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Analytical study of evaporator coil in humid environment
PERGAMON
Applied Thermal Engineering 19 (1999) 1129±1145
www.elsevier.com/locate/apthermeng
Analytical study of evaporator coil in humid environment S.Y. Liang, M. Liu, T.N. Wong *, G.K. Nathan School of Mechanical and Production Engineering, Nanyang Technological University, Singapore, 639798 Received 1 November 1997; accepted 20 October 1998
Abstract In this paper, a distributed simulation model for predicting steady-state performance of a directexpansion air-cooling coil is developed. This model uses a numerical method to calculate the partially wet and totally wet ®n eciency and takes into account the refrigerant pressure drop along the coil. The model simulation of a test coil is validated with experimental data collected under dierent air conditions using R134a as a refrigerant. On the basis of this model, a number of parameters which re¯ect the characteristics of evaporator coils are analyzed in diverse humid environments and it is found that the performance of the coil is signi®cantly aected by air relative humidity. Comparison of coil performance with R134a and R12 as refrigerants, whose dierence in properties is small, shows that the dierences in coil design for a speci®ed cooling load and in the heat exchange characteristics for a given coil cannot be ignored. # 1999 Elsevier Science Ltd. All rights reserved. Keywords: Air cooling; Evaporator modeling; Performance study; Humid environment; R134a
Nomenclature A Cp d h ifg k L Le m q
eective heat transfer area (m2) isobaric speci®c heat (kJ/kg8C) tube diameter (m) speci®c enthalpy (kJ/kg) latent heat of moisture condensation (kJ/kg) thermal conductivity (kW/m8C) tube length (m) Lewis number mass ¯ow rate (kg/s) heat transfer rate (kW)
* Corresponding author. Tel.: +65-7995587; fax: +65-7911859; e-mail: [email protected] 1359-4311/99/$ - see front matter # 1999 Elsevier Science Ltd. All rights reserved. PII: S 1 3 5 9 - 4 3 1 1 ( 9 8 ) 0 0 1 0 9 - 4
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r t T Uo W
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®n radius (m) ®n thickness (m) temperature (8C) overall heat transfer coecient (kW/m28C) air speci®c humidity (kg/kg)
Greek symbols a heat transfer coecient (kW/m28C) heat transfer coecient from tube outer surface to refrigerant (kW/m28C) ac d incremental element f ®n eciency Subscripts a air b ®n base d dew point f ®n i inner in inlet lat latent o outer out outlet r refrigerant s saturated sen sensible t tube w water 1 inlet 2 outlet Abbreviations HTC heat transfer coecient RH relative humidity
1. Introduction Finned-tube direct-expansion coils are widely used in refrigeration and air-conditioning applications. The increasing interest in heat recovery from refrigeration systems and the application of new environment-friendly working ¯uids have created a need for a detailed analysis of the coils with R134a as a refrigerant. Although extensive research related to the use of R134a in refrigeration systems has been done, to the authors' knowledge, the performance study of direct-expansion air-cooling coils using R134a as refrigerant is still lacking.
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Eckels and Pate [1] conducted an experimental comparison of evaporation and condensation heat transfer coecient of R134a and R12. Their results indicated that for similar mass ¯uxes, the evaporation heat transfer coecient of R134a are signi®cantly higher than those of R12 by 35±45% and that in single-phase ¯ow, heat transfer coecients of R134a are higher by 33%. The two refrigerants were also compared for equivalent cooling capacity; it was illustrated that even with a decreased mass ¯ow rate, R134a has a higher heat transfer coecient by 5±15%. The experimental data were compared with the predictions made from several existing correlations and it was found that the dierences for all correlations are within 225%. Hambraeus [2, 3] studied the evaporative heat transfer characteristics of R134a inside smooth, horizontal tubes with the pure refrigerant and oil±refrigerant mixtures. She concluded that R134a has better heat transfer coecients during evaporation than R22 at the same heat and mass ¯uxes. Recently, Torikoshi et al. [4] and Wattelet et al. [5] reported through experimental studies that the evaporative heat transfer coecients inside tubes of R134a without oil are about 25% higher than those for R12 at the same mass ¯uxes. Jung et al. [6] proposed an evaporative heat transfer coecient correlation based on the experimental data obtained with R22, R12, R152a and R114. This correlation was further validated by comparing it with more experimental data for R11 and R134a and it was reported that mean deviation was less than 7% for heat transfer of six pure refrigerants [7]. The work done by Jung et al. makes it possible for the analytical study of coil performance using R134a as a substitute to R12. In the study of an air-cooling evaporator, dierent numerical models are commonly used. Fischer and Rice [8] proposed an evaporator model as a part of a steady-state air-to-air heat pump computer design model. The evaporator simulation model divides coil into three regions, in two-phase region average refrigerant-side heat transfer coecient is obtained by integrating local values determined with correlation equation. Domanski [9] used the tube-to-tube computation approach while developing an evaporator model to study the eect of nonuniform air distribution on the performance of the heat exchanger. Oskarsson et al. [10] presented three dierent models for evaporators, which may operate with dry, wet and frosted ®nned surfaces. They have developed a ®nite element model to study the local behavior of heat transfer of ¯uids, as well as a three-region model and a parametric model. Recently, Judge et al. [11] and Ragazzi [12] investigated the tube-®nned evaporator coil and proposed their distributedparameter evaporator models, respectively. The kind of modeling methodology is used in the study of coils mainly depends on the compromise one is willing to make between the accuracy and the complexity of the model to be used. A distributed model is developed for this study because of its ¯exibility in predicting the total and local characteristics of the coils. It is well known that some key parameters, such as evaporating temperature, refrigerant mass ¯ow rate, coil face velocity and inlet air temperature, have signi®cant in¯uence on the evaporator performance. However, performance studies related to a humid environment have been very scarce, and will be pursued in this paper.
2. Simulation model There are three elements in this model: circuit, tube and control volume. The basic element for computation in the model is the control volume, shown in Fig. 1.
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Fig. 1. A control volume along a tube with ®ns.
2.1. Governing equations There are three heat transfer zones inside the tube corresponding to evaporation, transition and superheating of refrigerants, and there are two zones outside the tube due to unsaturated (dry) and saturated (wet) air. For simultaneous heat and mass transfer on the wet zone, the rates of heat gain due to mass transfer associated with the dehumidi®cation are calculated using a latent air-side convective heat transfer coecient. The eect of air dehumidi®cation is to increase the heat transfer. The driving potential for the simultaneous heat transfer is taken as the temperature dierence between air around the tube and the saturated water ®lm at the outer surface of the tube [13]. 2.1.1. Heat balance on air-side With the assumption that the leaving water temperature outside the tube is equal to the tube outer surface temperature, the heat balance equation for a control volume can be obtained: dma ha2 dma ha1 ÿ ÿ asen alat dAt f dAf Ta ÿ Tt;o Cp;w Tt;o dma Wa1 ÿ Wa2 asen ifg C Le Cp;a
1b
Wa ÿ Ws;t;o Ta ÿ Tt;o
1c
alat
C
1a
where A is the surface area, Cp isobaric speci®c heat, h enthalpy, m mass ¯ow rate, T temperature, W air speci®c humidity, a heat transfer coecient, d incremental element for a control volume, and f ®n eciency; subscripts a, f, lat, o, s, sen and t denote air, ®n, latent, outer, saturated, sensible and tube, respectively, subscripts 1 and 2 denote the inlet and outlet of a control volume, respectively. In Eq. (1a), parameter C is zero in case of no air moisture condensation and the equation reduces to a dry air-side equation.
