Engineering 41 Laboratory 6

Refrigeration Systems

Alexis Reedy, Zach Pezzementi, Adem Kader, Aron Dobos

December 10, 2004

 

Abstract

 

In this laboratory session, the functioning of a demonstration vapor cycle refrigeration system was observed, and its performance was characterized in various operating modes.  The thermal efficiency or coefficient of performance of the refrigeration system was determined for three different expansion valve configurations by logging pressure, temperature and mass flow data.  Direct observation of the phase changes of the refrigerant (R-11) was also possible due to the open glass construction of the demonstration unit.

 

Introduction and Theory

 

Refrigeration involves the transfer of heat from a lower temperature region to a higher temperature region.  Vapor-compression refrigeration methods are based on the Carnot cycle.  Since the Carnot cycle is fully reversible, it can be run as a heat pump instead of a heat engine.  As a result, it would be the most efficient refrigeration cycle possible.  However, many impracticalities are involved with true Carnot refrigeration, since it is very difficult to build a compressor that will work well with a two phase mixture.  Also, the expansion of high moisture content refrigerant introduces difficulties. 

 

The ideal vapor-compression refrigeration cycle is obtained from the Carnot cycle by ensuring that the refrigerant is completely vaporized before it is compressed, and by replacing the turbine with a throttling device.  The vapor-compression cycle achieves refrigeration by vaporizing the refrigerant to extract heat from the cold region, compressing it in the vapor phase, and then condensing it in the hot region. There are four stages to the vapor-compression cycle, as seen on a T-s diagram:

 

 

Figure 1. T-s Diagram for Ideal Vapor-Compression Refrigeration Cycle [Cengel 567]

 

1 – 2:  Compression

In this process, the compressor isentropically compresses the fully vaporized refrigerant to the condenser pressure.  The work input to the compressor is the net work input to the whole refrigeration system, and is thus the denominator of the coefficient of performance (COP) expression.

 

2 – 3:  Constant-pressure heat rejection in a condenser

            The refrigerant enters the condenser in superheated form.  Heat is rejected to the surroundings, and the refrigerant leaves the condenser at (3) in saturated liquid.  The temperature of the saturated liquid is still above that of the surroundings.

 

3 – 4:  Throttling in an expansion device

            The expansion device creates a rapid pressure drop, causing the temperature of the refrigerant to drop below that of the refrigerated region.  The refrigerant thus leaves the throttling valve as a low quality saturated mixture and enters the evaporator.

 

4 – 1:  Constant-pressure heat absorption

            The low pressure low quality refrigerant vaporizes in the evaporator by absorbing heat from the cold space.  By the time the refrigerant enters compressor, it is in saturated vapor form.

 

The vapor-compression refrigeration system is implemented in the following configuration:

Figure 2. Components of a Vapor-Compression Refrigeration System

 

The coefficient of performance (COP) for a refrigerator is the ratio of cooling effect to the amount of work input.

COP = Qevaporator / Wcompressor

 

Since Qevaporator = mflow* (h1 – h4), and Wcompressor = mflow * (h2 - h1) , and assuming a constant refrigerant flow rate throughout the refrigeration system, we can write the coefficient of performance as

COP = (h1 – h4) / (h2 - h1)

 

The actual vapor compression cycle differs from the ideal cycle due mainly to irreversibilities that show up in several components.  Heat transfer to or from the surroundings is one source that introduces irreversibilities, as well as fluid friction in the pipes.  In real life, it is usually unavoidable to have some pressure losses in the evaporator and condenser, meaning that more work input is required to maintain the performance.  As a result, the coefficient of performance drops.

 

 

Procedure

 

The demonstration described in the attached handout was observed.

 

 

Data and Results

 

Figure 3. Refrigeration Cycle with Capillary Expansion Valve

 

Figure 4. Refrigeration Cycle with Thermal Expansion Valve

 

Figure 5. Refrigeration Cycle with Hand Expansion Valve 1

 

Figure 6. Refrigeration Cycle with Hand Expansion Valve 2

 

Figure 7. Refrigeration Cycle with Hand Expansion Valve 3

 

Analysis, Discussion, and Conclusions

 

In the demonstration unit, the COP was found to vary linearly with the flowrate. The operating conditions under which the system exhibited the best performance were therefore when the flowrate was highest.

 

Table 1. COP and Flowrate with Different Expanders

Expander

Flowrate

COP

Capillary

0.39

0.3877

Thermal

0.23

0.9748

Hand 1

0.18

0.7838

Hand 2

0.28

1.2558

Hand 3

0.43

1.7371

 

Figure 8. COP vs Flowrate

 

There are several ways by which one can improve the performance of refrigerators. Since the COP is simply a ratio, the most obvious ways to improve it are by increasing its numerator or decreasing its denominator. Improving the efficiency of the pump or using a refrigerant with lower viscosity, but similar thermal properties, could both act to lower the power consumption of the refrigerator.

