Moran M.J., Shapiro H.N. Fundamentals of Engineering Thermodynamics

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compressor, the mass and energy rate balances for a control volume enclosing the

compressor give

5 h

2 h

(10.4)

where W

is the rate of power input per unit mass of refrigerant flowing.

c Next, the refrigerant passes through the condenser, where the refrigerant condenses

and there is heat transfer from the refrigerant to the cooler surroundings. For a

control volume enclosing the refrigerant side of the condenser, the rate of heat

transfer from the refrigerant per unit mass of refrigerant flowing is

out

5 h

2 h

(10.5)

c Finally, the refrigerant at state 3 enters the expansion valve and expands to the

evaporator pressure. This process is usually modeled as a throttling process for

which

5 h

(10.6)

The refrigerant pressure decreases in the irreversible adiabatic expansion, and

there is an accompanying increase in specific entropy. The refrigerant exits the

valve at state 4 as a two-phase liquid–vapor mixture.

In the vapor-compression system, the net power input is equal to the compressor

power, since the expansion valve involves no power input or output. Using the quan-

tities and expressions introduced above, the coefficient of performance of the vapor-

compression refrigeration system of Fig. 10.3 is

b 5

2 h

(10.7)

Provided states 1 through 4 are fixed, Eqs. 10.3 through 10.7 can be used to

evaluate the principal work and heat transfers and the coefficient of performance

of the vapor-compression system shown in Fig. 10.3. Since these equations have

been developed by reducing mass and energy rate balances, they apply equally for

actual performance when irreversibilities are present in the evaporator, compres-

sor, and condenser and for idealized performance in the absence of such effects.

Although irreversibilities in the evaporator, compressor, and condenser can have

a pronounced effect on overall performance, it is instructive to consider an ideal-

ized cycle in which they are assumed absent. Such a cycle establishes an upper

limit on the performance of the vapor-compression refrigeration cycle. It

is considered next.

10.2.2

Performance of Ideal Vapor-Compression Systems

If irreversibilities within the evaporator and condenser are ignored, there are

no frictional pressure drops, and the refrigerant flows at constant pressure

through the two heat exchangers. If compression occurs without irreversibili-

ties, and stray heat transfer to the surroundings is also ignored, the compres-

sion process is isentropic. With these considerations, the vapor-compression

refrigeration cycle labeled 1–2s–3–4–1 on the T–s diagram of Fig. 10.4 results.

The cycle consists of the following series of processes:

Process 1–2s: Isentropic compression of the refrigerant from state 1 to the

condenser pressure at state 2s.

Fig. 10.4 T–s diagram of an ideal vapor-

compression cycle.

10.2 Analyzing Vapor-Compression Refrigeration Systems 593

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594 Chapter 10 Refrigeration and Heat Pump Systems

Process 2s–3: Heat transfer from the refrigerant as it flows at constant pressure

through the condenser. The refrigerant exits as a liquid at state 3.

Process 3–4: Throttling process from state 3 to a two-phase liquid–vapor mixture

at 4.

Process 4–1: Heat transfer to the refrigerant as it flows at constant pressure through

the evaporator to complete the cycle.

All processes of the cycle shown in Fig. 10.4 are internally reversible except for the

throttling process. Despite the inclusion of this irreversible process, the cycle is com-

monly referred to as the ideal vapor-compression cycle.

The following example illustrates the application of the first and second laws of ther-

modynamics along with property data to analyze an ideal vapor-compression cycle.

ideal vapor-

compression cycle

Analyzing an Ideal Vapor-Compression Refrigeration Cycle

c c c c EXAMPLE 10.1 c

Refrigerant 134a is the working fluid in an ideal vapor-compression refrigeration cycle that communicates thermally

with a cold region at 08C and a warm region at 268C. Saturated vapor enters the compressor at 08C and saturated

liquid leaves the condenser at 268C. The mass flow rate of the refrigerant is 0.08 kg/s. Determine (a) the compressor

power, in kW, (b) the refrigeration capacity, in tons, (c) the coefficient of performance, and (d) the coefficient of

performance of a Carnot refrigeration cycle operating between warm and cold regions at 26 and 08C, respectively.

