REFRIGERATION & AIR CONDITIONING · 2. Refrigeration and Air conditioning by Manohar Prasad,New Age international publishers. 3. Refrigeration and Air conditioning by C.P. Arora,

REFRIGERATION & AIR CONDITIONING(Credit: 3-0-0, Code: PCME 4402)(As per BPUT, Rourkela, Syllabus)

MODULE - I

PreparedBy

Prof. (Dr.) Manmatha K. RoulProfessor and Principal

Gandhi Institute for Technological Advancement (GITA),Bhubaneswar – 752054

June 2016

Prepared By : Prof. (Dr.) Manmatha K. Roul || 1

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REFRIGERATION & AIR CONDITIONING(Credit: 3-0-0)

Module I (14 hours)

1. Air Refrigeration System : Introduction, Unit of refrigeration, Coefficient of performance,Reversed Carnot Cycle, Temperature limitations, maximum COP, Bell Coleman air cycle,Simple Air Cycle System for Air-craft with problems.

2. Vapour Compression System : Analysis of theoretical vapour compression cycle,Representation of cycle on T - S and p - h diagram, Simple saturation cycle, sub-cooled cycleand super-heated cycle, Effect of suction and discharge pressure on performance, Actualvapour compression cycle. Problem illustration and solution.

3. Multi-stage compression and Multi-evaporator systems : Different arrangements ofcompressors and inter-cooling, Multistage compression with inter-cooling, Multi-evaporatorsystem, Dual compression system. Simple problems

Module II (13 hours)

4. Vapour Absorption System : Simple Ammonia - absorption system, Improved absorptionsystem, Analysis of vapour absorption system (Specifically of analyzing coloumn andrectifier), Electrolux / Three fluid system, Lithium-bromide-water vapour absorption system,comparison of absorption system with vapour compression system. Simple Problems andsolution.

5. Thermoelectric Refrigeration: Basics and Principle. Defining the figure of Merit. (No Problem)6. Refrigerants : Classification of refrigerants and its degignation- Halocarbon (compounds,

Hydrocarbons, Inorganic compounds, Azeotropes, Properties of refrigerants, comparison ofcommon refrigerants, uses of important refrigerants, Brines. Alternative refrigerants (Organicand inorganic compounds).

Module III (13 hours)

7. Psychrometrics : Properties of air-vapour mixture, Law of water vapour-air mixture,Enthalpy of moisture, Psychrometric chart, simple heating and cooling, Humidification, De-humidification, Mixture of air streams. Review question and discussionsRequirements of comfort air conditioning : Oxygen supply, Heat removal, moisture removal,air motion, purity of air, Thermodynamics of human body, comfort and comfort chart, effectivetemperature, factors governing optimum effective temperature

8. Air Conditioning System : Process in air conditioning : Summer air conditioning, Winter airconditioning and year round air conditioning, Cooling load calculations. Review question anddiscussions.

Text Books :

1. Refrigeration and Air Conditioning by R.C. Arora , PHI Publication2. Refrigeration and Air Conditioning by S.C. Arora and S. Domkundwar, Dhanpat Rai & Sons.

Chapters ; 3,4,5,6,7,11,16,17,19,203. Refrigeration and Airconditioning Data book by Manohar Prasad

Reference Books :1. Refrigeration and Air conditioning by P.L. Balloney, Khanna Publishers.


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2. Refrigeration and Air conditioning by Manohar Prasad,New Age international publishers.3. Refrigeration and Air conditioning by C.P. Arora, Tata McGraw Hill.

AIR REFRIGERATION SYSTEM

Air Refrigeration System : Introduction, Unit of refrigeration, Coefficient of performance,Reversed Carnot Cycle, Temperature limitations, maximum COP, Bell Coleman air cycle, SimpleAir Cycle System for Air-craft with problems.

1.1INTRODUCTION

Refrigeration may be defined as the process of achieving and maintaining atemperature below that of the surroundings, the aim being to cool some product or space tothe required temperature.

Refrigeration is a process of removing heat from a low-temperature reservoir andtransferring it to a high-temperature reservoir. The work of heat transfer is traditionally drivenby or other means.

One of the most important applications of refrigeration has been the preservation ofperishable food products by storing them at low temperatures. Refrigeration systems are alsoused extensively for providing thermal comfort to human beings by means of air conditioning.Air Conditioning refers to the treatment of air so as to simultaneously control its temperature,moisture content, cleanliness, odour and circulation, as required by occupants, a process, orproducts in the space. The subject of refrigeration and air conditioning has evolved out ofhuman need for food and comfort, and its history dates back to centuries. The history ofrefrigeration is very interesting since every aspect of it, the availability of refrigerants, theprime movers and the developments in compressors and themethods of refrigeration all area part of it.

Refrigeration systems are also used for providing cooling and dehumidification insummer for personal comfort (air conditioning). The first air conditioning systems were usedfor industrial as well as comfort air conditioning. Eastman Kodak installed the first airconditioning system in 1891 in Rochester, New York for the storage of photographic films. Anair conditioning system was installed in a printing press in 1902 and in a telephone exchangein Hamburg in 1904. Many systems were installed in tobacco and textile factories around1900. The first domestic air conditioning system was installed in a house in Frankfurt in 1894.A private library in St Louis, USA was air conditioned in 1895, and a casino was airconditioned in Monte Carlo in 1901. Efforts have also been made to air condition passengerrail coaches using ice. The widespread development of air conditioning is attributed to theAmerican scientist and industrialist Willis Carrier. Carrier studied the control of humidity in1902 and designed a central air conditioning plant using air washer in 1904. Due to thepioneering efforts of Carrier and also due to simultaneous development of differentcomponents and controls, air conditioning quickly became very popular, especially after


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1923. At present comfort air conditioning is widely used in residences, offices, commercialbuildings, air ports, hospitals and in mobile applications such as rail coaches, automobiles,

Air cycle refrigeration systems belong to the general class of gas cycle refrigeration systems,in which a gas is used as the working fluid. The gas does not undergo any phase changeduring the cycle, consequently, all the internal heat transfer processes are sensible heattransfer processes. Gas cycle refrigeration systems find applications in air craft cabin coolingand also in the liquefaction of various gases. In the present chapter gas cycle refrigerationsystems based on air are discussed.

1.2 UNIT OF REFRIGERATION

The practical unit of refrigeration is expressed in the terms of “ Tonne of refrigeration ”.

A “tonne of refrigeration” is defined as the amount of refrigeration effect produced by the

uniform melting of one US ton of ice from and at 0°C in 24 hours.

1 US ton = 2000 lb = 2000 x 0.453592 kg =907.1847 kg

The latent heat of ice is 335kJ/kg and therefore one tonne of refrigeration

1TR = 907.1847 X 335 KJ / 24 hours

= 907.1847 X 335 / 24 X 60

= 211 kJ/min

In actual practice, one tonne of refrigeration is taken as a equivalent to 210kJ/min or

3.5 kW

1.3 COEFFICIENT OF PERFORMANCE

The coefficient of performance or COP of a heat pump, refrigerator or air conditioningsystem is a ratio of useful heating or cooling provided to work required. Higher COPs equateto lower operating costs. The COP usually exceeds 1, especially in heat pumps, because,instead of just converting work to heat (which, if 100% efficient, would be a COP of hp of 1), itpumps additional heat from a heat source to where the heat is required. For completesystems, COP calculations should include energy consumption of all power consumingauxiliaries. COP is highly dependent on operating conditions, especially absolute temperatureand relative temperature between sink and system, and is often graphed or averaged againstexpected conditions.

The equation is: COP =

where


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Q is the useful heat supplied or removed by the considered system. W is the work required by the considered system.The COP for heating and cooling are thus different, because the heat reservoir of interestis different. When one is interested in how well a machine cools, the COP is the ratio ofthe heat removed from the cold reservoir to input work. However, for heating, the COP isthe ratio of the heat removed from the cold reservoir plus the input work to the input work:

1.4 AIR STANDARD CYCLE ANALYSIS

Air cycle refrigeration system analysis is considerably simplified if one makes the followingassumptions:

i. The working fluid is a fixed mass of air that behaves as an ideal gasii. The cycle is assumed to be a closed loop cycle with all inlet and exhaust processes

of open loop cycles being replaced by heat transfer processes to or from theenvironment

iii. All the processes within the cycle are reversible, i.e., the cycle is internallyreversible

iv. The specific heat of air remains constant throughout the cycle

An analysis with the above assumptions is called as cold Air Standard Cycle (ASC) analysis.This analysis yields reasonably accurate results for most of the cycles and processesencountered in air cycle refrigeration systems. However, the analysis fails when oneconsiders a cycle consisting of a throttling process, as the temperature drop during throttlingis zero for an ideal gas, whereas the actual cycles depend exclusively on the real gasbehavior to produce refrigeration during throttling.

1.5 REVERSED CARNOT CYCLE EMPLOYING A GAS

Reversed Carnot cycle is an ideal refrigeration cycle for constant temperature external heatsource and heat sinks. Figure 1(a) shows the schematic of a reversed Carnot refrigerationsystem using a gas as the working fluid along with the cycle diagram on T-s and P-vcoordinates. As shown, the cycle consists of the following four processes:

Process 1-2: Reversible, adiabatic compression in a compressor

Process 2-3: Reversible, isothermal heat rejection in a compressor

Process 3-4: Reversible, adiabatic expansion in a turbine

Process 4-1: Reversible, isothermal heat absorption in a turbine


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Schematic of a reverse Carnot refrigeration system

Reverse Carnot refrigeration system in P-v and T-s coordinates

The heat transferred during isothermal processes 2-3 and 4-1 are given by:

3q2−3=∫T.ds=Th(s3 −s2) (5a)

21

q4−1=∫T.ds =Tl (s1 −s4) (5b)4

s = s2

and s3

= s4, hence s

2

- s3

= s - s4

(6)1 1

Applying first law of thermodynamics to the closed cycle,

∫δq=(q4−1 + q2−3) =∫δw =(w2−3 − w4−1) =− wnet (7)


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1.6 LIMITATIONS OF CARNOT CYCLE

Carnot cycle is an idealization and it suffers from several practical limitations. One of themain difficulties with Carnot cycle employing a gas is the difficulty of achieving isothermalheat transfer during processes 2-3 and 4-1. For a gas to have heat transfer isothermally, it isessential to carry out work transfer from or to the system when heat is transferred to thesystem (process 4-1) or from the system (process 2-3). This is difficult to achieve in practice.In addition, the volumetric refrigeration capacity of the Carnot system is very small leading tolarge compressor displacement, which gives rise to large frictional effects. All actualprocesses are irreversible, hence completely reversible cycles are idealizations only.

1.7 IDEAL REVERSE BRAYTON CYCLE

Schematic of a closed reverse Brayton cycle

This is an important cycle frequently employed in gas cycle refrigeration systems. This maybe thought of as a modification of reversed Carnot cycle, as the two isothermal processes ofCarnot cycle are replaced by two isobaric heat transfer processes. This cycle is also called asJoule or Bell-Coleman cycle. Figure 9.2(a) and(b) shows the schematic of a closed, reverseBrayton cycle and also the cycle on T-sdiagram. As shown in the figure, the ideal cycleconsists of the following four processes:


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Process 1-2: Reversible, adiabatic compression in a compressor

Process 2-3: Reversible, isobaric heat rejection in a heat exchanger

Process 3-4: Reversible, adiabatic expansion in a turbine

Process 4-1: Reversible, isobaric heat absorption in a heat exchanger

Reverse Brayton cycle in T-s plane

(Tl−T4 ) T4 (Tl −T4) γ−1 −1)−1 (9.16)COP = = = =(r γ(T −T ) −(T −T ) T −T γ−1 p

2 1 3 4 3 4 (T − T )(r γ −1)

1 4 p

1.8 AIRCRAFT COOLING SYSTEMS

In an aircraft, cooling systems are required to keep the cabin temperatures at a comfortablelevel. Even though the outside temperatures are very low at high altitudes, still cooling ofcabin is required due to:

i. Large internal heat generation due to occupants, equipment etc.

ii. Heat generation due to skin friction caused by the fast moving aircraft


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iii. At high altitudes, the outside pressure will be sub-atmospheric. When air at this lowpressure is compressed and supplied to the cabin at pressures close to atmospheric,the temperature increases significantly. For example, when outside air at a pressure of0.2 bar and temperature of 223 K (at 10000 m altitude) is compressed to 1 bar, itstemperature increases to about 353 K. If the cabin is maintained at 0.8 bar, thetemperature will be about 332 K. This effect is called as ram effect. This effect addsheat to the cabin, which needs to be taken out by the cooling system.

iv. Solar radiation

For low speed aircraft flying at low altitudes, cooling system may not be required, however,for high speed aircraft flying at high altitudes, a cooling system is a must.

Even though the COP of air cycle refrigeration is very low compared to vapour compressionrefrigeration systems, it is still found to be most suitable for aircraft refrigeration systems as:

i. Air is cheap, safe, non-toxic and non-flammable. Leakage of air is not a problem

ii. Cold air can directly be used for cooling thus eliminating the low temperature heatexchanger (open systems) leading to lower weight

iii. The aircraft engine already consists of a high speed turbo-compressor, henceseparate compressor for cooling system is not required. This reduces the weight perkW cooling considerably. Typically, less than 50% of an equivalent vapourcompression system

Design of the complete system is much simpler due to low pressures. Maintenance requiredis also less.


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Simple aircraft refrigeration cycle:

Figure 9.5 shows the schematic of a simple aircraft refrigeration system and the operatingcycle on T-s diagram. This is an open system. As shown in the T-s diagram, the outside lowpressure and low temperature air (state 1) is compressed due to ram effect to ram pressure(state 2). During this process its temperature increases from 1 to 2. This air is compressed inthe main compressor to state 3, and is cooled to state 4 in the air cooler. Its pressure isreduced to cabin pressure in the turbine (state 5), as a result its temperature drops from 4 to5. The cold air at state 5 is supplied to the cabin. It picks up heat as it flows through the cabinproviding useful cooling effect. The power output of the turbine is used to drive the fan, whichmaintains the required air flow over the air cooler. This simple system is good for groundcooling (when the aircraft is not moving) as fan can continue to maintain airflow over the aircooler.

By applying steady flow energy equation to the ramming process, the temperature rise at theend of the ram effect can be shown to be:

Bootstrap system:

Figure shows the schematic of a bootstrap system, which is a modification of the simplesystem. As shown in the figure, this system consists of two heat exchangers (air cooler and


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aftercooler), in stead of one air cooler of the simple system. It also incorporates a secondarycompressor, which is driven by the turbine of the cooling system. This system is suitable forhigh speed aircraft, where in the velocity of the aircraft provides the necessary airflow for theheat exchangers, as a result a separate fan is not required. As shown in the cycle diagram,ambient air state 1 is pressurized to state 2 due to the ram effect. This air is furthercompressed to state 3 in the main compressor. The air is then cooled to state 4 in the aircooler. The heat rejected in the air cooler is absorbed by the ram air at state 2. The air fromthe air cooler is further compressed from state 4 to state 5 in the secondary compressor. It isthen cooled to state 6 in the after cooler, expanded to cabin pressure in the cooling turbineand is supplied to the cabin at a low temperature T7. Since the system does not consist of aseparate fan for driving the air through the heat exchangers, it is not suitable for groundcooling. However, in general ground cooling is normally done by an external air conditioningsystem as it is not efficient to run the aircraft engine just to provide cooling when it isgrounded.

Other modifications over the simple system are: regenerative system and reduced ambientsystem. In a regenerative system, a part of the cold air from the cooling turbine is used forprecooling the air entering the turbine. As a result much lower temperatures are obtained atthe exit of the cooling turbine, however, this is at the expense of additional weight and designcomplexity. The cooling turbine drives a fan similar to the simple system. The regenerativesystem is good for both ground cooling as well as high speed aircrafts. The reduced ambientsystem is well-suited for supersonic aircrafts and rockets.


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Dry Air Rated Temperature (DART):

The concept of Dry Air Rated Temperature is used to compare different aircraft refrigerationcycles. Dry Air Rated Temperature is defined as the temperature of the air at the exit of thecooling turbine in the absence of moisture condensation. For condensation not to occurduring expansion in turbine, the dew point temperature and hence moisture content of the airshould be very low, i.e., the air should be very dry. The aircraft refrigeration systems arerated based on the mass flow rate of air at the design DART. The cooling capacity is thengiven by:

..

Q =m cp (Ti −TDART ).

where m is the mass flow rate of air, TDART and Ti are the dry air rated temperature andcabin temperature, respectively.

A comparison between different aircraft refrigeration systems based on DART at differentMach numbers shows that:

i. DART increases monotonically with Mach number for all the systems except thereduced ambient system

ii. The simple system is adequate at low Mach numbersiii. At high Mach numbers either bootstrap system or regenerative system should be

usediv. Reduced ambient temperature system is best suited for very high Mach number,

supersonic aircrafts



..










..









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VAPOUR COMPRESSION REFRIGERATIONSYSTEMS

Vapour Compression System : Analysis of theoretical vapour compression cycle,Representation of cycle on T - S and p - h diagram, Simple saturation cycle, sub-cooled cycle and super-heated cycle, Effect of suction and discharge pressure onperformance, Actual vapour compression cycle. Problem illustration and solution.

2.1 COMPARISON BETWEEN GAS CYCLES AND VAPORCYCLESThermodynamic cycles can be categorized into gas cycles and vapour cycles. Asmentioned in the previous chapter, in a typical gas cycle, the working fluid (a gas)does not undergo phase change, consequently the operating cycle will be awayfrom the vapour dome. In gas cycles, heat rejection and refrigeration take placeas the gas undergoes sensible cooling and heating. In a vapour cycle theworking fluid undergoes phase change and refrigeration effect is due to thevaporization of refrigerant liquid. If the refrigerant is a pure substance then itstemperature remains constant during the phase change processes. However, if azeotropic mixture is used as a refrigerant, then there will be a temperature glideduring vaporization and condensation. Since the refrigeration effect is producedduring phase change, large amount of heat (latent heat) can be transferred perkilogram of refrigerant at a near constant temperature. Hence, the required massflow rates for a given refrigeration capacity will be much smaller compared to agas cycle. Vapour cycles can be subdivided into vapour compression systems,vapour absorption systems, vapour jet systems etc. Among these the vapourcompression refrigeration systems are predominant.

2.2 VAPOUR COMPRESSION REFRIGERATION SYSTEMSAs mentioned, vapour compression refrigeration systems are the most commonlyused among all refrigeration systems. As the name implies, these systemsbelong to the general class of vapour cycles, wherein the working fluid(refrigerant) undergoes phase change at least during one process. In a vapourcompression refrigeration system, refrigeration is obtained as the refrigerantevaporates at low temperatures. The input to the system is in the form ofmechanical energy required to run the compressor. Hence these systems arealso called as mechanical refrigeration systems.


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VCR systems are available to suit almost all applications with the refrigerationcapacities ranging from few Watts to few megawatts. A wide variety ofrefrigerants can be used in these systems to suit different applications, capacitiesetc. The actual vapour compression cycle is based on Evans-Perkins cycle,which is also called as reverse Rankine cycle. Before the actual cycle isdiscussed and analysed, it is essential to find the upper limit of performance ofvapour compression cycles. This limit is set by a completely reversible cycle.

2.3 THE CARNOT REFRIGERATION CYCLECarnot refrigeration cycle is a completely reversible cycle, hence is used as amodel of perfection for a refrigeration cycle operating between a constanttemperature heat source and sink. It is used as reference against which the realcycles are compared. Figures (a) and (b) show the schematic of a Carnot vapourcompression refrigeration system and the operating cycle on T-s diagram.

