Experimental Thermal and Fluid Science 49 (2013) 152–159

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Experimental Thermal and Fluid Science

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Experimental research on the influence of the air humidity conditions inan air conditioning system

0894-1777/$ - see front matter � 2013 Elsevier Inc. All rights reserved.http://dx.doi.org/10.1016/j.expthermflusci.2013.04.013

⇑ Corresponding author.E-mail addresses: sobrinho@feg.unesp.br (P.M. Sobrinho), celso.tuna@feg.unesp.

br (C.E. Tuna).

Pedro Magalhães Sobrinho, Celso Eduardo Tuna ⇑UNESP – Univ Estadual Paulista, Engineering College, Energy Department, Av. Ariberto Pereira da Cunha 333, Guaratinguetá, São Paulo 12516-410, Brazil

a r t i c l e i n f o

Article history:Received 28 May 2012Received in revised form 18 April 2013Accepted 20 April 2013Available online 2 May 2013

Keywords:Air conditioningEnergy efficiencyRefrigeration

a b s t r a c t

The factors that have influence on the energy consumption of a small air conditioning system that areworth mentioning are the efficiencies of the compressor, evaporator and condenser, the form that therefrigerant flow is controlled, the fan model used, and climatic conditions. Within the climate issue, aninteresting factor is that the relative humidity when it comes to the effect that it causes, especially inthe performance of the air condenser, which generally is not considered in the projects. This study aimsto evaluate the influence of humidity on the coefficient of system performance (COP), seeking to quantifytheir influence when it happens. The tests were performed on a testing bench, mounted at the Laboratoryfor Energy Efficiency (LAMOTRIZ) UNESP-Campus Guaratinguetá. In the study, the wet bulb temperaturewas ranged, keeping the rotation of the scroll compressor with application of a frequency inverter in itsbest performance. The test bench is provided with a supervisory system of data collection that is also ableto control all functions of the bench. In the results, there was a significant influence, particularly whencomparing high humidity conditions with low humidity, noting that only over 65% relative humidity isthat significant changes are observed in the COP of the system.

� 2013 Elsevier Inc. All rights reserved.

1. Introduction

The importance of energy efficiency is fundamental, be for theworld economy, or to avoid rationing of electricity and even black-outs, as occurred in 2001 in Brazil, and also to combat globalwarming which is undoubtedly one of the biggest challenges ofthis century.

According to the report of the United Nations (UNs) released inThailand [1], Brazil needs to perform three actions to contain glo-bal warming as: ending the illegal deforestation, investing in cleanenergy sources like wind and solar, and apply techniques to reducethe waste of electricity, or improve energy efficiency.

The refrigeration is considered a branch of science which dealswith the processes of heat transfer and conservation to reduce thetemperature of a given volume control below the temperature ofthe surroundings [2] and therefore falls within the activities thatallow reducing the energy consumption.

The applications of refrigeration can be divided into fivecategories:

(a) Domestic: covers the manufacture of small refrigerators andfreezers.

(b) Commercial: covers the design, installation and mainte-nance of refrigerated facilities used in restaurants, hotelsand places of storage, display and processing of perishablefoods.

(c) Industrial: is varied and can be used in different industriesand manufacturing processes.

(d) Air conditioning: has as target to HVAC environments, aim-ing at the thermal comfort of its occupants.

(e) Refrigerated transport: includes temperature control in spe-cial vehicles.

An air conditioning system is a process that aims to controlsimultaneously, in an enclosed environment, purity, humidity,temperature and movement of the air. It is indispensable inmanufacturing processes that require the control of humidity,temperature and air purity, such as the manufacture ofpharmaceuticals, color printing, operating environments withflammable or toxics components, surgery rooms, in workplacesin order to increase comfort and productivity, in residences,among others.