SOLUTION

Known:

An ideal vapor-compression refrigeration cycle operates with Refrigerant 134a. The states of the refrig-

erant entering the compressor and leaving the condenser are specified, and the mass flow rate is given.

Find: Determine the compressor power, in kW, the refrigeration capacity, in tons, coefficient of performance, and

the coefficient of performance of a Carnot vapor refrigeration cycle operating between warm and cold regions

at the specified temperatures.

Schematic and Given Data:

Engineering Model:

Each component of the cycle is analyzed as a control volume at steady state. The control volumes are

indicated by dashed lines on the accompanying sketch.

Fig. E10.1

32s

Condenser

Evaporator

Compressor

out

Warm region T

= 26°C = 299 K

Cold region T

= 0°C = 273 K

Expansion

valve

Temperature of

warm region

Temperature of

cold region

26°C

0°C

VCRC

A.30 – Tab a

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2. Except for the expansion through the valve, which is a throttling process, all processes of the refrigerant

are internally reversible.

3. The compressor and expansion valve operate adiabatically.

4. Kinetic and potential energy effects are negligible.

5. Saturated vapor enters the compressor, and saturated liquid leaves the condenser.

Analysis: Let us begin by fixing each of the principal states located on the accompanying schematic and

T–s diagrams. At the inlet to the compressor, the refrigerant is a saturated vapor at 08C, so from Table A-10,

5 247.23 kJ/kg and s

5 0.9190 kJ/kg

The pressure at state 2s is the saturation pressure corresponding to 268C, or p

5 6.853 bar. State 2s is

fixed by p

and the fact that the specific entropy is constant for the adiabatic, internally reversible compres-

sion process. The refrigerant at state 2s is a superheated vapor with h

5 264.7 kJ/kg.

State 3 is saturated liquid at 268C, so h

5 85.75 kJ/kg. The expansion through the valve is a throttling process

(assumption 2), so h

5 h

(a) The compressor work input is

5 m

2 h

where m

is the mass flow rate of refrigerant. Inserting values

5 10.08 kg

s21264.7 2 247.232 kJ

1 kW

1 kJ

5 1

4 k

(b) The refrigeration capacity is the heat transfer rate to the refrigerant passing through the evaporator. This is

given by

5 m

2 h

5 10.08 kg

60 s

min

1247.23 2 85.752 kJ

1 ton

211 kJ

min

5 3.67 ton

b 5

2 h

247.23 2 85.75

264.7 2 247.23

5 9.24

(d) For a Carnot vapor refrigeration cycle operating at T

5 299 K and

5 273 K, the coefficient of performance determined from Eq. 10.1 is

➋

max

2 T

5 10.5

➊ The value for h

can be obtained by double interpolation in Table A-12 or

by using Interactive Thermodynamics: IT.

➋ As expected, the ideal vapor-compression cycle has a lower coefficient of

performance than a Carnot cycle operating between the temperatures of the

warm and cold regions. The smaller value can be attributed to the effects of

the external irreversibility associated with desuperheating the refrigerant in

the condenser (Process 2s–a on the T–s diagram) and the internal irrevers-

ibility of the throttling process.

Ability to…

❑

sketch the T–s diagram of

the ideal vapor compression

refrigeration cycle.

❑

fix each of the principal

states and retrieve

necessary property data.

❑

calculate refrigeration

capacity and coefficient

of performance.

❑

compare with the correspond-

ing Carnot refrigeration cycle.

✓

Skills Developed

Keeping all other given data the same, determine the mass

flow rate of refrigerant, in kg/s, for a 10-ton refrigeration capacity.

Ans. 0.218 kg/s.