As shown in Fig(a), the basic Carnot refrigeration system for purevapour consists of four components: compressor, condenser, turbine andevaporator. Refrigeration effect (q4-1 = qe) is obtained at the evaporator as therefrigerant undergoes the process of vaporization (process 4-1) and extracts thelatent heat from the low temperature heat source. The low temperature, lowpressure vapour is then compressed isentropically in the compressor to the heatsink temperature Tc. The refrigerant pressure increases from Pe to Pc during thecompression process (process 1- 2) and the exit vapour is saturated. Next thehigh pressure, high temperature saturated refrigerant undergoes the process ofcondensation in the condenser (process 2-3) as it rejects the heat ofcondensation (q2-3 = qc) to an external heat sink at Tc. The high pressuresaturated liquid then flows through the turbine and undergoes isentropicexpansion (process 3-4). During this process, the pressure and temperature fallfrom Pc,Tc to Pe, Te. Since a saturated liquid is expanded in the turbine, someamount of liquid flashes into vapour and the exit condition lies in the two-phaseregion. This low temperature and low pressure liquid-vapour mixture then entersthe evaporator completing the cycle. Thus as shown in Fig.10.1(b), the cycleinvolves two isothermal heat transfer processes (processes 4- 1 and 2-3) andtwo isentropic work transfer processes (processes 1-2 and 3-4). Heat is extractedisothermally at evaporator temperature Te during process 4-1, heat is rejectedisothermally at condenser temperature Tc during process 2-3. Work is supplied tothe compressor during the isentropic compression (1-2) of refrigerant vapourfrom evaporator pressure Pe to condenser pressure Pc, and work is produced bythe system as refrigerant liquid expands isentropically in the turbine fromcondenser pressure Pc to evaporator pressure Pe. All the processes are bothinternally as well as externally reversible, i.e., net entropy generation for thesystem and environment is zero.


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Heat sink

qc

C3 2

T Cwnet

4 1E

qe

Heat source

Schematic of a Carnot refrigeration system

Pc

PeT

Tc 3qc

2w3-4

Tew1-2

4 1

qe

sCarnot refrigeration cycle on T-s diagram

4 Versi on 1 ME, IIT Kharagpur


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now for the reversible, isothermal heat transfer processes 2-3 and 4-1, we canwrite:

3

− s3)q c =− q 2−3 = −∫T.ds =Tc(s2

21

−s4 )q e =q 4−1 = ∫T.ds =Te (s14

where Te and Tc are the evaporator and condenser temperatures, respectively, and,

s1=s 2 and s3=s 4

the Coefficient of Performance (COP) is given by:

refrigeration effect q e T (s 1 −s 4 ) TCOPCarnot

= e e

= = =net work input w net T (s 2 −s 3 ) −T (s 1−s 4 ) T −T

thus the COP of Carnot refrigeration cycle is a function of evaporator andcondenser temperatures only and is independent of the nature of the workingsubstance. This is the reason why exactly the same expression was obtained forair cycle refrigeration systems operating on Carnot cycle (Lesson 9). The CarnotCOP sets an upper limit for refrigeration systems operating between two constanttemperature thermal reservoirs (heat source and sink). From Carnot’s theorems,for the same heat source and sink temperatures, no irreversible cycle can haveCOP higher than that of Carnot COP.

T

Tc 3 2

Te

wnet

4 1

qe

b a s

Carnot refrigeration cycle represented in T-s planePrepared By : Prof. (Dr.) Manmatha K. Roul || 16

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It can be seen from the above expression that the COP of a Carnot refrigerationsystem increases as the evaporator temperature increases and condensertemperature decreases. This can be explained very easily with the help of the T-sdiagram (Fig.10.2). As shown in the figure, COP is the ratio of area a-1-4-b to thearea 1-2-3-4. For a fixed condenser temperature Tc, as the evaporatortemperature Te increases, area a-1-4-b (qe) increases and area 1-2-3-4 (wnet)decreases as a result, COP increases rapidly. Similarly for a fixed evaporatortemperature Te, as the condensing temperature Tc increases, the net work input(area 1-2-3-4) increases, even though cooling output remains constant, as aresult the COP falls. Figure 10.3 shows the variation of Carnot COP withevaporator temperature for different condenser temperatures. It can be seen thatthe COP increases sharply with evaporator temperatures, particularly at highcondensing temperatures. COP reduces as the condenser temperatureincreases, but the effect becomes marginal at low evaporator temperatures. It willbe shown later that actual vapour compression refrigeration systems also behavein a manner similar to that of Carnot refrigeration systems as far as theperformance trends are concerned.

Effects of evaporator and condenser temperatures on Carnot COP

PRACTICAL DIFFICULTIES WITH CARNOT REFRIGERATION SYSTEM

It is difficult to build and operate a Carnot refrigeration system due to thefollowing practical difficulties:i. During process 1-2, a mixture consisting of liquid and vapour have to becompressed isentropically in the compressor. Such a compression is known aswet compression due to the presence of liquid. In practice, wet compression isvery difficult especially with reciprocating compressors. This problem isparticularly severe in case of high speed reciprocating compressors, which getdamaged due to the presence of liquid droplets in the vapour. Even though sometypes of compressors can tolerate the presence of liquid in




It is difficult to build and operate a Carnot refrigeration system due to thefollowing practical difficulties:i. During process 1-2, a mixture consisting of liquid and vapour have to becompressed isentropically in the compressor. Such a compression is known aswet compression due to the presence of liquid. In practice, wet compression isvery difficult especially with reciprocating compressors. This problem isparticularly severe in case of high speed reciprocating compressors, which getdamaged due to the presence of liquid droplets in the vapour. Even though sometypes of compressors can tolerate the presence of liquid in




It is difficult to build and operate a Carnot refrigeration system due to thefollowing practical difficulties:i. During process 1-2, a mixture consisting of liquid and vapour have to becompressed isentropically in the compressor. Such a compression is known aswet compression due to the presence of liquid. In practice, wet compression isvery difficult especially with reciprocating compressors. This problem isparticularly severe in case of high speed reciprocating compressors, which getdamaged due to the presence of liquid droplets in the vapour. Even though sometypes of compressors can tolerate the presence of liquid inPrepared By : Prof. (Dr.) Manmatha K. Roul || 17

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vapour, since reciprocating compressors are most widely is refrigeration,traditionally drycompression (compression of vapour only) is preferred to wetcompression.

ii. The second practical difficulty with Carnot cycle is that using a turbine andextracting work from the system during the isentropic expansion of liquidrefrigerant is not economically feasible, particularly in case of small capacitysystems. This is due to the fact that the specific work output (per kilogram ofrefrigerant) from the turbine is given by:

w3−4

=Pc

(10.6)∫ v.dPPe

since the specific volume of liquid is much smaller compared to the specificvolume of a vapour/gas, the work output from the turbine in case of the liquid willbe small. In addition, if one considers the inefficiencies of the turbine, then thenet output will be further reduced. As a result using a turbine for extracting thework from the high pressure liquid is not economically justified in most of thecases1.

One way of achieving dry compression in Carnot refrigeration cycle is to havetwo compressors – one isentropic and one isothermal as shown in Fig.10.4.

qc qc

Condenser 3

Pc

Pi4 q2-3 C Pc> Pi> Pe Pew2-3

2 4 32

T Cw1-2

5Evaporator

15 1

qeqe

Carnot refrigeration system with dry compression

As shown in Fig.10.4, the Carnot refrigeration system with dry compressionconsists of one isentropic compression process (1-2) from evaporator pressurePe to an intermediate pressure Pi and temperature Tc, followed by an isothermal

compression process (2-3) from the intermediate pressure Pi to the condenser

pressure Pc. Though with this modification the problem of wet compression canbe avoided, still this modified system is not practical due to the difficulty inachieving true isothermal compression using high-speed compressors. Inaddition, use of two compressors in place of one is not economically justified.


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From the above discussion, it is clear that from practical considerations, theCarnot refrigeration system need to be modified. Dry compression with a singlecompressor is possible if the isothermal heat rejection process is replaced byisobaric heat rejection process. Similarly, the isentropic expansion process canbe replaced by an isenthalpic throttling process. A refrigeration system, whichincorporates these two changes is known as Evans-Perkins or reverse Rankinecycle. This is the theoretical cycle on which the actual vapour compressionrefrigeration systems are based.

qc3 Condenser

2

Exp.C wcDevice

4Evaporator

qe

1

T Pc

2 Pe

Tc3 2'

Te 4 1

S

Standard Vapour compression refrigeration system


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STANDARD VAPOUR COMPRESSION REFRIGERATIONSYSTEM (VCRS)

Figure 10.5 shows the schematic of a standard, saturated, single stage (SSS)vapour compression refrigeration system and the operating cycle on a T sdiagram. As shown in the figure the standard single stage, saturated vapourcompression refrigeration system consists of the following four processes:

Process 1-2: Isentropic compression of saturated vapour incompressor Process 2-3: Isobaric heat rejection in condenserProcess 3-4: Isenthalpic expansion of saturated liquid in expansiondevice Process 4-1: Isobaric heat extraction in the evaporator

By comparing with Carnot cycle, it can be seen that the standard vapourcompression refrigeration cycle introduces two irreversibilities: 1) Irreversibilitydue to non -isothermal heat rejection (process 2-3) and 2) Irreversibility due toisenthalpic throttling (process 3-4). As a result, one would expect the theoreticalCOP of standard cycle to be smaller than that of a Carnot system for the sameheat source and sink temperatures. Due to these irreversibilities, the coolingeffect reduces and work input increases, thus reducing the system COP. Thiscan be explained easily with the help of the cycle diagrams on T s charts. Figure10.6(a) shows comparison between Carnot and standard VCRS in terms ofrefrigeration effect.

T

2

Tc3 2'

2’’

Te4'

14

A2c d e S

Comparison between Carnot and standard VCRS

The heat extraction (evaporation) process is reversible for both the Carnot cycle andVCRS cycle. Hence the refrigeration effect is given by:

For Carnot refrigeration cycle (1-2’’-3-4’):1

−s4') =area e −1−4'−c −eq e,Carnot=q 4'−1= ∫T.ds =Te (s14'


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For VCRS cycle (1-2-3-4):1

−s4) =area e −1−4 −d −eq e,VCRS=q 4−1= ∫T.ds =Te (s14

thus there is a reduction in refrigeration effect when the isentropic expansionprocess of Carnot cycle is replaced by isenthalpic throttling process of VCRScycle, this reduction is equal to the area d-4-4’-c-d (area A2) and is known asthrottling loss. The throttling loss is equal to the enthalpy difference betweenstate points 3 and 4’, i.e,

q e,Carnot −q VCRS =area d −4 −4'−c −d =(h 3 −h 4' ) =(h 4 −h 4' ) =area A 2(10.9)

It is easy to show that the loss in refrigeration effect increases as the evaporatortemperature decreases and/or condenser temperature increases. A practicalconsequence of this is a requirement of higher refrigerant mass flow rate.

The heat rejection in case of VCRS cycle also increases when compared toCarnot cycle.

T A1

23 2'

2''

4'14

c d e S

Comparative evaluation of heat rejection rate of VCRS and Carnot cycle

As shown in Fig.10.6(b), the heat rejection in case of Carnot cycle (1-2’’-3-4’) isgiven by:

3−s3) =area e −2' '−3 −c −eq c,Carnot=−q 2''−3= − ∫T.ds =Tc (s 2''

2''

In case of VCRS cycle, the heat rejection rate is given by:3

q c,VCRS=−q 2−3= − ∫T.ds = area e −2−3 −c −e2

Hence the increase in heat rejection rate of VCRS compared to Carnot cycle isequal to the area 2’’-2-2’ (area A1). This region is known as superheat horn, andis due to thePrepared By : Prof. (Dr.) Manmatha K. Roul || 21

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replacement of isothermal heat rejection process of Carnot cycle by isobaric heatrejection in case of VCRS.

Since the heat rejection increases and refrigeration effect reduces when theCarnot cycle is modified to standard VCRS cycle, the net work input to the VCRSincreases compared to Carnot cycle. The net work input in case of Carnot andVCRS cycles are given by:

w net,Carnot=(q c−q e )Carnot=area 1−2' '−3 −4'−1

w net,VCRS=(q c−q e ) VCRS=area 1−2 −3 −4'−c −d −4 −1

As shown in Fig.10.6(c), the increase in net work input in VCRS cycle is given by:

w net,VCRS− w net,Carnot= area 2' '−2 −2' +area c −4'−4 −d −c =area A1+area A 2(10.14)

T

A12

3 2' 2’’

4'14

A2

c d e S

Fig.10.6(c). Figure illustrating the increase in net work input in VCRS cycle

To summarize the refrigeration effect and network input of VCRS cycle are given by:

wnet,VCRS

=wnet,Carnot

+area A1+area A

2

The COP of VCRS cycle is given by:

COPVCRS=

qe,VCRS

=

qe,Carnot

−area A2

wnet,VCRS w net,Carnot+area A1+area A 2

qe,VCRS

=qe,Carnot

−area A2


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If we define the cycle efficiency, ηR as the ratio of COP of VCRS cycle to theCOP of Carnot cycle, then:

area A21 −

COP qe,Carnot

ηR =VCRS

=COPCarnot 1 +

area A1 +area A 2

wnet,Carnot

The cycle efficiency (also called as second law efficiency) is a good indication ofthe deviation of the standard VCRS cycle from Carnot cycle. Unlike Carnot COP,the cycle efficiency depends very much on the shape of T s diagram, which inturn depends on the nature of the working fluid.

If we assume that the potential and kinetic energy changes during isentropiccompression process 1-2 are negligible, then the work input w1-2 is given by:

Now as shown in Fig.10.7, if we further assume that the saturated liquid line 3-fcoincides with the constant pressure line Pc in the subcooled region (which is areasonably good assumption), then from the 2nd Tds relation;

Tds =dh - v dP = dh; when P is constantf

∴(h 2−h f ) = ∫Tds =area e −2 −3 −f −g −e2


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fand, (h1−h f ) =∫Tds =area e −1 −f −g −e

1Substituting these expressions in the expression for net work input, we obtain thecompressor work input to be equal to area 1-2-3-f-1. Now comparing this with theearlier expression for work input (area 1-2-3-4’-c-d-4-1), we conclude that areaA2 is equal to area A3.

As mentioned before, the losses due to superheat (area A1) and throttling (areaA2 ≈ A3) depend very much on the shape of the vapor dome (saturation liquidand vapour curves) on T s diagram. The shape of the saturation curves dependson the nature of refrigerant. Figure 10.8 shows T s diagrams for three differenttypes of refrigerants.

Type 1 Type 2

22

T 3 2' T 3 2'2'' 2''

4'1

4'14 4

SType 3 S

3 2

T4'

41

S

Fig.10.8. T-s diagrams for three different types of refrigerants

Refrigerants such as ammonia, carbon di-oxide and water belong to Type 1.These refrigerants have symmetrical saturation curves (vapour dome), as aresult both the superheat and throttling losses (areas A1 and A3) are significant.That means deviation of VCRS cycle from Carnot cycle could be significant whenthese refrigerants are used as working fluids. Refrigerants such as CFC11,CFC12, HFC134a belong to Type 2, these refrigerants have small superheatlosses (area A1) but large throttling losses (area A3). High molecular weightrefrigerants such as CFC113, CFC114, CFC115, iso-butane belonging to Type 3,do not have any superheat losses, i.e., when the compression inlet condition issaturated (point 1), then the exit condition will be in the 2-phase region, as aresult it is not necessary to superheat the refrigerant. However, these refrigerants


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experience significant throttling losses. Since the compressor exit condition ofType 3 refrigerants may fall in the two -phase region, there is a danger of wetcompression leading to compressor damage. Hence for these refrigerants, thecompressor inlet condition is chosen such that the exit condition does not fall inthe two-phase region. This implies that the refrigerant at the inlet to thecompressor should be superheated, the extent of which depends on therefrigerant.

SUPERHEAT AND THROTTLING LOSSES:

It can be observed from the discussions that the superheat loss is fundamentallydifferent from the throttling loss. The superheat loss increases only the work inputto the compressor, it does not effect the refrigeration effect. In heat pumpssuperheat is not a loss, but a part of the useful heating effect. However, theprocess of throttling is inherently irreversible, and it increases the work input andalso reduces the refrigeration effect.

ANALYSIS OF STANDARD VAPOUR COMPRESSIONREFRIGERATION SYSTEMA simple analysis of standard vapour compression refrigeration system can becarried out by assuming a) Steady flow; b) negligible kinetic and potential energychanges across each component, and c) no heat transfer in connecting pipelines. The steady flow energy equation is applied to each of the four components.

.

Evaporator: Heat transfer rate at evaporator or refrigeration capacity, Qe is givenby:

Qe =mr

(h1 − h 4 ).

where mr is the refrigerant mass flow rate in kg/s, h1 and h4 are the specificenthalpies (kJ/kg) at the exit and inlet to the evaporator, respectively. (h1 − h 4 ) isknown as specificrefrigeration effect or simply refrigeration effect, which is equal to the heattransferred at the evaporator per kilogram of refrigerant. The evaporator pressurePe is the saturation pressure corresponding to evaporator temperature Te, i.e.,

Pe = Psat (Te ).

Compressor: Power input to the compressor, Wc is given by:. .

W c =m r (h 2 −h1 )

where h2 and h1 are the specific enthalpies (kJ/kg) at the exit and inlet to thecompressor, respectively. (h 2 −h1 ) is known as specific work of compression orsimply work ofcompression, which is equal to the work input to the compressor per kilogramofrefrigerant.Condenser: Heat transfer rate at condenser, Qc is given by:Prepared By : Prof. (Dr.) Manmatha K. Roul || 25

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. .

Qc =mr (h2 −h 3 )

where h3 and h2 are the specific enthalpies (kJ/kg) at the exit and inlet to thecondenser, respectively.

The condenser pressure Pc is the saturation pressure corresponding toevaporator temperature Tc, i.e.,

Pc = Psat (Tc )

Expansion device: For the isenthalpic expansion process, the kinetic energychange across the expansion device could be considerable, however, if we takethe control volume, well downstream of the expansion device, then the kineticenergy gets dissipated due to viscous effects, and

h 3 =h 4

The exit condition of the expansion device lies in the two-phase region, henceapplying the definition of quality (or dryness fraction), we can write:

h 4=(1− x 4 )h f ,e+ x 4 h g,e=h f+ x 4 h fg

where x4 is the quality of refrigerant at point 4, hf,e, hg,e, hfg are the saturatedliquid enthalpy, saturated vapour enthalpy and latent heat of vaporization atevaporatorpressure, respectively.

The COP of the system is given by:

. . −h (h −h )Q e m r (h 1 4 ) 1 4COP = = =

.

(h 2 −h1).

