Experimental investigation of heat transfer coefficient of a refrigerant mixture in a fin-and-tube evaporator.

Raj, M. Herbert ; Lal, D. Mohan

Introduction

HCFC has been effectively used in many refrigeration and airconditioning applications due to its better thermodynamic and material compatibility. The fully halogenated chlorofluorocarbons are responsible for destroying the stratospheric ozone layer. The hydrochlorofluorocarbon (HCFC) refrigerants are being replaced by hydrofluorocarbon (HFC) and HFC mixtures due to environmental concerns. R407C is recommended as an alternative to replace R22 in a window airconditioner but its oil miscibility nature is the main problem. Previous research [1, 2] reported that addition of HC blend (54.8% R600a and 45.2% R290a) to R407C will over come the oil miscibility issue with mineral oil and it was proved that the M20 refrigerant mixture (80% R407C and 20% HC blend by wt) could be viable drop-in substitute without changing the mineral oil.

The fin-and-tube evaporators are widely used in airconditioning and processing applications. The knowledge of overall heat transfer coefficients is of prime importance to optimize the design of heat exchangers for refrigeration and airconditioning applications since; most of the designs are based on overall heat transfer coefficients. Evaporation of pure and mixed refrigerant has been studied inside the horizontal and vertical tubes by large number of researchers, both the experimentally and analytically. A large number of correlations have been proposed based on experimental results to predict the heat transfer coefficient in forced convective boiling in horizontal and vertical smooth tube for pure refrigerant and mixture [3-8]. During the past few decades, many research works have been carried out on air-side performance of fin-and-tube evaporator and the heat transfer and friction characteristics of fin-and-tube evaporator with different fin configurations were measured [9-11]. Based on the experimental database heat transfer and friction correlations were predicted for air-side of fin-and-tube heat exchangers.

Horuz [12] investigated theoretically and experimentally the parameters affecting the cooling capacity and overall heat transfer coefficient of fin-and-tube heat exchanger. The air velocity, fin spacing, tube diameter, evaporator temperature, refrigerant type and frost height were varied during experimentation. The overall heat transfer coefficient of R717, R22, R502, R134a and R12 were calculated and compared with theoretical results. Somchai et al [13] studied the two-phase heat transfer coefficient characteristics of R134a, evaporating inside a plate fin- and-tube evaporator with plain fin geometry. The experiments were conducted at the different average saturated refrigerant temperatures. During experimentation the volumetric flow rate of air passing through the evaporator were varied between 0.25 [m.sup.3]/s and 0.5 [m.sup.3]/s and between 0.7 [m.sup.3]/s and 1.25 [m.sup.3]/s for the condenser.

Vapor and liquid compositions are different for a multi-component mixture, and will continuously change throughout the heat transfer process, which influences the properties of the two phases. A change in the saturation temperature during the evaporation and condensation processes at a constant pressure is known to have important effects on the heat exchanger design. To the best of the author's knowledge, there are no heat transfer study report currently available on the tube side evaporative heat transfer characteristics of M20 refrigerant mixture flowing through the fin-and-tube heat exchanger. In the present study, evaporative heat transfer characteristic of M20 refrigerant mixture in a fin-and-tube evaporator has been analyzed experimentally in a vapor compression system. The indoor and outdoor temperatures were varied in accordance with BIS [14] standards.

Experimentation

The experimental facility mainly consists of the psychrometric test room and the unit under test suitably instrumented to conduct the heat transfer study.

Psychrometric test room

An experimental setup was constructed as shown in fig.1 that would facilitate performance assessment of a window air-conditioner on various indoor and outdoor conditions in accordance with different standards (BIS and ASHRAE). The psychrometric room consists of two adjacent chambers to maintain indoor and outdoor conditions respectively. Both the rooms have separate AHUs with a cooling coil (Dehumidifier), air heater and steam injection facility (Humidifier) which are controlled / modulated by suitable feed back control system to maintain the required indoor and outdoor test conditions indicated in fig.1. The facility has been designed to maintain individual temperature readings within the tolerance prescribed in BIS- -1391-1992 and ASHRAE-116 -1995 standards ([+ or -] 0.5[degrees]C for dry bulb temperatures and [+ or -] 0.3[degrees]C for wet bulb temperatures). Six temperature sensors were strategically located in each room to confirm the uniformity in room temperature within [+ or -] 0.5[degrees]C.

