Experimental and Theoretical Study of Dew Point Evaporative Cooling System Suitable for Erbil Climate

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http://epu.edu.krd/ojs/index.php/Journal

https://doi.org/10.25156/ptj.2018.8.2.205

Experimental and Theoretical Study of Dew Point Evaporative

Cooling System Suitable for Erbil Climate

Bashir E. Kareem, Riyadh A. M. Al Hayes

Department of Mechanical and Energy, Technical engineering college, Erbil Polytechnic University, Iraq

ABSTRACT

Evaporative cooling is an efficient natural process. In this study a dew point evaporative cooling system has been designed, and constructed. A dew point evaporative cooler comprised of combination of multi-stage sensible water to air heat exchangers and evaporative media. In this study the performance of the dew point evaporative cooling system is studied as theoretical and experimental work which is carried out at Erbil Polytechnic University. This work includes the estimation of the effect of impact factors on the system. Based on the conducted experimental and numerical analysis, it is concluded that the performance of the cooling system significantly depends on the; circulating air mass flow rate to the pad, water spray mass flow rate, dry bulb temperature, humidity ratio of incoming air and using return air when evaporative media is having insufficient air. The performance of the system depends on both the effectiveness of water to air heat exchangers and evaporative media. The wet bulb effectiveness of the system is ranged from (45% to 73%) for a single stage. Using return air from cooled zone improves the wet bulb effectiveness. Exhaust humid air can be used in economizer for precooling incoming air. The economizer improves the wet bulb effectiveness of the system from (73% to 93%) with using return air from

cooled zone. The ASHRAE guidelines for thermal comfort recommend (20℃ to 24℃) in the winter

and (23℃ to 27℃) in the summer and a relative humidity (RH) of (30% to 60%) is recommended.

Key words: Dew point evaporative cooler, cooling tower and heat exchanger, Dew point temperature, Evaporative cooling.

NOMENCLATURE

α_fi Surface area per unit volume, m-1

A Area, m2

c Capacity Ratio

Cp Specific heat, 𝑘𝐽/𝑘𝑔. °𝑘

D Diameter, m

G Mass velocity, kg/m2.s

h Convection heat transfer, kJ/kg.k hd Mass transfer Coefficient, kg/m2.s

i Enthalpy, kJ/kg

ifgwo Latent heat of evaporation, kJ/kg

imasw Enthalpy of saturated air, kJ/kg

ma Air mass flow rate, kg/s

ma_in Inlet air mass flow rate, kg/s

ma_supply Mass flow rate of supply air, kg/s

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PTJ vol. 8 No.2 , 2018; doi: email: [email protected]

mw Water mass flow rate, kg/s

mw_in Inlet water mass flow rate, kg/s

mw_spray Water spray mass flow rate, kg/s

Q̇ Heat transfer, kW

Rh_supply Relative humidity of supply air, %

T Temperature, ℃

t Thickness of pad, m

Ta_in Inlet air temperature in pad, ℃

Ta_out Outlet air temperature in pad, ℃

Tdb_incoming Incoming air dry bulb temperature, ℃

Tdb_supply Dry bulb temperature of supply air, ℃

Tw_in tank Water temperature in tank, ℃

Tw_in Inlet water temperature in pad, ℃

Tw_out Outlet water temperature in pad, ℃

Twb_incoming Incoming air wet bulb temperature, ℃

U Overall heat transfer coefficient, W/m2.°k

Ẇ Power, kW

W Humidity ratio, kgwater/kgdryair

w Width of pad, m

W_in Inlet air humidity, kgwater/kgdryair

Wsw Humidity ratio of saturated air, kgwater/kgdryair

ε Effectiveness, %

∆T Temperature difference, ℃

SUBSCRIPTS

a Air

c Sensible heat, cold stream

ci , co inlet and outlet of cold stream in HX db Dry bulb

dp Dew point

hi , ho inlet and outlet of hot stream in HX

HX Heat exchanger m Latent ma Moist air v Water vapor w Water wb Wet bulb ABBREVIATIONS

Btu British thermal unit

COP Coefficient of Performance

CT/HX Cooling Tower and Heat exchanger DEC Direct Evaporative Cooling

DPEC Dew Point Evaporative Cooling EC Evaporative Cooling

e-NTU Effectiveness-Number of Transfer Unit HVAC Heating, Ventilation and Air Conditioning HX/EC Heat exchanger and Evaporative Cooler IEC Indirect Evaporative Cooling

