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

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9.124 An ideal gas mixture with k 5 1.31 and a molecular

weight of 23 is supplied to a converging nozzle at p

5 5 bar,

5 700 K, which discharges into a region where the

pressure is 1 bar. The exit area is 30 cm

. For steady isentropic

flow through the nozzle, determine

(a) the exit temperature of the gas, in K.

(b) the exit velocity of the gas, in m/s.

9.125 An ideal gas expands isentropically through a converging

nozzle from a large tank at 120 lbf/in.

, 6008R, and discharges

into a region at 60 lbf/in.

Determine the mass flow rate, in

lb/s, for an exit flow area of 1 in.

, if the gas is

(a) air, with k 5 1.4.

(b) carbon dioxide, with k 5 1.26.

9.126 Air at p

5 1.4 bar, T

5 280 K expands isentropically

through a converging nozzle and discharges to the atmosphere

at 1 bar. The exit plane area is 0.0013 m

(a) Determine the mass flow rate, in kg/s.

(b) If the supply region pressure, p

, were increased to 2 bar,

what would be the mass flow rate, in kg/s?

9.127 Air enters a nozzle operating at steady state at 45 lbf/in.

8008R, with a velocity of 480 ft/s, and expands isentropically

to an exit velocity of 1500 ft/s. Determine

(a) the exit pressure, in lbf/in.

(b) the ratio of the exit area to the inlet area.

converging–diverging in cross section.

9.128 Air as an ideal gas with k

5 1.4 enters a converging–

diverging nozzle operating at steady state and expands

isentropically as shown in Fig. P9.128. Using data from the

figure and from Table 9.2 as needed, determine

(a) the stagnation pressure, in lbf/in.

, and the stagnation

temperature, in 8R.

(b) the throat area, in in.

the nozzle is choked, and the diverging portion acts as a

supersonic nozzle, determine the mass flow rate, in kg/s, and

the Mach number, pressure, in bar, and temperature, in K, at

the exit. Repeat if the diverging portion acts as a supersonic

diffuser.

9.131 A converging–diverging nozzle operates at steady state.

Air as an ideal gas with k

5 1.4 enters the nozzle at 500 K,

6 bar, and a Mach number of 0.3. The air flows isentropically

to the exit plane, where a normal shock stands. The

temperature just upstream of the shock is 380.416 K.

Determine the back pressure, in bar.

9.132 A converging–diverging nozzle operates at steady state.

Air as an ideal gas with k

5 1.4 enters the nozzle at 500 K,

6 bar, and a Mach number of 0.3. A normal shock stands in

the diverging section at a location where the Mach number

1.40. The cross-sectional areas of the throat and the exit

plane are 4 cm

and 6 cm

, respectively. The flow is isentropic,

except where the shock stands. Determine the exit pressure,

in bar, and the mass flow rate, in kg/s.

9.133 Air as an ideal gas with k

5 1.4 enters a converging-

diverging duct with a Mach number of 2. At the inlet, the

pressure is 26 lbf/in.

and the temperature is 4458R. A normal

shock stands at a location in the converging section of the

duct, with M

5 1.5. At the exit of the duct, the pressure is

150 lbf/in.

The flow is isentropic everywhere except in the

immediate vicinity of the shock. Determine temperature, in

8R, and the Mach number at the exit.

9.134 Air as an ideal gas with k 5 1.4 undergoes a normal

shock. The upstream conditions are p

5 0.5 bar, T

5 280 K,

and M

5 1.8. Determine

(a) the pressure p

, in bar.

(b) the stagnation pressure p

, in bar.

, in K.

(d) the change in specific entropy across the shock, in

kJ/kg ? K.

(e) Plot the quantities of parts (a)–(d) versus M

ranging from

1.0 to 2.0. All other upstream conditions remain the same.

9.135 A converging–diverging nozzle operates at steady state.

Air as an ideal gas with k

5 1.4 flows through the nozzle,

discharging to the atmosphere at 14.7 lbf/in.

and 5208R. A

normal shock stands at the exit plane with M

5 1.5. The

exit plane area is 1.8 in.

Upstream of the shock, the flow is

isentropic. Determine

(a) the stagnation pressure p

, in lbf/in.

(b) the stagnation temperature T

, in 8R.

9.136 A converging–diverging nozzle operates at steady state.

Air as an ideal gas with k

5 1.4 flows through the nozzle,

discharging to the atmosphere at 14.7 lbf/in.

and 5108R. A

normal shock stands at the exit plane with p

5 9.714 lbf/in.

