Refrigeration Technology

Refrigeration and HVAC Technology

Department of Mechanical Engineering Technology

Yanbu Industrial College MET

DE

P A R T M E N

T

MEC

HAN

ICA

L EN

GINEERING TECHNOLO

GY

1

Chapter 1

Refrigerants

1.1- Introduction

The majority of domestic and commercial refrigerators and air conditionings in current use are charged with one of these refrigerants: R12, R22, R134a R404a,R410 A, etc. Examination of larger or more specialized installations will however reveal several others in common use. They include R717 (Ammonia), in block ice making plant or factory scale food freezing and cold – storage equipment.

The topic of ozone destruction has been included briefly in this chapter because many of the common refrigerants such as R11, R12 and R502 have already become unavailable or will disappear over the next few years due to the international Protocols such as Montreal 1987.

New refrigerants are developed as a result of global warming to get the best possible results from new types of equipment, or to boost the performance of existing compressors design and has no harmful effect on the environment.

1.2 Refrigerant Definition

Refrigerant is any substance which acts as a cooling agent by absorbing heat from another substance.

1.3 Classification of refrigerants

Refrigerant can be classified into two types of refrigerants, Primary such as Ammonia and Freon, which can be used in Direct Cooling System. Secondary such as Glycol and Brine, which are normally used in Indirect Cooling Systems. (see section 1.10).

1.4- Refrigerant Classification and Numbering

CFC (ChloroFluoroCarbon)

R111, R12, R13, R113, R114, R115, R500, R502

HCFC (HydroChloroFluoroCarbon)

R221, R123, R124, R401A, R401B, R401C, R402A, R402B, R403A, R403B, R405A, R406A, R408A, R409A, R409B, R411A, R411B, R412A AND R509A

HFC (HydroFluoroCarbon)

R125, R134a, R404A, R407A, R407B, R407C, R410A, R413A, R507, R507A

PFC (PerFluoroCarbons) and HC (HydroCarbon)

R218, R290, R6002, R602a

Inorganic (Natural Refrigerants)

1 ASHRAE designation system has been used in assigning a number to each refrigerant. Numbering is as follows:

The first digit on the right is the number of Fluorine (F) atoms in a molecule.

The second digit from the right is one more than the number of hydrogen (H) atoms in a molecule.

The third digit from the right is one less than the number of Carbon (C ) atoms in a molecule. If this is zero, then the number is omitted.

2 The 600 series has been assigned to miscellaneous organic compounds.

2

R7173, R718, R729, R702 ( OrthoHydrogen), R702 (Parahydrogen), R704, R720, R728, R732, R740.

Example 1.1

Determine the ASHRAE number designation for dichlorotetra-fluoroethane CClF2CClF2.

Solution

There are 4 fluorine atoms, no hydrogen atoms, and 2 carbon atoms per molecule. Thus the ASHRAE designation is;

(2-1) (0+1) (4) = 114 = R114

1.5 Refrigerant Properties

An Ideal refrigerant should possess certain chemical, physical and thermodynamic properties that make it safe for human and environment and economical to be used.

Table (1.1) shows the properties and standard performance comparisons for CFC‘s, HCFC‘s, PFC‘s, natural refrigerants and blends.

3 The 700 series has been assigned to inorganic compounds. The relative molecular mass of the compounds is added to 700 to arrive at the identifying refrigerant numbers.

C-1 H+1 F

3

Table 1.1

4

1.5.1- Properties of Ammonia

Ammonia is a colorless gas, easily soluble in water, which is chemically stable, has a characteristic odor, and is extremely corrosive, irritating and toxic. Ammonia is a compound of Nitrogen and Hydrogen with the chemical formula NH3.

Ammonia is lighter than air and forms an aerosol mist with air humidity and by its cooling during expansion in air. Ammonia concentration in air varies according to the nature of the land area. The air in cities may contain 7-35 ppm and that in the countryside, 2-6 ppm. In local areas where ammonia is in use, e.g., in agriculture and in cowsheds, this concentration may comprise 100-22 ppm. Such concentrations are directly noticeable because of characteristic odor of ammonia.

1 ppm (particles per million) = 0.7 mg/m3 at 25 oC at atmospheric pressure.

In practice, 1m3 of water can dissolve 120 kg of ammonia.

Risk of fire and explosive

Together with oxidizing gases, such as oxygen in the air, ammonia forms explosive mixtures. This hazard occurs only in enclosed areas to the limited upper and lower explosive limits and high ignition temperatures.

