρair,setpoint 3600s · h−1 · (1 + x setpoint )

(1)

The heat transferred to air is determined as: Q˙ = m ˙ da · (hair,out − hair,in )

(2)

3. Results The tested geometries are reported in Fig.2 : the tubes are in a staggered arrangement, with the same longitudinal pitch but with different transversal pitches. In this section the results for geometry with transversal pitch lower than 2 · do are discussed and they are compared with those referred to geometry with transversal pitch greater than 2 · do , which has been previously characterized [21]- [22].

Fig. 2. Tested geometries.

3.1. Influence of outer surface temperature The operating conditions for the test case A are summarized in Table 1.



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Table 1. Operating conditions (Case A). Pt G˙ air ,m3 h−1 T db,setpoint ,°C RH setpoint ,% G˙ water ,l·min−1 T water ,°C T wall ,°C

46/25 · do 800 23 50-60-70-80-90 1.2 26 28-30

The heat transfer rate trend with the relative humidity for different wall temperatures is represented in Fig.3. As expected the heat released to air decreases with the relative humidity and increases with the outer surface temperature. When the inlet relative humidity is lower than 75% the curves appear almost parallel, while for relative humidity higher than 75% the distance between them increases.

Fig. 3. Trend of the heat transfer rate with the inlet relative humidity and the wall temperature (Case A).

3.2. Influence of air flow rate The influence of the air flow rate on the heat transfer rate has been investigated too, by considering the operating conditions in Table 2. As observed in Fig.4, the heat released to air decreases with the air flow rate and this is due to the reduction of the water to air flow rate and the increasing air velocity. The maximum velocity occurring in the transverse plane is determined as: vair,max =

Pt vair,mean Pt − do

(3)

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350 6 Table 2. Operating conditions (Case B). Pt G˙ air ,m3 h−1 T db,setpoint ,°C RH setpoint ,% G˙ water ,l·min−1 T water ,°C T wall ,°C

46/25 · do 800-1100 23 50-60-70-80-90 1.2 26 30

where the mean velocity is referred to the cross section of the test chamber. The maximum velocities corresponding to the air flow rates of 800 and 1100 m3 h−1 are of 6.29 and 8.64ms−1 respectively. When the air velocity increases a portion of water is dragged by the air and another portion adheres to the vertical walls of the test chamber. As a consequence, the amount of evaporated water and the latent contribution to the heat transfer decreases with air velocity. An increase of 37.5% of the air flow rate leads to a decrease of the heat transfer rate of 29% if the relative humidity is lower than 70%, while it is reduced to half for relative humidity higher than 70%.

Fig. 4. Heat transfer rate versus inlet relative humidity for different air flow rates (Case B).

3.3. Influence of heat transfer geometry Two geometries with different transversal pitches have been compared in this section. The test conditions are summarized in Table 3. A reduction of the transversal pitch involves an increasing air velocity and this results in worse performance especially at higher relative humidity, as can be seen in Fig.5.



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Table 3. Operating conditions (Case C). Pt G˙ air ,m3 h−1 T db,setpoint ,°C RH setpoint ,% G˙ water ,l·min−1 T water ,°C T wall ,°C

46/25 · do -51/25 · do 800 23 50-60-70-80-90 1.2 26 28

Fig. 5. Heat transfer rate trend with relative humidity and transversal pitch (Case C).

4. Conclusions An experimental campaign on a small scale evaporative condenser is presented in this work. The bench is made up of a test chamber where condensing refrigerant is simulated by electrical heaters equipped with Pt100 sensors and PID controller for the outer surface temperature monitoring and regulation. Air at specified flow rate, dry bulb temperature and relative humidity is provided by a handling unit. The temperature and flow rate of water distributed over the electrical heaters is controlled too. The experimental tests on the geometry with transversal pitch equal to 46/25 of the outer diameter reveal that the heat transfer rate decreases with relative humidity and air velocity, while it increases with wall temperature of electrical heaters. The reduction of the transversal pitch leads to a maximum reduction of 50% of the system performance. The heat transfer geometry can be completely characterized by varying all the affecting parameters and a hybrid method will allow to extend the results at small scale to the whole evaporative condenser.

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