The exit plane area is 2 in.

Upstream of the shock, the flow

is isentropic. Determine

(a) the throat area, in in.

(b) the entropy produced in the nozzle, in Btu/8R per lb of

air flowing.

Problems: Developing Engineering Skills 583

Fig. P9.128

= 0.2

= 595°R

= 77.8 lbf/in.

= 1 lb/s

= 1.5

Throat

9.129 Air as an ideal gas with k 5 1.4 enters a diffuser operating

at steady state at 4 bar, 290 K, with a velocity of 512 m/s.

Assuming isentropic flow, plot the velocity, in m/s, the Mach

number, and the area ratio A/A* for locations in the flow

corresponding to pressures ranging from 4 to 14 bar.

9.130 A converging–diverging nozzle operating at steady state

has a throat area of 3 cm

and an exit area of 6 cm

. Air as an

ideal gas with k 5 1.4 enters the nozzle at 8 bar, 400 K, and a

Mach number of 0.2, and flows isentropically throughout. If

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584 Chapter 9 Gas Power Systems

9.137 Air at 3.4 bar, 530 K, and a Mach number of 0.4 enters

a converging–diverging nozzle operating at steady state. A

normal shock stands in the diverging section at a location

where the Mach number is M

5 1.8. The flow is isentropic,

except where the shock stands. If the air behaves as an ideal

gas with k

5 1.4, determine

(a) the stagnation temperature T

, in K.

(b) the stagnation pressure p

, in bar.

(d) the pressure p

, in bar.

(e) the stagnation pressure p

, in bar.

(f) the stagnation temperature T

, in K.

If the throat area is 7.6 3 10

, and the exit plane pressure

is 2.4 bar, determine the mass flow rate, in kg/s, and the exit

area, in m

9.138 Air as an ideal gas with k

5 1.4 enters a converging–

diverging channel at a Mach number of 1.6. A normal shock

stands at the inlet to the channel. Downstream of the shock

the flow is isentropic; the Mach number is unity at the throat;

and the air exits at 20 lbf/in.

, 7008R, with negligible velocity.

If the mass flow rate is 45 lb/s, determine the inlet and throat

areas, in ft

9.139 Derive the following expressions: (a) Eq. 9.55, (b) Eq.

9.56, (c) Eq. 9.57.

9.140 Using Interactive Thermodynamics: IT, generate tables

of the same isentropic flow functions as in Table 9.2 for

specific heat ratios of 1.2, 1.3, 1.4, and 1.67 and Mach numbers

ranging from 0 to 5.

9.141 Using Interactive Thermodynamics: IT, generate tables

of the same normal shock functions as in Table 9.3 for

specific heat ratios of 1.2, 1.3, 1.4, and 1.67 and Mach numbers

ranging from 1 to 5.

c DESIGN & OPEN-ENDED PROBLEMS: EXPLORING ENGINEERING PRACTICE

9.1D Congress has mandated the average fuel economy for

passenger cars sold in the United States to be 35 miles per

gallon beginning in 2020. In Europe, the goal is 47 miles per

gallon by 2012. In each case, identify the major factors

spurring legislative action, including, as appropriate, technical,

economic, societal, and political factors. Analyze the disparity

between these goals. Comment on their likely effectiveness

in achieving the respective legislative aims. Report your

findings in a PowerPoint presentation.

9.2D Automotive gas turbines have been under development

for decades but have not been commonly used in automobiles.

Yet helicopters routinely use gas turbines. Explore why

different types of engines are used in these respective

applications. Compare selection factors such as performance,

power-to-weight ratio, space requirements, fuel availability,

and environmental impact. Summarize your findings in a

report with at least three references.

9.3D The Annual Energy Outlook with Projections

report

released by the U.S. Energy Information Administration

projects annual consumption estimates for various fuel types

through the next 25 years. According to the report, biofuels

will play an increasing role in the liquid fuel supply over that

time period. Based on technologies commercially available

or reasonably expected to become available in the next

decade, identify the most viable options for producing

biofuels. Compare several options based on energy return on

energy invested (EROEI), water and land requirements, and

effects on global climate change. Draw conclusions based on

your study, and present your findings in a report with at least

three references.

9.4D Investigate the following technologies: plug-in hybrid

vehicles, all-electric vehicles, hydrogen fuel cell vehicles,

diesel-powered vehicles, natural gas–fueled vehicles, and

ethanol-fueled vehicles, and make recommendations on

which of these technologies should receive federal research,

development, and deployment support over the next decade.