➊

10.2 Analyzing Vapor-Compression Refrigeration Systems 595

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596 Chapter 10 Refrigeration and Heat Pump Systems

10.2.3

Performance of Actual Vapor-Compression Systems

Figure 10.5 illustrates several features exhibited by actual vapor-compression systems.

As shown in the figure, the heat transfers between the refrigerant and the warm and

cold regions are not accomplished reversibly: the refrigerant temperature in the evap-

orator is less than the cold region temperature, T

, and the refrigerant temperature

in the condenser is greater than the warm region temperature, T

. Such irreversible

heat transfers have a significant effect on performance. In particular, the coefficient

of performance decreases as the average temperature of the refrigerant in the evap-

orator decreases and as the average temperature of the refrigerant in the condenser

increases. Example 10.2 provides an illustration.

Fig. 10.5 T–s diagram of an actual

vapor-compression cycle.

Temperature of

warm region, T

Temperature of

cold region, T

Considering the Effect of Irreversible Heat Transfer on Performance

c c c c EXAMPLE 10.2 c

Modify Example 10.1 to allow for temperature differences between the refrigerant and the warm and cold regions

as follows. Saturated vapor enters the compressor at 210°C. Saturated liquid leaves the condenser at a pressure of

9 bar. Determine for the modified vapor-compression refrigeration cycle (a) the compressor power, in kW, (b) the

refrigeration capacity, in tons, (c) the coefficient of performance. Compare results with those of Example 10.1.

SOLUTION

Known:

An ideal vapor-compression refrigeration cycle operates with Refrigerant 134a as the working fluid. The

evaporator temperature and condenser pressure are specified, and the mass flow rate is given.

Find: Determine the compressor power, in kW, the refrigeration capacity, in tons, and the coefficient of perfor-

mance. Compare results with those of Example 10.1.

Schematic and Given Data:

Engineering Model:

Each component of the cycle is analyzed as a control volume at

steady state. The control volumes are indicated by dashed lines on

the sketch accompanying Example 10.1.

2. Except for the process through the expansion valve, which is a throt-

tling process, all processes of the refrigerant are internally reversible.

3. The compressor and expansion valve operate adiabatically.

4. Kinetic and potential energy effects are negligible.

5. Saturated vapor enters the compressor, and saturated liquid exits

the condenser.

Fig. E10.2

–

10°C

0°C

26°

9 bar

VCRC

A.30 – Tab b

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Referring again to Fig. 10.5, we can identify another key feature of actual vapor-

compression system performance. This is the effect of irreversibilities during com-

pression, suggested by the use of a dashed line for the compression process from

state 1 to state 2. The dashed line is drawn to show the increase in specific entropy

that accompanies an adiabatic irreversible compression. Comparing cycle 1–2–3–4–1

with cycle 1–2s–3–4–1, the refrigeration capacity would be the same for each, but

the work input would be greater in the case of irreversible compression than in the

ideal cycle. Accordingly, the coefficient of performance of cycle 1–2–3–4–1 is less

than that of cycle 1–2s–3–4–1. The effect of irreversible compression can be accounted

for by using the isentropic compressor efficiency, which for states designated as in

Fig. 10.5 is given by

2 h

Analysis: Let us begin by fixing each of the principal states located on the accompanying T–s diagram. Starting at

the inlet to the compressor, the refrigerant is a saturated vapor at 2108C, so from Table A-10, h

5 241.35 kJ/kg

and s

5 0.9253 kJ/kg

The superheated vapor at state 2s is fixed by p

5 9 bar and the fact that the specific entropy is constant for

the adiabatic, internally reversible compression process. Interpolating in Table A-12 gives h

5 272.39 kJ/kg.