Extensive investigations have shown that:

a- In the open air burning ammonia can be maintained only when it evaporates strongly and when the fire is sustained. When liquid ammonia at atmospheric pressure has achieved equilibrium with its surroundings, the ammonia gas cannot be ignited.

b- In an enclosed space it is possible to ignite a mixture of ammonia and air within the ignition limits with an electric spark or a naked flame.

c- The presence of humidity reduces the potential for ignition of an ammonia-air mixture. The combustion rate is 50 times slower than that of city gas. The force formed during combustion is around 1/7 of that in a city gas explosion.

The risk of ammonia exploding is thus limited in consequence of its high ignition temperature, limited ignition range in an air mixture, low force in conjunction with combustion and strong absorption of moisture from the air.

Safe Properties Thermal Properties

Nonflammable

Nontoxic (Depends on degree of concentration).

Non-explosive

Should not react unfavorably with lubricating oil.

Should not react unfavorably with moisture.

Should not have a harmful effect on product and environment. (see Tables 1.1 and 1.2)

Evaporating pressure

Condensation pressure

Freezing Temperature

Critical temperature and pressure.

Compression ratio.

Latent heat of evaporation

Viscosity

Specific heat

Specific Volume

Oil Miscibility

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Health Hazards

Its pungent smell is characteristic of ammonia. Since it is easily soluble in water, the gas settles on damp skin, in the mucous membranes of the respiratory passages, and in the eyes. In increasing concentrations this leads to a reflex closing of the eyes and a temporary loss of vision.

At high concentrations injuries cannot be excluded since ammonia has a penetrating burning and corrosive action.

Ammonia in liquid form or as a cold gas can cause severe frostbite and corrosion injuries if it comes into contact with unprotected skin.

Preventive measures

Face mask, breathing mask with a filter and protective gloves are to be worn when working on parts of the plant that contain ammonia.

Besides the above, helmet, hearing protectors, protective footwear and protective suit are to be worn when required.

Ventilation of the area where work is to be done is to in operation.

Emergency shower and eye shower are to be available in the vicinity.

1.6- Pressure/Temperature Relationship

The full physical properties of refrigerants can only be detailed in extensive tables.

Refer to Thermodynamics Tables.

1.7 Refrigerant Leak Detection

Leaks can not be tolerated in any refrigeration system, and leak detecting equipment must be well maintained and regularly used during maintenance checks as well as installation work.

1.7.1- Leak Detection Methods

The methods which can be used with specific refrigerants are listed below in increasing order of efficiency.

i- Sulphur Candles

When light exposed to air containing ammonia vapor, these give off a white cloud of ammonium chloride or ammonium sulphide. This method cannot be used to pinpoint leaks.

ii- Litmus Paper

Moist red litmus paper turn blue if exposed to ammonia vapor, but cannot be used with any of the halogen family.

iii- Bubble tests

Soapy water, a washing up liquid, or better still a purpose- developed leak indicator will indicate the locations of leaks by the formation of bubbles by escaping refrigerant. However, this type of test can only be made on piping or fittings known to be at a higher pressure than that of the atmosphere. Test solutions applied to low pressure suction lines could cause considerable damage because the liquid could be drawn into the pipes.

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iv- Halide test lamps

Detectors fuelled with propane, butane, or methylated spirits can be used to locate fluorocarbon refrigerant leaks. The detector includes a fuel tank which is, or can be , pressurized to supply fuel at a steady and controlled pressure and a jet to admit the fuel to a burner. When light, the burner flame is supported by oxygen in the air which is drawn through a tube used as a sensing probe. The probe is passed slowly over the joints or surfaces being leak tested. If any fluorocarbon refrigerants are drawn into the tube, the color of the lamp flame will change to green or blue, depending on the quantity of gas passed over the burner element.

This type of detector can only be used with non-flammable gases, and care must be taken to avoid igniting any other gases or materials (including pipeline insulation) or damaging heat –sensitive items of equipment.

v- Electronic leak detectors

A wide range of electronic detectors is available, and prices are not prohibitive. All are extremely sensitive-battery operated models for use on the site will pick up leaks which give as little as 14 grams (0.5 ounce) per year. Clearly, this is the most efficient tool for what can be a difficult and time consuming job.

The refrigerant is sensed by a plug-in element, exposed to air drawn through a probe or tube. Its presence will be indicated by a flashing lamp, an audible ―bleep‖ or buzz, or a meter reading, each increasing in speed or intensity as more refrigerant passes over the element.

Leak area

Sniffer hose

Valve

Opening to see flame color

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1.8 Refrigeration Oil

The lubricants used in refrigeration systems to do:

i- Protect moving compressor parts against wear.

ii- Creating a film of oil which seals the suction and discharge valves of the compressor or the shaft seal of an open type compressor.

iii- Act as a coolant, transferring mechanically generated heat from the crankcase to the shell of the compressor.

iv- It dampens noise.

The following definitions of oil are important and should be memorized

Oil Miscibility: Can be defined as the ability of given oil to mix with a given refrigerant.