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Fig. 2. Schematic of a circular ®n.
2.1.2. Fin eciency For circular partially wet or totally wet ®ns, shown in Fig. 2, the following heat balance equations are solved numerically to obtain the ®n temperature distribution. r
d2 Tf dTf 2asen ÿr Tf ÿ Ta 0 dr2 dr kf t
dry fin area Tf > Td;a 2a
" # d2 Tf dTf 2rasen ifg r 2 Ws;f ÿ Wa 0 ÿ Tf ÿ Ta dr dr kf t Le Cp;a wet fin area Tf RTd;a 2b Boundary conditions: Tf jrri Tt;o
dTf dr
0
2c
rro
where r is ®n radius, t ®n thickness, subscripts d and i denote air dew point and inner, respectively. The relationship between speci®c humidity and temperature at saturated conditions is given by a polynomial expression obtained from regression analysis. 0RTf R30 C 3 Ws;f 3:7444 0:3078Tf 0:0046T2f 0:0004T3f 10ÿ3 This is known as a boundary-value problem. The fourth-order Runge±Kutta method is used to solve the above nonlinear dierential equation. After the ®n temperature distribution is obtained, the ®n eciency can be calculated. ÿ 2 PN ÿ 2 a a ÿ T ÿ r T r sen lat;j a f;j j j1 j f 4 ÿ 2 asen alat Ta ÿ Tt;o ro ÿ r2i where alat is calculated according to Tt, o, and local alat, j according to Tf, j (see Fig. 2).
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2.1.3. Mass balance on air-side The speci®c humidity at the exit is given by the following equation: dma Wa2 dma Wa1 ÿ
asen C dAt f dAf Ta ÿ Tt;o Cp;a Le
5
2.1.4. Heat balance on refrigerant side Heat balance equation on the refrigerant side is relatively simple. dq dmr hr2 ÿ hr1
6
where q is heat transfer rate, and subscript r denotes refrigerant. 2.1.5. Heat transfer equation To solve the Eqs. (1) and (5), the expression of Tt, o is needed, which is obtained using heat transfer equations. The following heat transfer equation is used to determine ac, which is the heat transfer coecient from the tube outer surface to the refrigerant: ÿ p do di kt dL ÿ Tt;o ÿ Tt;i ac dAi Tt;o ÿ Tr ar dAi Tt;i ÿ Tr do ÿ di
7
where d is diameter, k thermal conductivity, and L tube length. Then, the heat balance between the air side and the refrigerant inside the tube is used to determine Tt, o with ac: ÿ ÿ ÿ 8a asen alat dAt fdAf Ta ÿ Tt;o ac dAi Tt;o ÿ Tr Tt;o
asen alat dAt f dAf Ta ac dAi Tr asen alat dAt f dAf ac dAi
8b
The overall HTC for a control volume is expressed as: 1 1 do ÿ di 1 Uo dAt dAf asen alat dAt f dAf p do di kt L dAi ar
9
where Uo is overall heat transfer coecient. 2.1.6. Heat and mass transfer coecient In the air side, correlation provided by Fisher and Rice [8] is used to obtain sensible HTC. The HTC on the refrigerant side depends on the ¯ow conditions. For superheated single-phase refrigerant and the region of evaporating two-phase ¯ow, the correlations proposed by Perry and Chilton [14] and Jung and Radermacher [7], respectively, are used. 2.1.7. Hydraulic losses Hydraulic loss for the air side is computed according to the friction factor correlation presented by Turaga et al. [15]. The pressure drop on the refrigerant side for a control volume
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is computed as the sum of friction, momentum and return bend pressure drops (in the case that after refrigerant leaving the control volume, there is a return bend). For two-phase ¯ow, friction is calculated according to correlation recommended by Paliwada [16] and the return bend pressure drop according to a local resistance coecient for single phase ¯ow and a twophase multiplier provided by Paliwoda [17]. 2.2. Computer simulation For a given coil con®guration, the simulation process begins with a set of input conditions of air and refrigerant. In the refrigerant side, it is known that the pressure drop along the coil would aect the refrigeration system performance. To isolate this eect to the evaporator coil, the refrigerant outlet pressure and temperature (i.e. the inlet state of the compressor neglecting the suction line eect) and the inlet enthalpy (i.e. the outlet enthalpy of throttling device) are speci®ed. The refrigerant mass ¯ow rate needs to be estimated initially. The governing equations, the correlations for heat and mass transfer coecients, and pressure loss equations are applied to a control volume and air and refrigerant properties at the exit of the control volume are calculated using an iterative procedure. The ®rst control volume is selected at the refrigerant outlet position and the above computation is repeated for each successive control volume along the opposite direction of refrigerant ¯ow until the refrigerant inlet position is reached. The refrigerant mass ¯ow rate is eventually obtained when the inlet refrigerant enthalpy converges. Alternatively, if the refrigerant mass ¯ow rate is known, the length of the coil and the heat transfer area are computed.
3. Experimental set-up A test rig has been set up to verify the proposed model of direct-expansion ®nned-tube evaporator coils. This test rig, the test section of which is shown in Fig. 3, mainly consists of four systems and they are a closed-circuit wind tunnel, an air property control system, a refrigeration circuit, and an instrumentation and data acquisition system. A wind tunnel provides required air ¯ow conditions, where uniform air speed is achieved at the test coil inlet; ¯ow can be varied using a variable speed blower and air ¯ow rate is measured using a standard nozzle. An air property control system provides the required air inlet thermodynamic conditions for the test coils, which consists of a variable output steam generator and a variable electric heater. Refrigeration system provides required refrigerant-side conditions for the coils. An inverter is used to control an open-type compressor capacity so as to match with the coil load variation under dierent testing conditions. A data acquisition system is used to collect all the information of temperature measurements and refrigerant-side pressure measurements. Air relative humidity is measured using a chilled-mirror hygrometer with an accuracy of 20.5%. The ¯ow rate of the refrigerant-side is measured by a mass ¯ow meter with a maximum error of 20.64% for the measurement. A three-row coil is tested under a wide range of air inlet relative humidity (from 38% to 88%), using R134a as refrigerant. The geometric parameters of the test coil are given in Table 1. Table 2 shows the steady-state ¯ow conditions (cases 1±10) and the comparison of
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Fig. 3. Schematic of coil test section.
measured data with model predictions. In the experiment, the air ¯ow rate (ma), air inlet temperature (Ta, in), air inlet relative humidity (RHin), air outlet average temperature (Ta, out), air outlet average relative humidity (RHout), refrigerant mass ¯ow rate (mr), condensing temperature (Tc), condenser outlet temperature (Tc, out), evaporator outlet temperature (Tr, out), and evaporator outlet pressure (Pr, out) are measured. For convenience, the air ¯ow rate and the measured refrigerant outlet pressure are converted to coil face velocity (Va) and refrigerant outlet saturated temperature (Te, out), respectively. The predicted refrigerant mass ¯ow rates are within 10% of the measured values. Since the heat transfer rate is the product of refrigerant mass ¯ow rate and enthalpy dierence between inlet and outlet (®xed for each case), hence the model predictions on coil heat transfer are in good agreement with the measured data. In addition, the model predictions on air outlet temperature and relative humidity are also close to the measured values.