W c m r (h 2−h

1)

.At any point in the cycle, the mass flow rate of refrigerant mr can be written interms of volumetric flow rate and specific volume at that point, i.e.,

. .

m r =V

v

applying this equation to the inlet condition of the compressor,. .

mr=V

1

v1.

where V1 is the volumetric flow rate at compressor inlet and v1 is the specificvolume at

.compressor inlet. At a given compressor speed, V1 is an indication of the size ofthe compressor. We can also write, the refrigeration capacity in terms ofvolumetric flow rate as:Prepared By : Prof. (Dr.) Manmatha K. Roul || 26

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. . . h1−h

4Q e =mr(h 1 −h 4 ) =V1

v1h

1−h

4 3 of refrigerant).where

v1

is called as volumetr ic refrigera tion effect ( kJ/m

Generally, the type of refrigerant, required refrigeration capacity, evaporatortemperature and condenser temperature are known. Then from the evaporatorand condenser temperature one can find the evaporator and condenserpressures and enthalpies at the exit of evaporator and condenser (saturatedvapour enthalpy at evaporator pressure and saturated liquid enthalpy atcondenser pressure). Since the exit condition of the compressor is in thesuperheated region, two independent properties are required to fix the state ofrefrigerant at this point. One of these independent properties could be thecondenser pressure, which is already known. Since the compression process isisentropic, the entropy at the exit to the compressor is same as the entropy at theinlet, s1 which is the saturated vapour entropy at evaporator pressure (known).Thus from the known pressure and entropy the exit state of the compressor couldbe fixed, i.e.,

h2 =h(Pc ,s2)

=h(Pc ,s1 )

s1 =s

2

The quality of refrigerant at the inlet to the evaporator (x4) could be obtained fromtheknown values of h3, hf,e and hg,e.Once all the state points are known, then from the required refrigeration capacityand various enthalpies one can obtain the required refrigerant mass flow rate,volumetric flow rate at compressor inlet, COP, cycle efficiency etc.Use of Pressure-enthalpy (P-h) charts:

Te Tc

Pc3 2' 2

P

Pe 4 1

h3 = h4 h1 h2hFig.10.9. Standard vapour compression refrigeration cycle on a P-h chartPrepared By : Prof. (Dr.) Manmatha K. Roul || 27

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Since the various performance parameters are expressed in terms of enthalpies,it is very convenient to use a pressure – enthalpy chart for property evaluationand performance analysis. The use of these charts was first suggested byRichard Mollier. Figure 10.9 shows the standard vapour compressionrefrigeration cycle on a P-h chart. As discussed before, in a typical P-h chart,enthalpy is on the x-axis and pressure is on y-axis. The isotherms are almostvertical in the subcooled region, horizontal in the two-phase region (for purerefrigerants) and slightly curved in the superheated region at high pressures, andagain become almost vertical at low pressures. A typical P-h chart also showsconstant specific volume lines (isochors) and constant entropy lines (isentropes)in the superheated region. Using P-h charts one can easily find variousperformance parameters from known values of evaporator and condenserpressures.

In addition to the P-h and T- s charts one can also use thermodynamic propertytables from solving problems related to various refrigeration cycles.


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Multi-stage compression and Multi-evaporatorsystems

Multi-stage compression and Multi-evaporator systems : Different arrangements ofcompressors and inter-cooling, Multistage compression with inter-cooling, Multi-evaporator system, Dual compression system. Simple problems

INTRODUCTION

A single stage vapour compression refrigeration system has one low sidepressure (evaporator pressure) and one high side pressure (condenserpressure). The performance of single stage systems shows that these systemsare adequate as long as the temperature difference between evaporator andcondenser (temperature lift) is small. However, there are many applicationswhere the temperature lift can be quite high. The temperature lift can becomelarge either due to the requirement of very low evaporator temperatures and/ordue to the requirement of very high condensing temperatures. For example, infrozen food industries the required evaporator can be as low as –40oC, while inchemical industries temperatures as low as –150oC may be required forliquefaction of gases. On the high temperature side the required condensingtemperatures can be very high if the refrigeration system is used as a heat pumpfor heating applications such as process heating, drying etc. However, as thetemperature lift increases the single stage systems become inefficient andimpractical. For example, Fig. 12.1 shows the effect of decreasing evaporatortemperatures on T s and P h diagrams. It can be seen from the T s diagrams thatfor a given condenser temperature, as evaporator temperature decreases:

i. Throttling losses increaseii. Superheat losses increaseiii. Compressor discharge temperature increasesiv. Quality of the vapour at the inlet to the evaporator increasesv. Specific volume at the inlet to the compressor increases

As a result of this, the refrigeration effect decreases and work of compressionincreases as shown in the P h diagram. The volumic refrigeration effect alsodecreases rapidly as the specific volume increases with decreasing evaporatortemperature. Similar effects will occur, though not in the same proportion whenthe condenser temperature increases for a given evaporator temperature. Due tothese drawbacks, single stage systems are not recommended when theevaporator temperature becomes very low and/or when the condensertemperature becomes high. In such cases multi-stage systems are used inpractice. Generally, for fluorocarbon and ammonia based refrigeration systems asingle stage system is used upto an evaporator temperature of –30o C. A two-stage system is used upto –60oC and a three-stage system is used fortemperatures below –60oC.

Apart from high temperature lift applications, multi-stage systems are alsoused in applications requiring refrigeration at different temperatures. ForPrepared By : Prof. (Dr.) Manmatha K. Roul || 29

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example, in a dairy plant refrigeration may be required at –30o C for making icecream and at 2oC for chilling milk. In such cases it may be advantageous to usea multi-evaporator system with the low temperature evaporator operating at –30oC and the high temperature evaporator operating at 2oC

Fig.12.1(a): Effect of evaporator temperature on cycle performance (T-sdiagram)


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MODULE - II

PreparedBy


Gandhi Institute for Technological Advancement (GITA),Bhubaneswar – 752054

June 2016


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HUBANESWAR

Vapour Absorption System : Simple Ammonia - absorption system, Improved absorptionsystem, Analysis of vapour absorption system (Specifically of analyzing coloumn andrectifier), Electrolux / Three fluid system, Lithium-bromide-water vapour absorption system,comparison of absorption system with vapour compression system. Simple Problems andsolution.

Thermoelectric Refrigeration: Basics and Principle. Defining the figure of Merit. (No Problem)

Refrigerants : Classification of refrigerants and its degignation- Halocarbon (compounds,Hydrocarbons, Inorganic compounds, Azeotropes, Properties of refrigerants, comparison ofcommon refrigerants, uses of important refrigerants, Brines. Alternative refrigerants (Organicand inorganic compounds).


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VAPOUR ABSORPTION SYSTEM

Simple Ammonia - absorption system, Improved absorption system, Analysis of vapourabsorption system (Specifically of analyzing coloumn and rectifier), Electrolux / Three fluidsystem, Lithium-bromide-water vapour absorption system, comparison of absorption systemwith vapour compression system. Simple Problems and solution.

4.1. INTRODUCTION

Vapour Absorption Refrigeration Systems (VARS) belong to the class of vapour cyclessimilar to vapour compression refrigeration systems. However, unlike vapour compressionrefrigeration systems, the required input to absorption systems is in the form of heat. Hencethese systems are also called as heat operated or thermal energy driven systems. Sinceconventional absorption systems use liquids for absorption of refrigerant, these are alsosometimes called as wet absorption systems. Similar to vapour compression refrigerationsystems, vapour absorption refrigeration systems have also been commercialized and arewidely used in various refrigeration and air conditioning applications. Since these systems runon low-grade thermal energy, they are preferred when low-grade energy such as waste heator solar energy is available. Since conventional absorption systems use natural refrigerantssuch as water or ammonia they are environment friendly.

In this lesson, the basic working principle of absorption systems, the maximum COP ofideal absorption refrigeration systems, basics of properties of mixtures and simple absorptionrefrigeration systems will be discussed.

4.2. BASIC PRINCIPLE

When a solute such as lithium bromide salt is dissolved in a solvent such as water, theboiling point of the solvent (water) is elevated. On the other hand, if the temperature of thesolution (solvent + solute) is held constant, then the effect of dissolving the solute is to reducethe vapour pressure of the solvent below that of the saturation pressure of pure solvent atthat temperature. If the solute itself has some vapour pressure (i.e., volatile solute) then thetotal pressure exerted over the solution is the sum total of the partial pressures of solute andsolvent. If the solute is non-volatile (e.g. lithium bromide salt) or if the boiling point differencebetween the solution and solvent is large (≥ 300oC), then the total pressure exerted over thesolution will be almost equal to the vapour pressure of the solvent only. In the simplestabsorption refrigeration system, refrigeration is obtained by connecting two vessels, with onevessel containing pure solvent and the other containing a solution. Since the pressure isalmost equal in both the vessels at equilibrium, the temperature of the solution will be higherthan that of the pure solvent. This means that if the solution is at ambient temperature, thenthe pure solvent will be at a temperature lower than the ambient. Hence refrigeration effect isproduced at the vessel containing pure solvent due to this temperature difference. Thesolvent evaporates due to heat transfer from the surroundings, flows to the vessel containingsolution and is absorbed by the solution. This process is continued as long as thecomposition and temperature of the solution are maintained and liquid solvent is available in


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the container.For example, Fig.4.1 shows an arrangement, which consists of two vessels A and B

connected to each other through a connecting pipe and a valve. Vessel A is filled with purewater, while vessel B is filled with a solution containing on mass basis 50 percent of waterand 50 percent lithium bromide (LiBr salt). Initially the valve connecting these two vessels isclosed, and both vessels are at thermal equilibrium with the surroundings, which is at 30o C.At 30oC, the saturation pressure of water is 4.24 kPa, and the equilibrium vapour pressure ofwater-lithium bromide solution (50 : 50 by mass) at 30oC is 1.22 kPa.


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Fig.4.1: Basic principle of vapour absorption systemsThus at initial equilibrium condition, the pressure in vessel A is 4.24 kPa, while it is

1.22 kPa in vessel B. Now the valve between vessels A and B is opened. Initially due topressure difference water vapour will flow from vessel A to vessel B, and this vapour will beabsorbed by the solution in vessel B. Since absorption in this case is exothermic, heat will bereleased in vessel B. Now suppose by some means the concentration and temperature ofvessel B are maintained constant at 50 % and 30oC, respectively. Then at equilibrium, thepressure in the entire system (vessels A and B) will be 1.22 kPa (equilibrium pressure of 50% LiBr solution at 30oC). Thetemperature of water in vessel A will be the saturationtemperature corresponding to 1.22 kPa, which is equal to about 10oC, as shown in the figure.Since the water temperature in A is lower than the surroundings, a refrigeration effect (Qe)can produced by transferring heat from the surroundings to water at 10oC. Due to this heattransfer, water vaporizes in A, flows to B and is absorbed by the solution in B. Theexothermic heat of absorption (Qa) is rejected to the surroundings.

Now for the above process to continue, there should always be pure water in vessel A,and vessel B must be maintained always at 50 percent concentration and 30oC. This is notpossible in a closed system such as the one shown in Fig.4.1. In a closed system with finitesized reservoirs, gradually the amount of water in A decreases and the solution in B becomesdiluted with water. As a result, the system pressure and temperature of water in A increasewith time. Hence the refrigeration effect at A reduces gradually due to the reducedtemperature difference between the surroundings and water. Thus refrigeration produced bysystems using only two vessels is intermittent in nature. In these systems, after a period, therefrigeration process has to be stopped and both the vessels A and B have to be broughtback to their original condition. This requires removal of water absorbed in B and adding itback to vessel A in liquid form, i.e., a process of regeneration as shown in Fig.4.1(c).

Assume that before regeneration is carried out, the valve between A and B is closedand both A and B are brought in thermal equilibrium with the surroundings (30oC), then duringthe regeneration process, heat at high temperature Tg is supplied to the dilute LiBr solution inB, as a result water vapour is generated in B. The vapour generated in B is condensed intopure water in A by rejecting heat of condensation to the surroundings. This process has to becontinued till all the water absorbed during the refrigeration process (4.1(b)) is transferredback to A. Then to bring the system back to its original condition, the valve has to be closedand solution in vessel B has to be cooled to 30oC. If we assume a steady-flow process ofregeneration and neglect temperature difference for heat transfer, then the temperature ofwater in A will be 30oC and pressure inside the system will be 4.24 kPa. Then thetemperature in vessel B, Tg depends on the concentration of solution in B. The amount ofheat transferred during refrigeration and regeneration depends on the properties of solutionand the operating conditions. It can be seen that the output from this system is therefrigeration obtained Qe and the input is heat supplied to vessel B during vapourregeneration process, Qg.

The system described may be called as an Intermittent Absorption RefrigerationSystem. The solvent is the refrigerant and the solute is called as absorbent. These simplesystems can be used to provide refrigeration using renewable energy such as solar energy inremote and rural areas. As already explained, these systems provided refrigerationintermittently, if solar energy is used for regenerating the refrigerant, then regenerationprocess can be carried out during the day and refrigeration can be produced during the night.

Though the intermittent absorption refrigeration systems discussed above are simplein design and inexpensive, they are not useful in applications that require continuousPrepared By : Prof. (Dr.) Manmatha K. Roul || 5

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refrigeration. Continuous refrigeration can be obtained by having a modified system with twopairs of vessels A and B and additional expansion valves and a solution pump.

Figs.14.2: a) Vapour compression refrigeration system (VCRS)b) Vapour Absorption Refrigeration System (VARS)

Figure 14.2(a) and (b) show a continuous output vapour compression refrigerationsystem and a continuous output vapour absorption refrigeration system. As shown in thefigure in a continuous absorption system, low temperature and low pressure refrigerant withlow quality enters the evaporator and vaporizes by producing useful refrigeration Qe. Fromthe evaporator, the low temperature, low pressure refrigerant vapour enters the absorberwhere it comes in contact with a solution that is weak in refrigerant. The weak solutionabsorbs the refrigerant and becomes strong in refrigerant. The heat of absorption is rejectedto the external heat sink at To. The solution that is now rich in refrigerant is pumped to highpressure using a solution pump and fed to the generator. In the generator heat at hightemperature Tg is supplied, as a result refrigerant vapour is generated at high pressure. Thishigh pressure vapour is then condensed in the condenser by rejecting heat of condensationto the external heat sink at To . The condensed refrigerant liquid is then throttled in theexpansion device and is then fed to the evaporator to complete the refrigerant cycle. On thesolution side, the hot, high-pressure solution that is weak in refrigerant is throttled to theabsorber pressure in the solution expansion valve and fed to the absorber where it comes incontact with the refrigerant vapour from evaporator. Thus continuous refrigeration isproduced at evaporator, while heat at high temperature is continuously supplied to thegenerator. Heat rejection to the external heat sink takes place at absorber and condenser. Asmall amount of mechanical energy is required to run the solution pump. If we neglectpressure drops, then the absorption system operates between the condenser and evaporatorpressures. Pressure in absorber is same as the pressure in evaporator and pressure ingenerator is same as the pressure in condenser.

It can be seen from Fig.14.2, that as far as the condenser, expansion valve and evaporatorsare concerned both compression and absorption systems are identical. However, thedifference lies in the way the refrigerant is compressed to condenser pressure. In vapour


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compression refrigeration systems the vapour is compressed mechanically using thecompressor, where as in absorption system the vapour is firstconverted into a liquid and thenthe liquid is pumped to condenser pressure using the solution pump. Since for the samepressure difference, work input required to pump a liquid (solution) is much less than thework required for compressing a vapour due to

Pcvery small specific volume of liquid ( w =− ∫v.dP ), the mechanical energy required to

Peoperate vapour absorption refrigeration system is much less than that required to operate acompression system. However, the absorption system requires a relatively large amount oflow-grade thermal energy at generator temperature to generate refrigerant vapour from thesolution in generator. Thus while the energy input is in the form of mechanical energy invapour compression refrigeration systems, it is mainly in the form of thermal energy in caseof absorption systems. The solution pump work is often negligible compared to the generatorheat input. Thus the COPs for compression and absorption systems are given by:

COP=Qe

(14.1)VCRS

Wc

COPVARS =Qe

≈Qe

(14.2)Qg + Wp Qg

Thus absorption systems are advantageous where a large quantity of low-gradethermal energy is available freely at required temperature. However, it will be seen that forthe refrigeration and heat rejection temperatures, the COP of vapour compressionrefrigeration system will be much higher than the COP of an absorption system as a highgrade mechanical energy is used in the former, while a low-grade thermal energy is used inthe latter. However, comparing these systems based on COPs is not fully justified, asmechanical energy is more expensive than thermal energy. Hence, sometimes the secondlaw (or exergetic) efficiency is used to compare different refrigeration systems. It is seen thatthe second law (or exergetic) efficiency of absorption system is of the same order as that of acompression system.

4.3 MAXIMUM COP OF IDEAL ABSORPTION REFRIGERATION SYSTEM

In case of a single stage compression refrigeration system operating betweenconstant evaporator and condenser temperatures, the maximum possible COP is given byCarnot COP:

COP = Te (4.3)Carnt

Tc −Te

If we assume that heat rejection at the absorber and condenser takes place at sameexternal heat sink temperature To, then a vapour absorption refrigeration system operates


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between three temperature levels, Tg, To and Te. The maximum possible COP of arefrigeration system operating between three temperature levels can be obtained by applyingfirst and second laws of thermodynamics to the system. Figure 4.3 shows the various energytransfers and the corresponding temperatures in an absorption refrigeration system.

Fig.4.3: Various energy transfers in a vapour absorption refrigeration system

From first law of thermodynamics,

Qe +Qg −Qc+a + Wp =0 (4.4)

where Qe is the heat transferred to the absorption system at evaporator temperature Te, Qgis the heat transferred to the generator of the absorption system at temperature Tg, Qa+c isthe heat transferred from the absorber and condenser of the absorption system attemperature To and Wp is the work input to the solution pump.

From second law of thermodynamics,

Stotal

= Ssys

+ Ssurr

≥0(4.5)

where Stotal is the total entropy change which is equal to the sum of entropy changeof the system Ssys and entropy change of the surroundings Ssurr. Since therefrigeration system operates in a closed cycle, the entropy change of the workingfluid of the system undergoing the cycle is zero, i.e., Ssys =0 . The entropy changeof the surroundings is given by:

Ssurr

= − Q e −

Qg

+ Q a+c ≥0 (4.6)Tg ToTe

Substituting the expression for first law of thermodynamics in the above equation

Tg −To T −Toe (4.7)TQ

g≥Q

e T −Wp

g e


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Neglecting solution pump work, Wp; the COP of VARS is given by:

Qe T Tg −ToeCOP

VARS=

Q −T T (4.8)g≤

To e g

An ideal vapour absorption refrigeration system is totally reversible (i.e., both internally andexternally reversible). For a completely reversible system the total entropy change(system+surroundings) is zero according to second law, hence for anideal VARSStotal,rev =0 Ssurr,rev =0 . Hence:

Ssurr,rev

=−

Qe

−

Qg

+

Q a+c

=0 (4.9)TgTe To

Hence combining first and second laws and neglecting pump work, the maximum possibleCOP of an ideal VARS system is given by:

Qe T Tg −Toe

COPideal VARS =Q −T T (4.10)=

T

g o e g

Thus the ideal COP is only a function of operating temperatures similar to Carnot system. Itcan be seen from the above expression that the ideal COP of VARS system is equal to theproduct of efficiency of a Carnot heat engine operating between Tg and To and COP of aCarnot refrigeration system operating between To and Te, i.e.,

Q e T Tg −Toe

COPideal VARS =Q −T T =COPCarnot .ηCarnot (4.11)=

T

g o e g

Thus an ideal vapour absorption refrigeration system can be considered to be a combinedsystem consisting of a Carnot heat engine and a Carnot refrigerator as shown in Fig.14.4.Thus the COP of an ideal VARS increases as generator temperature (Tg) and evaporatortemperature (Te) increase and heat rejection temperature (To) decreases. However, the COPof actual VARS will be much less than that of an ideal VARS due to various internal andexternal irreversibilities present in actual systems.