Lef Lewis Factor

REC Regenerative Evaporative Cooling

1 INTRODUCTION

Evaporative cooling is suitable for Erbil environment due to high dry bulb temperature and low relative humidity of the ambient air, and considering that it requires less power consumption. Evaporative cooling is a physical phenomenon in which evaporation of a liquid (usually water) into surrounding air cools an object or a liquid in contact with it. As the liquid turns to a gas, the phase change absorbs heat. Technically, this is called the “latent heat of evaporation”. There are some barriers of using evaporative cooling (EC), such as; high humidity ratio, water consumption and accessing of high supply air temperature. While heating ventilation and air conditioning (HVAC) system has some drawbacks, such as; high power consumption, using refrigerant is harmful to the environment when it's released to the atmosphere and it scores very low on the air change scale. The proposed system is good compromise between (EC) and (HVAC). The dew point evaporative cooling system with circulating air is modified version of indirect evaporative cooling system, which sensibly cools the air, keeping the same humidity ratio of the air, just by using water evaporation process. The water temperature can approach the wet bulb temperature of entering air. The system is installed in Erbil Polytechnic University, for performance investigation with Erbil climate. The minimum temperature of the supply air is equal to the wet bulb temperature of ambient air for direct evaporative cooling (DEC) and indirect evaporative cooling (IEC), usually it can't

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achieve thermal comfort, especially for humid regions because of inability to dehumidify air. (DEC) has better efficiency than (IEC) but it adds moisture to the air, which may cause discomfort; (DEC) can be used only when cooling and humidification is required, while (IEC) has the same humidity ratio. Theoretically, supply air temperature with this system (Dew point evaporative cooler) can approach the dew point temperature of the incoming air.

Many researchers have studied the evaporative cooling. In this section the important results and conclusions of the works have been summarized as follows; Heidarinejad et al. [1] investigated the performance of (IEC/DEC) systems. A wet bulb effectiveness value for the combination system is in the range of (1.08 to 1.11) which is higher than the wet bulb effectiveness of a single (IEC) system that was between (0.55 and 0.61). Elberling [2] tested the performance of a cross-flow (DPEC) unit based on the Maisotsenko cycle (M-cycle). The effects of inlet air dry/wet bulb temperatures, fan speed, airflow rate on the performance of cooling unit including effectiveness, cooling capacity, power consumption, energy efficiency and water evaporation rate were investigated experimentally. The test results showed that the wet-bulb effectiveness of the unit varied from (81% to 91%) over all test conditions, which is nearly (20-30%) higher than that of the typical (IEC) systems. Zhao et al. [3] found that the intake air velocity, amount of primary air extraction, and the dimensions of the heat exchanger flow passages were the primary factors affecting the overall effectiveness of the cooling system, while the temperature of the feeding water was not an important factor. Guo and Zhao [4] investigated the thermal performance of a cross-flow heat exchanger in (DPEC) system by analyzing the effects of primary and secondary air velocities, channel width, inlet relative humidity and wettability of plate. Their study suggested that smaller channel width, lower inlet relative humidity of the secondary air, higher wettability of plate and higher ratio of secondary to primary air, can improve effectiveness. Jun Xiong et al. [5] studied the (REC), The circulation volume of secondary air is one of the main factors impacting the effectiveness of (REC) system. The cooling effect was tested under different circulation volumes of secondary air in summer. They conclude that a proper range of the circulation volume is in range of (0.4 to 0.6). Jiang and Xie [6] developed a novel indirect evaporative chiller which can be used in buildings with the chilled water at a temperature closing on the dew point temperature of the outdoor air. It composed of the air-to-water heat exchanger and the air-to-water counter-current padding tower. Simulations were carried out to analyze the performance of the chiller, which include output water temperature, cooling efficiency, and the coefficient of Performance (COP). The results showed that the output water temperature provided by the prototype chiller was around (14–20℃), which was below the wet-bulb temperature and above the dew point of outdoor air. The (COP) of the prototype chiller was about (9). More than (40%) of energy can be saved by replacing conventional air conditioning systems (COP≈2-3) with this type of chiller.