4. Coil characteristic analysis A coil usually consists of a number of parallel refrigerant circuits. In this investigation, in order to recognize the general characteristics of evaporator coils, a single refrigerant circuit illustrated in Fig. 4 is simulated, the geometric parameters of the tube and ®ns of which are the same as for the test coil given in Table 1. Table 1 Coil geometric parameters Geometric parameters
Values
Geometric parameters
Values
Number of tube rows in air direction Tube number in each row Length of straight tube Transverse tube spacing Longitudinal tube spacing Outer tube diameter
3 8 0.2125 m 25 mm 21.6 mm 9.53 mm
Inner tube diameter Fin thickness Fin number Fin type Tube arrangement
8.83 mm 0.12 mm 84 wavy-®n stagger
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Table 2 Comparison of model predictions with experimental results Flow conditions Case Tc (8C) Tc, out (8C) Te, out (8C) Tr, out (8C) Ta, in (8C) RHin (%) Va (m/s)
1 43.5 37.5 2.0 12.6 25.5 61.0 1.96
2
3
4
5
6
7
8
9
10
42.5 36.7 ÿ 0.9 7.9 24.0 45.0 1.98
42.3 36.7 0.2 8.2 23.6 45.1 1.99
42.1 36.4 ÿ 1.8 6.2 24.3 38.6 1.62
46.0 40.9 5.2 15.1 26.4 76.4 1.56
49.1 45.3 11.2 23.2 30.9 87.2 2.23
46.0 41.0 5.3 15.5 25.8 72.7 1.96
48.5 44.2 10.2 14.5 30.2 88.4 1.93
45.0 41.2 7.1 17.4 30.0 76.7 2.22
47.2 40.6 6.0 16.6 30.7 66.2 2.20
9.39 15.2 100
8.33 13.4 85.5
7.88 13.7 87.1
7.50 11.9 84.0
9.62 18.8 95.1
12.20 26.2 100
9.62 19.0 100
12.35 25.4 100
11.52 23.4 92.9
11.89 22.6 94.3
9.91 14.8 100
8.64 13.5 77.8
7.87 13.9 75.6
7.86 12.2 74.0
9.22 18.9 90.7
12.25 26.0 94.9
9.21 19.2 96.8
11.56 24.7 100
12.61 23.5 88.9
12.31 23.1 85.3
Experimental results mr (g/s) Ta, out (8C) RHout (%) Predicted results mr (g/s) Ta, out (8C) RHout (%) Error mr % DTa, out (8C) DRHout (%)
5.5 ÿ 0.4 0
3.7 0.1 ÿ 7.7
ÿ 0.1 0.2 ÿ 11.5
4.8 0.3 ÿ 10.0
ÿ 4.1 0.1 ÿ 4.4
0.4 ÿ 0.2 ÿ 5.1
ÿ 4.3 0.2 ÿ 3.8
6.4 ÿ0.7 0
9.5 0.1 ÿ 4.0
ÿ 3.5 0.5 ÿ 9.0
4.1. Eect of air inlet relative humidity on coil performance In this study, the input parameters of air and refrigerant are: air inlet temperature (Ta, in = 288C), air inlet relative humidity (RH), coil face velocity (Va = 2 m/s), refrigerant outlet pressure (the corresponding saturated temperature Te, out is 108C), and refrigerant outlet
Fig. 4. The arrangement of a single circuit.
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Fig. 5. Refrigerant mass ¯ow rate and coil heat ¯ux with air inlet humidity.
temperature (Tr, out = 158C). The refrigerant inlet enthalpy is determined by an assumed condenser outlet conditions (condensing temperature 458C, subcooling temperature 408C). The local and overall heat transfer properties of the coil are determined from the calculated values after every control volume is computed. Figs. 5 and 6 show the changes of the refrigerant mass ¯ow rate, coil heat ¯ux, water condensing rate and air outlet temperature with air inlet RH. It is demonstrated that the refrigerant mass ¯ow rate, coil heat ¯ux, water condensing rate and air outlet temperature of the coil signi®cantly increase with air inlet RH, except when RH is below 40%, where all the above parameters remain unchanged. It can be seen that while the air inlet RH increases from
Fig. 6. Water condensing rate and air outlet temperature with air inlet relative humidity.
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Fig. 7. Coil load with air inlet relative humidity.
50% to 90% (from a relatively low humid environment to a high humid environment), the cooling load approximately doubles while the water condensing rate increases from 0.19 to 1.3 kg/h. It can be clearly observed from Fig. 7 that the increase of total cooling load is mainly due to the rapid increase in latent heat load while the sensible cooling load slightly decreases. In order to explain the above results, the distributions of heat transfer coecients within the coil are studied. Fig. 8 shows the air side HTC distributions along a refrigerant circuit for discrete control volumes, the relative length refers to the ratio of refrigerant ¯owing length measured from the inlet to the total path. The ®gure shows that for inlet relative humidity of
Fig. 8. Air side heat transfer coecient distribution.
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Fig. 9. Refrigerant side heat transfer coecient distribution.
30%, the air side HTC remains almost constant since no latent heat transfer takes place along the coil and the sensible heat transfer coecients are almost unchanged on account of constant air face velocities and air inlet temperature. Total air side HTC increases with increase in air inlet relative humidity as seen for 60% and 90% due to the intense mass transfer occurring along the coil caused by increased condensation of moisture (see Fig. 6), which causes increased ¯ow rate of refrigerant to meet the increasing cooling load (see Fig. 5). Increase in air side HTC is accompanied with increase in refrigerant side HTC as seen in Fig. 9, where the distribution of refrigerant side HTC along the circuit with dierent air inlet RH is presented. Fig. 9 shows that the refrigerant side HTC increases gradually until it reaches a maximum, there after the refrigerant side HTC drops rapidly. Fig. 10 shows the air average temperature, refrigerant temperature and tube wall temperature along the refrigerant circuit. The tube wall temperature is between the inside refrigerant and outside air temperatures, and the closeness of wall temperature to either one of them depends on the air side HTC and the refrigerant side HTC. It is observed in Fig. 10 that higher air inlet relative humidity results in higher tube wall temperature due to the fact that the eect of the air side HTC increase is more signi®cant than that of the refrigerant. 4.2. Comparison of coil performance with R134a and R12 In order to study how a given coil performance changes when the traditional refrigerant R12 is substituted by environmentally acceptable refrigerant R134a, the performance of a given circuit, shown in Fig. 4, is simulated with R134a and R12 as refrigerants. The geometric parameters of tube and ®ns are the same as the test coil, except that the straight tube length is 1.0 m. The simulations with both refrigerants are conducted at the same conditions: refrigerant superheating temperature dierence of 58C, coil face velocity of 2 m/s, air inlet relative humidity of 60% and air inlet temperature from 248C to 328C.
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Fig. 10. Distributions of air average temperature, tube wall temperature and refrigerant temperature.
Figs. 11±14 show the variations of refrigerant mass ¯ow rate, cooling load, water condensing rate and air outlet temperature, respectively, with refrigerant outlet saturated temperature (corresponding to refrigerant outlet pressure) for refrigerants R134a and R12. At the same refrigerant outlet saturated temperature, the cooling load (see Fig. 12), water condensing rate (see Fig. 13) and the air outlet temperature (see Fig. 14) remain about the same for both the refrigerants, but the mass ¯ow rate of R12 is signi®cantly greater than that of R134a and the dierence increases with the decreasing refrigerant outlet saturated temperature and the increasing air inlet temperatures. The dierence in refrigerant mass ¯ow rates between R12 and R134a is due to R134a having higher latent heat of vaporization.