In the early eighties, 81% of commercial buildings in the UnitedStates appealed to the use of air conditioning systems to promotethermal comfort. The estimated installed capacity was approxi-mately 100 million tons of cooling. Of these, 95% use the operatingvapor compression cycle. In some regions of Brazil, air conditioningaccounts for 7% of electricity consumption in residential use. In

Nomenclature

COP coefficient of performance, the effectiveness of therefrigeration cycle

(dh)EVA specific enthalpy change of the coolant in the evapora-tor (kJ/kg)

(dm/dt) refrigerant mass flow (kg/s)(dV/dt) volumetric flow rate of refrigerant (m3/s)ER refrigerating effect or the thermal load of the system

(kW)g acceleration due to gravity (m/s2)he specific enthalpy of the refrigerant at the evaporator in-

let (kJ/kg)hs specific enthalpy of the refrigerant at the evaporator

outlet (kJ/kg)Hp pressure loss in the suction line (mCar)HVAC heating, ventilation, and air conditioningn number of points of the samplep local atmospheric pressure (kPa)Pel electrical power drive of the motor-fanQ average air flow rate corrected to normal condition

(Nm3/s)_QR heat flow absorbed by the refrigerant fluid (kW)

R constant of air (kJ/kg K)S variance of the sample

T air temperature (K)TR tons of RefrigerationV corrected average velocity in the output section of the

installation (m/s)_WðcompÞ electrical power for compressor operation (kW)_WðfanÞ electrical power of the fan (kW)_Wðfan:condÞ electrical power of the condenser fan (kW)_WðtotalÞ sum of all the powers involved in the system (kW)

�X average of the sampleXi certain point in the sampleDpt total pressure difference between air intake and exhaust

of the motor-fan installation (Pa)Dpt(mCar) total pressure difference between air intake and exhaust

of the motor-fan installation (mCar)c specific weight of the air in the test temperature (N/m3)gtotal total efficiency of the motor-fan installationt specific volume of the refrigerant (m3/kg)qref specific mass of the refrigerant (kg/m3)q specific mass of air in the temperature of the test (kg/

m3)r standard deviation of the sample

P.M. Sobrinho, C.E. Tuna / Experimental Thermal and Fluid Science 49 (2013) 152–159 153

Minas Gerais, air conditioning accounts for 38–63% of the totalelectricity consumed in the banking sector [3].

Another estimate, published in the newspaper O Estado de SãoPaulo in 2012, is that the rate of electrical losses in Brazil is close to17%, as the world reaches the loss of 9% of the energy consumed[4].

Even the report of the IPCC (Intergovernmental Panel on Cli-mate Change) dedicated a chapter to energy efficiency and identi-fies opportunities to contain energy consumption. In addition, thereport adds that if adopted energy efficiency practices in new con-struction, emissions of gases that causes the greenhouse effectcould be reduced by 30% besides of improving air quality, socialwelfare and ensure energy security. Only with energy efficiencymeasures could reduce by about 20% of energy consumption inBrazil, and reduce CO2 emissions by 10% [1].

Due to its importance and the needy of studies on factors thathave effects on energy consumption in air conditioning systems,this study aimed to go deep in this area, emphasizing the influenceof relative humidity, using the resources available of a test bench,using refrigerant R-22, mounted on Electromotive Efficiency Labo-ratory of UNESP – Campus Guaratinguetá (LAMOTRIZ).

2. Daily behavior of the air humidity

In stable meteorological conditions, the concentration of waterin the atmosphere and so the pressure, remains constant duringthe day.

The temperature has a daily cycle, with a maximum in the earlyafternoon and a minimum during the night. Therefore, the satura-tion steam pressure, a function of temperature, presents respec-tively maximum and minimum values for the same periods asthe temperature.

The relative humidity is defined as the ratio of the partial pres-sure of water vapor in the mixture and the saturation pressure cor-responding to the dry bulb temperature in the mixture, thereforehas its minimum when the temperature is at maximum and viceversa.

These facts can be seen in Fig. 1 and 2 which show the variationin temperature and relative humidity between a maximum and aminimum during the year of 2012 in the city of Guaratinguetá-SP[5].

It is observed that the minimum relative humidity in the after-noon, come to be close to 20% in some seasons. During the night,due to the decrease of temperature, relative humidity increasesto values close to 100%.

These observations indicate that there is a tendency of changein behavior of the coefficient of performance (COP) of an air condi-tioning system, because since the conditions for heat exchangevary, certainly the performance also suffers variation.

3. Refrigeration system by steam compression

The air conditioning by steam compression can be done by theconditioning of the air through a system composed by a compres-sor, a condenser, an expansion valve or a capillary tube and anevaporator.

The steam compression cycle is the most used in practice.Among its main advantages is the possibility of equipment beingbuilt in small volumes compared to other systems and their highperformance, known in literature as the coefficient of performance(COP). Due to its portability and low cost is easily found in homesand businesses.