2 DEW POINT EVAPORATIVE COOLING THEORY

The dew point evaporative cooling system can be developed by coupling a cooling tower with a sensible water to air heat exchanger. A system

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schematic and process representation of this cooler is shown in Fig. (1). In this process, the inlet air at state (1) is sensibly cooled to state (2) while at the same time; the chilled tower water is warmed from state (𝑤₁) 𝑡𝑜 (𝑤₂), the air at state (2) is split into two streams. One is delivered as conditioned air to the building zone and the other is routed through the cooling tower. The portion of the inlet air that is channeled through the cooling tower is humidified and warmed to state (3) and then dumped to the ambient. And warm water returns to the cooling tower at state (𝑤₂), and the water cools in tower due to water evaporation then cooling tower delivers the water at state( 𝑤₁) to the heat exchanger. The Cooling Tower/Heat Exchanger (CT/HX) air cooler clearly has great potential for air conditioning applications. It has superior temperature reducing and cooling capacity effectiveness for any value of fraction delivered. The only possible drawback to this system, in comparison to the other evaporative coolers, is its complexity. This cooler uses a cooling tower instead of an evaporative cooler. It also requires a water circulating loop to couple the two components. A smart control system would be required to maintain the cooler near its optimum operating point. One further caution regarding this cooler is that even though the temperature effectiveness is very high, this cooler is still not able to dehumidify the inlet air stream, it would have to be used in conjunction with a dehumidifier to successfully meet a total air conditioning load [9].

3 THEORETICAL WORK

3.1 Mathematical model for water to air heat exchanger

Figure (1) Schematic diagram and process representation of a cooling tower/heat exchanger air

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The effectiveness–NTU method greatly

simplified heat exchanger analysis. This method is based on a dimensionless parameter called the heat transfer effectiveness ɛ, defined as [7].

ɛ = Qactual

Q_maximum=

actual heat transfer rate

maximum possible heat transfer rate (1)

Q_actual= C_c (T_c,out− T_c,in) = C_h (T_h,out− T_h,in) (2)

∆T_max= T_h,in− T_c,in (3) Q_max= C_min (T_h,in− T_c,in) (4)

𝐶𝑚𝑖𝑛 = (𝐶ℎ= 𝑚ℎ𝐶𝑝,ℎ) or (𝐶𝑐 = 𝑚𝑐 𝐶𝑝,𝑐) NTU =U As Cmin= U A_s (ṁ C_p) min (5) c = Cmin Cmax (6) ∈= function (U As Cmin , C_min Cmax) = f (NTU , c ) (7)

3.2 Mathematical model for wet aspen pad (heat and mass transfer)

To evaluate the performance of the evaporation media for cross flow, Merkel's theory is applied to the pads [8].

Assumptions of the Merkel's theory are:

The Lewis factor relating heat and mass transfer is unity. This assumption has small influence but affects results at low ambient temperature.

The air exiting the tower is saturated with water vapor and it is characterized only by its enthalpy. This assumption regarding saturation has a negligible influence above an ambient

Figure (2) Temperature distribution in counter flow heat exchanger

Air (hot stream)

Water (cold stream) Th,in

Tc,out Th,out

Tc,in

Length of heat exchanger

Temp er at ur e ima 𝑚_𝑎(1 + 𝑤) ima+ dima 𝑚𝑎(1 + 𝑤 + 𝑑𝑤) iw 𝑚_𝑤 iw+ diw mw+ dmw 𝑑𝑧

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temperature of 20℃ but it is significant at lower temperatures.

The reduction of water flow rate by evaporation is neglected in the energy balance. This energy balance simplification has a greater influence at elevated ambient temperature.

Governing equations for wet pad are;

∂G_w ∂z = −Ga h_da_fi Ga (wsw− w) (8) ∂w ∂x = h_da_fi Ga (wsw− w) (9) ∂i_ma ∂x = 1 Ga ∂q ∂x = h_da_fi

Ga [imasw− ima+ (Lef− 1){imasw− ima− iv(wsw− w}] (10)

∂Tw

∂z =

Gahdafi

CpwGwGa[

(w_sw− w)C_pwT_w− (i_masw− i_ma) −

(Le_f− 1){i_masw− i_ma− (w_sw− w)i_v}] (11) Me_m=hdafi

G_a (Merekel number from the mathematical model) (12) Me = Ga

GwMem (13)

Me =hdafi L

Gw (14)

A mathematical model is required to design and compute the performance of the system. The system consists of the heat exchanger (water to air heat exchanger) and the heat and mass exchanger (wet aspen pad). A mathematical model for both parts is required to accomplish the design of the system. In this section the Merkel's theory applies to the wet pad, a mathematical model for the wet aspen pad achieved by solving governing equations (8, 9, 10, 11, 12, 13, and 14). The Merkel number is a non-dimensional coefficient of performance, there are some unknown parameters in (Me), and it can be calculated with trial solution in the model.