Base your recommendation on the result of a decision matrix

method such as the Pugh method to compare the various

technologies. Clearly identify and justify the criteria used for

the comparison and the logic behind the scoring process.

Prepare a 15-minute briefing and an executive summary

suitable for a conference with your local congressperson.

9.5D Figure P9.5D shows a wheeled platform propelled by

thrust generated using an onboard water tank discharging

water through an elbow to which a nozzle is attached. Design

and construct such an apparatus using easily obtained materials

like a skateboard and one-gallon milk jug. Investigate the

effects of elbow angle and nozzle exit area on volumetric flow

and thrust. Prepare a report including results and conclusions

together with an explanation of the measurement techniques

and experimental procedures.

Fig. P9.5D

Nozzle

Water tank stand

Elbow (tubing)

Opening for atmosphere exposure

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9.6D Owing to its very low temperature relative to seawater,

liquid natural gas (LNG) arriving at U.S. ports by ship has

considerable thermomechanical exergy. Yet when LNG is

regasified in heat exchangers where seawater is the other

stream, that exergy is largely destroyed. Conduct a search of

the patent literature for methods to recover a substantial

portion of LNG exergy during regasification. Consider

patents both granted and pending. Critically evaluate the

technical merit and economic feasibility of two different

methods found in your search. Report your conclusions in

an executive summary and PowerPoint presentation.

9.7D Develop the preliminary specifications for a 160-MW

closed-cycle gas turbine power plant. Consider carbon

dioxide and helium as possible gas turbine working fluids

that circulate through a nuclear power unit where they

absorb energy by heat transfer. Sketch the schematic of

your proposed cycle. For each of the two working fluids,

determine key operating pressures and temperatures of the

gas turbine cycle and estimate the expected performance

of the cycle. Write a report that includes your analysis and

design, and recommend a working fluid. Include at least

three references.

9.8D As an engineer you must recommend whether or not to

purchase 2 MW of electric power from the local utility

company at a cost of $0.06 per kW ? h for a plant expansion

or to purchase and operate a diesel engine generator set that

runs on natural gas. Assume natural gas is priced at 6 dollars

per thousand cubic feet, each cubic foot of gas has a heating

value

of 1000 Btu, and the plant operates 7800 h per year.

Specify diesel-generator equipment that would supply the

required power and perform an economic analysis to determine

which alternative is best. Present your recommendation in a

memorandum with supporting analysis and calculations,

including at least three references.

9.9D A financial services company has a computer server

facility that requires very reliable electric power. The power

demand is 3000 kW. The company has hired you as a

consultant to study the feasibility of using 250-kW

microturbines

for this application. Write a report discussing

the pros and cons of such an arrangement compared to

purchasing power from the local utility.

9.10D Figure P9.10D shows a combined cycle formed by a

topping gas turbine and an organic Rankine bottoming

cycle. Steady-state operating data are labeled on the figure.

Owing to internal irreversibilities, the generator electricity

output is 95% of the input shaft power. The regenerator

preheats air entering the combustor. In the evaporator, hot

exhaust gas from the regenerator vaporizes the bottoming

Design & Open-Ended Problems: Exploring Engineering Practice 585

Fig. P9.10D

1 2

Saturated liquid

Compressor

Gas turbine cycle

Organic Rankine cycle

Turbine

GTC

Regenerator

Evaporator

= 100 kW

␩

gen

= 95%

␩

= 85%

␩

reg

= 80%

␩

= 85% ␩

= 85%

= 300 K

= 100 kPa

= 1200 K

= T

+ 20 K

= p

, T

Combustor

Generator

Turbine

ORC

␩

gen

= 95%

␩

= 85%

Generator

Condenser

out

Pump

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586 Chapter 9 Gas Power Systems

cycle working fluid. For each of three such working

fluids—propane, Refrigerant 22, and Refrigerant 134a—

specify appropriate ranges for p

, turbine inlet pressure,

and T

, turbine inlet temperature; also determine turbine

exit pressure p

. For each working fluid, investigate the

influence on net combined-cycle electricity production and

on combined-cycle thermal efficiency of varying p

, T

, and

compressor pressure ratio. Identify the bottoming working

fluid and operating conditions with greatest net combined-

cycle electricity production. Repeat for greatest combined-

cycle thermal efficiency. Apply engineering modeling

compatible with that used in the text for Rankine cycles

and air-standard analysis of gas turbines. Present your

analyses, results, and recommendations in a technical

article adhering to ASME standards with at least three

references.