Fig. 11. Refrigerant mass ¯ow rate with refrigerant outlet saturated temperature.
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Fig. 12. Cooling load with refrigerant outlet saturated temperature.
4.3. Comparison of coil design with R134a and R12 as refrigerants In this investigation, the design of a single circuit, shown in Fig. 4, is analyzed with refrigerants R134a and R12 so as to recognize the dierence in the whole coil design. Coil geometry and design parameters used are the same as in Section 4.1, except the length of straight tube is the parameter to be calculated. Fig. 15 shows the cooling load and the refrigerant circuit length against the refrigerant mass ¯ux for refrigerants R134a and R12. It can be seen in Fig. 15 that for the same cooling load R12 requires higher mass ¯ux and longer refrigerant circuit than R134a, but for the same mass ¯ux R134a provides higher cooling load and requires longer refrigerant circuit than R12.
Fig. 13. Water condensing rate with refrigerant outlet saturated temperature.
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Fig. 14. Air outlet temperature with refrigerant outlet saturated temperature.
Fig. 16 shows the average heat ¯ux (based on inner tube surface area) and the refrigerant pressure loss with refrigerant mass ¯ux. Here, the heat transfer mainly depends on two factors: the refrigerant side HTC and temperature dierence between refrigerant and air. Initially, with the increase of refrigerant mass ¯ux, the average heat ¯ux increases due to the increase of refrigerant side HTC. When the mass ¯ux is relatively high, the refrigerant pressure drop is large, the refrigerant inlet pressure and temperature would increase for a speci®ed outlet condition, which causes the signi®cant decrease in temperature dierence between refrigerant and air. Therefore, the coil average heat ¯ux drops after it reaches a maximum. It is observed from Fig. 16 that for a speci®ed refrigerant mass ¯ux from 80 to 140 kg/m2 s, the coil heat ¯ux of R134a is higher than that of R12 by around 7%. It means that for a given cooling load, the
Fig. 15. Refrigerant cooling load and circuit length with refrigerant mass ¯ux.
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Fig. 16. Coil heat ¯ux and refrigerant pressure loss with refrigerant mass ¯ux.
required overall heat transfer area of a coil using R134a will be smaller than that using R12 by around 7%. In coil design, in order to minimize the required heat transfer area, the mass ¯ux corresponding to the maximum heat ¯ux is usually used. It can be seen that the maximum heat ¯ux of R134a is signi®cantly higher than that of R12. 5. Conclusions A distributed simulation model is developed for the prediction of steady-state performance of a direct-expansion evaporator coil. This model uses a numerical method to calculate the partially wet and totally wet ®n eciency and takes into account the refrigerant pressure drop along the coil. The proposed model is validated by the comparison of experimental results with model predictions using R134a as a refrigerant. The predicted heat transfer rates are within 10% of the measured values under dierent air ¯ow conditions. The model predictions on air side outlet parameters are also in good agreement with the measured values. The analytical model and computer simulation for a given direct-expansion air-cooling coil have identi®ed that the air relative humidity is a very signi®cant parameter in determining the energy requirement and the quality of air conditioning in a humid environment. A high humid environment will require much higher cooling load and hence energy requirement compared with a low humid environment, which is mainly the results of higher latent heat load due to condensation of moisture. A comparison of performance of coils with R134a as a substitute to R12 for a given direct expansion coil results in slight increase of moisture condensation rate and cooling load of the coil, decrease of refrigerant mass ¯ow rate and nearly constant air outlet temperature. In coil design, if the cooling load and refrigerant mass ¯uxes (within the range from 80 to 140 kg/m2 s) are the same, the required overall heat transfer area of a coil using R134a will be less than that using R12 by approximately 7% due to higher refrigerant HTC of R134a. But, the
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refrigerant side pressure loss of R134a is greater than that of R12 partly due to longer refrigerant circuits. The approach proposed in this study gives a general method to analyze the performances of direct expansion air-cooling coils. Extending the approach to dierent geometry of coils and optimization of the design parameters for dierent applications would give a better tool for design and analysis of coils. References [1] S.J. Eckels, M.B. Pate, An experimental comparison of evaporation and condensation heat transfer coecients for HFC-134a and CFC-12, Int. J. Refrig. 14 (1991) 70±77. [2] K. Hambraeus, Heat transfer coecient during two-phase ¯ow boiling of HFC, Int. J. Refrig. 14 (1991) 357± 362. [3] K. Hambraeus, Heat transfer of oil-contaminated HFC 134a in a horizontal evaporator, Int. J. Refrig. 18 (1995) 87±99. [4] K. Torikoshi, T. Ebisu, Heat transfer and pressure drop characteristics of R134a, R32, and a mixture of R32/ R134a inside a horizontal tube, ASHRAE Trans. 99 (2) (1993) 90±96. [5] J.P. Wattelet, J.C. Chato, A.L. Souza, B.R. Christoersen, Evaporative characteristics of R12, R134a, and a mixture at low mass ¯uxes, ASHRAE Trans. 100 (1) (1994) 603±615. [6] D.S. Jung, M. McLinden, R. Radermacher, D. Didion, A study of ¯ow boiling heat transfer with refrigerant mixtures, Int. J. Heat Mass Transfer. 32 (1989) 1751±1764. [7] D.S. Jung, R. Radermacher, Prediction of heat transfer coecients of various refrigerants during evaporation, ASHRAE Trans. 97 (2) (1991) 48±53. [8] S.K. Fisher, C.K. Rice, The Oak Ridge heat pump models: a steady state computer design model for air-to-air heat pumps. Oak Ridge National Laboratory, Oak Ridge, Tennessee: ORNL/CON 80/R1, 1983. [9] P.A. Domanski, Simulation of an evaporator with nonuniform one-dimensional air distribution, ASHRAE Trans. 97 (1) (1991) 793±802. [10] S.P. Oskarsson, K.I. Krakow, S. Lin, Evaporator models for operating with dry, wet and frosted ®nned surfaces. Part I: Heat transfer and ¯uid ¯ow theory, ASHRAE Trans. 96 (1) (1990) 373±380. [11] J. Judge, R. Radermacher, A heat exchanger model for mixtures and pure refrigerant cycle simulations, Int. J. Refrig. 20 (1997) 244±255. [12] F. Ragazzi, Thermodynamic optimization of evaporators with zeotropic refrigerant mixtures. Ph.D. Dissertation, University of Illinois at Urbana-Champaign. [13] J.L. Threlkeld, Thermal Environmental Engineering. 2nd ed. Prentice Hall, Englewood Clis, NJ., 1970. [14] R.H. Perry, C.H. Chilton, Chemical Engineer's Handbook, 5th ed.. Section 10, McGraw-Hill, New York, 1973. [15] M. Turaga, S. Lin, P. Fazio, Correlations for heat transfer and pressure drop factors for direct expansion air cooling and dehumidifying coils, ASHRAE Trans. 94 (2) (1988) 616±629. [16] A. Paliwoda, Generalized method of pressure drop and tube length calculation with boiling and condensing refrigerants within the entire zone of saturation, Int. J. Refrig. 12 (1989) 314±322. [17] A. Paliwoda, Generalized method of pressure drop calculation across pipe components containing two-phase ¯ow of refrigeration, Int. J. Refrig. 15 (1992) 41±47.