4 EXPERIMENTAL WORK

The dew point evaporative cooler with circulating air in a form of combination between sensible water to air heat exchanger and cooling tower has been designed, constructed and tested in Erbil Polytechnic University. It is a modified indirect evaporative cooler. The system consists of two main parts, which are water to air compact sensible heat exchanger with the dimensions (60*40*6) cm, and evaporative cooling media (aspen pads) with dimensions (60*60*15) cm. Three layers of aspen pad were used. The incoming (ambient) air was sensibly cooled through water to air heat exchanger, utilizing the cooled water which is pumped from atmospheric tank placed underneath the evaporative media. Part of the cooled air was diverted to the room. The flow rate of this air was

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controlled by using duct dampers and variable speed fan. The other part of the cooled air was diverted to the evaporative media after been mixed with return air from the room. The ratio of mixing was controlled by the duct dampers and variable speed fan. Water was pumped from the atmospheric tank to the heat exchangers and then pumped to evaporative media to be cooled again. The water was sprayed on wet pad and cool air passed through the pad, some water evaporates and goes into the air, the air will be humid and cool. Because high humidity causes discomfort, it was dumped to outdoor. The water molecules that evaporated had higher kinetic energy than water remain, that is why remaining water can be cooled even if both incoming air and water have higher temperatures. Water and air flow rates were measured by rotameter and anemometer respectively. Water flow rate was regulated via globe valves. Also the temperature of air and water were measured by resistance temperature detector (RTD). The system can theoretically approach the dew point temperature of the incoming air, which is the lowest temperature that can be achieved by evaporative cooling without using vapor compression refrigeration system.

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PTJ vol. 8 No.2 , 2018; doi: email: [email protected] Figure (6 b) schematic of dew point evaporative cooling system

Figure (6 c) water flow diagram of the dew point evaporative cooling system

5 RESULTS AND DISCUSSION

Analyzing theoretical and experimental results in section (5), the effect of all important parameters on water evaporation through the pad is studied and results shown in section (5.1).

5.1 Theoretical results of the heat and mass balance in wet pad

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PTJ vol. 8 No.2 , 2018; doi: email: [email protected] kg/ s % 0.05 0.1 0.15 0.2 0.25 0.3 0.35 0.5 1 1.5 2 2.5 3 3.5 4 0 eva por at ed w at er m as s f low r at e , G w _e va por at ed

water mass flow rate , mw_in

ma=0.1

ma=0.2

ma=0.3 ma=0.4

ma=0.5 Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

Tdb_in=25℃ Win=0.008kgv/kga Tw_in=30℃ C C 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 16 18 20 22 24 26 28 30 32 34 36 38 40 42 W at er out le t t em pe ra tur e , T w _out

Water temperature , Tw_in

ma=0.1

ma=0.2

ma=0.3

ma=0.4

ma=0.5

Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/ Tdb_in=25℃ Win=0.008kgv/kga mw_in=0.24 kg/s C C 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 14 16 18 20 22 24 26 28 30 W at er out le t t em pe ra tur e , T w _out

Air inlet temperature , Ta_in

mw=0.05

mw=0.1

mw=0.2

mw=0.3

Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/ ma_in=0.35 kg/s

Win=0.008kgv/kga

Tw_in=30℃

kg_water vapor/ kg_air C 0 0.002 0.004 0.006 0.008 0.01 0.012 0.014 0.016 0.018 0.02 0.022 20 21 22 23 24 25 26 27 28 29 30 W at er out le t t em pe ra tur e , T w _out

Inlet air Humidity ratio , W

ma=0.1

ma=0.2

ma=0.3

ma=0.4

ma=0.5

Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/ Tdb_in=25℃ Tw_in=30℃ mw_in=0.24 kg/s C C 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 W at er out le t t em pe ra tur e , T w _out

Water temperature , Tw_in

mw=0.05

mw=0.1

mw=0.2

mw=0.3

Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/ Tdb_in=25℃ Win=0.008 kgv/kga ma_in=0.35 kg/s kg/ s C 0 0.05 0.1 0.15 0.2 0.25 0.3 0.35 0.4 0.45 0.5 0.55 0.6 20 21 22 23 24 25 26 27 28 29 30 31 32 A ir out le t t em pe ra tur e , T a_out

Air mass flow rate , m_a

mw=0.05

mw=0.1

mw=0.2

mw=0.3

Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/ Tdb_in =36℃ Win=0.008kgv/kga Tw_in=30℃ C C 20 22.5 25 27.5 30 32.5 35 37.5 40 42.5 45 21 22 23 24 25 26 27 28 29 30 W at er out le t t em pe ra tur e , T w _out

Air inlet temperature , Ta_in

ma=0.1

ma=0.2

ma=0.3 ma=0.4

ma=0.5

Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

mw_in=0.24 kg/s Win=0.008kgv/kga Tw_in=30℃ kg/ s C 0 0.05 0.1 0.15 0.2 0.25 0.3 0.35 0.4 0.45 0.5 0.55 0.6 0.65 0.7 0.75 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 W at er out le t t em pe ra tur e , T w _out

Air mass flow rate , ma_in

mw=0.05

mw=0.1 mw=0.2

mw=0.3 mw=0.01

Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/ Tdb_in=25℃

Win=0.008kgv/kga

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PTJ vol. 8 No.2 , 2018; doi: email: [email protected] Figure (7) Effect of thermodynamic parameters on heat and mass transfer in wet pad

5.2 Experimental results of the system

All the experimental results are shown in section (5.2). The effect of important parameters is experimentally shown in this section.