9.11D Micro-CHP (combined heat and power) units capable of

producing up to 1.8 kW of electric power are now commercially

available for use in the home. Such units contribute to

domestic space or water heating needs while providing

electricity as a by-product. They operate on an internal

combustion engine fueled by natural gas. By hybridizing a

micro-CHP unit with a gas furnace, all

domestic heating needs

can be met while generating a substantial portion of the

annual electric power requirement. Evaluate this hybrid form

for application to a typical single-family dwelling in your

locale having natural gas service. Consider on-site battery

High-temperature

superheater

Engine exhaust

Low-temperature superheater,

High-temperature vapor generator

High-temperature

condenser,

Low-temperature

vapor generator

Radiator,

Low-temperature

condenser

Low-temperature

vapor generator

1-2-3-4-1:

5-6-7-8-5:

a-b-c:

d-e:

High-temperature cycle

Low-temperature cycle

Engine coolant

ine exhaust

Low-temperature

expander

High-temperature

expander

Pump

Four-cylinder

engine

Exhaust

gas

Coolant

Fuel

Air

Pump

Fig. P9.12D

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Design & Open-Ended Problems: Exploring Engineering Practice 587

storage of excess electricity generated and the possibility of net

metering. Specify equipment and determine costs, including

initial cost and installation cost. Estimate the annual cost for

heating and power using the hybrid and compare it to the

annual cost for heating and power with a stand-alone gas

furnace and grid power. Using your findings, recommend the

better approach for the dwelling.

9.12D Figure P9.12D provides the schematic of an internal

combustion automobile engine fitted with two Rankine

vapor power bottoming cycles: a high-temperature cycle

1-2-3-4-1 and a low-temperature cycle 5-6-7-8-5. These

cycles develop additional power using waste heat derived

from exhaust gas and engine coolant. Using operating

data for a commercially available car having a conventional

four-cylinder internal combustion engine with a size of 2.5

liters or less, specify cycle working fluids and state data

at key points sufficient to produce at least 15 hp more

power. Apply engineering modeling compatible with that

used in the text for Rankine cycles and air-standard

analysis of internal combustion engines. Write a final

report justifying your specifications together with

supporting calculations. Provide a critique of the use of

such bottoming cycles on car engines and a recommendation

about whether such technology should be actively pursued

by automakers.

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588

Refrigeration systems commonly used for preserving food are introduced in Sec. 10.1.

ENGINEERING CONTEXT Refrigeration systems for food preservation and air conditioning play

prominent roles in our everyday lives. Heat pumps also are used for heating buildings and for producing indus-

trial process heat. There are many other examples of commercial and industrial uses of refrigeration, including

air separation to obtain liquid oxygen and liquid nitrogen, liquefaction of natural gas, and production of ice.

To achieve refrigeration by most conventional means requires an electric power input. Heat pumps also

require power to operate. Referring again to Table 8.1, we see that in the United States electricity is obtained

today primarily from coal, natural gas, and nuclear, all of which are nonrenewable. These nonrenewables

have significant adverse effects on human health and the environment associated with their use. Depending

on the type of resource, such effects are related to extraction from the earth, processing and distribution,

emissions during power production, and waste products.

Ineffective refrigeration and heat pump systems, excessive building cooling and heating, and other wasteful

practices and lifestyle choices not only misuse increasingly scarce nonrenewable resources but also endanger

our health and burden the environment. Accordingly, refrigeration and heat pump systems is an area of appli-

cation where more effective systems and practices can significantly improve our national energy posture.

The objective of this chapter is to describe some of the common types of refrigeration and heat pump systems

presently in use and to illustrate how such systems can be modeled thermodynamically. The three principal

types described are the vapor-compression, absorption, and reversed Brayton cycles. As for the power sys-

tems studied in Chaps. 8 and 9, both vapor and gas systems are considered. In vapor systems, the refrigerant

is alternately vaporized and condensed. In gas refrigeration systems, the refrigerant remains a gas.

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Refrigeration and

Heat Pump Systems

When you complete your study of this chapter, you will be able to...

Demonstrate understanding of basic vapor-compression refrigeration and heat pump

systems.

Develop and analyze thermodynamic models of vapor-compression systems and their

modifications, including

sketching schematic and accompanying T–s diagrams.

evaluating property data at principal states of the systems.

applying mass, energy, entropy, and exergy balances for the basic processes.

determining refrigeration and heat pump system performance, coefficient of performance

and capacity.