Pour Point: Is the lowest temperature at which the oil will flow or ‗pour‘.

Cloud point: Since all lubricating oils contain a certain amount of paraffin, wax will precipitate from any oil if the temperature of the oil is reduced to a sufficiently low level. Because the oil becomes cloudy at this point, the temperature at which the wax begins to precipitate from the oil is called the Cloud Point of the oil.

Compatible oils for each refrigerant are listed in Tables 1.5 and 1.6.

Flash or Fire Points: Flash or fire points express the boiling point and vapor pressure data for oil. You might think this unnecessary for oil used in a hermetic refrigeration system, but such data does have to be taken into account when designing systems with high compression ratios or other characteristics which might lead to unsuitable oil being carbonized.

Oil selection

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System capacity

Evaporating and condensing temperatures

Refrigerant type

1.9 Conversion options

Some conversion options are summarized in Table 1. 2. Note that the options depending on the application, the existing refrigerant, and the compressor type.

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Table 1. 1 Alternative refrigerants suggested in 1995

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1.10 Secondary Refrigerants

In some large installations it is not possible to use direct expansion systems-often because pipes are too long to permit the efficient return of oil being carried with fluorocarbon refrigerants, or have rises too high to permit liquid to be lifted without ―flashing‖ into a mixture of gas and liquid as the result of drops in pressure. In some cases, particularly where R717 is used booster pumps are installed to overcome the resistance of long lines; but this is not the safest of refrigerants to have circulating around large, occupied building; and secondary refrigerants are frequently used instead.

Types of secondary refrigerant

i- Glycol:

It is sometimes not realized that two types of glycol are available-ethene based and propene based. Ethylene glycol has the better physical properties, especially at low temperatures; but where the secondary refrigerant might come into contact with foods or drink, toxicity considerations may require the use of propylene glycol. Their leading physical properties are listed in Table 1.3.

The glycols are normally used as solutions in water and freezing point of solution decreases as the percentage by weight of glycol is increased. As shown in Table 1.4.

Table 1.3

Properties of Glycols Ethylene Glycol Propylene Glycol

Relative molecular weight Specific Gravity at 20 oC Boiling Point at 1 bar Vapor Pressure at 20 oC Freezing Point Specific heat capacity 20 oC (kJ/kg K) Heat of Fusion at -13 oC Heat of vaporization at 1 bar (kJ/kg)

62.07 1.113 197 0.05

-13 oC 2.35 187 846

76.1 1.036 187.4 0.07

-60 oC 2.48

- 688

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Table 1.4

Freezing points of glycols

% by weight Ethylene Glycol

Freezing point oC Propylene Glycol Freezing point oC

0 10 20 30 40 50 60

0 -5 -9

-16 -24 -36 -57

0 -3 -7

-13 -22 -32 -57

ii- Brines:

Once more, two types of brine are used that made with calcium chloride (CaCl2) for use in ice plants and ice rinks, and sodium chloride (NaCl) where the brine might come into contact with foodstuffs. For the majority of refrigeration applications, the lower freezing point of calcium chloride makes its use commercially popular.

It must not forgotten that brines are very corrosive, and must be inhibited before use with sufficient sodium chromate to produce an alkaline solution (pH 7.0to 8.5). The pH level is subsequently adjusted by adding sodium hydroxide (caustic soda) to correct acidity (pH below 7.0) or adding sodium dichromate to correct excessive alkalinity (pH above 8.5).

1.11- Global Warming and Ozone Depletion Indices

Both global warming and ozone depletion involve a range of chemicals. A number of indices have been developed to quantify the relative effect of each chemical on the environment. Tables 1.5 and 1.6 give atmospheric lifetimes and values for some of these indices for CFC‘s, HCFC‘s and some possible alternatives.

Ozone Depletion Potential (ODP)

Ozone Depletion Potential (ODP) is a measure of the relative damage caused to the ozone layer by chemicals (over100 years lifetime). Generally speaking, the more chlorine in the chemical structure and the longer the lifetime of the chemical in the atmosphere, the greater the ODP. Atmospheric lifetime is highly related to the presence or otherwise of hydrogen and fluorine in the chemical structure. The more fluorine presents the greater the chemical stability of the molecule and hence greater amounts will reach the outer atmosphere. If a hydrogen molecule is present this forms a site for chemical attack so the molecule is less stable.

The chemical CFC-11 and CFC-12, also known as Refrigerant 11 and 12 or sometimes as Freons 11 and 12,(although it should be noted that ―Freon‖ is merely the brand-name of one of the largest manufacturers) are given an ODP of 1.0 to establish a basis.

Table 1.5

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Table 1.6

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1.12 Revision questions

1- Explain in details the refrigerant leak detection methods

2- Explain why oil used in refrigeration systems and how can be selected.