State 3 is a saturated liquid at 9 bar, so h

5 99.56 kJ/kg. The expansion through the valve is a throttling

process; thus, h

5 h

(a) The compressor power input is

5 m

2 h

where m

is the mass flow rate of refrigerant. Inserting values

5 10.08 kg

s21272.39 2 241.352 kJ

kg `

1 kW

1 kJ

2.48 kW

(b) The refrigeration capacity is

5 m

2 h

5 10.08 kg

60 s

min

1241.35 2 99.562 kJ

1 ton

211 kJ

min

b 5

2 h

241.35 2 99.56

272.39 2 241.35

5 4.57

Comparing the results of the present example with those of Example 10.1,

we see that the power input required by the compressor is greater in the pres-

ent case. Furthermore, the refrigeration capacity and coefficient of performance

are smaller in this example than in Example 10.1. This illustrates the considerable

influence on performance of irreversible heat transfer between the refrigerant

and the cold and warm regions.

Ability to…

❑

sketch the T– s diagram of

the ideal vapor compression

refrigeration cycle.

❑

fix each of the principal

states and retrieve

necessary property data.

❑

calculate compressor power,

refrigeration capacity, and

coefficient of performance.

✓

Skills Developed

TAKE NOTE...

The isentropic compressor

efficiency is introduced in

Sec. 6.12.3. See Eq. 6.48.

Determine the rate of heat transfer from the refrigerant pass-

ing through the condenser to the surroundings, in kW. Ans. 13.83 kW.

10.2 Analyzing Vapor-Compression Refrigeration Systems 597

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598 Chapter 10 Refrigeration and Heat Pump Systems

Additional departures from ideality stem from frictional effects that result in pressure

drops as the refrigerant flows through the evaporator, condenser, and piping connect-

ing the various components. These pressure drops are not shown on the T–s diagram

of Fig. 10.5 and are ignored in subsequent discussions for simplicity.

Finally, two additional features exhibited by actual vapor-compression systems are

shown in Fig. 10.5. One is the superheated vapor condition at the evaporator exit

(state 1), which differs from the saturated vapor condition shown in Fig. 10.4. Another

is the subcooling of the condenser exit state (state 3), which differs from the saturated

liquid condition shown in Fig. 10.4.

Example 10.3 illustrates the effects of irreversible compression and condenser exit

subcooling on the performance of the vapor-compression refrigeration system.

Analyzing an Actual Vapor-Compression Refrigeration Cycle

c c c c EXAMPLE 10.3 c

Reconsider the vapor-compression refrigeration cycle of Example 10.2, but include in the analysis that the com-

pressor has an isentropic efficiency of 80%. Also, let the temperature of the liquid leaving the condenser be 308C.

Determine for the modified cycle (a) the compressor power, in kW, (b) the refrigeration capacity, in tons, (c) the

coefficient of performance, and (d) the rates of exergy destruction within the compressor and expansion valve,

in kW, for T

5 299 K (268C).

SOLUTION

Known:

A vapor-compression refrigeration cycle has an isentropic compressor efficiency of 80%.

Find: Determine the compressor power, in kW, the refrigeration capacity, in tons, the coefficient of performance,

and the rates of exergy destruction within the compressor and expansion valve, in kW.

Schematic and Given Data:

Analysis:

Let us begin by fixing the principal states. State 1 is the same as in Example 10.2, so h

5 241.35 kJ/kg

and s

5 0.9253 kJ/kg

Owing to the presence of irreversibilities during the adiabatic compression process, there is an increase in

specific entropy from compressor inlet to exit. The state at the compressor exit, state 2, can be fixed using the

isentropic compressor efficiency

2 h

Engineering Model:

Each component of the cycle is analyzed as a control

volume at steady state.

2. There are no pressure drops through the evaporator and

condenser.

3. The compressor operates adiabatically with an isentropic

efficiency of 80%. The expansion through the valve is a

throttling process.

4. Kinetic and potential energy effects are negligible.

5. Saturated vapor at 2108C enters the compressor, and

liquid at 308C leaves the condenser.

6. The environment temperature for calculating exergy is

5 299 K (268C).