Explain the effects on vapor-compression system performance of varying key parameters.

Demonstrate understanding of the operating principles of absorption and gas refrigeration

systems, and perform thermodynamic analysis of gas systems.

LEARNING OUTCOMES

589

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590 Chapter 10

Refrigeration and Heat Pump Systems

10.1 Vapor Refrigeration Systems

The purpose of a refrigeration system is to maintain a cold region at a temperature

below the temperature of its surroundings. This is commonly achieved using the vapor

refrigeration systems that are the subject of the present section.

10.1.1

Carnot Refrigeration Cycle

To introduce some important aspects of vapor refrigeration, let us begin by considering

a Carnot vapor refrigeration cycle. This cycle is obtained by reversing the Carnot vapor

power cycle introduced in Sec. 5.10. Figure 10.1 shows the schematic and accompanying

T–s diagram of a Carnot refrigeration cycle operating between a region at temperature

and another region at a higher temperature T

. The cycle is executed by a refriger-

ant circulating steadily through a series of components. All processes are internally

reversible. Also, since heat transfers between the refrigerant and each region occur with

no temperature differences, there are no external irreversibilities. The energy transfers

shown on the diagram are positive in the directions indicated by the arrows.

Let us follow the refrigerant as it passes steadily through each of the components in

the cycle, beginning at the inlet to the evaporator. The refrigerant enters the evaporator

as a two-phase liquid–vapor mixture at state 4. In the evaporator some of the refrigerant

changes phase from liquid to vapor as a result of heat transfer from the region at tem-

perature T

to the refrigerant. The temperature and pressure of the refrigerant remain

constant during the process from state 4 to state 1. The refrigerant is then compressed

adiabatically from state 1, where it is a two-phase liquid–vapor mixture, to state 2,

where it is a saturated vapor. During this process, the temperature of the refrigerant

increases from T

to T

, and the pressure also increases. The refrigerant passes from

the compressor into the condenser, where it changes phase from saturated vapor to

saturated liquid as a result of heat transfer to the region at temperature T

. The

temperature and pressure remain constant in the process from state 2 to state 3. The

refrigerant returns to the state at the inlet of the evaporator by expanding adiabati-

cally through a turbine. In this process, from state 3 to state 4, the temperature

decreases from T

to T

, and there is a decrease in pressure.

TAKE NOTE...

See Sec. 6.13.1 for the area

interpretation of heat trans-

fer on a T–s diagram for the

case of internally reversible

flow though a control volume

at steady state.

Fig. 10.1 Carnot vapor refrigeration cycle.

Condenser

Evaporator

Compressor

out

4 1

Turbine

Cold region at T

Warm region at T

Refrig_Cycle

A.10 – All Tabs

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Since the Carnot vapor refrigeration cycle is made up of internally reversible pro-

cesses, areas on the T–s diagram can be interpreted as heat transfers. Area 1–a–b–4–1

is the heat added to the refrigerant from the cold region per unit mass of refrigerant

flowing. Area 2–a–b–3–2 is the heat rejected from the refrigerant to the warm region

per unit mass of refrigerant flowing. The enclosed area 1–2–3–4–1 is the net heat

transfer from the refrigerant. The net heat transfer from the refrigerant equals the net

work done on the refrigerant. The net work is the difference between the compressor

work input and the turbine work output.

The coefficient of performance b of any refrigeration cycle is the ratio of the refrig-

eration effect to the net work input required to achieve that effect. For the Carnot

vapor refrigeration cycle shown in Fig. 10.1, the coefficient of performance is

max

2 W

area 1–a–b–4–

area 1–2–3–4–1

2 s

2 T

21s

2 s

which reduces to

max

2 T

(10.1)

This equation, which corresponds to Eq. 5.10, represents the maximum theoretical

coefficient of performance of any refrigeration cycle operating between regions at T

and T

10.1.2

Departures from the Carnot Cycle

Actual vapor refrigeration systems depart significantly from the Carnot cycle con-

sidered above and have coefficients of performance lower than would be calcu-

lated from Eq. 10.1. Three ways actual systems depart from the Carnot cycle are

considered next.

c One of the most significant departures is related to the heat trans-

fers between the refrigerant and the two regions. In actual systems,

these heat transfers are not accomplished reversibly as presumed

above. In particular, to achieve a rate of heat transfer sufficient to

maintain the temperature of the cold region at T

with a practi-

cal-sized evaporator requires the temperature of the refrigerant in

the evaporator, T9

, to be several degrees below T

. This is illus-

trated by the placement of the temperature T9

on the T–s diagram

of Fig. 10.2. Similarly, to obtain a sufficient heat transfer rate from

the refrigerant to the warm region requires that the refrigerant

temperature in the condenser, T9

, be several degrees above T

This is illustrated by the placement of the temperature T9

on the

T–s diagram of Fig. 10.2.