Before an evaluation of the performance of a refrigeration cyclecan be done, efficiency should be defined. However, the index isnot called efficiency because this term is usually reserved to de-scribe the ratio between the amount of energy obtained by theamount of energy expended. This ratio can lead to misinterpreta-tions if applied to a refrigeration system, since the outgoing energyin the condenser, is generally lost. The concept of the performanceindex of a refrigeration cycle is the same as the efficiency in thesense that it represents, as shown in the following equation:

COP ¼ Amount of what you wantAmount spent to obtain the desire amount

ð1Þ

Fig. 1. Minimun and maximum temperature of the city of Guaratinguetá-Sp in 2012.

Fig. 2. Maximum and minimum relative humidity of the of Guaratinguetá-Sp in 2012.

154 P.M. Sobrinho, C.E. Tuna / Experimental Thermal and Fluid Science 49 (2013) 152–159

The performance of a refrigeration cycle is denominated coeffi-cient of efficacy or coefficient of performance (COP) and is definedby the following equation: [6].

COP ¼ useful coolinguseful work

ð2Þ

The heat exchangers are important elements in the definition ofenergy efficiency and size of refrigeration and air conditioning sys-tems. In these applications, the heat exchangers used are calledcondensers and evaporators. Many researchers have been workingin technology development and in order to increase the perfor-mance of heat exchangers, particularly the air side [7].

There are many publications on heat exchangers, in which theypresent correlations to determine the coefficient of convective heattransfer and the friction factor of the flow. Normally the Nusseltnumber (convective heat transfer) is correlated with the Reynolds

number (flow) and Prandtl number (properties of the fluid), whilethe friction factor are correlated only with the Reynolds number.Shah and Focke [8] and Saunders [9] offer some correlations for dif-ferent geometries of plates which are correlated various parameters.

Kim et al. [10], studied the effects of humidity in the coefficientof finned heat transfer, where the inlet air temperature was 12 �Cand relative humidity ranging from 60% to 90%. In this study it isconcluded that the coefficient of heat transfer varies with thevelocity, angle of the fin and the relative humidity of the inlet airflow.

As the fins are important to the increasing of the heat transferand heat exchange, the tube-fin are commonly used for cooling,several researchers have investigated the effect of the thermalproperties of these variables on performance.

Threlkeld [11] studied the efficacy of fins, which has developedan analytical expression to the overall efficiency of the fin using the

Fig. 3. Variation of the heat transfer coefficient inside a pipe where occurs completecondensation of the overheated steam.

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potential of the enthalpies for a combined heat and mass transfer.He also assumed a linear relationship between the ambient airtemperature and saturated air temperature. The model showedthat the fin’s efficiency was affected by relative air humidity.

In other studies made by Webb and Wu [12], Threlkeld [11],Wang et al. [13] and McLaughin et al. [14], concluded that thetransfer coefficient of sensible heat transfer of a finned surface incontact with humid air and with the presence of condensation isthe same or similar that when there is no condensation.

Bourabaa et al. [15], in their work on finding the influence of in-let conditions on the air side of an evaporator, found that the heattransfer coefficients increase with the rise of relative humidityindicating that the latent heat transfer is a very significant portionunder dehumidifying conditions.

The relationship between weather conditions and frost layergrowth and the relationship between frost layer growth, weather,and overall heat transfer coefficients were studied by Monaghanet al. [16]. They found that normal fluctuations in weather condi-tions (wind, specific humidity and temperature) have a stronger ef-fect on heat transfer coefficients than the state of the frost layer.

Enshen [17] noted in a study of the influence of outside airhumidity that the annual heating or cooling energy consumptionby temperature humidity control is higher than or equal to thatby temperature control and there exists a close relationship be-tween the absolute increment of energy consumption and absolutehumidity of outside air. The absolute increase in cooling energyconsumption increases with increasing air humidity, whereasthere is a decrease in the consumption of heating energy and bothhave good linear relationship.

Yoo and Lee [18] in an experimental study on performance ofautomotive condenser and evaporator observed that the overallheat transfer coefficient of evaporator increases as air flow rate,air temperature and relative humidity increases. The coolingcapacity at relative humidity of 70% is 23% higher than that at rel-ative humidity of 50%. High relative humidity causes high air tem-perature across the evaporator, and consequently high refrigerantflow. Wang et al. [19] obtained similar results in their experimentswith an automotive air conditioning system, wherein the coolingcapacity and COP of the system increases with the increase ofthe evaporator air inlet temperature.