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Fig.4.4: Vapour absorption refrigeration system as a combination of a heat Vapourabsorption system as a combination of Heat Engine and engine and a refrigerator

4.4 Basic Vapour Absorption Refrigeration SystemFigure 4.5 shows a basic vapour absorption refrigeration system with a solutionheat exchanger on a pressure vs temperature diagram. As shown in the figure, lowtemperature and low pressure refrigerant vapour from evaporator at state 1 entersthe absorber and is absorbed by solution weak in refrigerant (state 8). The heat ofabsorption (Qa) is rejected to an external heat sink at T∞. The solution, rich in

refrigerant (state 2) is pumped to the generator pressure (Pg) by the solution pump(state 3) . The pressurized solution gets heated up sensibly as it flows through thesolution heat exchanger by extracting heat from hot solution coming from generator(state 4). Heat is supplied to this solution from an external heat source in thegenerator (Qg at Tg), as a result refrigerant vapour is generated (absorbent mayalso boil to give off vapour in case of ammonia-water systems) at state 5. This high-pressure refrigerant vapour condenses in the condenser by rejecting heat ofcondensation to the external heat sink (Qc at T∞) and leaves the condenser as ahigh pressure liquid (state 9). This high pressure refrigerant liquid is throttled in theexpansion device to evaporator pressure P e (state 10) from where it enters the

evaporator, extracts heat from low temperature heat source (Qe at Te) and leavesthe evaporator as vapour at state 1, completing a cycle. The hot solution that isweak in refrigerant (state 6) leaves the generator at high temperature and is cooledsensibly by rejecting heat to the solution going to the generator in the solution heatexchanger (state 7). Then it is throttled to the evaporator pressure in the throttlevalve (state 8), from where it enters the absorber to complete the cycle. It can beseen that though not an essential component, the solution heat exchanger is usedin practical systems to improve the COP by reducing the heat input in the generator.A solution heat exchanger as shown in Fig.4.5 is a counterflow heat exchanger inwhich the hot solution coming from the generator comes in thermal contact with thecold solution going to the generator. As a result of this heat exchange, less heatinput is required in the generator and less heat is rejected in the absorber, thusimproving the system performance significantly.

The thermodynamic performance of the above system can be evaluated byapplying mass and energy balance to each component assuming a steady flowprocess. In simple theoretical analyses, internal irreversibilities such as pressuredrops between the components are generally neglected. To find the performancefrom the mass and energy balance equations one needs to know inputs such as thetype of refrigerant-absorbent mixtures used in the system, operating temperatures,composition of solution at the entry and exit of absorber, effectiveness of solutionheat exchanger etc. A simple steady flow analysis of the system will be presented inlater sections.


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Fig.4.5: Basic vapour absorption refrigeration system with a solution heat exchangeon apressure vs temperature diagram

4.5 Refrigerant-absorbent combinations for VARS

The desirable properties of refrigerant-absorbent mixtures for VARS are:

1. The refrigerant should exhibit high solubility with solution in the absorber. This is to saythat it should exhibit negative deviation from Raoult’s law at absorber.

2. There should be large difference in the boiling points of refrigerant and absorbent(greater than 200oC), so that only refrigerant is boiled-off in the generator. Thisensures that only pure refrigerant circulates through refrigerant circuit (condenser-expansion valve-evaporator) leading to isothermal heat transfer in evaporator andcondenser

3. It should exhibit small heat of mixing so that a high COP can be achieved. However,this requirement contradicts the first requirement. Hence, in practice a trade-off isrequired between solubility and heat of mixing.

4. The refrigerant-absorbent mixture should have high thermal conductivity and lowviscosity for high performance.

5. It should not undergo crystallization or solidification inside the system.

6. The mixture should be safe, chemically stable, non-corrosive, inexpensive and shouldbe available easily.

The most commonly used refrigerant-absorbent pairs in commercial systems are:


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i. Water-Lithium Bromide (H2O-LiBr) system for above 0oC applications such as airconditioning. Here water is the refrigerant and lithium bromide is the absorbent.

ii. Ammonia-Water (NH3-H2O) system for refrigeration applications with ammonia asrefrigerant and water as absorbent.

Of late efforts are being made to develop other refrigerant-absorbent systems using bothnatural and synthetic refrigerants to overcome some of the limitations of (H2O-LiBr) and(NH3-H2O) systems.

Currently, large water-lithium bromide (H2O-LiBr) systems are extensively used in airconditioning applications, where as large ammonia-water (NH3-H2O) systems are used inrefrigeration applications, while small ammonia-water systems with a third inert gas are usedin a pumpless form in small domestic refrigerators (triple fluid vapour absorption systems).

4.6 VARS based on Water- Lithium-Bromide Pair

Vapour absorption refrigeration systems using water-lithium bromide pair areextensively used in large capacity air conditioning systems. In these systems water is usedas refrigerant and a solution of lithium bromide in water is used as absorbent. Since water isused as refrigerant, using these systems it is not possible to provide refrigeration at sub-zerotemperatures. Hence it is used only in applications requiring refrigeration at temperaturesabove 0oC. Hence these systems are used for air conditioning applications. The analysis ofthis system is relatively easy as the vapour generated in the generator is almost purerefrigerant (water), unlike ammonia-water systems where both ammonia and water vapourare generated in the generator

Fig.15 H2O-LiBr-VARS frigeration withsolution system heat with exchanger solution


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4.6.1 Steady flow analysis of Water-Lithium Bromide Systems

Figure 15.5 shows the schematic of the system indicating various state points. Asteady flow analysis of the system is carried out with the following assumptions:

i. Steady state and steady flowii. Changes in potential and kinetic energies across each component are negligibleiii. No pressure drops due to frictioniv. Only pure refrigerant boils in the generator.

The nomenclature followed is:.m = mass flow rate of refrigerant, kg/s.mss = mass flow rate of strong solution (rich in LiBr), kg/s.m ws = mass flow rate of weak solution (weak in LiBr), kg/s

The circulation ratio (λ) is defined as the ratio of strong solution flow rate to refrigerant flowrate. It is given by:

.

λ = mss (15.7).

m

this implies that the strong solution flow rate is given by:. . (15.8)


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mss =λm

The analysis is carried out by applying mass and energy balance across each component.Condenser:

. . .

(15.9)m1=m 2 =m.

Qc =m(h1 −h 2 ) (15.10)Pc =Psat (Tc

) (15.11)

where Tc is the condenser temperature

Expansion valve (refrigerant):. . .

m 2 =m3 =mh2 = h3Evaporator:. . .

m3 =m4 =m.

Qe =m(h4 −h3 )Pe =Psat (Te )

where Te is the evaporator temperature Absorber:

From total mass balance:m +mss = m ws

. . . .

mss =λm m ws =(1 +λ) m

From mass balance for pure water:

. . .

m + (1 − ξss)mss =(1− ξws)mws

λ =

ξws

ξss

− ξws

. . .

Qa =m h 4 +λmh10 −(1+λ) m h 5

.

or, Qa =m[(h 4 −h 5 ) +λ(h10 −h 5 )]

The first term in the above equation m(h 4 −h 5 ) represents the enthalpy change of water aschanges its state from vapour at state 4 to liquid at state 5. The second term

.

m λ(h10 −h 5 ) represents the sensible heat transferred as solution at state 10 is cooled tosolution at state 5.


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Solution pump:

. . .

(15.21)m5 =m 6 =m ws. .

WP =m ws (h 6 −h 5 )=(1 +λ) m(h 6 −h 5 ) (15.22)

however, if we assume the solution to be incompressible, then:


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THERMOELECTRIC REFRIGERATIONThermoelectric Refrigeration: Basics and Principle. Defining the figure of Merit. (NoProblem)

COMPARISON BETWEEN GAS CYCLES AND VAPOR CYCLES:Thermodynamic


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MODULE - III

PreparedBy


Gandhi Institute for Technological Advancement(GITA),

Bhubaneswar – 752054June 2016


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REFRIGERANTS

Refrigerants : Classification of refrigerants and its degignation- Halocarbon (compounds,Hydrocarbons, Inorganic compounds, Azeotropes, Properties of refrigerants, comparisonof common refrigerants, uses of important refrigerants, Brines. Alternative refrigerants(Organic and inorganic compounds).

6.1. INTRODUCTION

The thermodynamic efficiency of a refrigeration system depends mainly onits operating temperatures. However, important practical issues such as thesystem design, size, initial and operating costs, safety, reliability, andserviceability etc. depend very much on the type of refrigerant selected for agiven application. Due to several environmental issues such as ozone layerdepletion and global warming and their relation to the various refrigerants used,the selection of suitable refrigerant has become one of the most important issuesin recent times. Replacement of an existing refrigerant by a completely newrefrigerant, for whatever reason, is an expensive proposition as it may call forseveral changes in the design and manufacturing of refrigeration systems. Henceit is very important to understand the issues related to the selection and use ofrefrigerants. In principle, any fluid can be used as a refrigerant. Air used in an aircycle refrigeration system can also be considered as a refrigerant. However, inthis lecture the attention is mainly focused on those fluids that can be used asrefrigerants in vapour compression refrigeration systems only.

6.2. PRIMARY AND SECONDARY REFRIGERANTS

Fluids suitable for refrigeration purposes can be classified into primary andsecondary refrigerants. Primary refrigerants are those fluids, which are useddirectly as working fluids, for example in vapour compression and vapourabsorption refrigeration systems. When used in compression or absorptionsystems, these fluids provide refrigeration by undergoing a phase changeprocess in the evaporator. As the name implies, secondary refrigerants are thoseliquids, which are used for transporting thermal energy from one location to other.Secondary refrigerants are also known under the name brines or antifreezes. Ofcourse, if the operating temperatures are above 0oC, then pure water can also beused as secondary refrigerant, for example in large air conditioning systems.Antifreezes or brines are used when refrigeration is required at sub-zerotemperatures. Unlike primary refrigerants, the secondary refrigerants do notundergo phase change as they transport energy from one location to other. Animportant property of a secondary refrigerant is its freezing point. Generally, thefreezing point of a brine will be lower than the freezing point of its constituents.The temperature at which freezing of a brine takes place its depends on itsconcentration. The concentration at which a lowest temperature can be reachedwithout solidification is called as eutectic point. The commonly used secondaryrefrigerants are the solutions of water and ethylene glycol, propylene glycol orcalcium chloride. These solutions are known under the general name of brines.

In this lecture attention is focused on primary refrigerants used mainly invapour compression refrigeration systems. As discussed earlier, in an absorptionrefrigeration system, a refrigerant and absorbent combination is used as theworking fluid.Prepared By : Prof. (Dr.) Manmatha K. Roul || 2

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6.3. REFRIGERANT SELECTION CRITERIA

Selection of refrigerant for a particular application is based on thefollowing requirements:

i. Thermodynamic and thermo-physical propertiesii. Environmental and safety properties, andiii. Economics

6.3.1. THERMODYNAMIC AND THERMO-PHYSICAL PROPERTIES

The requirements are:

a) Suction pressure: At a given evaporator temperature, the saturation pressureshould be above atmospheric for prevention of air or moisture ingress into thesystem and ease of leak detection. Higher suction pressure is better as it leads tosmaller compressor displacement

b) Discharge pressure: At a given condenser temperature, the dischargepressure should be as small as possible to allow light-weight construction ofcompressor, condenser etc.

c) Pressure ratio: Should be as small as possible for high volumetric efficiencyand low power consumption

d) Latent heat of vaporization: Should be as large as possible so that therequired mass flow rate per unit cooling capacity will be small

The above requirements are somewhat contradictory, as the operatingpressures, temperatures and latent heat of vaporization are related by Clausius-Clapeyron Equation:

ln(P ) =−h

fg +s

fg (26.1)sat RT R

In the above equation, Psat is the saturation pressure (in atm.) at a temperatureT(in Kelvin), hfg and sfg are enthalpy and entropy of vaporization and R is the gasconstant. Since the change in entropy of vaporization is relatively small, from theabove equation it can be shown that:

P hfg 1 1c =exp − (26.2)

P R T Te e c

In the above equation, Pc and Pe are the condenser and evaporatorpressures, Tc and Te are condenser and evaporator temperatures. From theabove equation, it can be seen that for given condenser and evaporatortemperatures as the latent heat of vaporization increases, the pressure ratio alsoincreases. Hence a trade-off is required between the latent heat of vaporizationand pressure ratio.

In addition to the above properties; the following properties are alsoimportant:Prepared By : Prof. (Dr.) Manmatha K. Roul || 3

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e) Isentropic index of compression: Should be as small as possible so that thetemperature rise during compression will be small

f) Liquid specific heat: Should be small so that degree of subcooling will be largeleading to smaller amount of flash gas at evaporator inlet

g) Vapour specific heat: Should be large so that the degree of superheating willbe small

h) Thermal conductivity: Thermal conductivity in both liquid as well as vapourphase should be high for higher heat transfer coefficients

i) Viscosity: Viscosity should be small in both liquid and vapour phases forsmaller frictional pressure drops

The thermodynamic properties are interrelated and mainly depend onnormal boiling point, critical temperature, molecular weight and structure.

The normal boiling point indicates the useful temperature levels as it isdirectly related to the operating pressures. A high critical temperature yieldshigher COP due to smaller compressor superheat and smaller flash gas losses.On the other hand since the vapour pressure will be low when criticaltemperature is high, the volumetric capacity will be lower for refrigerants with highcritical temperatures. This once again shows a need for trade-off between highCOP and high volumetric capacity. It is observed that for most of the refrigerantsthe ratio of normal boiling point to critical temperature is in the range of 0.6 to 0.7.Thus the normal boiling point is a good indicator of the critical temperature of therefrigerant.

The important properties such as latent heat of vaporization and specificheat depend on the molecular weight and structure of the molecule. Trouton’srule shows that the latent heat of vaporization will be high for refrigerants havinglower molecular weight. The specific heat of refrigerant is related to the structureof the molecule. If specific heat of refrigerant vapour is low then the shape of thevapour dome will be such that the compression process starting with a saturatedpoint terminates in the superheated zone (i.e, compression process willbedry). However, a small value of vapour specific heat indicates higher degreeofsuperheat. Since vapour and liquid specific heats are also related, a large valueof vapour specific heat results in a higher value of liquid specific heat, leading tohigher flash gas losses. Studies show that in general the optimum value of molarvapour specific heat lies in the range of 40 to 100 kJ/kmol.K.

The freezing point of the refrigerant should be lower than the lowestoperating temperature of the cycle to prevent blockage of refrigerant pipelines.

6.3.2. ENVIRONMENTAL AND SAFETY PROPERTIES

Next to thermodynamic and thermophysical properties, the environmentaland safety properties are very important. In fact, at present the environmentfriendliness of the refrigerant is a major factor in deciding the usefulness of aparticular refrigerant. The important environmental and safety properties are:

a) Ozone Depletion Potential (ODP): According to the Montreal protocol, the ODPof refrigerants should be zero, i.e., they should be non-ozone depletingsubstances. Refrigerants having non-zero ODP have either already been


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phased-out (e.g. R 11, R 12) or will be phased-out in near-future(e.g. R22). SinceODP depends mainly on the presence of chlorine or bromine in the molecules,refrigerants having either chlorine (i.e., CFCs and HCFCs) or bromine cannot beused under the new regulations

b) Global Warming Potential (GWP): Refrigerants should have as low a GWPvalue as possible to minimize the problem of global warming. Refrigerants withzero ODP but a high value of GWP (e.g. R134a) are likely to be regulated infuture.

c) Total Equivalent Warming Index (TEWI): The factor TEWI considers bothdirect (due to release into atmosphere) and indirect (through energyconsumption) contributions of refrigerants to global warming. Naturally,refrigerants with as a low a value of TEWI are preferable from global warmingpoint of view.

d) Toxicity: Ideally, refrigerants used in a refrigeration system should be non-toxic. However, all fluids other than air can be called as toxic as they will causesuffocation when their concentration is large enough. Thus toxicity is a relativeterm, which becomes meaningful only when the degree of concentration and timeof exposure required to produce harmful effects are specified. Some fluids aretoxic even in small concentrations. Some fluids are mildly toxic, i.e., they aredangerous only when the concentration is large and duration of exposure is long.Some refrigerants such as CFCs and HCFCs are non-toxic when mixed with airin normal condition. However, when they come in contact with an open flame oran electrical heating element, they decompose forming highly toxic elements (e.g.phosgene-COCl2). In general the degree of hazard depends on:

- Amount of refrigerant used vs total space- Type of occupancy- Presence of open flames- Odor of refrigerant, and- Maintenance condition

Thus from toxicity point-of-view, the usefulness of a particular refrigerantdepends on the specific application.

e) Flammability: The refrigerants should preferably be non-flammable and non-explosive. For flammable refrigerants special precautions should be taken toavoid accidents.

Based on the above criteria, ASHRAE has divided refrigerants into six safetygroups (A1 to A3 and B1 to B3). Refrigerants belonging to Group A1 (e.g. R11,R12, R22, R134a, R744, R718) are least hazardous, while refrigerants belongingto Group B3 (e.g. R1140) are most hazardous.

Other important properties are:

f) Chemical stability: The refrigerants should be chemically stable as long as theyare inside the refrigeration system.

g) Compatibility with common materials of construction (both metals and non-metals)

h) Miscibility with lubricating oils: Oil separators have to be used if the refrigerantis not miscible with lubricating oil (e.g. ammonia). Refrigerants that are


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completely miscible with oils are easier to handle (e.g. R12). However, forrefrigerants with limited solubility (e.g. R 22) special precautions should be takenwhile designing the system to ensure oil return to the compressori) Dilelectric strength: This is an important property for systems using hermeticcompressors. For these systems the refrigerants should have as high a dielectricstrength as possible

j) Ease of leak detection:Inthe event of leakage of refrigerant from the system,itshould be easy to detect the leaks.

6.3.3. ECONOMIC PROPERTIES

The refrigerant used should preferably be inexpensive and easilyavailable.

6.4. DESIGNATION OF REFRIGERANTS

Figure 26.1 shows the classification of fluids used as refrigerants in vapourcompression refrigeration systems. Since a large number of refrigerants havebeen developed over the years for a wide variety of applications, a numberingsystem has been adopted to designate various refrigerants. From the numberone can get some useful information about the type of refrigerant, its chemicalcomposition, molecular weight etc. All the refrigerants are designated by Rfollowed by a unique number.

i) Fully saturated, halogenated compounds: These refrigerants are derivativesof alkanes (CnH2n+2) such as methane (CH4), ethane (C2H6). These refrigerantsare designated by R XYZ, where:

X+1 indicates the number of Carbon (C) atoms Y-

1 indicates number of Hydrogen (H) atoms, and

Z indicates number of Fluorine (F) atoms

The balance indicates the number of Chlorine atoms. Only 2 digits indicates thatthe value of X is zero.

Ex: R 22X = 0 No. of Carbon atoms = 0+1 = 1 derivative of methane (CH4)

Y = 2 No. of Hydrogen atoms = 2-1 = 1Z = 2 No. of Fluorine atoms = 2

The balance = 4 – no. of (H+F) atoms = 4-1-2 = 1 No. of Chlorine atoms = 1

The chemical formula of R 22 = CHClF2

Similarly it can be shown that the chemical formula of:

R12 = CCl2F2

R134a = C2H2F4 (derivative of ethane)

(letter a stands for isomer, e.g. molecules having same chemical composition butdifferent atomic arrangement, e.g. R134 and R134a)Prepared By : Prof. (Dr.) Manmatha K. Roul || 6

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ii) Inorganic refrigerants: These are designated by number 7 followed bythemolecular weight of the refrigerant (rounded-off).