Maintaining the refrigerant temperatures in the heat exchangers at T9

and T9

rather than at T

and T

, respectively, has the effect of reducing the coefficient of

performance. This can be seen by expressing the coefficient of performance of the

refrigeration cycle designated by 19–29–39–49–19 on Fig. 10.2 as

b¿ 5

area 1¿–a–b–4¿–1

area 1¿–2¿–3¿–4¿–1¿

T¿

2 T¿

(10.2)

Comparing the areas underlying the expressions for b

max

and b9 given above, we con-

clude that the value of b9 is less than b

max

. This conclusion about the effect of refriger-

ant temperature on the coefficient of performance also applies to other refrigeration

cycles considered in the chapter.

Temperature of warm

region, T

Temperature of cold

region, T

Evaporator

temperature, T

Condenser

temperature, T

′

4′ 1′

3′ 2

′

Fig. 10.2 Comparison of the condenser and

evaporator temperatures with those of the warm

and cold regions.

10.1 Vapor Refrigeration Systems 591

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592 Chapter 10

Refrigeration and Heat Pump Systems

c Even when the temperature differences between the refrigerant and

warm and cold regions are taken into consideration, there are other

features that make the vapor refrigeration cycle of Fig. 10.2 impractical

as a prototype. Referring again to the figure, note that the compres-

sion process from state 19 to state 29 occurs with the refrigerant as a

two-phase liquid–vapor mixture. This is commonly referred to as wet

compression. Wet compression is normally avoided because the pres-

ence of liquid droplets in the flowing liquid–vapor mixture can dam-

age the compressor. In actual systems, the compressor handles vapor

only. This is known as dry compression.

c Another feature that makes the cycle of Fig. 10.2 impractical is the expan-

sion process from the saturated liquid state 39 to the low-quality, two-

phase liquid–vapor mixture state 49. This expansion typically produces a

relatively small amount of work compared to the work input in the com-

pression process. The work developed by an actual turbine would be

smaller yet because turbines operating under these conditions have low

isentropic efficiencies. Accordingly, the work output of the turbine is nor-

mally sacrificed by substituting a simple throttling valve for the expansion

turbine, with consequent savings in initial and maintenance costs. The

10.2 Analyzing Vapor-Compression

Refrigeration Systems

Vapor-compression refrigeration systems are the most common refrigeration systems in

use today. The object of this section is to introduce some important features of sys-

tems of this type and to illustrate how they are modeled thermodynamically.

10.2.1

Evaluating Principal Work and Heat Transfers

Let us consider the steady-state operation of the vapor-compression system illustrated

in Fig. 10.3. Shown on the figure are the principal work and heat transfers, which are

positive in the directions of the arrows. Kinetic and potential energy changes are

neglected in the following analyses of the components. We begin with the evaporator,

where the desired refrigeration effect is achieved.

c As the refrigerant passes through the evaporator, heat transfer from the refrigerated

space results in the vaporization of the refrigerant. For a control volume enclosing

the refrigerant side of the evaporator, the mass and energy rate balances reduce to

give the rate of heat transfer per unit mass of refrigerant flowing as

5 h

2 h

(10.3)

where m

is the mass flow rate of the refrigerant. The heat transfer rate

is referred

to as the refrigeration capacity. In the SI unit system, the capacity is normally expressed

in kW. In the English unit system, the refrigeration capacity may be expressed in Btu/h.

Another commonly used unit for the refrigeration capacity is the ton of refrigeration,

which is equal to 200 Btu/min or about 211 kJ/min.

c The refrigerant leaving the evaporator is compressed to a relatively high pressure

and temperature by the compressor. Assuming no heat transfer to or from the

vapor-compression

refrigeration

refrigeration capacity

ton of refrigeration

Condenser

Evaporator

Compressor

out

Saturated or

superheated vapor

Expansion

valve

Fig. 10.3 Components of a vapor-

compression refrigeration system.

components of the resulting cycle are illustrated in Fig. 10.3, where dry compression is

presumed. This cycle, known as the vapor-compression refrigeration cycle, is the subject

of the section to follow.

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