5.2.1 Effect of supply air mass flow rate on the system

Figure (8) Effect of supply air mass flow rate on the system kg_water vapor/ kg_air

C 0 0.002 0.004 0.006 0.008 0.01 0.012 0.014 0.016 0.018 0.02 0.022 10 12 14 16 18 20 22 24 26 28 30 32 34 W a te r out le t t e m pe ra tur e , T w _out

Air Humidity ratio , W

mw=0.05

mw=0.1

mw=0.2 mw=0.3 Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

Tdb_in=25℃

Tw_in=30℃

ma_in=0.35 kg/s

kg_water vapor/ kg_air C 0 0.003 0.006 0.009 0.012 0.015 0.018 0.021 0.024 22 23 24 25 26 27 28 29 30 31 W a te r out le t t e m pe ra tur e , T w _out

Air Humidity ratio , W_in

Ta=20

Ta=25

Ta=30

Ta=35

Ta=40 Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

ma_in=0.35 kg/s mw_in=0.24 kg/s Tw_in=30℃ kg/ s C 0.25 0.3 0.35 0.4 0.45 0.5 0.55 0.6 0.65 0.7 0.75 20 20.5 21 21.5 22 22.5 23 23.5 24 24.5 25 T em pe ra tur e of w at er i n t ank

mass flow rate of supply air

experimental result Y(x)=24.0400292*x^0.0769554 Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

Tdb_incoming=36℃ Win=0.006kgv/kga mw_in=0.24 kg/s 24 26 28 30 32 34 36 38 % R el at ive h um id ity of s up pl y ai r, R h kg/ s C 0.25 0.3 0.35 0.4 0.45 0.5 0.55 0.6 0.65 0.7 0.75 22 24 26 28 30 32 34 36 38 dr y bul b t em pe ra tur e of s uppl y a ir

supply air relative humidity supply air temperature Y(x)=3.056338*x+23.9070423 Y(x)=-21.8169014*x+42.1478873 Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

Tdb_incoming=36℃ Win=0.006kgv/kga mw_in=0.24 kg/s kg/ s % 0.25 0.3 0.35 0.4 0.45 0.5 0.55 0.6 0.65 0.7 0.75 20 25 30 35 40 45 50 55 60 65 70 75 80 ef fe ct ive ne ss

wet bulb effectiveness dew point effectiveness Y(x)=-25.4225352*x+73.7971831 Y(x)=-15.4225352*x+45.5471831 Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

Tdb_incoming=36℃

Win=0.006kgv/kga

mw_in=0.24 kg/s

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5.2.2 Effect of water spray mass flow rate on the system

Figure (9) Effect of water spray mass flow rate on the system

5.2.3 Effect of incoming air dry bulb temperature on the system

Figure (10) Effect of incoming air dry bulb temperature on the system

5.2.4 Effect of dry bulb temperature of incoming air on wet bulb effectiveness with different supply air mass flow rate, and effect of using return air

kg/ s C 0.06 0.08 0.1 0.12 0.14 0.16 0.18 0.2 0.22 0.24 20 21 22 23 24 25 26 27 28 29 30 T em pe ra tur e of w at er i n t ank

mass flow rate of water spray

Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/ Tdb_incoming=36℃ Win=0.006kgv/kga ma_supply=0.5 kg/s ma_incoming=1.2 kg/s kg/ s C 0 0.02 0.04 0.06 0.08 0.1 0.12 0.14 0.16 0.18 0.2 0.22 0.24 0.26 0.28 0.3 15 20 25 30 35 40 45 50 55 60 dr y bul b t e m pe ra tur e of s uppl y a ir

suppy air temperature supply air relative humidity Y(x)=72.4288525*x^0.1633851 Y(x)=20.9228048*x^(-0.1260702) Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