Ex.: Ammonia: Molecular weight is 17, the designation is R 717

Carbon dioxide: Molecular weight is 44, the designation is R 744

Water: Molecular weight is 18, the designation is R 718


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Refrigerants

Pure fluids Mixtures- Azeotropic- Zeotropic

Synthetic Natural- CFCs - Organic (HCs)- HCFCs - Inorganic- HFCs o NH3

o CO2

o H2O

Fig.26.1: Classification of fluids used as refrigerants

iii) Mixtures: Azeotropic mixtures are designated by 500 series, whereaszeotropic refrigerants (e.g. non-azeotropic mixtures) are designated by 400series.

Azeotropic mixtures:

R 500: Mixture of R 12 (73.8 %) and R 152a (26.2%)

R 502: Mixture of R 22 (48.8 %) and R 115 (51.2%)

R503: Mixture of R 23 (40.1 %) and R 13 (59.9%)

R507A: Mixture of R 125 (50%) and R 143a (50%)

Zeotropic mixtures:

R404A : Mixture of R 125 (44%), R 143a (52%) and R 134a (4%)

R407A : Mixture of R 32 (20%), R 125 (40%) and R 134a (40%)

R407B : Mixture of R 32 (10%), R 125 (70%) and R 134a (20%)

R410A : Mixture of R 32 (50%) and R 125 (50%)


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iv) Hydrocarbons:

Propane (C3H8) n-

butane (C4H10)

iso-butane (C4H10)

: R 290

: R 600

: R 600a

Unsaturated Hydrocarbons: R1150 (C2H4)

R1270 (C3H6)

6.5. COMPARISON BETWEEN DIFFERENTREFRIGERANTS

Synthetic refrigerants that were commonly used for refrigeration, cold storageand air conditioning applications are: R 11 (CFC 11), R 12 (CFC 12), R 22(HCFC 22), R 502 (CFC 12+HCFC 22) etc. However, these refrigerants have tobe phased out due to their Ozone Depletion Potential (ODP). The syntheticreplacements for the older refrigerants are: R-134a (HFC-134a) and blends ofHFCs. Generally, synthetic refrigerants are non-toxic and non-flammable.However, compared to the natural refrigerants the synthetic refrigerants offerlower performance and they also have higher Global Warming Potential (GWP).As a result, the synthetic refrigerants face an uncertain future. The mostcommonly used natural refrigerant is ammonia. This is also one of the oldestknown refrigerants. Ammonia has good thermodynamic, thermophysical andenvironmental properties. However, it is toxic and is not compatible with some ofthe common materials of construction such as copper, which somewhat restrictsits application. Other natural refrigerants that are being suggested arehydrocarbons (HCs) and carbon di-oxide (R-744). Though these refrigerantshave some specific problems owing to their eco-friendliness, they are beingstudied widely and are likely to play a prominent role in future.

Prior to the environmental issues of ozone layer depletion and globalwarming, the most widely used refrigerants were: R 11, R 12, R 22, R 502 andammonia. Of these, R 11 was primarily used with centrifugal compressors in airconditioning applications. R 12 was used primarily in small capacity refrigerationand cold storage applications, while the other refrigerants were used in largesystems such as large air conditioning plants or cold storages. Among therefrigerants used, except ammonia, all the other refrigerants are syntheticrefrigerants and are non-toxic and non-flammable. Though ammonia is toxic, ithas been very widely used due to its excellent thermodynamic andthermophysical properties. The scenario changed completely after the discoveryof ozone layer depletion in 1974. The depletion of stratospheric ozone layer wasattributed to chlorine and bromine containing chemicals such as Halons, CFCs,HCFCs etc. Since ozone layer depletion could lead to catastrophe on a globallevel, it has been agreed by the global community to phase out the ozonedepleting substances (ODS). As a result except ammonia, all the otherrefrigerants used in cold storages had to be phased-out and a search for suitablereplacements began in earnest. At the same time, it was also observed that inaddition to ozone layer depletion, most of the conventional synthetic refrigerantsalso cause significant global warming. In view of the environmental problemscaused by the synthetic refrigerants, opinions differed on replacements forconventional refrigerants. The alternate refrigerants can be classified into twobroad groups:Prepared By : Prof. (Dr.) Manmatha K. Roul || 9

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i) Non-ODS, synthetic refrigerants based on Hydro-Fluoro-Carbons(HFCs) and their blends

ii) Natural refrigerants including ammonia, carbon dioxide, hydrocarbonsand their blends

It should be noted that the use of natural refrigerants such as carbon dioxide,hydrocarbons is not a new phenomena, but is a revival of the once-used-and-discarded technologies in a much better form. Since the natural refrigerants areessentially making a comeback, one advantage of using them is that they arefamiliar in terms of their strengths and weaknesses. Another important advantageis that they are completely environment friendly, unlike the HFC basedrefrigerants, which do have considerable global warming potential. The alternatesynthetic refrigerants are normally non-toxic and non-flammable. It is alsopossible to use blends of various HFCs to obtain new refrigerant mixtures withrequired properties to suit specific applications. However, most of these blendsare non-azeotropic in nature, as a result there could be significant temperatureglides during evaporation and condensation, and it is also important takeprecautions to prevent leakage, as this will change the composition of themixture. Table 26.1 shows a list of refrigerants being replaced and theirreplacements.


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Refrigerant Application Substitute suggestedRetrofit(R)/New (N)

R 11(CFC) Large air conditioning systems R 123 (R,N)NBP = 23.7oC Industrial heat pumps

R 141b (N)hfg at NBP=182.5 kJ/kg As foam blowing agentTcr =197.98oC R 245fa (N)Cp/Cv = 1.13

n-pentane (R,N)ODP = 1.0GWP = 3500R 12 (CFC) Domestic refrigerators R 22 (R,N)NBP = -29.8oC Small air conditioners R 134a (R,N)hfg at NBP=165.8 kJ/kg Water coolers R 227ea (N)Tcr =112.04oC Small cold storages R 401A,R 401B (R,N)Cp/Cv = 1.126 R 411A,R 411B (R,N)ODP = 1.0 R 717 (N)GWP = 7300R 22 (HCFC) Air conditioning systems R 410A, R 410B (N)NBP = -40.8oC Cold storages R 417A (R,N)hfg at NBP=233.2 kJ/kg R 407C (R,N)Tcr =96.02oC R 507,R 507A (R,N)Cp/Cv = 1.166 R 404A (R,N)ODP = 0.05 R 717 (N)GWP = 1500R 134a (HFC) Used as replacement for R 12 No replacement requiredNBP = -26.15oC in domestic refrigerators, waterhfg at NBP=222.5 kJ/kg coolers, automobile A/Cs etc * Immiscible in mineral oilsTcr =101.06oC * Highly hygroscopicCp/Cv = 1.102ODP = 0.0GWP = 1200R 717 (NH3) Cold storages No replacement requiredNBP = -33.35oC Ice plantshfg at NBP=1368.9 kJ/kg Food processing * Toxic and flammableTcr =133.0oC Frozen food cabinets * Incompatible with copperCp/Cv = 1.31 * Highly efficientODP = 0.0 * Inexpensive and availableGWP = 0.0R 744 (CO2) Cold storages No replacement requiredNBP = -78.4oC Air conditioning systems * Very low critical temperaturehfg at 40oC=321.3 kJ/kg Simultaneous cooling and * Eco-friendlyTcr =31.1oC heating (Transcritical cycle) * Inexpensive and availableCp/Cv = 1.3ODP = 0.0GWP = 1.0

Table 26.1 : Refrigerants, their applications and substitutes


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Refrigerant Application Substitute suggestedRetrofit(R)/New (N)

R718 (H2O) Absorption systems No replacement requiredNBP = 100.oC Steam jet systems * High NBPhfg at NBP=2257.9 kJ/kg * High freezing pointTcr =374.15oC * Large specific volumeCp/Cv = 1.33 * Eco-friendlyODP = 0.0 * Inexpensive and availableGWP = 1.0R600a (iso-butane) Replacement for R 12 No replacement requiredNBP = -11.73oC Domestic refrigerators * Flammablehfg at NBP=367.7 kJ/kg Water coolers * Eco-friendlyTcr =135.0oCCp/Cv = 1.086ODP = 0.0GWP = 3.0

Table 26.1: Refrigerants, their applications and substitutes (contd.)


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PSYCHROMETRICSPsychrometrics : Properties of air-vapour mixture, Law of water vapour-air mixture,Enthalpy of moisture, Psychrometric chart, simple heating and cooling, Humidification, De-humidification, Mixture of air streams. Review question and discussionsRequirements of comfort air conditioning : Oxygen supply, Heat removal, moistureremoval, air motion, purity of air, Thermodynamics of human body, comfort and comfortchart, effective temperature, factors governing optimum effective temperature

7.1. Introduction

Atmospheric airmakes up the environment in almost every type ofairconditioning system. Hence a thorough understanding of the properties ofatmospheric air and the ability to analyze various processes involving air isfundamental to air conditioning design.

Psychrometryis the study of the properties of mixtures of air and watervapour.

Atmospheric air is a mixture of many gases plus water vapour and a number ofpollutants (Fig.27.1). The amount of water vapour and pollutants vary from place toplace. The concentration of water vapour and pollutants decrease with altitude, andabove an altitude of about 10 km, atmospheric air consists of only dry air. Thepollutants have to be filtered out before processing the air. Hence, what we process isessentially a mixture of various gases that constitute air and water vapour. This mixtureis known as moist air.

The moist air can be thought of as a mixture of dry air and moisture. For allpractical purposes, the composition of dry air can be considered as constant. In1949, a standard composition of dry air was fixed by the International JointCommittee on Psychrometric data. It is given in Table 27.1.

Constituent Molecular weight Mol fractionOxygen 32.000 0.2095Nitrogen 28.016 0.7809Argon 39.944 0.0093Carbon dioxide 44.010 0.0003

Table 27.1: Composition of standard air

Based on the above composition the molecular weight of dry air is found to be28.966 and the gas constant R is 287.035 J/kg.K.

As mentioned before the air to be processed in air conditioning systems is amixture of dry air and water vapour. While the composition of dry air is constant, theamount of water vapour present in the air may vary from zero to a maximumdepending upon the temperature and pressure of the mixture (dry air + watervapour).

At a given temperature and pressure the dry air can only hold a certainmaximum amount of moisture. When the moisture content is maximum, then the airis known as saturated air , which is established by a neutral equilibrium betweenthemoist air and the liquid or solid phases of water.

For calculation purposes, the molecular weight of water vapour is taken as18.015andits gas constant is461.52 J/kg.K.


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Atmospheric air

Mixture of permanent Water vapour

gases (N2,O2,Ar,H2,…)Dust particles,

fumes etc

After filtration

Mixture of permanentgases (N2,O2,Ar,H2,…)

Water vapour

Moist air forconditioning

Fig.27.1: Atmospheric air

7.2. METHODS FOR ESTIMATING PROPERTIES OF MOISTAIR

In order to perform air conditioning calculations, it is essential first to estimatevarious properties of air. It is difficult to estimate the exact property values of moistair as it is a mixture of several permanent gases and water vapour. However, moistair upto 3 atm. pressure is found to obey perfect gas law with accuracy sufficient forengineering calculations. For higher accuracy Goff and Gratch tables can be usedfor estimating moist air properties. These tables are obtained using mixture modelsbased on fundamental principles of statistical mechanics that take into account thereal gas behaviour of dry air and water vapour. However, these tables are valid for abarometric pressure of 1 atm. only. Even though the calculation procedure is quitecomplex, using the mixture models it is possible to estimate moist air properties at

Atmospheric air



fumes etc

After filtration


Water vapour




In order to perform air conditioning calculations, it is essential first to estimatevarious properties of air. It is difficult to estimate the exact property values of moistair as it is a mixture of several permanent gases and water vapour. However, moistair upto 3 atm. pressure is found to obey perfect gas law with accuracy sufficient forengineering calculations. For higher accuracy Goff and Gratch tables can be usedfor estimating moist air properties. These tables are obtained using mixture modelsbased on fundamental principles of statistical mechanics that take into account thereal gas behaviour of dry air and water vapour. However, these tables are valid for abarometric pressure of 1 atm. only. Even though the calculation procedure is quitecomplex, using the mixture models it is possible to estimate moist air properties at

Atmospheric air



fumes etc

After filtration


Water vapour




In order to perform air conditioning calculations, it is essential first to estimatevarious properties of air. It is difficult to estimate the exact property values of moistair as it is a mixture of several permanent gases and water vapour. However, moistair upto 3 atm. pressure is found to obey perfect gas law with accuracy sufficient forengineering calculations. For higher accuracy Goff and Gratch tables can be usedfor estimating moist air properties. These tables are obtained using mixture modelsbased on fundamental principles of statistical mechanics that take into account thereal gas behaviour of dry air and water vapour. However, these tables are valid for abarometric pressure of 1 atm. only. Even though the calculation procedure is quitecomplex, using the mixture models it is possible to estimate moist air properties atPrepared By : Prof. (Dr.) Manmatha K. Roul || 14

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other pressures also. However, since in most cases the pressures involved are low,one can apply the perfect gas model to estimate psychrometric properties.

7.2.1. BASIC GAS LAWS FOR MOIST AIR

According to the Gibbs-Dalton law for a mixture of perfect gases, the totalpressure exerted by the mixture is equal to the sum of partial pressures of theconstituent gases. According to this law, for a homogeneous perfect gas mixtureoccupying a volume V and at temperature T, each constituent gas behaves asthough the other gases are not present (i.e., there is no interaction between thegases). Each gas obeys perfect gas equation. Hence, the partial pressures exertedby each gas, p1,p2,p3 … and the total pressure pt are given by:

p1=n

1R

uT

; p2=n

2R

uT

; p3=n

3R

uT

......V V V

pt= p1+p2+p3+ ......

where n1,n2,n3,… are the number of moles of gases 1,2,3,…

Applying this equation to moist air.

p = pt= pa+pv

where p = pt = total barometric pressure pa =partial pressure of dry airpv = partial pressure of water vapour

27.2.2. Important psychrometric properties:

(27.1)

(27.2)

Dry bulb temperature (DBT) is the temperature of the moist air as measured by astandard thermometer or other temperature measuring instruments.

Saturated vapour pressure (psat) is the saturated partial pressure of water vapour atthe dry bulb temperature. This is readily available in thermodynamic tables andcharts. ASHRAE suggests the following regression equation for saturated vapourpressure of water, which is valid for 0 to 100oC.

ln(psat ) =c1

+ c2 + c3T + c4T2

+ c5T3

+ c6ln(T) (27.3)Twhere psat = saturated vapor pressure of water in kiloPascals

T = temperature in KThe regression coefficients c1 to c6 are given by:

c1 = -5.80022006E+03, c2 = -5.516256E+00, c3 = -4.8640239E-02c4 = 4.1764768E-05, c5 = -1.4452093E-08, c6 = 6.5459673E+00

Relative humidity (Φ)is defined as the ratio of the mole fraction of water vapourinmoist air to mole fraction of water vapour in saturated air at the same temperatureand pressure. Using perfect gas equation we can show that:


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φ = partial pressure of water vapour = pvsaturation pressure of pure water vapour at same temperature

psat

(27.4)

Relative humidity is normally expressed as a percentage. When Φ is 100percent, the air is saturated.

Humidity ratio (W): The humidity ratio (or specific humidity) W is the mass of waterassociated with each kilogram of dry air1. Assuming both water vapour and dry air tobe perfect gases2, the humidity ratio is given by:

W = kg of water vapour =pv V / RvT

=pv / Rv (27.5)

kg of dry air pa V / RaT (pt−pv ) / RaSubstituting the values of gas constants of water vapour and air Rv and Ra in

the above equation; the humidity ratio is given by:

pv

W = 0.622 pt−pv (27.6)For a given barometric pressure pt, given the DBT, we can find the saturated vapourpressure psat from the thermodynamic property tables on steam. Then using theabove equation, we can find the humidity ratio at saturated conditions, Wsat.

It is to be noted that, W is a function of both total barometric pressure andvapor pressure of water.

Dew-point temperature: If unsaturated moist air is cooled at constant pressure, thenthe temperature at which the moisture in the air begins to condense is known asdew-point temperature (DPT)of air. An approximate equation for dew-pointtemperature is given by:

DPT = 4030(DBT + 235) − 235 (27.7)4030 − (DBT + 235)lnφ

where Φ is the relative humidity (in fraction). DBT & DPT are in oC. Of course, sincefrom its definition, the dew point temperature is the saturation temperaturecorresponding to the vapour pressure of water vapour, it can be obtained from steamtables or using Eqn.(27.3).

1 Properties such as humidity ratio, enthalpy and specific volume are based on 1 kg of dryair. This is useful as the total mass of moist air in a process varies by the addition/removal ofwater vapour, but the mass of dry air remains constant.2 Dry air is assumed to be a perfect gas as its temperature is high relative to its saturationtemperature, and water vapour is assumed to be a perfect gas because its pressure is lowrelative to its saturation pressure. These assumptions result in accuracies, that are, sufficientfor engineering calculations (less than 0.7 percent as shown by Threlkeld). However, moreaccurate results can be obtained by using the data developed by Goff and Gratch in 1945.


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Degree of saturationμ: The degree of saturation is the ratio of the humidity ratio W tothe humidity ratio of a saturated mixture Ws at the same temperature and pressure,i.e.,

μ= W (27.8)Ws t,P

Enthalpy: The enthalpy of moist air is the sum of the enthalpy of the dry air and theenthalpy of the water vapour. Enthalpy values are always based on some referencevalue. For moist air, the enthalpy of dry air is given a zero value at 0oC, and forwater vapour the enthalpy of saturated water is taken as zero at 0oC.The enthalpy of moist air is given by:

h=ha+ Whg=cpt +W(hfg+ cpw t) (27.9)

where cp = specific heat of dry air at constant pressure, kJ/kg.Kcpw = specific heat of water vapor, kJ/kg.Kt = Dry-bulb temperature of air-vapor mixture, oC

W = Humidity ratio, kg of water vapor/kg of dry airha = enthalpy of dry air at temperature t, kJ/kghg = enthalpy of water vapor3 at temperature t, kJ/kg

hfg = latent heat of vaporization at 0oC, kJ/kgThe unit of h is kJ/kg of dry air. Substituting the approximate values of cp and hg, weobtain:

h =1.005 t + W(2501 + 1.88t) (27.10)

Humid specific heat: From the equation for enthalpy of moist air, the humid specificheat of moist air can be written as:

cpm= cp+ W.cpw (27.11)

where cpm = humid specific heat, kJ/kg.Kcp = specific heat of dry air, kJ/kg.Kcpw = specific heat of water vapor, kJ/kgW = humidity ratio, kg of water vapor/kg of dry air

Since the second term in the above equation (w.cpw) is very small comparedto the first term, for all practical purposes, the humid specific heat of moist air, cpmcan be taken as 1.0216 kJ/kg dry air.K

Specific volume: The specific volume is defined as the number of cubic meters ofmoist air per kilogram of dry air. From perfect gas equation since the volumesoccupied by the individual substances are the same, the specific volume is alsoequal to the number of cubic meters of dry air per kilogram of dry air, i.e.,

3 Though the water vapor in moist air is likely to be superheated, no appreciable error resultsif we assume it to be saturated. This is because of the fact that the constant temperaturelines in the superheated region on a Mollier chart (h vs s) are almost horizontal.