Fig. E10.3

30°C

–10°

= 26°C = 299 K

= 9 bar

VCRC

A.30 – Tabs c & d

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where h

is the specific enthalpy at state 2s, as indicated on the accompanying T–s diagram. From the solution

to Example 10.2, h

5 272.39 kJ/kg. Solving for h

and inserting known values

2 h

1 h

272.39 2 241.35

0.80

1 241.35 5 280.15 kJ

State 2 is fixed by the value of specific enthalpy h

and the pressure, p

5 9 bar. Interpolating in Table A-12, the

specific entropy is s

5 0.9497 kJ/kg

The state at the condenser exit, state 3, is in the liquid region. The specific enthalpy is approximated using

Eq. 3.14, together with saturated liquid data at 308C, as follows: h

< h

5 91.49 kJ

. Similarly, with Eq. 6.5,

< s

5 0.3396 kJ

? K.

The expansion through the valve is a throttling process; thus, h

5 h

. The quality and specific entropy at state

4 are, respectively

2 h

91.49 2 36.97

204.39

5 0.2667

and

5 s

1 x

2 s

5 0.1486 1

0.2667

0.9253 2 0.1486

5 0.3557 kJ

kg ?

(a) The compressor power is

5 m

2 h

5 10.08 kg

s21280.15 2 241.352 kJ

1 kW

1kJ

5 3.1 kW

(b) The refrigeration capacity is

5 m

2 h

5 10.08 kg

60 s

min

1241.35 2 91.492 kJ

1 ton

211 k

min

ton

b 5

2 h

241.35 2 91.49

280.15 2 241.35

5 3.86

(d) The rates of exergy destruction in the compressor and expansion valve can be

found by reducing the exergy rate balance or using the relationship E

5 T

where s

is the rate of entropy production from an entropy rate balance. With

either approach, the rates or exergy destruction for the compressor and valve are,

respectively

5 m

2 s



and





valve

5 m

2 s

Substituting values

5 a0.08

1299 K210.9497 2 0.92532

? K

1 kW

1 kJ

5 0.58 kW

and

valve

0.08

299

0.3557 2 0.3396

5 0.39 kW

➊ While the refrigeration capacity is greater than in Example 10.2, irrevers-

ibilities in the compressor result in an increase in compressor power com-

pared to isentropic compression. The overall effect is a lower coefficient of

performance than in Example 10.2.

Ability to…

❑

sketch the T– s diagram of

the vapor compression

refrigeration cycle with

irreversibilities in the

compressor and subcooled

liquid exiting the condenser.

❑

fix each of the principal

states and retrieve neces-

sary property data.

❑

calculate compressor power,

refrigeration capacity, and

coefficient of performance.

❑

calculate exergy destruc-

tion in the compressor and

expansion valve.

✓

Skills Developed

➊

➋

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600 Chapter 10 Refrigeration and Heat Pump Systems

10.3 Selecting Refrigerants

Refrigerant selection for a wide range of refrigeration and air-conditioning applica-

tions is generally based on three factors: performance, safety, and environmental

impact. The term performance refers to providing the required cooling or heating

capacity reliably and cost effectively. Safety refers to avoiding hazards

such as toxicity and flammability. Finally, environmental impact pri-

marily refers to using refrigerants that do not harm the stratospheric

ozone layer or contribute significantly to global climate change. We

begin by considering some performance aspects.

The temperatures of the refrigerant in the evaporator and con-

denser of vapor-compression cycles are governed by the temperatures

of the cold and warm regions, respectively, with which the system

interacts thermally. This, in turn, determines the operating pressures

in the evaporator and condenser. Consequently, the selection of a

refrigerant is based partly on the suitability of its pressure–tempera-

ture relationship in the range of the particular application. It is gen-

erally desirable to avoid excessively low pressures in the evaporator

and excessively high pressures in the condenser. Other considerations

in refrigerant selection include chemical stability, corrosiveness, and

cost. The type of compressor also affects the choice of refrigerant.