 

 

Figure 9. P-h Diagram for a Refrigeration Cycle

 

Increasing the numerator of the COP can be accomplished by several changes. From the refrigeration cycle, one can see that the step which determines the measured performance is from state 1 to state 2. The difference in h between these points multiplied by the mass flowrate gives the heat absorbed by the evaporator. It is therefore desirable that points 1 and 2 be as far apart as possible in the h-dimension. Both states must, however, remain in the two-phase region to allow the rest of the evaporator and compressor to function. The heat absorbed in the evaporator could potentially be increased by the use of more thermally conductive materials and by increasing the surface area of contact with the environment.

 

The expander has a significant effect on the amount of heat absorbed since it determines the position of phase 1. In the ideal expander, the fluid undergoes a constant-enthalpy process, generating a vertical line on the graph above. In reality, these expanders are usually allowed to be non-reversible so that they can be cheaply made, and there is actually a change in entropy and in enthalpy associated with the process. The expander should be designed to place state 1 as far to the left as possible, while remaining in the two-phase region. From the Gibbs relation, Tds = dh – vdP, one can see that decreasing the change in entropy for this process, if T and dP are unaltered, would also decrease dh for the process, meaning that the line on the P-h diagram for state 1-4 would be as close to vertical as possible. Small gains in performance can be made by decreasing the temperature of the refrigerant, since fluids at lower temperatures require a greater change in h to transition from liquid to gas. Increasing the insulation of the compressor and expander can prevent the system from absorbing heat in those parts of the cycle, allowing the refrigerant to be as cold as possible during the evaporation stage. Whether consumers would consider the cost of implementing these changes to justify the increase in performance remains to be seen.

 

 

 

 

 

 

 

Refrigeration and Heating Terminology

 

Window air conditioners are rated in terms of heat removal capacity in BTUH and an Energy Efficiency Ratio, or EER.  The EER value is determined by the following equation:

 

, so in BTU/(W-h)

 

A related measure is the seasonal energy efficiency ratio SEER:

 

 

The coefficient of performance is defined as:

 

 

Obtained from reference (2).

 

According to the Ontario Ministry of Energy, SEER values range from 8 to 17.

 

Heating Degree Days (HDD) and Cooling Degree Days (CDD) are measured as Fahrenheit differences from 65 degrees each day. The typical monthly CDD of Philadelphia is given by

 

 

 

NORMALS

1971-2000

YRS

JAN

FEB

MAR

APR

MaY

JUN

JUL

AUG

SEP

OCT

NOV

DEC

ANN

PHILADELPHIA, PA

30

0

0

2

10

70

234

395

351

152

19

2

0

1235

 

 

Example:

A 10,000 Btu system with a SEER of 10.00 located in Philadelphia, Pennsylvania, with an electrical rate of 14 cents per kwh (kilowatt hour).

Formula:

Capacity (Btuh)/SEER X Cooling Load Hrs/1000 X Elec. Rate = Annual Cost of Operation

10000/10 X 2250/1000 X .14 = 1000 X 2.25 X .14 = $315.00

 

For an SEER of 8, this cost rises to $393.75, or for an SEER of 15, the cost drops to $210, a savings of 33%!!

EXTRA

 

A)

Table 2. Fuel Required for Transfer of 1 BTU of Heat in Hicks with Different Expanders

Expander

COP

BTU burned at Eddystone

Capillary

0.97

2.51

Hand 1

1.96

1.24

Thermal

2.44

1.00

Hand 2

3.14

0.78

Hand 3

4.34

0.56

 

For calculating these values, we assumed best measured heat rate was approximately 8,530 Btu/kWh. Then, to convert our measurements from BTU to kWh, we multiplied by the conversion factor .0002928. The power required to transfer 1 BTU of heat was taken to be 1/COP.
 
B)

Assuming an air conditioner with an EER of 10, which is fairly good for a window unit according to …, 1 BTU requires 10 watt-hours of energy, which in turn requires 10/8530 = 0.0012 BTU of fuel to be burned at Eddystone.

Resources

 

1. Ontario Ministry of Energy

http://www.energy.gov.on.ca/index.cfm?fuseaction=conservation.guide12

 

2. City Water, Light & Power of Springfield, Illinois

http://www.cwlp.com/Energy_services/efficiency_ratings.htm

 

3. ABC – 123

http://www.leeric.lsu.edu/bgbb/7/ecep/hvac/a/a.htm

 

4. USA Today

http://www.usatoday.com/weather/resources/askjack/waskdays.htm

 

5. Normal Monthly Cooling Degree Days (Base 65) – Golden Gate Weather Services

http://ggweather.com/ccd/nrmcdd.htm