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v = RaT = RaT m3

/kg dry air (27.12)

pt−pvpa7.2.3. PSYCHROMETRIC CHART

A Psychrometric chart graphically represents the thermodynamic properties ofmoist air. Standard psychrometric charts are bounded by the dry-bulb temperatureline (abscissa) and the vapour pressure or humidity ratio (ordinate). The Left HandSide of the psychrometric chart is bounded by the saturation line. Figure 27.2 showsthe schematic of a psychrometric chart. Psychrometric charts are readily availablefor standard barometric pressure of 101.325 kPa at sea level and fornormaltemperatures (0-50oC). ASHRAE has also developed psychrometric chartsfor other temperatures and barometric pressures (for low temperatures: -40 to 10oC,high temperatures 10 to 120oC and very high temperatures 100 to 120oC)

Lines ofLines ofconstant RH constantsp.volume

Saturation curve(RH = 100%)

Lines ofconstant Wenthalpy (kgw/kgda)

DBT, oC

Fig.27.2: Schematic of a psychrometric chart for a given barometric pressure

7.3. MEASUREMENT OF PSYCHROMETRIC PROPERTIES

Based on Gibbs’ phase rule, the thermodynamic state of moist air is uniquelyfixed if the barometric pressure and two other independent properties are known.This means that at a given barometric pressure, the state of moist air can bedetermined by measuring any two independent properties. One of them could be thedry-bulb temperature (DBT), as the measurement of this temperature is fairly simpleand accurate. The accurate measurement of other independent parameters such ashumidity ratio is very difficult in practice. Since measurement of temperatures is


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easier, it would be convenient if the other independent parameter is also atemperature. Of course, this could be the dew-point temperature (DPT), but it isobserved that accurate measurement of dew-point temperature is difficult. In thiscontext, a new independent temperature parameter called the wet-bulb temperature(WBT) is defined. Compared to DPT, it is easier to measure the wet-bulbtemperature of moist air. Thus knowing the dry-bulb and wet-bulb temperatures frommeasurements, it is possible to find the other properties of moist air.

To understand the concept of wet-bulb temperature, it is essential tounderstand the process of combined heat and mass transfer.

7.3.1. COMBINED HEAT AND MASS TRANSFER; THE STRAIGHT LINE LAW

The straight line law states that “when air is transferring heat and mass(water)to or from a wetted surface, the condition of air shown on a psychrometric chartdrives towards the saturation line at the temperature of the wetted surface”.

For example, as shown in Fig.27.3, when warm air passes over a wettedsurface its temperature drops from 1 to 2. Also, since the vapor pressure of air at 1 isgreater than the saturated vapor pressure at tw, there will be moisture transfer fromair to water, i.e., the warm air in contact with cold wetted surface cools anddehumidifies. According to the straight line law, the final condition of air (i.e., 2) lieson a straight line joining 1 with tw on the saturation line. This is due to the value ofunity of the Lewis number, that was discussed in an earlier chapter on analogybetween heat and mass transfer.

Fig.27.3: Principle of straight-line law for air-water mixtures

7.3.2. Adiabatic saturation and thermodynamic wet bulb temperature

Adiabatic saturation temperature is defined as that temperature at whichwater, by evaporating into air, can bring the air to saturation at the same temperatureadiabatically. An adiabatic saturator is a device using which one can measuretheoretically the adiabatic saturation temperature of air.

As shown in Fig.27.4, an adiabatic saturator is a device in which air flowsthrough an infinitely long duct containing water. As the air comes in contact with


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water in the duct, there will be heat and mass transfer between water and air. If theduct is infinitely long, then at the exit, there would exist perfect equilibrium betweenair and water at steady state. Air at the exit would be fully saturated and itstemperature is equal to that of water temperature. The device is adiabatic as thewalls of the chamber are thermally insulated. In order to continue the process, make-up water has to be provided to compensate for the amount of water evaporated intothe air. The temperature of the make-up water is controlled so that it is the same asthat in the duct.

After the adiabatic saturator has achieved a steady-state condition, thetemperature indicated by the thermometer immersed in the water is thethermodynamic wet-bulb temperature. The thermodynamic wet bulb temperaturewillbe less than the entering air DBT but greater than the dew point temperature.

Certain combinations of air conditions will result in a given sump temperature,and this can be defined by writing the energy balance equation for the adiabaticsaturator. Based on a unit mass flow rate of dry air, this is given by:

h1=h2− (W2− W1)hf (27.13)

where hf is the enthalpy of saturated liquid at the sump or thermodynamic wet-bulbtemperature, h1 and h2 are the enthalpies of air at the inlet and exit of the adiabaticsaturator, and W1 and W2 are the humidity ratio of air at the inlet and exit of theadiabatic saturator, respectively.

It is to be observed that the thermodynamic wet-bulb temperature is athermodynamic property, and is independent of the path taken by air. Assuming thehumid specific heat to be constant, from the enthalpy balance, the thermodynamicwet-bulb temperature can be written as:

t 2 = t 1−h

fg,2 (w 2 − w ) (27.14)c 1

pmwhere hfg,2 is the latent heat of vaporization at the saturated condition 2. Thusmeasuring the dry bulb (t1) and wet bulb temperature (t2) one can find the inlethumidity ratio (W1) from the above expression as the outlet saturated humidity ratio(W2) and latent heat heat of vaporizations are functions of t2 alone (at fixedbarometric pressure).

On the psychrometric chart as shown in Fig.27.4, point 1 lies below the line ofconstant enthalpy that passes through the saturation point 2. t2 = f(t1,W1) is not aunique function, in the sense that there can be several combinations of t1 and W1which can result in the same sump temperature in the adiabatic saturator. A linepassing through all these points is a constant wet bulb temperature line. Thus allinlet conditions that result in the same sump temperature, for example point 1’ havethe same wet bulb temperature. The line is a straight line according to the straight-line law. The straight-line joining 1 and 2 represents the path of the air as it passesthrough the adiabatic saturator.


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Normally lines of constant wet bulb temperature are shown on thepsychrometric chart. The difference between actual enthalpy and the enthalpyobtained by following constant wet-bulb temperature is equal to (w2-w1)hf.

Perfect insulation

Moist air Moist air

t1,W1,p

Make-up water(W2-W1) per kgda

Fig.27.4: The process of adiabatic saturation of air

W2

W1Prepared By : Prof. (Dr.) Manmatha K. Roul || 21

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t2 t1

Fig.27.5: Adiabatic saturation process 1-2 on psychrometric chart


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7.3.3. Wet-Bulb Thermometer

In practice, it is not convenient to measure the wet-bulb temperature using anadiabatic saturator. In stead, a thermometer with a wetted wick is used to measurethe wet bulb temperature as shown in Fig.27.6. It can be observed that since thearea of the wet bulb is finite, the state of air at the exit of the wet bulb will not besaturated, in stead it will be point 2 on the straight line joining 1 and i, provided thetemperature of water on the wet bulb is i. It has been shown by Carrier, that this is avalid assumption for air-water mixtures. Hence for air-water mixtures, one canassume that the temperature measured by the wet-bulb thermometer is equal to thethermodynamic wet-bulb temperature4. For other gas-vapor mixtures, there can beappreciable difference between the thermodynamic and actual wet-bulbtemperatures.

W

Wet wick

DBT

Fig.27.6: Schematic of a wet-bulb thermometer and theprocesson psychrometric chart

4 By performing energy balance across the wet-bulb, it can be shown that, thetemperature measured by the wet-bulb thermometer is:t2= t1− (kw / hc )hfg(wi− w); where kw is the mass transfer coefficientfor air-water mixtures, the ratio (hc/kwcpm ) = Lewis number is ≈1, hence, the wicktemperature is approximately equal to the thermodynamic wet-bulb temperature.It should be noted that, unlike thermodynamic WBT, the WBT of wet bulbthermometer is not a thermodynamic property as it depends upon the rates of heatand mass transfer between the wick and air.


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7.4. Calculation of psychrometric properties from p, DBTand WBT

As mentioned before, to fix the thermodynamic state of moist air, we need toknow three independent properties. The properties that are relatively easier tomeasure, are: the barometric pressure, dry-bulb temperature and wet-bulbtemperature. For a given barometric pressure, knowing the dry bulb and wet bulbtemperatures, all other properties can be easily calculated from the psychrometricequations. The following are the empirical relations for the vapor pressure of water inmoist air:

i) Modified Apjohn equation:

pv= p,v−

1.8p(t − t,

)2700

ii) Modified Ferrel equation:p = p

,− 0.00066p(t − t

,) 1+ 1.8t

v v

iii) Carrier equation:1571

1.8(p −p,v )(t − t

,),

pv= pv − 2800 −1.3(1.8t + 32)

(27.15)

(27.16)

(27.17)

where t = dry bulb temperature, oCt’ = wet bulb temperature, oCp = barometric pressurepv = vapor pressurepv’ = saturation vapor pressure at wet-bulb temperature

The units of all the pressures in the above equations should be consistent.

Once the vapor pressure is calculated, then all other properties such asrelative humidity, humidity ratio, enthalpy, humid volume etc. can be calculated fromthe psychrometric equations presented earlier.

7.5. Psychrometer

Any instrument capable of measuring the psychrometric state of air is called apsychrometer. As mentioned before, in order to measure the psychrometric state ofair, it is required to measure three independent parameters. Generally two of theseare the barometric pressure and air dry-bulb temperature as they can be measuredeasily and with good accuracy.

Two types of psychrometers are commonly used. Each comprises of twothermometers with the bulb of one covered by a moist wick. The two sensing bulbsare separated and shaded from each other so that the radiation heat transferbetween them becomes negligible. Radiation shields may have to be used over thebulbs if the surrounding temperatures are considerably different from the airtemperature.


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The sling psychrometer is widely used for measurements involving room air orother applications where the air velocity inside the room is small. The slingpsychrometer consists of two thermometers mounted side by side and fitted in a framewith a handle for whirling the device through air. The required air circulation (≈ 3 to 5m/s) over the sensing bulbs is obtained by whirling the psychrometer (≈ 300 RPM).Readings are taken when both the thermometers show steady-state readings.

In the aspirated psychrometer, the thermometers remain stationary, and a smallfan, blower or syringe moves the air across the thermometer bulbs.

The function of the wick on the wet-bulb thermometer is to provide a thin film ofwater on the sensing bulb. To prevent errors, there should be a continuous film of wateron the wick. The wicks made of cotton or cloth should be replaced frequently, and onlydistilled water should be used for wetting it. The wick should extend beyond the bulb by1 or 2 cms to minimize the heat conduction effects along the stem.

Other types of psychrometric instruments:

1. Dunmore Electric Hygrometer2. DPT meter3. Hygrometer (Using horse’s or human hair)

7.6. IMPORTANT PSYCHROMETRIC PROCESSES

a) Sensible cooling:

During this process, the moisture content of air remains constant but itstemperature decreases as it flows over a cooling coil. For moisture content to remain


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constant, the surface of the cooling coil should be dry and its surface temperatureshould be greater than the dew point temperature of air. If the cooling coil is 100%effective, then the exit temperature of air will be equal to the coil temperature. However,in practice, the exit air temperature will be higher than the cooling coil temperature.Figure 28.1 shows the sensible cooling process O-A on a psychrometric chart. The heattransfer rate during this process is given by:

Qc= ma (hO−hA ) = macpm (TO− TA ) (28.1)

ho

hA

A OW

DBT

Fig.28.1: Sensible cooling process O-A on psychrometric chart

b) Sensible heating (Process O-B):

During this process, the moisture content of air remains constant and itstemperature increases as it flows over a heating coil. The heat transfer rate during thisprocess is given by:

Qh= ma (hB−hO ) = macpm (TB− TO ) (28.2)


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where cpm is the humid specific heat (≈1.0216 kJ/kg dry air) and ma is the mass flowrate of dry air (kg/s). Figure 28.2 shows the sensible heating process on apsychrometric chart.

hB

W

O B

DBTFig.28.2: Sensible heating process on psychrometric chart

c) Cooling and dehumidification (Process O-C):

When moist air is cooled below its dew-point by bringing it in contact with a coldsurface as shown in Fig.28.3, some of the water vapor in the air condenses and leavesthe air stream as liquid, as a result both the temperature and humidity ratio of airdecreases as shown. This is the process air undergoes in a typical air conditioningsystem. Although the actual process path will vary depending upon the type of coldsurface, the surface temperature, and flow conditions, for simplicity the process line isassumed to be a straight line. The heat and mass transfer rates can be expressed interms of the initial and final conditions by applying the conservation of mass andconservation of energy equations as given below:

By applying mass balance for the water:

ma.wO= ma.wC+ mw (28.3)


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hw

hOCooling coil

hC O Wo

ma ma C Wc

ho hC

Wo WC

Qt mwTs TC To

Fig.28.3: Cooling and dehumidification process (O-C)

By applying energy balance:

ma .hO= Q t+ mw .hw+ ma .hC (28.4)

from the above two equations, the load on the cooling coil, Qt is given by:

Qt= ma (hO−hC ) − ma (w O− w C )hw (28.5)

the 2nd term on the RHS of the above equation is normally small compared to the otherterms, so it can be neglected. Hence,

Q t= ma (hO−hC ) (28.6)

It can be observed that the cooling and de-humidification process involves both latentand sensible heat transfer processes, hence, the total, latent and sensible heat transferrates (Qt, Ql and Qs) can be written as:

Qt= Ql+ Qswhere Ql= ma (hO−hw ) = ma .hfg (w O− w C ) (28.7)

Qs= ma (hw−hC ) = ma .cpm (TO− TC )

By separating the total heat transfer rate from the cooling coil into sensible andlatent heat transfer rates, a useful parameter called Sensible Heat Factor (SHF) isdefined. SHF is defined as the ratio of sensible to total heat transfer rate, i.e.,

SHF = Qs / Qt= Qs /(Qs+ Ql ) (28.8)

From the above equation, one can deduce that a SHF of 1.0 corresponds to nolatent heat transfer and a SHF of 0 corresponds to no sensible heat transfer. A SHF of0.75 to 0.80 is quite common in air conditioning systems in a normal dry-climate. A


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lower value of SHF, say 0.6, implies a high latent heat load such as that occurs in ahumid climate.

From Fig.28.3, it can be seen that the slope of the process line O-C is given by:

tan c = w (28.9)T

From the definition of SHF,

1− SHF = Q l =m

ah

fg w = 2501 w = 2451 w (28.10)

SHFm

ac

pm T 1.0216 T TQsFrom the above equations, we can write the slope as:

tan c = 1 1− SHF (28.11)2451 SHF

Thus we can see that the slope of the cooling and de-humidification line is purelya function of the sensible heat factor, SHF. Hence, we can draw the cooling and de-humidification line on psychrometric chart if the initial state and the SHF are known. Insome standard psychrometric charts, a protractor with different values of SHF isprovided. The process line is drawn through the initial state point and in parallel to thegiven SHF line from the protractor as shown in Fig.28.4.

S H Fh

o-h

c

-hc

hw

Wo-

Wc

c

Fig.28.4: A psychrometric chart with protractor for SHF lines

In Fig.28.3, the temperature Ts is the effective surface temperature of the coolingcoil, and is known as apparatus dew-point (ADP) temperature. In an ideal situation,when all the air comes in perfect contact with the cooling coil surface, then the exittemperature of air will be same as ADP of the coil. However, in actual case the exittemperature of air will always be greater than the apparatus dew-point temperature dueto boundary layer development as air flows over the cooling coil surface and also due to


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temperature variation along the fins etc. Hence, we can define a by-pass factor (BPF)as:

BPF = TC− TS (28.12)T − T

O SIt can be easily seen that, higher the by- pass factor larger will be the difference

between air outlet temperature and the cooling coil temperature. When BPF is 1.0, allthe air by-passes the coil and there will not be any cooling or de-humidification. Inpractice, the by-pass factor can be increased by increasing the number of rows in acooling coil or by decreasing the air velocity or by reducing the fin pitch.

Alternatively, a contact factor(CF) can be defined which is given by:

CF = 1−BPF (28.13)

d) Heating and Humidification (Process O-D):

During winter it is essential to heat and humidify the room air for comfort. Asshown in Fig.28.5., this is normally done by first sensibly heating the air and thenadding water vapour to the air stream through steam nozzles as shown in the figure.

Heating coil Steam nozzleshD

hO DwDma ma

TO TDwO wD

O

wO

hO hD

Qhmw

TO TD

Fig.28.5: Heating and humidification process

Mass balance of water vapor for the control volume yields the rate at which steam hasto be added, i.e., mw:

mw= ma (wD− w O ) (28.14)


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where ma is the mass flow rate of dry air.

From energy balance:

Qh= ma (hD− hO ) − mwhw (28.15)

where Qh is the heat supplied through the heating coil and hw is the enthalpy of steam.

Since this process also involves simultaneous heat and mass transfer, we candefine a sensible heat factor for the process in a way similar to that of a coolind anddehumidification process.

e) Cooling & humidification (Process O-E):

As the name implies, during this process, the air temperature drops and itshumidity increases. This process is shown in Fig.28.6. As shown in the figure, this canbe achieved by spraying cool water in the air stream. The temperature of water shouldbe lower than the dry-bulb temperature of air but higher than its dew-point temperatureto avoid condensation (TDPT< Tw< TO).

Cold water spray ora wetted surface

ma ma wETO TEwO wE wO

hO hE

Tw TDPT TE TO

Fig.28.6: Cooling and humdification process

It can be seen that during this process there is sensible heat transfer from air towater and latent heat transfer from water to air. Hence, the total heat transfer dependsupon the water temperature. If the temperature of the water sprayed is equal to the wet-bulb temperature of air, then the net transfer rate will be zero as the sensible heattransfer from air to water will be equal to latent heat transfer from water to air. If thewater temperature is greater than WBT, then there will be a net heat transfer from waterto air. If the water temperature is less than WBT, then the net heat transfer will be fromair to water. Under a special case when the spray water is entirely recirculated and isneither heated nor cooled, the system is perfectly insulated and the make-up water issupplied at WBT, then at steady-state, the air undergoes an adiabatic saturationprocess, during which its WBT remains constant. This is the process of adiabaticsaturation discussed in Chapter 27. The process of cooling and humidification isencountered in a wide variety of devices such as evaporative coolers, cooling towersetc.


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f) Heating and de-humidification (Process O-F):

This process can be achieved by using a hygroscopic material, which absorbs oradsorbs the water vapor from the moisture. If this process is thermally isolated, then theenthalpy of air remains constant, as a result the temperature of air increases as itsmoisture content decreases as shown in Fig.28.7. This hygroscopic material can be asolid or a liquid. In general, the absorption of water by the hygroscopic material is anexothermic reaction, as a result heat is released during this process, which istransferred to air and the enthalpy of air increases.

Hygroscopicmaterial

O WOO F

F WF

TO TF

Fig.28.7. Chemical de-humidification process

g) Mixing of air streams:

Mixing of air streams at different states is commonly encountered in manyprocesses, including in air conditioning. Depending upon the state of the individualstreams, the mixing process can take place with or without condensation of moisture.

i) Without condensation: Figure 28.8 shows an adiabatic mixing of two moist airstreams during which no condensation of moisture takes place. As shown in the figure,when two air streams at state points 1 and 2 mix, the resulting mixture condition 3 canbe obtained from mass and energy balance.