Centrifugal compressors are best suited for low evaporator pressures and refriger-

ants with large specific volumes at low pressure. Reciprocating compressors perform

better over large pressure ranges and are better able to handle low specific volume

refrigerants.

Refrigerant Types and Characteristics

Prior to the 1930s, accidents were prevalent among those who worked closely with

refrigerants due to the toxicity and flammability of most refrigerants at the time.

Because of such hazards, two classes of synthetic refrigerants were developed, each

containing chlorine and possessing highly stable molecular structures: CFCs (chloro-

fluorocarbons) and HCFCs (hydrochlorofluorocarbons). These refrigerants were

widely known as “freons,” the common trade name.

In the early 1930s, CFC production began with R-11, R-12, R-113, and R-114. In

1936, the first HCFC refrigerant, R-22, was introduced. Over the next several decades,

p–h diagram

10.2.4

The p–h Diagram

A thermodynamic property diagram widely used in the refrigeration field is the pressure–

enthalpy or p–h diagram. Figure 10.6 shows the main features of such a property diagram.

The principal states of the vapor-compression cycles of Fig. 10.5 are located on this

p–h diagram. It is left as an exercise to sketch the cycles of Examples 10.1, 10.2, and

10.3 on p–h diagrams. Property tables and p–h diagrams for many refrigerants are

given in handbooks dealing with refrigeration.

Fig. 10.6 Principal features of the pressure–

enthalpy diagram for a typical refrigerant, with

vapor-compression cycles superimposed.

Condenser

pressure

Evaporator

pressure

Constant s

Constant T

➋ The exergy destruction rates calculated in part (d) measure the effect of irreversibilities as the refrigerant

flows through the compressor and valve. The percentages of the power input (exergy input) to the compres-

sor destroyed in the compressor and valve are 18.7 and 12.6%, respectively.

What would be the coefficient of performance if the isentropic

compressor efficiency were 100%? Ans. 4.83.

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nearly all of the synthetic refrigerants used in the U.S. were either CFCs or HCFCs,

with R-12 being most commonly used.

To keep order with so many new refrigerants having complicated names, the “R”

numbering system was established in 1956 by DuPont and persists today as the industry

standard. Table 10.1 lists information, including refrigerant number, chemical composi-

tion, and global warming potential for selected refrigerants.

Environmental Considerations

After decades of use, compelling scientific data indicating that release of chlorine-

containing refrigerants into the atmosphere is harmful became widely recognized.

Concerns focused on released refrigerants depleting the stratospheric ozone layer and

contributing to global climate change. Because of the molecular stability of the CFC

and HCFC molecules, their adverse effects are long-lasting.

In 1987, an international agreement was adopted to ban production of certain chlorine-

containing refrigerants. In response, a new class of chlorine-free refrigerants was developed:

the HFCs (hydrofluorocarbons). One of these, R-134a, has been used for over 20 years as

the primary replacement for R-12. Although R-134a and other HFC refrigerants do not

contribute to atmospheric ozone depletion, they do contribute to global climate change.

Owing to a relatively high Global Warming Potential of about 1430 for R-134a, we may

soon see reductions in its use in the United States despite widespread deployment in refrig-

eration and air-conditioning systems, including automotive air conditioning. Carbon dioxide

(R-744) and R-1234yf are potential replacements for R-134a in automotive systems. See

Sec. 10.7.3 for discussion of CO

-charged automotive air-conditioning systems.

Another refrigerant that has been used extensively in air-conditioning and refrigeration

systems for decades, R-22, is being phased out under a 1995 amendment to the international

agreement on refrigerants because of its chlorine content. Effective in 2010, R-22 cannot

be installed in new systems. However, recovered and recycled R-22 can be used to service

existing systems until supplies are no longer available. As R-22 is phased out, replacement

refrigerants are being introduced, including R-410A and R-407C, both HFC blends.