The influence of condenser airflow and its temperature on theperformance of an air conditioning unit and its compressor powerconsumption has been investigated and presented at differentevaporator cooling load by Elsayed and Hariri [20]. It has beenfound that a 10% reduction in compressor power consumption isachieved by increasing the condenser air flow by about 50%.

Studying the coefficient of performance (COP) of air-cooledchillers, Yang et al. [21] found in their results that the COP of thechiller can be changed by various degrees depending on the weath-er and load conditions.

The convective heat transfer from a cylinder to a humid airstream flowing normal to the cylinder was investigated experi-mentally at atmospheric pressure over a range of variables by Stillet al. [22]. In this work it was found that the heat transfer increasedwith increasing humidity. The ratio of heat transfer rates in humidair and dry air is a unique function of the molar fraction of watervapor, independent of the air temperature and flow velocity.

Xu and Yang [23] investigated the influence of outdoor air tem-perature and humidity on a heat-pump air conditioning systemdriven by gas engine in order to improve its energy performanceby using the waste heat from the gas engine. They found that withthe rising of the outdoor air dry-bulb temperature and the relativehumidity, the energy saving percentage also rises.

In the case of this study, air-cooled condenser is used, which ischaracterized by the transfer of heat absorbed directly to the out-side air.

In a normal condition and operation of the project (maximumload of the system), the refrigerant is approximately 14–16 �C war-mer than the air outside.

Compared with a water condenser, the air cooled system re-quires a greater difference in temperature between the refrigerantand outside air. Although this characteristic makes them less en-ergy efficient, its simple design allows lower installation costsand maintenance. For this reason is that the vast majority of resi-dential equipment from 5 TRs (17,585 kW, 60.000 Btu/h) to com-mercial 50 TRs (17,585 kW, 600.000 Btu/h) uses air condensers.

With manufacturers increasingly seeking to build compactmodels, the air condenser, with the fins very close, retains a lotof dust, dirt, hair, etc. If not cleaned regularly come to work withhigh discharge pressures and temperatures, the main cause ofthe defects of the compressors in these devices.

For commercial equipment, it is necessary to increase the circu-lation of air through the condenser, due to the higher frequency ofdoor opening. This is achieved by an engine forcing air against thefinned.

In air condensers, the condensation occurs inside the tubes,according to a relatively complex process. The variation of the coef-ficient of heat transfer along a tube in which complete condensa-tion occurs is shown in Fig. 3.

At the entrance, the refrigerant is in overheated steam state,presenting a relatively low coefficient typical of gas flow. The coef-ficient increases significantly as the condensation proceeds in theinner surface of the tube. However, from a particular section, thereis a progressive reduction of the coefficient of heat transfer due toincreased thickness of the condensate at the surface of the tubeand the consequent reduction of speed of fluid flow in the crosssection [6].

4. Experimental procedures

This study presents the results of an experimental study con-ducted in a laboratory (LAMOTRIZ) test bench, which demon-strated the influence of relative humidity on the coefficient ofperformance of the refrigeration cycle using a compressor rotatingspiral (scroll) with frequency inverter, in which gas compression ispossible due to the eccentricity of the mobile spiral which forms acompression chamber within the compressor housing.

The test bench (Fig. 4) consists of the following equipment:

Fig. 4. Bench test (LAMOTRIZ, UNESP-campus Guaratinguetá).

Fig. 5. Window of the supervisory control of the testing bench.

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02 frequency inverters;01 CLP:01 meter multifunction, precision: 0.5%;01 three-phase asynchronous motor – 2 poles, 1 ½ hp (1134 W),220/380 V – type high performance;01 Fan, axial flow 5000 Nm3/h, pressure 24 mmCA, eight bladesdistributed symmetrically, design pressure coefficient of 0.18,about 0.43 cube, outside diameter of 350 mm and the tip clear-ance of 10 mm, driven directly by an electric motor;01 differential pressure transducer, pressure range 0–300 mmH2O, accuracy: 0.5%;

01 pressure transmitter manifold capacitive type, pressurerange 0–300 psi, 24 VDC, 0.5% accuracy, protection IP 65, theoutput signal 4–20 mA with digital local indication;01 pressure transmitter manifold capacitive type, pressurerange 0–500 psi, 24 VDC, 0.5% accuracy, protection IP 65, theoutput signal 4–20 mA with digital local indication;02 Pressure gauge, reversible contact with 5A, range 0–100 psi;02 Pressure gauge, reversible contact with 5A, range 0–485 psi;01 air conditioning compressor, hermetic piston type with acapacity of around 26,297 BTU/h (7700 W), three-phase motorof 1.49 kW (nominal) for refrigerant R-22;

Table 1Experimental uncertainties (l).