Applied Thermal Engineering 19 (1999) 1129±1145
www.elsevier.com/locate/apthermeng
Analytical study of evaporator coil in humid environment S.Y. Liang, M. Liu, T.N. Wong *, G.K. Nathan School of Mechanical and Production Engineering, Nanyang Technological University, Singapore, 639798 Received 1 November 1997; accepted 20 October 1998
Abstract In this paper, a distributed simulation model for predicting steady-state performance of a directexpansion air-cooling coil is developed. This model uses a numerical method to calculate the partially wet and totally wet ®n eciency and takes into account the refrigerant pressure drop along the coil. The model simulation of a test coil is validated with experimental data collected under dierent air conditions using R134a as a refrigerant. On the basis of this model, a number of parameters which re¯ect the characteristics of evaporator coils are analyzed in diverse humid environments and it is found that the performance of the coil is signi®cantly aected by air relative humidity. Comparison of coil performance with R134a and R12 as refrigerants, whose dierence in properties is small, shows that the dierences in coil design for a speci®ed cooling load and in the heat exchange characteristics for a given coil cannot be ignored. # 1999 Elsevier Science Ltd. All rights reserved. Keywords: Air cooling; Evaporator modeling; Performance study; Humid environment; R134a
Nomenclature A Cp d h ifg k L Le m q
eective heat transfer area (m2) isobaric speci®c heat (kJ/kg8C) tube diameter (m) speci®c enthalpy (kJ/kg) latent heat of moisture condensation (kJ/kg) thermal conductivity (kW/m8C) tube length (m) Lewis number mass ¯ow rate (kg/s) heat transfer rate (kW)
* Corresponding author. Tel.: +65-7995587; fax: +65-7911859; e-mail: [email protected] 1359-4311/99/$ - see front matter # 1999 Elsevier Science Ltd. All rights reserved. PII: S 1 3 5 9 - 4 3 1 1 ( 9 8 ) 0 0 1 0 9 - 4
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r t T Uo W
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®n radius (m) ®n thickness (m) temperature (8C) overall heat transfer coecient (kW/m28C) air speci®c humidity (kg/kg)
Greek symbols a heat transfer coecient (kW/m28C) heat transfer coecient from tube outer surface to refrigerant (kW/m28C) ac d incremental element f ®n eciency Subscripts a air b ®n base d dew point f ®n i inner in inlet lat latent o outer out outlet r refrigerant s saturated sen sensible t tube w water 1 inlet 2 outlet Abbreviations HTC heat transfer coecient RH relative humidity
1. Introduction Finned-tube direct-expansion coils are widely used in refrigeration and air-conditioning applications. The increasing interest in heat recovery from refrigeration systems and the application of new environment-friendly working ¯uids have created a need for a detailed analysis of the coils with R134a as a refrigerant. Although extensive research related to the use of R134a in refrigeration systems has been done, to the authors' knowledge, the performance study of direct-expansion air-cooling coils using R134a as refrigerant is still lacking.
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Eckels and Pate [1] conducted an experimental comparison of evaporation and condensation heat transfer coecient of R134a and R12. Their results indicated that for similar mass ¯uxes, the evaporation heat transfer coecient of R134a are signi®cantly higher than those of R12 by 35±45% and that in single-phase ¯ow, heat transfer coecients of R134a are higher by 33%. The two refrigerants were also compared for equivalent cooling capacity; it was illustrated that even with a decreased mass ¯ow rate, R134a has a higher heat transfer coecient by 5±15%. The experimental data were compared with the predictions made from several existing correlations and it was found that the dierences for all correlations are within 225%. Hambraeus [2, 3] studied the evaporative heat transfer characteristics of R134a inside smooth, horizontal tubes with the pure refrigerant and oil±refrigerant mixtures. She concluded that R134a has better heat transfer coecients during evaporation than R22 at the same heat and mass ¯uxes. Recently, Torikoshi et al. [4] and Wattelet et al. [5] reported through experimental studies that the evaporative heat transfer coecients inside tubes of R134a without oil are about 25% higher than those for R12 at the same mass ¯uxes. Jung et al. [6] proposed an evaporative heat transfer coecient correlation based on the experimental data obtained with R22, R12, R152a and R114. This correlation was further validated by comparing it with more experimental data for R11 and R134a and it was reported that mean deviation was less than 7% for heat transfer of six pure refrigerants [7]. The work done by Jung et al. makes it possible for the analytical study of coil performance using R134a as a substitute to R12. In the study of an air-cooling evaporator, dierent numerical models are commonly used. Fischer and Rice [8] proposed an evaporator model as a part of a steady-state air-to-air heat pump computer design model. The evaporator simulation model divides coil into three regions, in two-phase region average refrigerant-side heat transfer coecient is obtained by integrating local values determined with correlation equation. Domanski [9] used the tube-to-tube computation approach while developing an evaporator model to study the eect of nonuniform air distribution on the performance of the heat exchanger. Oskarsson et al. [10] presented three dierent models for evaporators, which may operate with dry, wet and frosted ®nned surfaces. They have developed a ®nite element model to study the local behavior of heat transfer of ¯uids, as well as a three-region model and a parametric model. Recently, Judge et al. [11] and Ragazzi [12] investigated the tube-®nned evaporator coil and proposed their distributedparameter evaporator models, respectively. The kind of modeling methodology is used in the study of coils mainly depends on the compromise one is willing to make between the accuracy and the complexity of the model to be used. A distributed model is developed for this study because of its ¯exibility in predicting the total and local characteristics of the coils. It is well known that some key parameters, such as evaporating temperature, refrigerant mass ¯ow rate, coil face velocity and inlet air temperature, have signi®cant in¯uence on the evaporator performance. However, performance studies related to a humid environment have been very scarce, and will be pursued in this paper.
2. Simulation model There are three elements in this model: circuit, tube and control volume. The basic element for computation in the model is the control volume, shown in Fig. 1.
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Fig. 1. A control volume along a tube with ®ns.
2.1. Governing equations There are three heat transfer zones inside the tube corresponding to evaporation, transition and superheating of refrigerants, and there are two zones outside the tube due to unsaturated (dry) and saturated (wet) air. For simultaneous heat and mass transfer on the wet zone, the rates of heat gain due to mass transfer associated with the dehumidi®cation are calculated using a latent air-side convective heat transfer coecient. The eect of air dehumidi®cation is to increase the heat transfer. The driving potential for the simultaneous heat transfer is taken as the temperature dierence between air around the tube and the saturated water ®lm at the outer surface of the tube [13]. 2.1.1. Heat balance on air-side With the assumption that the leaving water temperature outside the tube is equal to the tube outer surface temperature, the heat balance equation for a control volume can be obtained: dma ha2 dma ha1 ÿ ÿ asen alat dAt f dAf Ta ÿ Tt;o Cp;w Tt;o dma Wa1 ÿ Wa2 asen ifg C Le Cp;a
1b
Wa ÿ Ws;t;o Ta ÿ Tt;o
1c
alat
C
1a
where A is the surface area, Cp isobaric speci®c heat, h enthalpy, m mass ¯ow rate, T temperature, W air speci®c humidity, a heat transfer coecient, d incremental element for a control volume, and f ®n eciency; subscripts a, f, lat, o, s, sen and t denote air, ®n, latent, outer, saturated, sensible and tube, respectively, subscripts 1 and 2 denote the inlet and outlet of a control volume, respectively. In Eq. (1a), parameter C is zero in case of no air moisture condensation and the equation reduces to a dry air-side equation.