From the mass balance of dry air and water vapor:

ma,1

w1+m

a,2w

2=m

a,3w

3=(m

a,1+m

a,2)w

3 (28.16)From energy balance:

ma,1

h1+m

a,2h

2=m

a,3h

3=(m

a,1+m

a,2)h

3 (28.17)From the above equations, it can be observed that the final enthalpy and

humidity ratio of mixture are weighted averages of inlet enthalpies and humidity ratios.A generally valid approximation is that the final temperature of the mixture is the


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HUBANESWAR

weighted average of the inlet temperatures. With this approximation, the point on thepsychrometric chart representing the mixture lies on a straight line connecting the twoinlet states. Hence, the ratio of distances on the line, i.e., (1-3)/(2-3) is equal to the ratioof flow rates ma,2/ma,1. The resulting error (due to the assumption that the humidspecific heats being constant) is usually less than 1 percent.

ma,1

ma,1+ma,2=ma,3

ma,2

Fig.28.8. Mixing of two air streams without condensation

ii) Mixing with condensation:

As shown in Fig.28.9, when very cold and dry air mixes with warm air at highrelative humidity, the resulting mixture condition may lie in the two-phase region, as aresult there will be condensation of water vapor and some amount of water will leavethe system as liquid water. Due to this, the humidity ratio of the resulting mixture (point3) will be less than that at point 4. Corresponding to this will be an increase intemperature of air due to the release of latent heat of condensation. This process rarelyoccurs in an air conditioning system, but this is the phenomenon which results in theformation of fog or frost (if the mixture temperature is below 0oC). This happens inwinter when the cold air near the earth mixes with the humid and warm air, whichdevelops towards the evening or after rains.

Fig.28.9. Mixing of two air streams with condensation


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7.7. AIR WASHERSAn air washer is a device for conditioning air. As shown in Fig.28.10, in an air

washer air comes in direct contact with a spray of water and there will be an exchangeof heat and mass (water vapour) between air and water. The outlet condition of airdepends upon the temperature of water sprayed in the air washer. Hence, bycontrolling the water temperature externally, it is possible to control the outlet conditionsof air, which then can be used for air conditioning purposes.

EliminatorPlates

Air in Air out

Make-upwater

PumpCooler/heater

Fig.28.10: Air washer

In the air washer, the mean temperature of water droplets in contact with airdecides the direction of heat and mass transfer. As a consequence of the 2nd law, theheat transfer between air and water droplets will be in the direction of decreasingtemperature gradient. Similarly, the mass transfer will be in the direction of decreasingvapor pressure gradient. For example,

3 Cooling and dehumidification: tw< tDPT. Since the exit enthalpy of air is less than itsinlet value, from energy balance it can be shown that there is a transfer of total energyfrom air to water. Hence to continue the process, water has to be externally cooled.Here both latent and sensible heat transfers are from air to water. This is shown byProcess O-A in Fig.28.11.

4 Adiabatic saturation: tw = tWBT. Here the sensible heat transfer from air to water isexactly equal to latent heat transfer from water to air. Hence, no external cooling orheating of water is required. That is this is a case of pure water recirculation. This is


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HUBANESWAR

shown by Process O-B in Fig.28.11. This the process that takes place in a perfectlyinsulated evaporative cooler.

4. Cooling and humidification: tDPT< tw< tWBT. Here the sensible heat transfer is from air towater and latent heat transfer is from water to air, but the total heat transfer is from air towater, hence, water has to be cooled externally. This is shown by Process O-C inFig.28.11.

5. Cooling and humidification: tWBT< tw< tDBT. Here the sensible heat transfer is fromair to water and latent heat transfer is from water to air, but the total heat transfer isfrom water to air, hence, water has to be heated externally. This is shown by ProcessO-D in Fig.28.11. This is the process that takes place in a cooling tower. The air streamextracts heat from the hot water coming from the condenser, and the cooled water issent back to the condenser.

6. Heating and humidification: tw> tDBT. Here both sensible and latent heat transfersare from water to air, hence, water has to be heated externally. This is shown byProcess O-E in Fig.28.11.

Thus, it can be seen that an air washer works as a year-round air conditioningsystem. Though air washer is a and extremely useful simple device, it is not commonlyused for comfort air conditioning applications due to concerns about health resultingfrom bacterial or fungal growth on the wetted surfaces. However, it can be used inindustrial applications.

D E

B WC

O

A

DBT

Fig.28.11: Various psychrometric processes that can take place in an air washer


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7.8. ENTHALPY POTENTIALAs shown in case of an air washer, whenever water (or a wetted surface) and air

contact each other, there is possibility of heat and moisture transfer between them. Thedirections of heat and moisture transfer depend upon the temperature and vaporpressure differences between air and water. As a result, the direction of the total heattransfer rate, which is a sum of sensible heat transfer and latent heat transfers alsodepends upon water and air conditions. The concept of enthalpy potential is very usefulin quantifying the total heat transfer in these processes and its direction.

The sensible (QS) and latent (QL) heat transfer rates are given by:

QS= hCAS (ti − ta)(28.18).

= hD.AS(wi − wa).hfgQL= mw .hfgthe total heat transfer QT is given by:

QT= QS+ QL= hCAS (ti− ta ) +hD .AS (wi− w a ).hfg (28.19)

where ta = dry-bulb temperature of air, oC

ti = temperature of water/wetted surface, oC

wa = humidity ratio of air, kg/kg

wi = humidity ratio of saturated air at ti, kg/kg

hc = convective heat transfer coefficient, W/m2.oC

hD = convective mass transfer coefficient, kg/m2

hfg = latent heat of vaporization, J/kg

Since the transport mechanism that controls the convective heat transferbetween air and water also controls the moisture transfer between air and water, thereexists a relation between heat and mass transfer coefficients, hc and hD as discussed inan earlier chapter. It has been shown that for air-water vapor mixtures,

hD≈

hC

orhc

=Lewis number ≈ 1.0 (28.20)cpm

hD

.cpm

where cpm is the humid specific heat ≈ 1.0216 kJ/kg.K.

Hence the total heat transfer is given by:

QT= QS+ QL=hCAS[(ti− ta ) + (wi− w a ).hfg](28.21)

cpm


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by manipulating the term in the parenthesis of RHS, it can be shown that:

QT= QS+ QL=h

CA

S [(hi−ha)] (28.22)c

pmthus the total heat transfer and its direction depends upon the enthalpy difference (orpotential) between water and air (hi-ha).

if hi> ha; then the total heat transfer is from water to air and water gets cooled

if hi< ha; then the total heat transfer is from air to water and water gets heated

if hi = ha; then the net heat transfer is zero, i.e., the sensible heat transfer rate isequalto but in the opposite direction of latent heat transfer. Temperature of waterremains at its wet bulb temperature value

The concept of enthalpy potential is very useful in psychrometric calculationsand is frequently used in the design and analysis of evaporative coolers, coolingtowers, air washers etc.


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REFRIGERATION & AIR CONDITIONING (Credit: 3-0-0, Code: PCME 4402)

(As per BPUT, Rourkela, Syllabus)

MODULE - IV

Prepared

By

Prof. (Dr.) Manmatha K. Roul

Professor and Principal

Gandhi Institute for Technological Advancement

(GITA),

Bhubaneswar – 752054

June 2016


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AIR CONDITIONING SYSTEM

Air Conditioning System : Process in air conditioning : Summer air conditioning, Winter air

conditioning and year round air conditioning, Cooling load calculations. Review question and

discussions. 8.1. INTRODUCTION

Generally from the building specifications, inside and outside design conditions; the latent and sensible cooling or heating loads on a building can be estimated. Normally, depending on the ventilation requirements of the building, the required outdoor air (fresh air) is specified. The topic of load estimation will be discussed in a later chapter. From known loads on the building and design inside and outside conditions, psychrometric calculations are performed to find:

1. Supply air conditions (air flow rate, DBT, humidity ratio & enthalpy)

2. Coil specifications (Latent and sensible loads on coil, coil ADP & BPF)

In this chapter fixing of supply air conditions and coil specifications for

summer air conditioning systems are discussed. Since the procedure is similar for

winter air conditioning system, the winter air conditioning systems are not discussed

here.

8.2. SUMMER AIR CONDITIONING SYSTEMS 8.2.1. SIMPLE SYSTEM WITH 100 % RE-CIRCULATED AIR

In this simple system, there is no outside air and the same air is recirculated as shown in Fig.30.1. Figure 30.2 also shows the process on a psychrometric chart. It can be seen that cold and dry air is supplied to the room and the air that leaves the condition space is assumed to be at the same conditions as that of the conditioned space. The supply air condition should be such that as it flows through the conditioned space it can counteract the sensible and latent heat transfers taking place from the outside to the conditioned space, so that the space can be maintained at required low temperature and humidity. Assuming no heat gains in the supply and return ducts and no energy addition due to fans, and applying energy balance across

the room; the Room Sensible Cooling load (Qs,r), Room Latent Cooling Load (Ql,r)

and Room Total Cooling load (Qt,r) are given by:

Qs,r=msCpm (ti− t s ) (30.1)

Ql,r=mshfg (Wi− Ws ) (30.2)

Q t,r=Q s,r+ Ql,r=ms (hi−hs ) (30.3)

From cooling load calculations, the sensible, latent and total cooling loads on

the room are obtained. Hence one can find the Room Sensible Heat Factor (RSHF)

from the equation:


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HUBANESWAR

RSHF line

i

s

ADP ts ti

Qs,c, Q

Fig.30.1: A simple, 100% re

From the RSHF value one can calculate the slope of the process undergone

by the air as it flows through the conditioned space (process s

slope of process line s

Since the condition i is known say, from thermal comfort criteria, knowing the

slope, one can draw the process line s

the saturation curve gives the ADP of the cooling coil as shown in Fig.30.1. It should be noted that for the given room sensible and latent cooling loads, supplycondition must always lie on this line so that the it can extract the sensible and latent loads on the conditioned space in the required proportions

Since the case being considered is one of 100 % re

that the air undergoes as it flows through the cooling coil (i.e. process iexactly opposite to the process undergone by air as it flows through the room (process s-i). Thus, the temperature and humidity ratio of air through the cooling coil and temperature and humidity ratio increase as air flows through the conditioned space. Assuming no heat transfer due to the ducts and fans, the sensible and latent heat transfer rates at the cooling coil are esensible and latent heat transfer rates to the conditioned space; i.e.,

Q s,r

=Q s,c

& Q

ti,Wi,hi

Return t

Fan

ms

Supply

, Ql,c Cooling Fan

coil

: A simple, 100% re-circulation type air conditioning system

value one can calculate the slope of the process undergone

by the air as it flows through the conditioned space (process s-i) as:

slope of process line s − i, tan θ = 1 1 − RSHF

2451 RSHF

is known say, from thermal comfort criteria, knowing the slope, one can draw the process line s-i through i. The intersection of this line with

the saturation curve gives the ADP of the cooling coil as shown in Fig.30.1. It should ven room sensible and latent cooling loads,

supplycondition must always lie on this line so that the it can extract the sensible and latent loads on the conditioned space in the required proportions

Since the case being considered is one of 100 % re-circulation, the process that the air undergoes as it flows through the cooling coil (i.e. process iexactly opposite to the process undergone by air as it flows through the room

i). Thus, the temperature and humidity ratio of air decrease as it flows through the cooling coil and temperature and humidity ratio increase as air flows through the conditioned space. Assuming no heat transfer due to the ducts and fans, the sensible and latent heat transfer rates at the cooling coil are exactly equal to the sensible and latent heat transfer rates to the conditioned space; i.e.,

& Ql,r

=Ql,c (30.6)

ti,Wi,hi

ms

ts,Ws,hs

circulation type air conditioning system

value one can calculate the slope of the process undergone

(30.5)

is known say, from thermal comfort criteria, knowing the i through i. The intersection of this line with

the saturation curve gives the ADP of the cooling coil as shown in Fig.30.1. It should ven room sensible and latent cooling loads, the

supplycondition must always lie on this line so that the it can extract the sensible and latent loads on the conditioned space in the required proportions.

circulation, the process that the air undergoes as it flows through the cooling coil (i.e. process i-s) will be exactly opposite to the process undergone by air as it flows through the room

decrease as it flows through the cooling coil and temperature and humidity ratio increase as air flows through the conditioned space. Assuming no heat transfer due to the ducts and fans,

xactly equal to the

(30.6)

Qs,r Ql r


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Fixing of supply condition:

The supply condition has to be fixed using Eqns.(30.1) to (30.3). However,

since there are 4 unknowns (ms , ts, Ws and hs) and 3 equations, (Eqns.(30.1) to (30.3)), one parameter has to be fixed to find the other three unknown parameters from the three equations.

If the by-pass factor (X) of the cooling coil is known, then, from room

conditions, coil ADP and by-pass factor, the supply air temperature ts is obtained using the definition of by-pass factor as:

t s − t

ADP

X =

⇒ts=t ADP+ X(ti− t ADP ) (30.7)

ti

−t

ADP

Once the supply temperature ts is known, then the mass flow rate of supply air is obtained from Eqn.(30.1) as:

ms=

Qs,r

=

Qs,r

(30.8)

Cpm (ti− ts ) Cpm (ti− t ADP )(1− X)

From the mass flow rate of air and condition i, the supply air humidity ratio and

enthalpy are obtained using Eqns.(30.2) and (30.3) as:

W =W −

Ql,r

(30.9)

msh

fg

s i

hs =hi −

Qt,r

(30.10)

ms

From Eqn.(30.8), it is clear that the required mass flow rate of supply air

decreases as the by-pass factor X decreases. In the limiting case when the by-pass

factor is zero, the minimum amount of supply air flow rate required is:

ms,min

=

Q s,r

(30.11)

Cpm (ti− t ADP )

Thus with 100 % re-circulated air, the room ADP is equal to coil ADP and the

load on the coil is equal to the load on the room.

8.2.2. SYSTEM WITH OUTDOOR AIR FOR VENTILATION

In actual air conditioning systems, some amount of outdoor (fresh) air is added to take care of the ventilation requirements. Normally, the required outdoor air for ventilation purposes is known from the occupancy data and the type of the building (e.g. operation theatres require 100% outdoor air). Normally either the quantity of outdoor air required is specified in absolute values or it is specified as a fraction of the re-circulated air.


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Fixing of supply condition: Case i) By-pass factor of the cooling coil is zero:

Figure 30.2 shows the schematic of the summer air conditioning system with outdoor air and the corresponding process on psychrometric chart, when the by-pass factor X is zero. Since the sensible and latent cooling loads on the conditioned space are assumed to be known from cooling load calculations, similar to the earlier case, one can draw the process line s-i, from the RSHF and state i. The intersection of this line with the saturation curve gives the room ADP. As shown on the psychrometric chart, when the by-pass factor is zero, the room ADP is equal to coil ADP, which in turn is equal to the temperature of the supply air. Hence from the supply temperature one can calculate the required supply air mass flow rate (which is the minimum required as X is zero) using the equation:

ms=

Qs,r

=

Qs,r

(30.12)

Cpm (ti− t s ) Cpm (ti− t ADP )

From the supply mass flow rate, one can find the supply air humidity ratio and

enthalpy using Eqns.(30.9) and (30.10).

o

Qs,r

m ti,Wi,hi Ql r

me

ti,Wi,hi

s i m rc

ms=mo+mrc

ti,Wi,hi

mo

ADP = ts ti tm to Supply

Cooling Fan

Qs,c, Ql,c

coil

Fig.30.2: A summer air conditioning system with outdoor air for ventilation and a

zero by-pass factor From mass balance of air;

ms=mrc+ mo (30.13)

Where mrc is the re-circulated air flow rate and mo is the outdoor air flow rate. Since either mo or the ratio mo : mrc are specified, one can calculate the amount of re-circulated air from Eqn.(30.13).


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Calculation of coil loads:

From energy balance across the cooling coil; the sensible, latent and total

heat transfer rates, Qs,c, Ql,c and Qt,c at the cooling coil are given by:

Q s,c=msCpm (tm− t s )

(30.14)

Ql,c=mshfg (Wm− Ws )

Q t,c=Q s,c+ Ql,c=ms (hm−hs )

Where ‘m’ refers to the mixing condition which is a result of mixing of the recirculated

air with outdoor air. Applying mass and energy balance to the mixing process one

can obtain the state of the mixed air from the equation:

mo =Wm− Wi =hm−hi ≈ tm− ti (30.15)

m s

W − W h o

−h t o

− t i

o i i

Since (m o/ms) > 0, from the above equation it is clear that Wm> Wi, hm> hi

and tm> ti. This implies that ms(hm- hs) > ms(hi- hs), or the load on the cooling coil is greater than the load on the conditioned space . This is of course due to the fact that during mixing, some amount of hot and humid air is added and the same

amount of relative cool and dry air is exhausted (mo = me).

From Eqn.(30.1) to (30.3) and (30.14), the difference between the cooling

load on the coil and cooling load on the conditioned space can be shown to be equal

to:

Qs,c− Qs,r= moCpm (to− ti )

(30.16) Q

l,c −Q

l,r =m

oh

fg (W

o −W

i )

Qt,c− Qt,r= mo (ho−hi )

From the above equation it is clear that the difference between cooling coil

and conditioned space increases as the amount of outdoor air (mo) increases and/or the outdoor air becomes hotter and more humid.

The line joining the mixed condition ‘m’ with the coil ADP is the process line

undergone by the air as it flows through the cooling coil. The slope of this line

depends on the Coil Sensible Heat Factor (CSHF) given by:

CSHF=

Qs,c

=

Qs,c

(30.17)

Qs,c

+Ql,c

Qt,c

Case ii: Coil by-pass factor, X > 0:

For actual cooling coils, the by-pass factor will be greater than zero, as a

result the air temperature at the exit of the cooling coil will be higher than the coil

ADP. This is shown in Fig.30.3 along with the process on psychrometric chart. It can


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HUBANESWAR

be seen from the figure that when X > 0, the room ADP will be different from the coil ADP. The system shown in Fig.30.3 is adequate when the RSHF is high ( > 0.75).

Return duct losses

Return fan

Qs,r, Ql,r

me= mo A/C Room

i

ti, Wi, hi

mrc

Cooling coil s

Supply fan ms=mo+ mrc

mo m

o

Supply duct losses Qt,c=Qs,c+Ql,c

By-pass

o

m

i

s

ADP ts ti tm to

Fig.30.3: A summer air conditioning system with outdoor air for ventilation and a

non-zero by-pass factor


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HUBANESWAR

Normally in actual systems, either the supply temperature (ts) or the

temperature rise of air as it flows through the conditioned space (ti-ts) will be

specified. Then the step-wise procedure for finding the supply air conditions and the coil loads are as follows: i. Since the supply temperature is specified one can calculate the required supply air

flow rate and supply conditions using Eqns. (30.8) to (30.10).

ii. Since conditions ‘i’, supply air temperature ts and RSHF are known, one can draw

the line i-s. The intersection of this line with the saturation curve gives the room ADP.

iii. Condition of air after mixing (point ‘m’) is obtained from known values of ms and

mo using Eqn.(30.15). iv. Now joining points ‘m’ and ‘s’ gives the process line of air as it flows through the

cooling coil. The intersection of this line with the saturation curve gives the coil ADP.

It can be seen that the coil ADP is lower than the room ADP. v. The capacity of the cooling coil is obtained from Eqn.(30.14). vi. From points ‘m’, ‘s’ and coil ADP, the by-pass factor of the cooling coil can be

calculated.