20 30 40 50 % R el at ive h um id ity o f s up pl y a ir, R h Tdb_incoming=36℃ Win=0.006kgv/kga ma_supply=0.5 kg/s ma_incoming=1.2 kg/s kg/ s % 0.06 0.07 0.08 0.09 0.1 0.11 0.12 0.13 0.14 0.15 0.16 0.17 0.18 0.19 0.2 0.21 0.22 0.23 0.24 0.25 25 30 35 40 45 50 55 60 65 ef fe ct ive ne ss

wet bulb effectiveness dew point effectiveness Y(x)=77.8514527*x+21.524298 Y(x)=133.7657059*x+34.4806001 Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

Tdb_incoming=36℃ Win=0.006kgv/kga ma_supply=0.5 kg/s ma_incoming=1.2 kg/s C C 22 24 26 28 30 32 34 36 38 40 42 44 46 17 18 19 20 21 22 23 24 25 26 27 T em pe ra tur e of w at er in t ank

incoming air dry bulb temperature Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

mw_in=0.24 kg/s Win=0.006kgv/kga ma_supply=0.4 kg/s ma_incoming=1.0 kg/s C C 20 22 24 26 28 30 32 34 36 38 40 42 44 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 dr y bul b t em pe ra tur e of s uppl y a ir

incoming air dry bulb temperature

dry bulb temperature of supply air relative humidity of supply air Y(x)=-0.8370741*x+64.0903061 Y(x)=0.5801499*x+5.6366005 Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

25 % R el at ive h um id ity of s up pl y ai r, R h 20 30 35 40 45 mw_in=0.24 kg/s Win=0.006kgv/kga ma_supply=0.4 kg/s ma_incoming=1.0 kg/s kg/ s % 22 24 26 28 30 32 34 36 38 40 42 44 46 20 25 30 35 40 45 50 55 60 65 70 ef fe ct ive ne ss

wet bulb effectiveness dew point effectiveness Y(x)=1.1383041*x-3.0657988 Y(x)=1.2737801*x+13.8076738 Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

mw_in=0.24 kg/s

Win=0.006kgv/kga

ma_supply=0.4 kg/s

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PTJ vol. 8 No.2 , 2018; doi: email: [email protected] Figure (11) Effect of using return air on the system at different dry bub temperature

5.2.5 Effect of supply air mass flow rate on wet bulb effectiveness at different incoming air dry bulb temperature, and effect of using return air

Figure (12) Effect of using return air on the system at different supply air

5.3 Experimental and theoretical results of the system

A theoretical and experimental result is analyzing and comparing in this section, the figures shows the deviation of the model with experimental system.

Figure (13) Comparison between theoretical and experimental results C % 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 52 54 56 58 60 62 64 66 68 70 72 74 76 78 W et bul b e ff ec tive ne ss

Incoming air dry bulb temperature

m_supply=0.3 kg/ s , without return air m_supply=0.3 kg/ s , with using return air Y(x)=16.6039145*x^0.3795428 Y(x)=3.0592846*x^0.8512966 Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

Win=0.006kgv/kga mw_in=0.24 kg/s ma_incoming=0.7kg/s ma_retrn=0.2kg/s C % 28 30 32 34 36 38 40 42 44 44 46 48 50 52 54 56 58 60 62 64 66 68 70 72 74 w et bul b e ff ec ti ve ne ss

m_supply=0.7 kg/ s , without return air m_supply=0.7 kg/ s , with resturn air used Y(x)=7.7599638*x^0.5578519 Y(x)=2.6090844*x^0.8785354 Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

Win=0.006kgv/kga mw_in=0.24 kg/s ma_incoming=0.9 kg/s ma_retrn=0.5 kg/s kg/ s % 0.3 0.35 0.4 0.45 0.5 0.55 0.6 0.65 54.4 55.2 56 56.8 57.6 58.4 59.2 60 60.8 61.6 w et bul b e ff ec tive ne ss

without using return air @ incoming air Temp.=30C with using Return air @ incoming air Temp.=30C Y(x)=52.7813408*x^(-0.1226484) Y(x)=54.3286961*x^(-0.0804767) Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

Win=0.006kgv/kga mw_in=0.24 kg/s kg/ s % 0.25 0.3 0.35 0.4 0.45 0.5 0.55 0.6 0.65 0.7 0.75 54 56 58 60 62 64 66 68 70 72 w et bul b e ff ec tive ne ss

with using Return air @ incoming air Temp.=38C without using Return air @ incoming air Temp.=38C Y(x)=52.4556372*x^(-0.1893787) Y(x)=63.2789516*x^(-0.0841027) Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

Win=0.006kgv/kga mw_in=0.24 kg/s C C 22 24 26 28 30 32 34 36 38 40 42 44 46 48 17 18 19 20 21 22 23 24 25 26 27 T em pe ra tur e of w at er in t ank

incoming air dry bulb temperature experimental theoritical

Y(x)=0.4243166*x+8.0890739 Y(x)=0.4537931*x+6.8019296 Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

