March 31, 2007

Laboratory Report: MEDE2005 Thermofluid for Medical Engineers

Applied Thermodynamics Laboratory

Test on domestic Air-Conditioner (Mod. TAC/82)

Wong Yan Tak

BEng in Medical Engineering ProgrammeFaculty of EngineeringThe University of Hong KongPokfulam Road, Hong Kong

University No.: 2005205784Lab Session: G4, March 31, 2007

I. BackgroundThe test unit TAC/82 was a domestic air-conditioner. It used Freon-22 as its working

fluid and was powered by a 0.5HP Hermetic Compressor. The condenser and evaporator were two forced-air finned heat exchangers cooled by a two-wheel fan. The expansion of the working fluid mainly took place in the capillary tubes. The unit was equipped with peripheral instruments such as pressure gauge, digital thermometer, floating flowmeter, thermostat and sight glass, etc. These instruments helped indicating the status of the unit, including its power consumption, temperature and pressures in some important parts of the unit. By using data on these instruments, the actual performance of the unit could be quantitatively expressed.

II. ObjectivesThe objective of this laboratory session was to determine the actual coefficient of

performance of the refrigeration unit.

III. Results and CalculationsThe experiments were carried out with two settings of the TAC/82, one with ‘*’

indicating high fan speed, and another with ‘***’ indicating low fan speed.

Figure 1: The TAC/82 Air-Conditioning Unit

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The room temperature recorded was 24oC, and the pressure was 1bar (100kPa).Fan Setting '*' '***'Low Gauge Pressure /bar 2.6 3High Gauge Pressure /bar 12.6 13.8Wattmeter Reading /W 625 600Voltmeter Reading /V 215 215Ammeter Reading /A 2.8 2.7Air Temperature at Evaporator Outlet /oC 19.7 17.8Refrigerant Flowrate /(l/min) 0.25 0.26Compressor Inlet Temperature, t1/oC 26.6 28.2Compressor Outlet Temperature, t2/oC 72 90.5Condenser Outlet Temperature, t31/oC 29 31.9Capillary Inlet Temperature, t3/oC 28.7 31.1Evaporator Temperatures, te1/oC te2/oC te3/oC te4/oC

019.322.824.8

0.715.62325.7

The colour of the refrigerant in the sight class was blue throughout the experiment, indicating it was functioning properly.

F igure 2: Colour indicator of the Refrigerant

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After taking the atmospheric pressure into account, the actual low and high gauge pressures could be obtained. Then the corresponding enthalpies in different temperatures could be found from the Pressure-Enthalpy Diagram for the Air-Conditioning Model.

From the diagrams, the enthalpies obtained were as follows:Fan Setting Enthalpy /(kJ/kg)‘*’ H1, at P=3.6bar and T1=t1

H2, at P=13.6bar and T2=t2

H3, at P=13.6bar and T3=t3

H4 = H3

426447235235

‘***’ H1, at P=4bar and T1=t1

H2, at P=14.8bar and T2=t2

H3, at P=14.8bar and T3=t3

H4 = H3

425461238238

Heat taken up in evaporator = H1 – H4 (kJ/kg)Heat given up in condenser = H2 – H3 (kJ/kg)Actual Compression Work = H2 – H1 (kJ/kg)

Actual coefficient of performance, COPactual =

For Fan Setting at ‘*’, H1 – H4 = 191kJ/kg H2 – H3 = 212kJ/kg H2 – H1 = 21kJ/kg

∴for this setting, COPactual =

For Fan Setting at ‘***’, H1 – H4 = 187kJ/kg H2 – H3 = 223kJ/kg H2 – H1 = 36kJ/kg

∴for this setting, COPactual =

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IV. Discussions and ConclusionsThe circuit diagram of the air-conditioner was as follows. For different fan settings,

the temperatures in the parts varied.

Figure 3: Circuit Diagram with temperature labels for Fan Setting ‘ * ’

Figure 4: Circuit Diagram with temperature labels for Fan Setting ‘ *** ’

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T1=26.6oC

T2=72oC

t31=29oC

T3=28.7oC

T4=22.8oC

T1=28.2oC

T2=90.5oC

t31=31.9oC

T3=31.1oC

T4=23oC

In this experiment, all processes should be reversible in theory. However, there were some constraints to make these processes irreversible. Some of them were fluid friction between pipe walls and the refrigerant as well as heat transfer to and from surroundings through temperature difference.

In the ideal cycle, the refrigerant leaves the evaporator and enters the compressor in saturated vapour state. However, in reality, it may not be possible to control the state of the refrigerant so precisely. Instead, the system was designed so that the refrigerant was slightly superheated at the compressor inlet. This slight modification ensured that the refrigerant was completely vaporized when it entered the compressor. Also, the capillary tubules connecting the evaporator to the compressor was usually very long, resulting in the pressure drop caused by fluid friction and heat transfer from the surroundings to the refrigerant could not be neglected. These factors contributed to the increase in the specific volume of the refrigerant, thus an increase in the power input requirement to the compressor since steady-flow work was proportional to the specific volume.

The compression process in the ideal cycle is internally reversible and adiabatic. The actual compression process, however, included frictional effects, which increased the entropy. The heat transfer might alter the entropy depending on its direction. Therefore, the entropy of the refrigerant varied during actual compression process. This variation might be beneficial if the specific volume of the refrigerant could be made smaller as well as the work input requirement. A suggestion was to cool the refrigerant during the compression process whenever practical and economical.

In ideal case, the refrigerant is assumed to leave the condenser as saturated liquid at the compressor exit pressure. In actual case, however, it was unavoidable to have some pressure drop in the condenser as well as in the pipes connecting the condenser to the compressor. Besides, it was difficult to execute the condensation process with such precision that the refrigerant was a saturated liquid at the end, and it was undesirable to route the refrigerant to the expansion valve before the refrigerant had completely condensed. Therefore, the refrigerant was irreversibly cooled somewhat before it entered the throttling valve.

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