If the coil ADP and coil by-pass factor are given instead of the supply air

temperature, then a trial-and-error method has to be employed to obtain the supply

air condition. 8.2.3. HIGH LATENT COOLING LOAD APPLICATIONS (LOW RSHF)

When the latent load on the building is high due either to high outside humidity or due to large ventilation requirements (e.g. hospitals) or due to high internal latent loads (e.g. presence of kitchen or laundry), then the simple system discussed above leads to very low coil ADP. A low coil ADP indicates operation of the refrigeration system at low evaporator temperatures. Operating the system at low evaporator temperatures decreases the COP of the refrigeration system leading to higher costs. Hence a reheat coil is sometimes used so that the cooling coil can be operated at relatively high ADP, and at the same time the high latent load can also be taken care of. Figure 30.4 shows an air conditioning system with reheat coil along with the psychrometric representation of the process. As shown in the figure, in a system with reheat coil, air is first cooled and dehumidified from point ‘m’ to point ’c’ in the cooling

coil and is then reheated sensibly to the required supply temperature ts using the

reheat coil. If the supply temperature is specified, then the mass flow rate and state of the supply air and condition of the air after mixing can be obtained using equations given above. Since the heating process in the reheat coil is sensible, the process line c-s will be horizontal. Thus if the coil ADP is known, then one can draw the coil condition line and the intersection of this line with the horizontal line drawn from supply state ‘s’ gives the condition of the air at the exit of the cooling coil. From this condition, one can calculate the load on the cooling coil using the supply mass flow rate and state of air after mixing. The capacity of the reheat coil is then obtained from energy balance across it, i.e.,


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HUBANESWAR

m

o

Cooling coil

Fig.30.4: A summer air conditioning system with reheat coil for high latent

i i

c s

Cooling coil Reheat coil

o

m

i

c s

: A summer air conditioning system with reheat coil for high latent

coolingload applications

: A summer air conditioning system with reheat coil for high latent


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HUBANESWAR

Advantages and disadvantages of reheat coil: a) Refrigeration system can be operated at reasonably high evaporator

temperatures leading to high COP and low running cost. b) However, mass flow rate of supply air increases due to reduced

temperature rise (ti-ts) across the conditioned space c) Wasteful use of energy as air is first cooled to a lower temperature and then heated. Energy is required for both cooling as well as reheat coils. However, this can be partially offset by using waste heat such as heat rejected at the condenser for reheating of air.

Thus the actual benefit of reheat coil depends may vary from system.

8.3. Guidelines for selection of supply state and cooling coil i. As much as possible the supply air quantity should be minimized so that

smaller ducts and fans can be used leading savings in cost of space, material and power. However, the minimum amount should be sufficient to prevent the feeling of stagnation. If the required air flow rate through the cooling coil is insufficient, then it is possible to mix some amount of re-circulated air with this air so that amount of air supplied to the conditioned space increases. This merely increases the supply air flow rate, but does not affect sensible and cooling

loads on the conditioned space. Generally, the temperature rise (ti-ts) will be

in the range of 8 to 15oC.

ii. The cooling coil should have 2 to 6 rows for moderate climate and 6 to 8

rows in hot and humid climate. The by-pass factor of the coil varies from 0.05

to 0.2. The by-pass factor decreases as the number of rows increases and

vice versa. The fin pitch and air velocity should be suitable. iii. If chilled water is used for cooling and dehumidification, then the coil ADP

will be higher than about 4oC.

8.4. INTRODUCTION TO EVAPORATIVE AIR CONDITIONING SYSTEMS

Summer air conditioning systems capable of maintaining exactly the required conditions in the conditioned space are expensive to own and operate. Sometimes, partially effective systems may yield the best results in terms of comfort and cost. Evaporative air conditioning systems are inexpensive and offer an attractive alternative to the conventional summer air conditioning systems in places, which are hot and dry. Evaporative air conditioning systems also find applications in hot industrial environments where the use of conventional air conditioning systems becomes prohibitively expensive.

Evaporative cooling has been in use for many centuries in countries such as India for cooling water and for providing thermal comfort in hot and Prepared By : Prof. (Dr.) Manmatha K. Roul || 10

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dry regions. This system is based on the principle that when moist but unsaturated air comes in contact with a wetted surface whose temperature is higher than the dew point temperature of air, some water from the wetted surface evaporates into air. The latent heat of evaporation is taken from water, air or both of them. In this process, the air loses sensible heat but gains

latent heat due to transfer of water vapour. Thus the air gets cooled and

humidified. The cooled and humidified air can be used for providing thermal

comfort.

8.5. CLASSIFICATION OF EVAPORATIVE COOLING SYSTEMS

The principle of evaporative cooling can be used in several ways.

Cooling can be provided by:

3. Direct evaporation process

4. Indirect evaporation process, or

5. A combination or multi-stage systems 8.5.1. DIRECT EVAPORATIVE COOLING SYSTEMS

In direct evaporative cooling, the process or conditioned air comes in direct contact with the wetted surface, and gets cooled and humidified. Figure 31.1 shows the schematic of an elementary direct, evaporative cooling system and the process on a psychrometric chart. As shown in the figure, hot and dry outdoor air is first filtered and then is brought in contact with the wetted surface or spray of water droplets in the air washer. The air gets cooled and dehumidified due to simultaneous transfer of sensible and latent heats between air and water (process o-s). The cooled and humidified air is supplied to the conditioned space, where it extracts the sensible and latent heat from the conditioned space (process s-i). Finally the air is exhausted at state i. In an ideal case when the air washer is perfectly insulated and an infinite amount of contact area is available between air and the wetted surface, then the cooling and humidification process follows the constant wet bulb temperature line and the temperature at the exit of the air washer is equal to the wet bulb temperature of the entering air (to,wbt), i.e., the process becomes an adiabatic saturation process. However, in an actual system the temperature at the exit of the air washer will be higher than the inlet wet bulb temperature due to heat leaks from the surroundings and also due to finite contact area. One can define the saturation efficiency or effectiveness of the evaporative cooling system ε as:

ε =

(to− t s )

(31.1)

(to

−to,wbt

)


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Exhaust air(i)

Condition ed Qt = Q s+Q l

space

Water spray or

wetted surface

Air washer

OD air(o)

Supply air (s) Blower

Filter

Water pump

i

s w

o

to,wbt ts to t

Fig.31.1: A direct, evaporative cooling system


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Depending upon the design aspects of the evaporative cooling system, the

effectiveness may vary from 50% (for simple drip type) to about 90% (for

efficient spray pads or air washers).

. The amount of supply air required ms can be obtained by writing

energy balance equation for the conditioned space, i.e.,

. Qt

ms= (31.2)

(hi−hs )

where Qt is the total heat transfer rate (sensible + latent) to the building, hi

and hs are the specific enthalpies of return air and supply air, respectively.

Compared to the conventional refrigeration based air conditioning systems, the amount of airflow rate required for a given amount of cooling is much larger in case of evaporative cooling systems. As shown by the above equation and also from Fig.30.1, it is clear that for a given outdoor dry bulb temperature, as the moisture content of outdoor air increases, the required amount of supply air flow rate increases rapidly. And at a threshold moisture content value, the evaporative coolers cannot provide comfort as the cooling and humidification line lies above the conditioned space condition ‘i’. Thus evaporative coolers are very useful essentially in dry climates, whereas the conventional refrigeration based air conditioning systems can be used in any type of climate. 8.5.2. INDIRECT EVAPORATIVE COOLING SYSTEM

Figure 30.2 shows the schematic of a basic, indirect evaporative cooling system and the process on a psychrometric chart. As shown in the figure, in an indirect evaporative cooling process, two streams of air - primaryand secondary are used. The primary air stream becomes cooled andhumidified by coming in direct contact with the wetted surface (o-o’), while the secondary stream which is used as supply air to the conditioned space, decreases its temperature by exchanging only sensible heat with the cooled and humidified air stream (o-s). Thus the moisture content of the supply air remains constant in an indirect evaporative cooling system, while its temperature drops. Obviously, everything else remaining constant, the temperature drop obtained in a direct evaporative cooling system is larger compared to that obtained in an indirect system, in addition the direct evaporative cooling system is also simpler and hence, relatively inexpensive. However, since the moisture content of supply air remains constant in an indirect evaporation process, this may provide greater degree of comfort in regions with higher humidity ratio. In modern day indirect evaporative coolers, the conditioned air flows through tubes or plates made of non-corroding plastic materials such as polystyrene (PS) or polyvinyl chloride (PVC). On the outside of the plastic tubes or plates thin film of water is maintained. Water from the liquid film on the outside of the tubes or plates evaporates into the air blowing over it (primary air) and cools the conditioned air flowing through the tubes or plates sensibly. Even though the plastic materials used in these


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coolers have low thermal conductivity, the high external heat transfer

coefficient due to evaporation of water more than makes up for this. The

commercially available indirect evaporative coolers have saturation efficiency

as high as 80%.

Supply air to

conditioned

space (s) Fan

Primary (o’)

Blower Air-to-air heat

air (o)

exchanger

Water Exhaust Secondary

air (o)

pump air(e)

o’ e

w

s o

t

Fig.31.2: An indirect, evaporative cooling system

8.5.3: MULTI-STAGE EVAPORATIVE COOLING SYSTEMS

Several modifications are possible which improve efficiency of the

evaporative cooling systems significantly. One simple improvement is to

sensibly cool the outdoor air before sending it to the evaporative cooler by exchanging heat with the exhaust air from the conditioned space. This is possible since the temperature of the outdoor air will be much higher than the


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exhaust air. It is also possible to mix outdoor and return air in some proportion so that the temperature at the inlet to the evaporative cooler can be reduced, thereby improving the performance. Several other schemes of increasing complexity have been suggested to get the maximum possible benefit from the evaporative cooling systems. For example, one can use multistage evaporative cooling systems and obtain supply air temperatures lower than the wet bulb temperature of the outdoor air. Thus multistage systems can be used even in locations where the humidity levels are high.

Figure 30.3 shows a typical two-stage evaporative cooling system and the process on a psychrometric chart. As shown in the figure, in the first stage the primary air cooled and humidified (o -o’) due to direct contact with a wet surface cools the secondary air sensibly (o -1) in a heat exchanger. In the second stage, the secondary air stream is further cooled by a direct evaporation process (1-2). Thus in an ideal case, the final exit temperature of

the supply air (t2) is several degrees lower than the wet bulb temperature of

the inlet air to the system (to’).

8.6. ADVANTAGES AND DISADVANTAGES OF

EVAPORATIVE COOLING SYSTEMS

Compared to the conventional refrigeration based air conditioning

systems, the evaporative cooling systems offer the following advantages:

vii. Lower equipment and installation costs viii. Substantially lower operating and power costs. Energy savings can be

as high as 75 % ix. Ease of fabrication and installation x. Lower maintenance costs xi. Ensures a very good ventilation due to the large air flow rates involved,

hence, are very good especially in 100 % outdoor air applications xii. Better air distribution in the conditioned space due to higher flow rates xiii. The fans/blowers create positive pressures in the conditioned space,

so that infiltration of outside air is prevented xiv.Very environment friendly as no harmful chemicals are used

Compared to the conventional systems, the evaporative cooling systems

suffer from the following disadvantages:

d) The moisture level in the conditioned space could be higher, hence, direct evaporative coolers are not good when low humidity levels in the conditioned space is required. However, the indirect evaporative cooler can be used without increasing humidity

e) Since the required air flow rates are much larger, this may create draft and/or high noise levels in the conditioned space

f) Precise control of temperature and humidity in the conditioned space is not possible

g) May lead to health problems due to micro-organisms if the water used

is not clean or the wetted surfaces are not maintained properly.


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Primary air

in (o)

Fig.31.3: A two

8.7. APPLICABILITY OF EVAPORATIVE COOLING SYSTEMS

As mentioned before, evaporative cooling systems are ideal in hot and dry places, i.e., in places where the dry bulb temperature is high and the coincident wet bulb temperature is low. However, there are no clearas to where these systems can or provide some measure of comfort in where the humidity levels are very high, standsystems cannot be used for providing thermal comfort especially in residences, office buildings etc. One of the older rules

(1)

Primary air Secondary

air in (o)

out (e)

o’ e

2

w

1 o

t2to’t

: A two-stage evaporative cooling system


As mentioned before, evaporative cooling systems are ideal in hot and dry places, i.e., in places where the dry bulb temperature is high and the coincident wet bulb temperature is low. However, there are no clear-as to where these systems can or cannot be used. Evaporative cooling can provide some measure of comfort in any location. However, in many locations where the humidity levels are very high, stand-alone evaporative cooling systems cannot be used for providing thermal comfort especially in residences, office buildings etc. One of the older rules-of-thumb used in USA

Secondary

air out (2)


As mentioned before, evaporative cooling systems are ideal in hot and dry places, i.e., in places where the dry bulb temperature is high and the

-cut rules cannot be used. Evaporative cooling can any location. However, in many locations

alone evaporative cooling systems cannot be used for providing thermal comfort especially in

thumb used in USA


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specifies that evaporative cooling systems can be used wherever the averagenoon relative humidity during July is less than 40%. However, experienceshows that evaporative coolers can be used even in locations where the relative humidity is higher than 40%. A more recent guideline suggests that evaporative cooling can be used in locations where the

summer design wetbulb temperatures are less than about 24oC (75

oF). It

is generally observedthat evaporative coolers can compete with conventional systems when the noon relative humidity during July is less than 40%, hence should definitely be considered as a viable alternative, whereas these systems can be used in places where the noon relative humidity is higher tha

40% but the design WBT is lower than 24oC, with a greater sacrifice of

comfort. It should be mentioned that both these guidelines have been developed for direct evaporative cooling systems. Indirect evaporative coolers can be used over a slightly broader range. Evaporative air conditioning systems can also be used over a broader range of outdoor conditions in factories, industries and commercial buildings, where the comfort criteria is

not so rigid (temperatures as high as 30oC in the conditioned space are

acceptable). Evaporative air conditioning systems are highly suitable in applications requiring large amounts of ventilation and/or high humidity in the conditioned space such as textile mills, foundries, dry cleaning plants etc.

Evaporative cooling can be combined with a conventional refrigeration based air conditioning systems leading to substantial savings in energy consumption, if the outside conditions are favorable. Again, a number of possibilities exist. For example, the outdoor air can be first cooled in an evaporative cooler and then mixed with the re-circulating air from the conditioned space and then cooled further in the conventional refrigerant or chilled water coil.

8.8. WINTER AIR CONDITIONING SYSTEMS

In winter the outside conditions are cold and dry. As a result, there will be a continuous transfer of sensible heat as well as moisture (latent heat) from the buildings to the outside. Hence, in order to maintain required comfort conditions in the occupied space an air conditioning system is required which can offset the sensible and latent heat losses from the building. Air supplied to the conditioned space is heated and humidified in the winter air conditioning system to the required level of temperature and moisture content depending upon the sensible and latent heat losses from the building. In winter the heat losses from the conditioned space are partially offset by solar and internal heat gains. Thus in a conservative design of winter A/C systems, the effects of solar radiation and internal heat gain are not considered.

Heating and humidification of air can be achieved by different schemes. Figure 31.4 shows one such scheme along with the cycle on psychrometric chart. As shown in the figure, the mixed air (mixture of return and outdoor air) is first pre-heated (m-1) in the pre-heater, then humidified using a humidifier or an air washer ( 1-2) and then finally reheated in the re-heater (2-s). The reheated air at state ‘s’ is supplied to the conditioned space.

The flow rate of supply air should be such that when released into the

conditioned space at state ‘s’, it should be able to maintain the conditioned


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space at state I and offset the sensible and latent heat losses (Qs and Ql). Pre-heating of air is advantageous as it ensures that water in the humidifier/air washer does not freeze. In addition, by controlling the heat supplied in the pre-heater one can control the moisture content in the conditioned space.

Return air (i)

Qs,Ql

Exhaust air

Conditioned space

Recirculated

air (i)

Supply air (s)

OD air (o) (m) (1) (2)

Pre-heater Humidifier Re-heater

2 s

w i

m 1

o

t

Fig.31.4: A winter air conditioning system with a pre-heater Prepared By : Prof. (Dr.) Manmatha K. Roul || 18

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The humidification of air can be achieved in several ways, e.g. by bringing the air in contact with a wetted surface, or with droplets of water as in an air washer, by adding aerosol sized water droplets directly to air or by direct addition of dry saturated or superheated steam. Humidification by direct contact with a wetted surface or by using an air washer are not recommended for comfort applications or for other applications where people are present in the conditioned space due to potential health hazards by the presence of micro-organisms in water. The most common method of humidifying air for these applications is by direct addition of dry steam to air. When air is humidified by contact with wetted surface as in an air washer, then temperature of air decreases as its humidity increases due to simultaneous transfer of sensible and latent heat. If the air washer functions as an adiabatic saturator, then humidification proceeds along the constant wet bulb temperature line. However, when air is humidified by directly adding dry, saturated steam, then the humidification proceeds close to the constant dry bulb temperature line. The final state of air is always obtained by applying conservation of mass (water) and conservation of energy equations to the humidification process.

By applying energy balance across the conditioned space, at steady

state, the sensible and latent heat losses from the building can be written as:

.

Qs=ms cpm (t s− ti ) (31.3).

Ql=ms hfg (w s− wi ) (31.4).

where ms is the mass flow rate of supply air, cpm is the specific heat of air, hfg is the latent heat of vapourization of water, ws and wi are the supply and

return air humidity ratios and ts, ti are the supply and return temperatures of air. By applying mass and/or energy balance equations across individual components, the amount of sensible heat transfer rate to the pre-heater and re-heater and the amount of moisture to be added in the humdifier can easily be calculated.

Figure 31.5 shows another scheme that can also be used for heating and humidification of air as required in a winter air conditioning system. As shown in the figure, this system does not consist of a pre-heater. The mixed air is directly humidified using an air washer ( m-1) and is then reheated (1-s) before supplying it to the conditioned space. Though this system is simpler compared to the previous one, it suffers from disadvantages such as possibility of water freezing in the air washer when large amount of cold outdoor air is used and also from health hazards to the occupants if the water used in the air washer is not clean. Hence this system is not recommended for comfort conditioning but can be used in applications where the air

temperatures at the inlet to the air washer are above 0oC and the conditioned

space is used for products or processes, but not for providing personnel comfort.


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Return air

Qs,Ql

Exhaust air(i)

Conditioned space

Recirculated

air(i)

OD air(o) (m) (1) Supply air(s)

Air washer Re-heater

1s

w

i

m

o

t

Fig.31.5: A winter air conditioning system without a pre-heater

Actual winter air conditioning systems, in addition to the basic components shown above, consist of fans or blowers for air circulation and filters for purifying air. The fan or blower introduces sensible heat into the air stream as all the electrical power input to the fan is finally dissipated in the form of heat.


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8.9. All year (complete) air conditioning systems

Figure 30.6 shows a complete air conditioning system that can be used for providing air conditioning throughout the year, i.e., during summer as well as winter. As shown in the figure, the system consists of a filter, a heating coil, a cooling & dehumidifying coil, a re-heating coil, a humidifier and a blower. In addition to these, actual systems consist of several other accessories such as dampers for controlling flow rates of re-circulated and outdoor (OD) air, control systems for controlling the space conditions, safety devices etc. Large air conditioning systems use blowers in the return air stream also. Generally, during summer the heating and humidifying coils remain inactive, while during winter the cooling and dehumidifying coil remains inactive. However, in some applications for precise control of conditions in the conditioned space all the coils may have to be made active. The blowers will remain active throughout the year, as air has to be circulated during summer as well as during winter. When the outdoor conditions are favourable, it is possible to maintain comfort conditions by using filtered outdoor air alone, in which case only the blowers will be running and all the coils will be inactive leading to significant savings in energy consumption. A control system is required which changes-over the system from winter operation to summer operation or vice versa depending upon the outdoor conditions.

Exhaust

Return air

Conditioned Cooling/

Recirculated Heating

space load

air

Supply air

OD air

F

RH B

H CC Hu

F: Filter; H: Heating coil; CC: Cooling & dehumidifying coil

RH: Re-heating coil;Hu: Humidifier;B: Blower

Fig.31.6: An all year air conditioning system


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REFRIGERATION & AIR CONDITIONING · 2. Refrigeration and Air conditioning by Manohar Prasad,New Age international publishers. 3. Refrigeration and Air conditioning by C.P. Arora,

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