Greatness l

Atmospheric pressure ±0.5 mmHgCross sectional area of the workbench ±0.0005 m2

Power of electric motors ±3.5 WDifference in air pressure ±0.09 mm H2ODifferential pressure of the refrigerant ±0.09 psiAir velocity ±0.03 m/sAirflow ±0.03 Nm3/sFlow rate of refrigerant ±0.012 Nm3/sAir temperature ±0.3 �CRefrigerant temperature ±0.1 �CSpecific mass of air ±0.58 kg/m3

Specific mass of refrigerant ±0.3 kg/m3

Table 2Standard deviation (r), variance and average for the motor-fan.

Flow rate (Nm3/h) Pressure difference (mmH2O) Power demand (W)

r 21.00 0.29 1.58S 447.00 0.08 2.50�X 1185.00 25.00 1338.00

P.M. Sobrinho, C.E. Tuna / Experimental Thermal and Fluid Science 49 (2013) 152–159 157

01 air conditioning compressor, scroll type (rotary), with acapacity of around 26,297 BTU/h (7700 W), three-phase motorof 1.62 kW (nominal) for refrigerant R-22;02 Transmitters of temperature range 0–200 �C, type PT inputsignal 100 and output signal of 4–20 mA;01 Condensing Unit, 24,000 BTU/h (7034.4 W);01 Evaporator, 24,000 BTU/h (7034.4 W);01 Flow Meter;01 dry bulb thermometer01 wet bulb thermometer01 Expansion valve.01 transducer speed and air temperature, speed range: 0.1–20 m/s – accuracy ± 0.03 m/s, temperature range: 0–50 �C –Accuracy ± 0.3 �C, analog output 4–20 mA;01 Capillary tube.

In the study, the wet bulb temperature was ranged, keeping therotation of the scroll compressor by applying a frequency inverterin its best performance. All tests were performed under conditionswhere they were kept constant barometric local pressure(712 mmHg (95 kPa)), the dry bulb temperature of the air rangedfrom 22 �C to 27 �C, and relative humidity in a range between40% and 92%.

The air flow system, derived from an axial fan, was kept con-stant with the engine at 60 Hz, providing an average air flow of1250 Nm3/h. The test bench is provided with supervisory system(InduSoft) data collection capable of measuring up to 102 variablesfour times per second. Fig. 3 shows the test bench which includesthe air conditioning installation.

4.1. Software used to control the supervisory system

All functions of the bench, even a simple command on and off,to vary the speed of the engines, the opening of the damper, the

Table 3Standard deviation (r), variance and average for scroll compressor.

System COP Cycle COP Power demand (W)

r 0.12 0.82 349.00S 0.0143 0.6806 121783�X 2.35 4.79 1455.86

compressors on and off, opening and closing a relief valve, etc.,Everything is done by a supervisory system. These are made possi-ble thanks to a software that is behind the figure simple andfriendly of the supervisory system, that displays on the computerscreen, performs complex mathematical functions, multiple timesper second, so that everything goes as desired.

Fig. 5 illustrates the command window of the supervisory sys-tem (InduSoft) of the test bench.

4.2. Methodology

The electrical power, such as compressors and fans, are all di-rectly measured by instruments previously described. The volu-metric flow rate of refrigerant of the air conditioning system isdirectly measured by a flow meter and the coefficient of perfor-mance (COP) of the cycle of vapor compression refrigeration is gi-ven by Eq. (3) [24].

COP ¼ ER_Wðcomp:Þ

¼_QR

_WðcompÞð3Þ

For the system studied, considering all the energy flows in-volved, the COP is given by Eq. (4) [25].

COP ¼ ER_Wðcomp:Þ þ _WðfanÞ þ _Wðfan:cond:Þ

¼_QR

_WðtotalÞð4Þ

The cooling effect or thermal load, that is the flow of heat ab-sorbed by the refrigerant is calculated from Eq. (5) [26].