To measure the supply air flow rate from the unit under test a code tester is available in the indoor room. This code tester design is based on ASHRAE standard 41.2--1987 [15]. This consists of a set of nozzles that can be suitably selected to allow the air to flow through the selected nozzle. The pressure drop across the nozzle is measured using a differential pressure transducer. There is an auxiliary blower driven with a VFD to maintain zero gauge pressure at the receiving chamber. This is done so that the UUT does not experience any resistance to throw the supply air due to upstream surging effect of the flow through nozzle. The code tester was connected with a suitable leak proof duct to the supply grill of the UUT. The air flow rate was varied using a selector switch generally available in window airconditioner appliance.

[FIGURE 1 OMITTED]

Unit Under Test

The unit under test was suitably modified to connect temperature sensors (RTDs PT100-class A [+ or -] 0.15[degrees]C accuracy) and pressure sensors ([+ or -] 0.1% accuracy) across each component. The mass flow rate of refrigerant was measured by a Coriolis type mass flow meter ([+ or -] 0.1% accuracy) connected in the liquid line. The entire flow lines along with components were properly insulated to avoid heat infiltration. The power consumed by the compressor was measured by a separate power meter ([+ or -] 0.25% accuracy). Same diameter and different length of capillaries were used for R22 and M20. Suitable hand shutoff valves were used to select the required capillary to be included in the circuit. One sight glass was provided in the liquid line to check the condition of the condensed refrigerant in the circuit. The supply and return air temperatures (DBT and WBT) were measured by suitable RTDs (class--A) fixed at appropriate locations. The evaporator with plain fin geometry are made from aluminum plate finned, copper tube. The external dimension of the fin-and-tube evaporator is 410 x 365 x 90mm (Height x width x Depth). The outer diameter of the copper tube is 9.52 mm.

Data Logging

All measured data are logged into a PC through a suitable data logging system. Once steady state condition is achieved all the data will be automatically logged in to the system. The steady state condition is manually confirmed by checking the uniformity in temperature indicated by all the room temperature sensors in accordance with the test conditions. After stabilization of indoor and outdoor air temperature, refrigerant temperatures and pressures across the each component, flow rates of refrigerant and air were scanned and recorded in the computer through a data acquisition system.

Experimental Procedure

The performance of the system was determined in accordance with the BIS--1391- 1992 test conditions (given in fig. 1) for residential sized air conditioner. In air enthalpy test method, refrigerating capacity is determined from the difference in enthalpies obtained against DBT and WBT of air entering and leaving the unit under test (UUT) and the associated air flow rate under specified test conditions. The mass of air was calculated using the measured pressure drop across the nozzle and DBT and WBT of sample air in the code tester based on ASHRAE standard 41.2-1987 [18]. Refrigerant side measurements were also made to ensure that the maximum difference between the air side and refrigerant side capacity was less than 6% as prescribed in the standards. Properties of refrigerant were extracted from REFPROP [17].

To have a realistic comparison of the performance of the M20 refrigerant mixtures with conventional refrigerant the experiment was carried out initially with the conventional refrigerant R22. At steady state refrigerant mass flow rate, pressure and temperature across the evaporator and condenser, DBT / WBT of return as well as supply air and mass flow rate of air were measured for the various indoor and outdoor room conditions as given in fig. 1.

Before starting the experiment with mixture, the mixture was prepared separately in a cylinder. For the mixture, the equivalent charge quantity for the considered mass of R22 was obtained, along with the mass of R407C and the HC blend, making use of the specific volume ratios at suction condition. Each mixture component was weighed individually in an electronic balance with an accuracy of [+ or -] 0.1 g and filled in the cylinder with the help of a suitable charging manifold. After completing all the performance tests with R22, the refrigerant was recovered and equivalent quantity of M20 mixture was charged. The same tests for performance, pull down and per day energy consumption were repeated for M20.

Test condition

In general mass flux, heat flux, inlet quality and saturation temperature and pressure are the major factors which affect evaporative heat transfer in all heat exchanger. But in practical cases, the heat transfer coefficient is mainly affected by evaporator inlet air temperature and condenser inlet air temperature in air cooled heat exchangers. In this study, the evaporator air inlet and condenser air inlet temperatures are varied in accordance with BIS [14] standards.