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Fig. 2. Schematic of a circular ®n.
2.1.2. Fin eciency For circular partially wet or totally wet ®ns, shown in Fig. 2, the following heat balance equations are solved numerically to obtain the ®n temperature distribution. r
d2 Tf dTf 2asen ÿr Tf ÿ Ta 0 dr2 dr kf t
dry fin area Tf > Td;a 2a
" # d2 Tf dTf 2rasen ifg r 2 Ws;f ÿ Wa 0 ÿ Tf ÿ Ta dr dr kf t Le Cp;a wet fin area Tf RTd;a 2b Boundary conditions: Tf jrri Tt;o
dTf dr
0
2c
rro
where r is ®n radius, t ®n thickness, subscripts d and i denote air dew point and inner, respectively. The relationship between speci®c humidity and temperature at saturated conditions is given by a polynomial expression obtained from regression analysis. 0RTf R30 C 3 Ws;f 3:7444 0:3078Tf 0:0046T2f 0:0004T3f 10ÿ3 This is known as a boundary-value problem. The fourth-order Runge±Kutta method is used to solve the above nonlinear dierential equation. After the ®n temperature distribution is obtained, the ®n eciency can be calculated. ÿ 2 PN ÿ 2 a a ÿ T ÿ r T r sen lat;j a f;j j j1 j f 4 ÿ 2 asen alat Ta ÿ Tt;o ro ÿ r2i where alat is calculated according to Tt, o, and local alat, j according to Tf, j (see Fig. 2).
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2.1.3. Mass balance on air-side The speci®c humidity at the exit is given by the following equation: dma Wa2 dma Wa1 ÿ
asen C dAt f dAf Ta ÿ Tt;o Cp;a Le
5
2.1.4. Heat balance on refrigerant side Heat balance equation on the refrigerant side is relatively simple. dq dmr hr2 ÿ hr1
6
where q is heat transfer rate, and subscript r denotes refrigerant. 2.1.5. Heat transfer equation To solve the Eqs. (1) and (5), the expression of Tt, o is needed, which is obtained using heat transfer equations. The following heat transfer equation is used to determine ac, which is the heat transfer coecient from the tube outer surface to the refrigerant: ÿ p do di kt dL ÿ Tt;o ÿ Tt;i ac dAi Tt;o ÿ Tr ar dAi Tt;i ÿ Tr do ÿ di
7
where d is diameter, k thermal conductivity, and L tube length. Then, the heat balance between the air side and the refrigerant inside the tube is used to determine Tt, o with ac: ÿ ÿ ÿ 8a asen alat dAt fdAf Ta ÿ Tt;o ac dAi Tt;o ÿ Tr Tt;o
asen alat dAt f dAf Ta ac dAi Tr asen alat dAt f dAf ac dAi
8b
The overall HTC for a control volume is expressed as: 1 1 do ÿ di 1 Uo dAt dAf asen alat dAt f dAf p do di kt L dAi ar
9
where Uo is overall heat transfer coecient. 2.1.6. Heat and mass transfer coecient In the air side, correlation provided by Fisher and Rice [8] is used to obtain sensible HTC. The HTC on the refrigerant side depends on the ¯ow conditions. For superheated single-phase refrigerant and the region of evaporating two-phase ¯ow, the correlations proposed by Perry and Chilton [14] and Jung and Radermacher [7], respectively, are used. 2.1.7. Hydraulic losses Hydraulic loss for the air side is computed according to the friction factor correlation presented by Turaga et al. [15]. The pressure drop on the refrigerant side for a control volume
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is computed as the sum of friction, momentum and return bend pressure drops (in the case that after refrigerant leaving the control volume, there is a return bend). For two-phase ¯ow, friction is calculated according to correlation recommended by Paliwada [16] and the return bend pressure drop according to a local resistance coecient for single phase ¯ow and a twophase multiplier provided by Paliwoda [17]. 2.2. Computer simulation For a given coil con®guration, the simulation process begins with a set of input conditions of air and refrigerant. In the refrigerant side, it is known that the pressure drop along the coil would aect the refrigeration system performance. To isolate this eect to the evaporator coil, the refrigerant outlet pressure and temperature (i.e. the inlet state of the compressor neglecting the suction line eect) and the inlet enthalpy (i.e. the outlet enthalpy of throttling device) are speci®ed. The refrigerant mass ¯ow rate needs to be estimated initially. The governing equations, the correlations for heat and mass transfer coecients, and pressure loss equations are applied to a control volume and air and refrigerant properties at the exit of the control volume are calculated using an iterative procedure. The ®rst control volume is selected at the refrigerant outlet position and the above computation is repeated for each successive control volume along the opposite direction of refrigerant ¯ow until the refrigerant inlet position is reached. The refrigerant mass ¯ow rate is eventually obtained when the inlet refrigerant enthalpy converges. Alternatively, if the refrigerant mass ¯ow rate is known, the length of the coil and the heat transfer area are computed.
3. Experimental set-up A test rig has been set up to verify the proposed model of direct-expansion ®nned-tube evaporator coils. This test rig, the test section of which is shown in Fig. 3, mainly consists of four systems and they are a closed-circuit wind tunnel, an air property control system, a refrigeration circuit, and an instrumentation and data acquisition system. A wind tunnel provides required air ¯ow conditions, where uniform air speed is achieved at the test coil inlet; ¯ow can be varied using a variable speed blower and air ¯ow rate is measured using a standard nozzle. An air property control system provides the required air inlet thermodynamic conditions for the test coils, which consists of a variable output steam generator and a variable electric heater. Refrigeration system provides required refrigerant-side conditions for the coils. An inverter is used to control an open-type compressor capacity so as to match with the coil load variation under dierent testing conditions. A data acquisition system is used to collect all the information of temperature measurements and refrigerant-side pressure measurements. Air relative humidity is measured using a chilled-mirror hygrometer with an accuracy of 20.5%. The ¯ow rate of the refrigerant-side is measured by a mass ¯ow meter with a maximum error of 20.64% for the measurement. A three-row coil is tested under a wide range of air inlet relative humidity (from 38% to 88%), using R134a as refrigerant. The geometric parameters of the test coil are given in Table 1. Table 2 shows the steady-state ¯ow conditions (cases 1±10) and the comparison of
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Fig. 3. Schematic of coil test section.
measured data with model predictions. In the experiment, the air ¯ow rate (ma), air inlet temperature (Ta, in), air inlet relative humidity (RHin), air outlet average temperature (Ta, out), air outlet average relative humidity (RHout), refrigerant mass ¯ow rate (mr), condensing temperature (Tc), condenser outlet temperature (Tc, out), evaporator outlet temperature (Tr, out), and evaporator outlet pressure (Pr, out) are measured. For convenience, the air ¯ow rate and the measured refrigerant outlet pressure are converted to coil face velocity (Va) and refrigerant outlet saturated temperature (Te, out), respectively. The predicted refrigerant mass ¯ow rates are within 10% of the measured values. Since the heat transfer rate is the product of refrigerant mass ¯ow rate and enthalpy dierence between inlet and outlet (®xed for each case), hence the model predictions on coil heat transfer are in good agreement with the measured data. In addition, the model predictions on air outlet temperature and relative humidity are also close to the measured values.