ER ¼ dmdtðdhÞEVA ¼ qref

dVdtðhs � heÞEVA ¼

1t

dVdtðhs � heÞEVA ð5Þ

The variation of enthalpy of the refrigerant is obtained indi-rectly. For this purpose, measurements are made of temperatureand pressure of the fluid before and after the evaporator and theninsert these data into a software called CATT3 (Computer AidedThermodynamic Table 3) or in thermodynamic tables to obtainthe values of specific enthalpy [26]. The temperatures and pres-sures were obtained by instruments described in Section 4.

The specific volume is also obtained in CATT3, analogously tothe enthalpy values.

The electrical power drive of the motor-fan can be expressed byEq. (6) [27]. This parameter was determined and compared withthe one that has been measured directly by the instruments ofthe bench tests to estimate the total performance of all.

Pel ¼c � Q � DptðmCarÞ

gtotalð6Þ

The average flow of the air was calculated by multiplying theaverage velocity and cross-sectional area of flow. The maximumair flow was measured by a transducer that is placed in the centerof the cross section located in the exhaust of bench tests. The profilesection of the exhaust duct is a square with an area of 0.1681 m2.From this we calculated the average speed, which for the range ofReynolds number between 1.4 � 104 and 7.8 � 104 correction fac-tor calculated average was 0.81 ([28]). Also, for the universalizationof results, since the tests were performed with dry bulb tempera-ture of the environment around 23 �C and barometric pressureaveraging 712 mmHg (95 kPa), the volumetric flow rate averagewas reduced to the normal condition (0 �C and 760 mmHg). TheEq. (7) defines the difference in total pressure of the air.

Dpt ¼ q � g � Hpþ q:V2

2ð7Þ

The specific mass of the fluid considered ideal gas, a function oftemperature and pressure was determined by the followingequation:

Fig. 7. Variation of the COP of the system due to the variation of relative humidityof inlet air.

158 P.M. Sobrinho, C.E. Tuna / Experimental Thermal and Fluid Science 49 (2013) 152–159

q ¼ pR � T ð8Þ

Based on Holman [29] and Cruz et al. [30] were determined theuncertainty (l) of each primary measure, i.e., obtained without cal-culation, directly from the instrument, and from these were calcu-lated the uncertainty of results, all shown in Table 1.

From ten trials, referring to a type of experiment was calculatedthe median of the data and generated a single point. This procedurewas repeated until all the results of the experiment were obtained.

The average standard deviation (r) as shown by Eq. (9), and var-iance, as shown by Eq. (10) were calculated based on Costa Neto[31]. Table 2 presents the results for the motor–fan and Table 3for the compressor.

r ¼

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiPni�1ðXi � �XÞ2

ðn� 1Þ

sð9Þ

S ¼Pn

i¼1ðXi � �XÞ2

ðn� 1Þ ð10Þ

Fig. 8. Variation of the heat removed from air to the variation of relative humidityof inlet air.

5. Results

According to Stoecker [6], in equipments with air condensation,the cooling capacity depends mainly on the dry bulb temperatureof ambient air and the air flow through the condenser. As the con-densing temperature increases with the increasing of the ambienttemperature, the cooling capacity decreases with the increase ofthe evaporation temperature.

However, in the test results obtained and presented in Figs. 6and 7 it can be noted that the relative humidity also has a signifi-cant influence on the system, because it has influence on the heattransfer coefficient responsible for the performance of the system,as indicated in the results obtained by Bourabaa et al. [15], Still etal. [22], and Yoo and Lee [18].

In addition, Figs. 8 and 9 show that the relative humidity influ-ence on the amount of heat removed from the airflow passingthrough the evaporator and not on the amount of work done bythe compressor. Fig. 10 shows that the dry bulb temperature of in-let air does not influence the COP of the system.

Tests were performed from June to November, where it waspossible to characterize the behavior of an air conditioning systemduring the winter period, which has a low relative humidity, andduring rains period, where has a high relative humidity. The drybulb temperature and wet bulb and relative humidity varieddepending on the environmental conditions of each day ofmeasurement.

Fig. 6. Variation of the COP of the system due to the wet bulb temperature of inletair.

Fig. 9. Variation of the compressor work due to the variation of relative humidity ofinlet air.