Natural Refrigerants

Nonsynthetic, naturally occurring substances also can be used as refrigerants. Called

natural refrigerants, they include carbon dioxide, ammonia, and hydrocarbons. As

Refrigerant Data Including Global Warming Potential (GWP)

Refrigerant Number Type Chemical Formula Approx. GWP

R-12 CFC CCl

10900

R-11 CFC CCl

F 4750

R-114 CFC CClF

CClF

10000

R-113 CFC CCl

FCCIF

6130

R-22 HCFC CHClF

1810

R-134a HFC CH

FCF

1430

R-1234yf HFC CF

CF5CH

R-410A HFC blend R-32, R-125 1725

(50/50 Weight %)

R-407C HFC blend R-32, R-125, R-134a 1526

(23/25/52 Weight %)

R-744 (carbon dioxide) Natural CO

R-717 (ammonia) Natural NH

R-290 (propane) Natural C

R-50 (methane) Natural CH

R-600 (butane) Natural C

TABLE 10.1

The Global Warming Potential (GWP) depends on the time period over which the potential influence on global

warming is estimated. The values listed are based on a 100 year time period, which is an interval favored by

some regulators.

TAKE NOTE...

Global warming refers to an

increase in global average

temperature due to a combi-

nation of natural phenomena

and human industrial, agricul-

tural, and lifestyle activities.

The Global Warming Potential

(GWP) is a simplified index

that aims to estimate the

potential future influence on

global warming of different

gases when released to the

atmosphere. The GWP of a

gas refers to how much that

gas contributes to global

warming in comparison to

the same amount of carbon

dioxide. The GWP of carbon

dioxide is taken to be 1.

10.3 Selecting Refrigerants 601

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602 Chapter 10

Refrigeration and Heat Pump Systems

indicated by Table 10.1, natural refrigerants typically have low Global Warming

Potentials.

Ammonia (R-717), which was widely used in the early development of vapor-

compression refrigeration, continues to serve today as a refrigerant for large systems

used by the food industry and in other industrial applications. In the past two decades,

ammonia has been increasingly used because of the R-12 phase out and is receiving

even greater interest today due to the R-22 phase out. Ammonia is also used in the

absorption systems discussed in Sec. 10.5.

Hydrocarbons, such as propane (R-290), are used worldwide in various refrigera-

tion and air-conditioning applications including commercial and household appli-

ances. In the United States, safety concerns limit propane use to niche markets like

industrial process refrigeration. Other hydrocarbons—methane (R-50) and butane

(R-600)—are also under consideration for use as refrigerants.

Refrigeration with No Refrigerant Needed

Alternative cooling technologies aim to achieve a refrigerating effect without use of

refrigerants, thereby avoiding adverse effects associated with release of refrigerants

to the atmosphere. One such technology is thermoelectric cooling. See the box.

New Materials May Improve Thermoelectric Cooling

You can buy a thermoelectric cooler powered from the cigarette lighter outlet of your

car. The same technology is used in space flight applications and in power amplifiers

and microprocessors.

Figure 10.7 shows a thermoelectric cooler separating a cold region at temperature T

and a warm region at temperature T

. The cooler is formed from two n-type and two

p-type semiconductors with low thermal conductivity, five metallic interconnects with high

electrical conductivity and high thermal conductivity, two electrically insulating ceramic

substrates, and a power source. When power is provided by the source, current flows

through the resulting electric circuit, giving a refrigeration effect: a heat transfer of energy

from the cold region. This is known as the Peltier effect.

The p-type semiconductor material in the right leg of the cooler shown in Fig. 10.7 has

electron vacancies, called holes. Electrons move through this material by filling individual

holes, slowing electron motion. In the adjacent n-type semiconductor, no holes exist in its

Fig. 10.7 Schematic of a thermoelectric cooler.

Negatively

charged

electron

Ceramic

substrate

Ceramic

substrate

Positively

charged

hole

Metallic

interconnect

–

+–

–

+ +

out

Cold Region at T

Warm Region at T

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