Win=0.006kgv/kga ma_supply=0. 4 kg/s mw_in=0.24 kg/s ma_incoming=1.0 kg/s C C 20 22 24 26 28 30 32 34 36 38 40 42 44 46 17 18 19 20 21 22 23 24 25 26 27 28 29 30 dr y bul b t em pe ra tur e of s uppl y a ir

incoming air dry bulb temperature experimental theoritical

Y(x)=1.7884279*x^ 0.7516717 Y(x)=2.2110563*x^ 0.6726383 Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

Win=0.006kgv/kga

ma_supply=0. 5 kg/s

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PTJ vol. 8 No.2 , 2018; doi: email: [email protected] Figure (14) Effect of return/incoming air ratio on the system effectiveness

5.4 Theoretical results of the system

All the theoretical result is analyzing in this section, controlling some parameter is difficult in the system to get results experimentally, so the effect of these parameters is taken theoretically.

5.4.1 Effect of humidity ratio of incoming air on the system

Figure (15) Effect of incoming air humidity ratio on the system

5.4.2 Effect of using air to air heat exchanger as an economizer in the system

% % 0 10 20 30 40 50 60 70 80 90 100 51 54 57 60 63 66 69 72 w et bul b e ff ec tive ne ss

Return air to incoming air mass ratio

T_incoming air=30 C , experimental T_incoming air=38 C , experimental T_incoming air=38 C, theoretical T_incoming air=30 C , theoretical Y(x)=0.0588462*x+55.224359 Y(x)=0.0694231*x+55.3788462 Y(x)=0.1223077*x+59.7512821 Y(x)=0.1323077*x+59.9512821 Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

Win=0.006kgv/kga ma_supply=0. 5 kg/s mw_in=0.24 kg/s kg_v/ kg_a C 0 0.0025 0.005 0.0075 0.01 0.0125 0.015 0.0175 0.02 20 21 22 23 24 25 26 27 28 29 30 T em pe ra tur e of w at er in t ank

Humidity ratio of incoming air

Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/ Ta_incoming=36 ℃

ma_supply=0. 4 kg/s

mw_in=0.24 kg/s

ma_incoming=0.95 kg/s

kg_water vapor/ kg_air C 0.003 0.004 0.005 0.0060.007 0.008 0.009 0.01 0.011 0.012 0.013 0.014 0.015 0.016 0.017 24 25 26 27 28 29 30 31 32 33 34 35 36 dr y bul b t em pe ra tur e of s uppl y a ir

supply air relative humidity supply air temperature Y(x)=305.0943396*x+24.4366038 Y(x)=-559.4339623*x+36.4396226 Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

Ta_in=36 ℃ ma_supply=0. 4 kg/s mw_in=0.24 kg/s ma_incoming=0.95 kg/s R el at ive hum idi ty of s uppl y a ir , 30 35 25 % kg_v/ kg_a % 0.005 0.0075 0.01 0.0125 0.015 20 25 30 35 40 45 50 55 60 65 70 75 80 ef fe ct ive ne ss

wet bulb effectiveness dew point effectiveness Y(x)=-1135.8490566*x+41.990566 Y(x)=-3705.6603774*x+92.6037736 Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

Ta_in=36 ℃

ma_supply=0. 4 kg/s

mw_in=0.24 kg/s

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PTJ vol. 8 No.2 , 2018; doi: email: [email protected] Figure (16) Temperature reduction with using an economizer

5.4.3 Effect of using both return air from the room and air to air heat exchanger as an economizer in the system

Figure (17) Temperature reduction with using return air and economizer

Theoretical results in section (5.1) show that, the rate of water evaporation increases and the air outlet temperature decrease with increasing water mass flow rate in the pad. The water outlet temperature increase with increasing; inlet air temperature, inlet water temperature and inlet air humidity ratio but the water outlet temperature decrease with increasing inlet air mass flow rate to the pad. When the air reaches saturation condition, the evaporation process will stop because the air cannot hold water anymore. Experimental results in section (5.2) shows that, with increasing the supply air mass flow rate in the system the water temperature in the tank increases and this increasing in water temperature will cause decrease in the system effectiveness. Also it shows that, with increasing the water spray mass flow rate, the water temperature decrease and system effectiveness increase. With increasing incoming air temperature the water temperature increase and system effectiveness decrease. Using return air will increase system effectiveness when there is insufficient air in the pad. The experimental and theoretical results are compared in section (5.3). In section (5.4), the theoretical results of