Analyzing the results, it can be identified that there is an influ-ence of humidity on the COP of the system, identifying that for lowair humidity condition the performance of air conditioning systemhas a 40% drop in its efficiency. When it is considered a wet bulbtemperature at the air inlet, it has the same falling of only 3 �Capart.

This reflects that the performance of the air condenser, whichsuffers direct influence of the variation of the film coefficient as afunction of the relative humidity, is contrary to the behavior ofevaporative condensers, where the increase in wet bulb tempera-ture reduces the efficiency of the system [6], because the heat to

Fig. 10. Variation of the COP of the system due to the variation of the dry bulbtemperature of inlet air.

P.M. Sobrinho, C.E. Tuna / Experimental Thermal and Fluid Science 49 (2013) 152–159 159

be dissipated by the condenser becomes larger due to the effect ofwater in the system. In this case, there is a correction factor for thecapacity of the condenser which varies depending on the condens-ing temperature and wet bulb temperature reference, reducing theefficiency of the system, if the wet bulb temperature exceeds thetemperature of the project.

It should be noted that in the present study was not consideredthe effect of altitude on the relative humidity and on the perfor-mance of the air conditioning system. This influence is describedby Fumo et al. [32], which show equations relating the altitudewith the air density, with the relative humidity and the wet bulbtemperature. It also equates the influence of altitude in combus-tion process and heat exchange in thermal equipments.

6. Conclusion

In this paper an experimental analysis was performed on an airconditioning system, where was evaluated the influence of relativeair humidity on system performance, as in the literature, for airconditioning systems where air condensers are employed, it isnot considered this issue but only the dry bulb temperature inthe analysis of those systems.

In the results obtained in bench, was identified that within therange 40–65% RH does not have large variations in system perfor-mance. However, from this value the modifying of the system per-formance is quite significant, reaching the range of 70%, for thecases where the test is performed on rainy days where the RHwas high.

For the case that consider for analysis the wet bulb temperaturein the evaluation of system performance, it was not identified largevariations in the range of 16–20 �C, however, above this tempera-ture, it was noted that system performance becomes greater.

Considering that in the region and period that the analysis wasperformed, it was possible to identify that the RH have some influ-ence on the system performance, however, as most of the time theclimatic characteristics are not more than 65% of RH, it can concludethat the consideration of using only the dry bulb temperature toevaluate the performance of a system cannot be disqualified.

It is suggested as a continuation of this work, that tests should becarried out during the summer where humidity is still higher, sothat the cycle is completed and it can be possible to raise the perfor-mance of an air conditioning system during a whole year. Anothersuggestion is to develop an analysis of system performance, bythe installation of an air humidifier in front of the condenser fan.

References

[1] IPCC – Intergovernmental Panel on Climate Change, NU, Washington, 2007.[2] R.J. Dossat, Principles of Refrigeration, Emus, São Paulo, 2004. pp. 896 (in

Portuguese).[3] A.M. Oliveira, Climatization by Evaporative Cooling: Theoretical Study and

Experimental Prototypes (in Portuguese), PhD thesis, Federal University ofMinas Gerais, Belo Horizonte, 2007.

[4] O ESTADO DE SÃO PAULO, Brazilian energy loss is almost twice the worldaverage (in Portuguese).<http://www.estadao.com.br/noticias/impresso,perda-brasileira-de-energia-e-quase-o-dobro-da-media-mundial-,834727,0.htm/>. accessed on 21.02.2013.

[5] National Institute of Space Research – INPE, Brasília, Ministry of Science andTechnology, 2013

[6] W. Stoecker, J.M.S. Jabardo, Industrial Refrigeration, Edgard Blucher, São Paulo,2002 (in Portuguese).

[7] A.M. Jacobi, R.K. Shah, Air-side flow and heat transfer in compact heatexchangers: a discussion of enhancement mechanisms, Heat TransferEngineering 19 (4) (1998) 29–41.

[8] R.K. Shah, W.W. Focke, Plate heat exchangers and their design theory, in: R.K.Shah, E.C. Subbarao, R.A. Mashelkar (Eds.), Heat Transfer Equipment Design,Routledge, New York, 1988, pp. 227–254.

[9] E.A.D. Saunders, Heat Exchangers: Selection, Halsted Press, Design andConstruction, 1989. pp. 568.