4. Coil characteristic analysis A coil usually consists of a number of parallel refrigerant circuits. In this investigation, in order to recognize the general characteristics of evaporator coils, a single refrigerant circuit illustrated in Fig. 4 is simulated, the geometric parameters of the tube and ®ns of which are the same as for the test coil given in Table 1. Table 1 Coil geometric parameters Geometric parameters
Values
Geometric parameters
Values
Number of tube rows in air direction Tube number in each row Length of straight tube Transverse tube spacing Longitudinal tube spacing Outer tube diameter
3 8 0.2125 m 25 mm 21.6 mm 9.53 mm
Inner tube diameter Fin thickness Fin number Fin type Tube arrangement
8.83 mm 0.12 mm 84 wavy-®n stagger
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Table 2 Comparison of model predictions with experimental results Flow conditions Case Tc (8C) Tc, out (8C) Te, out (8C) Tr, out (8C) Ta, in (8C) RHin (%) Va (m/s)
1 43.5 37.5 2.0 12.6 25.5 61.0 1.96
2
3
4
5
6
7
8
9
10
42.5 36.7 ÿ 0.9 7.9 24.0 45.0 1.98
42.3 36.7 0.2 8.2 23.6 45.1 1.99
42.1 36.4 ÿ 1.8 6.2 24.3 38.6 1.62
46.0 40.9 5.2 15.1 26.4 76.4 1.56
49.1 45.3 11.2 23.2 30.9 87.2 2.23
46.0 41.0 5.3 15.5 25.8 72.7 1.96
48.5 44.2 10.2 14.5 30.2 88.4 1.93
45.0 41.2 7.1 17.4 30.0 76.7 2.22
47.2 40.6 6.0 16.6 30.7 66.2 2.20
9.39 15.2 100
8.33 13.4 85.5
7.88 13.7 87.1
7.50 11.9 84.0
9.62 18.8 95.1
12.20 26.2 100
9.62 19.0 100
12.35 25.4 100
11.52 23.4 92.9
11.89 22.6 94.3
9.91 14.8 100
8.64 13.5 77.8
7.87 13.9 75.6
7.86 12.2 74.0
9.22 18.9 90.7
12.25 26.0 94.9
9.21 19.2 96.8
11.56 24.7 100
12.61 23.5 88.9
12.31 23.1 85.3
Experimental results mr (g/s) Ta, out (8C) RHout (%) Predicted results mr (g/s) Ta, out (8C) RHout (%) Error mr % DTa, out (8C) DRHout (%)
5.5 ÿ 0.4 0
3.7 0.1 ÿ 7.7
ÿ 0.1 0.2 ÿ 11.5
4.8 0.3 ÿ 10.0
ÿ 4.1 0.1 ÿ 4.4
0.4 ÿ 0.2 ÿ 5.1
ÿ 4.3 0.2 ÿ 3.8
6.4 ÿ0.7 0
9.5 0.1 ÿ 4.0
ÿ 3.5 0.5 ÿ 9.0
4.1. Eect of air inlet relative humidity on coil performance In this study, the input parameters of air and refrigerant are: air inlet temperature (Ta, in = 288C), air inlet relative humidity (RH), coil face velocity (Va = 2 m/s), refrigerant outlet pressure (the corresponding saturated temperature Te, out is 108C), and refrigerant outlet
Fig. 4. The arrangement of a single circuit.
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Fig. 5. Refrigerant mass ¯ow rate and coil heat ¯ux with air inlet humidity.
temperature (Tr, out = 158C). The refrigerant inlet enthalpy is determined by an assumed condenser outlet conditions (condensing temperature 458C, subcooling temperature 408C). The local and overall heat transfer properties of the coil are determined from the calculated values after every control volume is computed. Figs. 5 and 6 show the changes of the refrigerant mass ¯ow rate, coil heat ¯ux, water condensing rate and air outlet temperature with air inlet RH. It is demonstrated that the refrigerant mass ¯ow rate, coil heat ¯ux, water condensing rate and air outlet temperature of the coil signi®cantly increase with air inlet RH, except when RH is below 40%, where all the above parameters remain unchanged. It can be seen that while the air inlet RH increases from
Fig. 6. Water condensing rate and air outlet temperature with air inlet relative humidity.
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Fig. 7. Coil load with air inlet relative humidity.
50% to 90% (from a relatively low humid environment to a high humid environment), the cooling load approximately doubles while the water condensing rate increases from 0.19 to 1.3 kg/h. It can be clearly observed from Fig. 7 that the increase of total cooling load is mainly due to the rapid increase in latent heat load while the sensible cooling load slightly decreases. In order to explain the above results, the distributions of heat transfer coecients within the coil are studied. Fig. 8 shows the air side HTC distributions along a refrigerant circuit for discrete control volumes, the relative length refers to the ratio of refrigerant ¯owing length measured from the inlet to the total path. The ®gure shows that for inlet relative humidity of
Fig. 8. Air side heat transfer coecient distribution.
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Fig. 9. Refrigerant side heat transfer coecient distribution.
30%, the air side HTC remains almost constant since no latent heat transfer takes place along the coil and the sensible heat transfer coecients are almost unchanged on account of constant air face velocities and air inlet temperature. Total air side HTC increases with increase in air inlet relative humidity as seen for 60% and 90% due to the intense mass transfer occurring along the coil caused by increased condensation of moisture (see Fig. 6), which causes increased ¯ow rate of refrigerant to meet the increasing cooling load (see Fig. 5). Increase in air side HTC is accompanied with increase in refrigerant side HTC as seen in Fig. 9, where the distribution of refrigerant side HTC along the circuit with dierent air inlet RH is presented. Fig. 9 shows that the refrigerant side HTC increases gradually until it reaches a maximum, there after the refrigerant side HTC drops rapidly. Fig. 10 shows the air average temperature, refrigerant temperature and tube wall temperature along the refrigerant circuit. The tube wall temperature is between the inside refrigerant and outside air temperatures, and the closeness of wall temperature to either one of them depends on the air side HTC and the refrigerant side HTC. It is observed in Fig. 10 that higher air inlet relative humidity results in higher tube wall temperature due to the fact that the eect of the air side HTC increase is more signi®cant than that of the refrigerant. 4.2. Comparison of coil performance with R134a and R12 In order to study how a given coil performance changes when the traditional refrigerant R12 is substituted by environmentally acceptable refrigerant R134a, the performance of a given circuit, shown in Fig. 4, is simulated with R134a and R12 as refrigerants. The geometric parameters of tube and ®ns are the same as the test coil, except that the straight tube length is 1.0 m. The simulations with both refrigerants are conducted at the same conditions: refrigerant superheating temperature dierence of 58C, coil face velocity of 2 m/s, air inlet relative humidity of 60% and air inlet temperature from 248C to 328C.
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Fig. 10. Distributions of air average temperature, tube wall temperature and refrigerant temperature.
Figs. 11±14 show the variations of refrigerant mass ¯ow rate, cooling load, water condensing rate and air outlet temperature, respectively, with refrigerant outlet saturated temperature (corresponding to refrigerant outlet pressure) for refrigerants R134a and R12. At the same refrigerant outlet saturated temperature, the cooling load (see Fig. 12), water condensing rate (see Fig. 13) and the air outlet temperature (see Fig. 14) remain about the same for both the refrigerants, but the mass ¯ow rate of R12 is signi®cantly greater than that of R134a and the dierence increases with the decreasing refrigerant outlet saturated temperature and the increasing air inlet temperatures. The dierence in refrigerant mass ¯ow rates between R12 and R134a is due to R134a having higher latent heat of vaporization.