Data reduction

The inside area, outside surface area and air side area are calculated using equations from Kuppan [16] based on the measured diameter of tube, thickness of the tube, number of fins, fin thickness. The enthalpy of supply and return air are calculated using the measured DBT and WBT of the supply and return air. Refrigerant side enthalpies are calculated based on measured inlet and outlet temperatures and pressures. The refrigerant and refrigerant mixture properties were calculated from REFPROP version 7.1 [17].

The overall heat transfer coefficient, U can be calculated using eqn. 1

[U.sub.o] = [Q.sub.ave]/F[A.sub.o][DELTA][T.sub.lm] ... (1)

Where,

[Q.sub.ave] = ([Q.sub.a] + [Q.sub.r])/2 ... (2)

[Q.sub.a] = [??]([DELTA][h.sub.a]) ... (3)

[Q.sub.r] = [??]([DELTA][h.sub.r]) ... (4)

Where, [DELTA][T.sub.lm]) is the logarithmic mean temperature difference, [A.sub.o] is the outside surface area and F is the correction factor which is equal to 1 [13] for this kind of fin-and-tube evaporator.

The temperature difference, LMTD was determined from the inlet and exit temperatures (DBT) of the air flowing through the evaporator and from the inlet and outlet saturation temperature of refrigerant flowing in the test section. The refrigerant temperature at the inlet of the test section was calculated by considering isenthalpic process in expansion device. The inlet two-phase temperatures were calculated based on the calculated inlet quality and measured saturation pressure. The outlet temperature of refrigerant mixture was calculated based on the dew point temperature at the corresponding measured pressure and wall temperature.

Result and discussion

Experiments were performed for R22 and M20 refrigerant mixture at the test conditions as given in fig.1. Based on measured parameters the evaporator pressure, heat flux, overall heat transfer coefficient, LMTD and mass flow rate for M20 refrigerant mixture were discussed with the baseline results of R22 at two different air velocity.

Figure 2 shows the variation of refrigerant pressure with velocity of air passing over the evaporator with different test conditions for R22 and M20. It was observed that as the air velocity increases the refrigerant pressure decreases for all test conditions. The maximum pressure reached in ETA_M compared to other test conditions due to higher operating conditions. However at higher velocity for all test condition the evaporator inlet pressure was lower for all test conditions. But all test conditions the refrigerant pressure of M20 was higher than that of R22. It was measured that the evaporator inlet pressure of M20 refrigerant was higher in the range 0.5% to 3.66% than R22 at all operating conditions.

[FIGURE 2 OMITTED]

Figure 3 shows the variation of mass flow rate for R22 and M20 at two different velocities with various test conditions. As the velocity increases the mass flow rate decreases for all test conditions. However at higher velocity the mass flow rate was lower for all test conditions. It was found that the mass flow rate of M20 refrigerant mixture was lower than that of R22 for a same test condition due to lower liquid density than R22 as evidence from REFPROP [17]. The mass flow rate of M20 refrigerant mixture was lower in the range 8.58% to 13.1% than R22 at all studied operating conditions.

[FIGURE 3 OMITTED]

Figure 4 shows the variation of heat flux for R22 and M20 refrigerant mixture at two different velocity of air with different test conditions. It was observed that as the velocity increases heat flux increases for all test conditions. The test ETA has higher heat flux than the other test conditions. The heat flux of R22 was varied from 7.716 to 9.428 kW/[m.sup.2] and M20 was varied from 7.450 to 8.585 kW/[m.sup.2] at all tested conditions. It was observed that the heat flux of M20 refrigerant mixture was lower in the range 1.8% to 9.82% than R22 at the studied operating conditions.

[FIGURE 4 OMITTED]

Figure 5 the variation of logarithmic mean temperature difference with velocity of air passing through the evaporator with different test conditions for R22 and M20 refrigerant mixture. Due to lower refrigerant outlet condition, the LMTD decreases as the velocity increases for all considered operating conditions. The test ETB has high LMTD than the other test conditions due to higher operating condition for R22 and M20 refrigerant mixture. However at higher velocity LMTD is lower for all test conditions. It was observed that the M20 refrigerant mixture LMTD was higher in the range 1.15% to 8.67% than R22. This was attributed due to higher air outlet temperatures and higher inlet and outlet temperatures of M20 refrigerant mixture.