[10] Man.-Hoe Kim, Sumin Song, Clark W. Bullard, Effect of inlet humiditycondition on the air-side performance of an inclined brazed aluminumevaporator, International Journal of Refrigeration 25 (2002) 611–620.

[11] James L. Threlkeld, Thomas H. Kuehn, James W. Ramsey, Thermal EnvironmentEngineering, third ed., Prentice Hall, New Jersey, 1998. pp. 740.

[12] Ralph.L. Webb, X.M. Wu, Thermal and hydraulic analysis of brazed aluminumevaporator, Applied Thermal Engineering 22 (2002) 1369–1390.

[13] C.C. Wang, Y.T. Lin, C.J. Lee, Heat and momentum transfer for compact louversfin-and-tube-heat exchangers in wet conditions, Intrenational Journal Heatand Mass Transfer 43 (2000) 3443–3452.

[14] W. McLaughlin, R. Webb, Wet Air Side Performance of Louver Fin AutomotiveEvaporators, SAE Technical Paper 2000-01-0574, 2000.

[15] A. Bourabaa, M. Saighi, I. Belal, The influence of the inlet conditions on the airside heat transfer performance of plain finned evaporator, World Academy ofScience, Engineering and Technology 59 (2011).

[16] P.F. Monaghan, F. Grealish, P.H. Oosthuizen, Frost growth and heat transfer fora row of vertical cylindrical tubes located outdoors-general trends,Experimental Thermal and Fluid Science 4 (1991) 406–417.

[17] Long Enshen, Research on the influence of air humidity on the annual heatingor cooling energy consumption, Building and Environment 40 (2005) 571–578.

[18] S.Y. Yoo, D.W. Lee, An experimental study on performance of automotivecondenser and evaporator, international refrigeration and air conditioningconference, Paper 641, 2004.

[19] S. Wang, J. Gu, T. Dickson, J. Dexter, I. McGregor, Vapor quality andperformance of an automotive air conditioning system, ExperimentalThermal and Fluid Science 30 (2005) 59–66.

[20] A. O. Elsayed, A. S. Hariri, Effect of Condenser Air Flow on the Performance ofSplit Air Conditioner, World Renewable Energy Congress, Linköping, Sweden,2011.

[21] Jia Yang, K.T. Chan, X. Wu, F.W. Yu, X. Yang, An analysis on the energyefficiency of air-cooled chillers with water mist system, Energy and Buildings55 (2012) 273–284.

[22] M. Still, H. Venzke, F. Durst, A. Melling, Influence of humidity on the convectiveheat transfer from small cylinders, Experiments in Fluids 24 (1998) 141–150.

[23] Z. Xu, Z. Yang, Saving energy in the heat-pump air conditioning system drivenby gas engine, Energy and Buildings 41 (2009) 206–211.

[24] H. Creder, Air Conditioning Installations, sixth ed., LTC, Rio de Janeiro, 2004 (inPortuguese).

[25] R. Moreira, P.M. Sobrinho, T.M. Souza, The influence of a frequency inverter inrefrigeration systems by steam compression, in: 12th Brazilian Congress ofThermal Engineering and Sciences – ENCIT 2008, Belo Horizonte, CD_ROM,2008.

[26] G. Van Wylen, R. Sonntag, C. Borgnakke, Fundamentals of Thermodynamics,seventh ed., Wiley, 2008.

[27] A.N.C. Viana, Industrial Energy Efficiency, Fans and Exhaust Fans – AdvancedGuide, Eletrobrás, Rio de Janeiro, 2004 (in Portuguese).

[28] R.W Fox, A.T Mcdonald, P.J Pritchard, Introduction to Fluid Mechanics, seventhed., Wiley, 2008.

[29] J.P. Holman, Experimental Methods for Engineers, 7th ed., McGraw-Hill,Boston, 2001.

[30] C.H.B. Cruz; H. Frangnito, I.F. Costa, B.A. Mello, Experimental Physics -Appendices and Supplements (in Portuguese), Campinas: IFGW-Unicamp,1997. <http://www.ifi.unicamp.br/~brito/graferr.pdf>.

[31] P.L.O. Costa Neto, Statistics, second ed., BLÜCHER, São Paulo, 2002.[32] N. Fumo, P.J. Mago, K. Jacobs, Design considerations for combined cooling,

heating, and power systems at altitude, Energy Conversion and Management52 (2011) 1459–1469.