Fig. 11. Refrigerant mass ¯ow rate with refrigerant outlet saturated temperature.
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Fig. 12. Cooling load with refrigerant outlet saturated temperature.
4.3. Comparison of coil design with R134a and R12 as refrigerants In this investigation, the design of a single circuit, shown in Fig. 4, is analyzed with refrigerants R134a and R12 so as to recognize the dierence in the whole coil design. Coil geometry and design parameters used are the same as in Section 4.1, except the length of straight tube is the parameter to be calculated. Fig. 15 shows the cooling load and the refrigerant circuit length against the refrigerant mass ¯ux for refrigerants R134a and R12. It can be seen in Fig. 15 that for the same cooling load R12 requires higher mass ¯ux and longer refrigerant circuit than R134a, but for the same mass ¯ux R134a provides higher cooling load and requires longer refrigerant circuit than R12.
Fig. 13. Water condensing rate with refrigerant outlet saturated temperature.
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Fig. 14. Air outlet temperature with refrigerant outlet saturated temperature.
Fig. 16 shows the average heat ¯ux (based on inner tube surface area) and the refrigerant pressure loss with refrigerant mass ¯ux. Here, the heat transfer mainly depends on two factors: the refrigerant side HTC and temperature dierence between refrigerant and air. Initially, with the increase of refrigerant mass ¯ux, the average heat ¯ux increases due to the increase of refrigerant side HTC. When the mass ¯ux is relatively high, the refrigerant pressure drop is large, the refrigerant inlet pressure and temperature would increase for a speci®ed outlet condition, which causes the signi®cant decrease in temperature dierence between refrigerant and air. Therefore, the coil average heat ¯ux drops after it reaches a maximum. It is observed from Fig. 16 that for a speci®ed refrigerant mass ¯ux from 80 to 140 kg/m2 s, the coil heat ¯ux of R134a is higher than that of R12 by around 7%. It means that for a given cooling load, the
Fig. 15. Refrigerant cooling load and circuit length with refrigerant mass ¯ux.
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Fig. 16. Coil heat ¯ux and refrigerant pressure loss with refrigerant mass ¯ux.
required overall heat transfer area of a coil using R134a will be smaller than that using R12 by around 7%. In coil design, in order to minimize the required heat transfer area, the mass ¯ux corresponding to the maximum heat ¯ux is usually used. It can be seen that the maximum heat ¯ux of R134a is signi®cantly higher than that of R12. 5. Conclusions A distributed simulation model is developed for the prediction of steady-state performance of a direct-expansion evaporator coil. This model uses a numerical method to calculate the partially wet and totally wet ®n eciency and takes into account the refrigerant pressure drop along the coil. The proposed model is validated by the comparison of experimental results with model predictions using R134a as a refrigerant. The predicted heat transfer rates are within 10% of the measured values under dierent air ¯ow conditions. The model predictions on air side outlet parameters are also in good agreement with the measured values. The analytical model and computer simulation for a given direct-expansion air-cooling coil have identi®ed that the air relative humidity is a very signi®cant parameter in determining the energy requirement and the quality of air conditioning in a humid environment. A high humid environment will require much higher cooling load and hence energy requirement compared with a low humid environment, which is mainly the results of higher latent heat load due to condensation of moisture. A comparison of performance of coils with R134a as a substitute to R12 for a given direct expansion coil results in slight increase of moisture condensation rate and cooling load of the coil, decrease of refrigerant mass ¯ow rate and nearly constant air outlet temperature. In coil design, if the cooling load and refrigerant mass ¯uxes (within the range from 80 to 140 kg/m2 s) are the same, the required overall heat transfer area of a coil using R134a will be less than that using R12 by approximately 7% due to higher refrigerant HTC of R134a. But, the
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refrigerant side pressure loss of R134a is greater than that of R12 partly due to longer refrigerant circuits. The approach proposed in this study gives a general method to analyze the performances of direct expansion air-cooling coils. Extending the approach to dierent geometry of coils and optimization of the design parameters for dierent applications would give a better tool for design and analysis of coils. References [1] S.J. Eckels, M.B. Pate, An experimental comparison of evaporation and condensation heat transfer coecients for HFC-134a and CFC-12, Int. J. Refrig. 14 (1991) 70±77. [2] K. Hambraeus, Heat transfer coecient during two-phase ¯ow boiling of HFC, Int. J. Refrig. 14 (1991) 357± 362. [3] K. Hambraeus, Heat transfer of oil-contaminated HFC 134a in a horizontal evaporator, Int. J. Refrig. 18 (1995) 87±99. [4] K. Torikoshi, T. Ebisu, Heat transfer and pressure drop characteristics of R134a, R32, and a mixture of R32/ R134a inside a horizontal tube, ASHRAE Trans. 99 (2) (1993) 90±96. [5] J.P. Wattelet, J.C. Chato, A.L. Souza, B.R. Christoersen, Evaporative characteristics of R12, R134a, and a mixture at low mass ¯uxes, ASHRAE Trans. 100 (1) (1994) 603±615. [6] D.S. Jung, M. McLinden, R. Radermacher, D. Didion, A study of ¯ow boiling heat transfer with refrigerant mixtures, Int. J. Heat Mass Transfer. 32 (1989) 1751±1764. [7] D.S. Jung, R. Radermacher, Prediction of heat transfer coecients of various refrigerants during evaporation, ASHRAE Trans. 97 (2) (1991) 48±53. [8] S.K. Fisher, C.K. Rice, The Oak Ridge heat pump models: a steady state computer design model for air-to-air heat pumps. Oak Ridge National Laboratory, Oak Ridge, Tennessee: ORNL/CON 80/R1, 1983. [9] P.A. Domanski, Simulation of an evaporator with nonuniform one-dimensional air distribution, ASHRAE Trans. 97 (1) (1991) 793±802. [10] S.P. Oskarsson, K.I. Krakow, S. Lin, Evaporator models for operating with dry, wet and frosted ®nned surfaces. Part I: Heat transfer and ¯uid ¯ow theory, ASHRAE Trans. 96 (1) (1990) 373±380. [11] J. Judge, R. Radermacher, A heat exchanger model for mixtures and pure refrigerant cycle simulations, Int. J. Refrig. 20 (1997) 244±255. [12] F. Ragazzi, Thermodynamic optimization of evaporators with zeotropic refrigerant mixtures. Ph.D. Dissertation, University of Illinois at Urbana-Champaign. [13] J.L. Threlkeld, Thermal Environmental Engineering. 2nd ed. Prentice Hall, Englewood Clis, NJ., 1970. [14] R.H. Perry, C.H. Chilton, Chemical Engineer's Handbook, 5th ed.. Section 10, McGraw-Hill, New York, 1973. [15] M. Turaga, S. Lin, P. Fazio, Correlations for heat transfer and pressure drop factors for direct expansion air cooling and dehumidifying coils, ASHRAE Trans. 94 (2) (1988) 616±629. [16] A. Paliwoda, Generalized method of pressure drop and tube length calculation with boiling and condensing refrigerants within the entire zone of saturation, Int. J. Refrig. 12 (1989) 314±322. [17] A. Paliwoda, Generalized method of pressure drop calculation across pipe components containing two-phase ¯ow of refrigeration, Int. J. Refrig. 15 (1992) 41±47.