Figure 6 shows the variation of overall heat transfer coefficient for R22and M20 refrigerant mixture at different velocity of air passing through the evaporator with different test conditions. However at highest air velocity for all test condition the overall heat transfer coefficient is lower for both the refrigerant. The test ETA has higher overall heat transfer coefficient than other test conditions due to lower LMTD. The overall heat transfer coefficient was varied for R22 in the range 143.73 to 193.81 W/[m.sup.2]K and for M20 refrigerant mixture 131.56 to 174.45 W/[m.sup.2]K. The overall heat transfer coefficient of M20 was lower than R22 for all test conditions due higher LMTD and lower heat transfer rate for the same geometric conditions. The overall heat transfer coefficient of M20 was lower in the range 6.72% to 20.76% than R22 at all studied operating conditions.

[FIGURE 5 OMITTED]

[FIGURE 6 OMITTED]

Uncertainty Analysis

Experimentation using R22 and M20 refrigerant mixture included the measurement of air temperatures (DBT and WBT), pressure difference across the nozzles, refrigerant temperature and pressures at various locations of the systems and refrigerant mass flow rate. Air-side heat transfer rate, refrigerant-side heat transfer rate, overall heat transfer coefficient were calculated based on the measured parameters. The uncertainties in air-side heat transfer rate, refrigerant-side heat transfer rate and overall heat transfer coefficient were in the range 2.6% to 3.5%, 1.2% to 2.0% and 3.2% to 4.7% respectively. The uncertainties for the R22 and M20 refrigerant mixture were calculated from the equations listed below,

[U.sub.o] = [Q.sub.ave]/[DELTA][T.sub.lm] ... (1)

[Q.sub.ave] = [Q.sub.r,ac] + [Q.sub.a,ac] ... (2)

[Q.sub.r,ac] = [[??].sub.r] ([DELTA][h.sub.r]) ... (3)

[Q.sub.a,ac] = [[??].sub.a] ([DELTA][h.sub.a]) (4)

[DELTA][h.sub.a] = f([DELTA][P.sub.nozzle], [T.sub.WBT], [T.sub.DBT]) ... (5)

[DELTA][h.sub.r] = f([T.sub.e,in], [T.sub.e,out], [P.sub.e,in], [P.sub.e,out], [T.sub.WBT], [T.sub.DBT]) ... (6)

[DELTA][T.sub.lm] = f([T.sub.e,in], [T.sub.e,out], [T.sub.DBT,supply], [T.sub.DBT,return]) ... (7)

[delta][Q.sub.a,ac] = [[([delta][[??].sub.a]/[[??].sub.a]).sup.2] + [([delta][DELTA][h.sub.a]/[DELTA][h.sub.a]).sup.2].sup.1/2] ... (8)

[delta][[??].sub.a] = [[([delta][DELTA][P.sub.nozzle]/[DELTA][P.sub.nozzle]).sup.2] + [([delta][T.sub.WBT]/[T.sub.WBT]).sup.2] + [([delta][T.sub.DBT]/[T.sub.DBT]).sup.2].sup.1/2] ... (9)

[delta][DELTA][h.sub.a] = [[([delta][h.sub.a,supply]/[h.sub.a,supply]).sup.2] + [([delta][h.sub.a,return]/[h.sub.a,return]).sup.2].sup.1/2] ... (10)

[delta][h.sub.a,supply] = [[([delta][T.sub.WBT]/[T.sub.WBT]).sup.2] + [([delta][T.sub.DBT]/[T.sub.DBT]).sup.2].sup.1/2] ... (11)

[MATHEMATICAL EXPRESSION NOT REPRODUCIBLE IN ASCII] ... (12)

[delta][DELTA][T.sub.lm] = [[([delta][T.sub.e,in]/[T.sub.e,in]).sup.2] + [([delta][T.sub.e,out]/[T.sub.e,out]).sup.2] + [([delta][T.sub.DBT,supply]/[T.sub.DBT,supply]).sup.2] + [([delta][T.sub.DBT,return]/[T.sub.DBT,return]).sup.2].sup.1/2] ... (13)

Conclusions

The heat-transfer characteristics of M20 refrigerant mixture and R22 were experimentally investigated in a fin-and-tube evaporator as a function of evaporator air inlet temperature, condenser air inlet temperature and air velocity. A study was carried out in a 9.525mm smooth tube fin-and-tube evaporator with optimum charge quantity of R22 and M20 refrigerant mixture. Based on the measurement the following conclusions were drawn,

(1) The inlet pressure of M20 was higher in the range 0.5% to 3.66%.