C C 28 30 32 34 36 38 40 42 44 46 48 50 2 3 4 5 6 7 8 9 10 ai r t em pe ra tur e di ff er enc e

incoming air temperature

Table: 16 items Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

C C 28 30 32 34 36 38 40 42 44 46 48 50 2 4 6 8 10 12 14 ai r t em pe ra tur e di ff er enc e

incoming air temperature

Table: 16 items Created with a trial version of Advanced Grapher - http:/ / www.alentum.com/ agrapher/

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the system indicate that, with increasing inlet air humidity ratio the water temperature in the tank increase and system effectiveness decrease. Also it shows that, the using air to air heat exchanger as an economizer will decrease the water temperature in the tank. Therefore, adding economizer to the system and using return air will increase the system effectiveness.

6 CONCLUSIONS

A theoretical model has been used to test the performance of the system which concludes; • Experimental and theoretical results are in good agreement as presented graphically in section

(5) with (4% to 10%) deviation.

• Using the return air from the cooled zone can improve the wet bulb effectiveness by (9%). • Using return air from the room in the system depends on both, the incoming air dry bulb

temperature and the supply air mass flow rate in the system. When incoming air temperature is (30℃), using return air, improves the performance of the system only when supply air mass flow rate is higher than (0.5 kg/s). But when incoming air temperature is (36℃), using return air, improves the performance of the system for all range of the supply air flow rate.

• Using exhaust humid air of the system (air leaving the evaporative media) in air to air heat exchanger as an economizer can improve the wet bulb effectiveness by (10%).

• Using both return air from the cooled zone and economizer for exhaust humid air from discharge of evaporative media can improve the wet bulb effectiveness in total by (21%) for (45℃) of incoming air temperature.

• Experimentally, the system achieves wet bulb effectiveness up to (73%) with using return air. • The wet bulb effectiveness of the system can be improved to (93%) by using both the return air

from the cooled zone and the economizer for precooling incoming air via exhaust humid air. • The dry bulb temperature of the supply air in this system ranges from (21℃ to 28℃) with (30%

to 50%) of relative humidity for a range of (25℃ to 45℃) of incoming air, without using the economizer.

• The results are within the thermal comfort criteria. The (COP) of the system is (8), while the (COP) of conventional air conditioning cooling system is (2 to 3).

REFERENCES

[1] G. Heidarinejad, M. Bozorgmehr, S. Delfani, And J. Esmaeelian, “Experimental investigation of two-stage indirect/direct evaporative cooling system in various climatic conditions”, building and environment, vol. 44, pp. 2073-2079, 2009.

[2] L. Elberling,“Laboratory evaluation of the coolerado cooler indirect evaporative cooling unit”, pacific gas and electric company, 2006.

[3] X. Zhao, J. M. Li, And S. B. Riffat, “Numerical study of a novel counter-flow heat and mass exchanger for dew point evaporative cooling”, applied thermal engineering, vol. 28, pp. 1942-1951, 2008.

[4] X. C. Guo And T. S. Zhao, “A parametric study of an indirect evaporative air cooler,” international communications in heat and mass transfer”, vol. 25, pp. 217-226, 1998.

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[5] Jun Xiong, Zehua Liu, Chao Wang And Guojie Chen, “Experimental study of the circulation air volume of recirculation evaporative cooling”, icebo2006, shenzhen, china hvac technologies for energy efficiency, vol. 4, pp. 113-130, 2006.

[6] Jiang, Y. and X. Xie, “Theoretical and testing performance of an innovative indirect evaporative chiller”, solar energy, vol. 84, pp. 2041-2055, 2010.

[7] Daniel R. Crum. 1986. Open cycle desiccant air conditioning systems. Master thesis, Wisconsin-madison University.

[8] Zhiyin Duan. 2011. Investigation of a novel dew point indirect evaporative air conditioning system for buildings. Ph.D. Dissertation, Nottingham University.

[9] Kloppers, J.C., D.G. Kroger. 2003. A critical evaluation and refinement of the performance prediction of wet-cooling towers. Ph.D. Dissertation, Stellenbosch University.

[10] Ali Pakari. 2013. Investigation of a regenerative indirect evaporative cooling system. Master thesis, Qatar University.

[11] Yngvi gudmundsson. 2012. Performance evaluation of wet-cooling tower fills with computational fluid dynamics. Master thesis, Stellenbosch University.

[12] Lorenzo Bellemo. 2011. New desiccant cooling system using the regenerative indirect evaporative process. master thesis, technical university of Denmark.

[13] Mu’azu musa. 2008. novel evaporative cooling systems for building applications. Ph.D. Dissertation, Nottingham University.