(2) Mass flow rate of M20 was 8.58% to 13.1% lower than R22

(3) Heat flux of M20 was lower in the range 1.8% to 9.82%.

(4) LMTD of M20 was higher in the range 1.4% to 8.67%.

(5) Overall heat transfer coefficient of M20 was lower in the range 6.28% to 20.76%.

The fact that POE oil can be dispensed with by using R407C/HC blend refrigerant mixture in the place of R407C is a significant finding in this work. This mixture would be the better choice to retrofit the existing window air-conditioners while R22 has to be phased out. When R22 is to be phased out, there may be millions of airconditioners operating with R22 in both developed and developing countries. Retrofitting with M20 is an option to extend the life of such units without changing the mineral oil in the compressor, although the performance is slightly poorer.

Nomenclature

A               Surface area ([m.sub.2])
AHU             Air handling unit
ASHRAE          American society of Heating Refrigerating
                Air-Conditioning Engineers
BIS             Bureau of Indian Standards
DBT             Dry bulb Temperature ([degrees]C)
h               Heat transfer coefficient (W/[m.sup.2]/K)
HC              Hydrocarbon
HFC             Hydroflurocarbon
k               Thermal conductivity (W/m/K)
L               Length of the tube (m)
LMTD            Logarithmic mean Temperature difference (K)
m               mass flow rate
[m.sub.a]       mass the air (kg of dry air/hr)
MFM             Mass Flow Meter
NF              Normal fan speed
Q               Heat transfer rate (W)
RTD             Resistance temperature detector
SG              Sight Glass
SQ              Super Quiet fan speed
T               Temperature
U               Overall heat transfer coefficient (W/[m.sup.2]/K)
UUT             Unit under test
WBT             Wet bulb Temperature ([degrees]C)

Subscripts

a               Air
ave             Average
f               Fin
I               Refrigerant side
lm              LMTD
in              Inlet
m               mean
o               Outer
out             Outlet
r               Refrigerant
s               Saturation
w               wall

Greek Symbol

[delta]         Thickness of wall
[[eta].sub.o]   Effectiveness
[eta]           Surface efficiency

References

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[2] Jabaraj, D.B., Narendran, A., Mohanlal, D., and Renganarayanan, S., 2007, "Evolving an optimal composition of HFC407C/HC290/HC600a mixture as an alternative to HCFC22 in window air conditioners", International Journal of Thermal Sciences, 46, pp 276-283.

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[11] Webb, R.L., 1990, "Air-side heat transfer correlations for flat and wavy plate fin-and-tube geometrics", ASHRAE Transactions, 96, pp. 445-449.

[12] Horuz, I., Kurem, E., and Yamankaradeniz, R., 1998, "Experimental and theoretical performance analysis of air cooled plate finned tube evaporators", International Communication in Heat and Mass Transfer, 25, pp. 787-798.

[13] Wongwises, S., Disawas, S., Kaewon, J., and Onurai, C., 2000, Two-phase evaporative heat transfer co-efficient of refrigerant HFC-134a under forced flow condition in a small horizontal tube, International Communication in Heat Mass Transfer, 27, pp. 35-48.

[14] Indian Standard 1391, (1992) Part-I "Specification of Unitary Air Conditioners".

[15] ANSI/ASHRAE 41.2-1987, "Standard methods for laboratory air flow measurement".

[16] Kuppan, T., Heat Exchanger Design Hand book, pp 159-228, Marcel Dekker Publications, 2000.

[17] National Institute of Standards and Technology, NIST Thermodynamic properties of refrigerants and their mixtures, data base 23, (REFPROP V 7.1).

[18] ASHRAE standards. Methods for Testing for Rating Seasonal Efficiency of Unitary Air Conditioners and Heat Pumps. ANSI/ASHRAE 116-1995, Atlanta.

M. Herbert Raj * and D. Mohan Lal

Refrigeration and Air Conditioning Division, Department of Mechanical Engineering, Anna University, Sardar Patel road, Chennai, India.

* Corresponding author E-mail address: herbertraj09@gmail.com