3- State six of the refrigerant safe and thermal properties

4- What are the differences between the two types of Glycol.

5- Determine the ASHRAE number designation for;

a- Chlorodifluoromethane CHClF2

b- Ammonia

c- Dichlorodifluoromethane CCl2F2

6- Explain why oil used in refrigeration systems should not be exposed to the atmosphere.

7- Define the following

a- Refrigerant

b- Oil miscibility

c- Oil Pour point

d- Oil flash point

e- Oil cloud point

f- ODP

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Chapter 2

Simple Vapor Compression Cycle

2.1- Important Definitions

Refrigeration

In general, refrigeration is defined as any process of heat removal. More specifically, refrigeration is defined as the branch of science that deals with the process of reducing and maintaining the temperature of space or material below the temperature of the surroundings.

Refrigeration Load

The rate at which heat must be removed from the refrigerated space or material in order to produce and maintain the desired temperature conditions in called refrigeration load, the cooling load, or the heating load.

Cycle

As the refrigerant circulates through the system it passes through a number of changes in state or condition each of which is called a process. The refrigerant starts at some initial state or condition, passes through a series of processes in a definite sequence, and return to the initial condition. This series of processes is called a cycle.

Refrigerating Effect (Cooling effect)

The quantity of heat that each unit mass of refrigerant absorbs from the refrigerated and/or air conditioned space is known as the refrigerating effect or cooling load.

System Capacity

The capacity of any refrigerating system is the rate at which it will remove heat from the refrigerated space. It is usually expressed in kW.

2.2 How Can Cooling Be Accomplished?

The method most often used for cooling is evaporation of a liquid. When a liquid evaporates it takes up the latent heat of evaporation. Try wetting your finger and then blowing on it. Your finger feels cold – as evaporates it takes up heat which cools your finger down.‖ Try to blowing on your dry fingers and feel the difference‖.

In general if one wishes to cool something down, this can be accomplished by making a liquid evaporate at a temperature. The temperature at which a liquid evaporates is controlled by the pressure over it.

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How Can Heat Be Rejected

In contrast once the liquid has evaporated it is a vapor. The process of condensation is simply the reverse of boiling. In evaporation the liquid takes up heat to become a vapor and thus accomplishes cooling. In condensation a vapor gives up heat to become a liquid. It thus rejects heat and actually heats the substance the heat is rejected to. If the vapor is taken up to a high pressure the temperature at which it will boil or condense rises so the heat can be rejected at a higher temperature than that at which it was taken up.

The Mollier or ph Diagram

The diagrams frequently used in the analysis of the refrigeration cycle are the pressure-enthalpy ph diagram and the temperature-enthalpy (T-S diagram). Of the two, the ph diagram seems to be the most useful. You will be given ph diagram for R22, R134a, R12 and R717 and try to photocopy them for exercises.

The ph diagram is a plot of pressure verses enthalpy and is based on a 1 kg mass of the refrigerant, hence all its properties is per kilogram (kg). The envelope represents mixture of liquid and vapor. If you look at the diagrams they appear rather complicated. This is because four other sets of lines are also plotted on them.

Isotherms

A line of constant temperature is horizontal inside the envelope (Liquid-Vapor mixture region). To the left of the envelope a line of constant temperature is vertical. This is because liquids are essentially incompressible so that a change in the pressure over them neither increases their temperature, or changes their volume. The definition of enthalpy therefore means that for a sub-cooled liquid (i.e. lying to the left of the envelope) temperature does not change with pressure if the enthalpy remains constant. On many ph diagrams vertical temperature lines are not drawn to the left of the envelope.

To the right of the envelope the refrigerant is present as a gas which is compressible (superheated region) . Hence the temperature change with pressure is complex, and the isotherm bend –they are always drawn in as a result.

Iso-Volume Lines

Lines showing the volume occupied by 1 kg of refrigerant are drawn on the ph diagram. These sweep upwards at an angle. For the refrigeration engineer, his chief interest in these is in the region to the right of the envelope (compressor suction condition).

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Isentropic Lines

Lines of constant entropy are also drawn. The reasons for these will become obvious when compression is studied-the ideal reversible compression is one in which entropy remains constant so the isentropic lines establish a way of tracing the behaviour of perfect compressor.

Iso-Quality Lines

Within the envelope lines representing the refrigerant quality ( as discussed in the thermodynamic subject) are also present. These are particularly useful for quickly determining the percentage of vapor in a liquid/ vapor mixture. ‘ x ‘ is used to denote liquid refrigerant quality. If 1 kg of the refrigerant mixture is taken then there will be x kg of vapor and (1-x) kg of liquid.

2.4.1- Sub-cooled or compressed region

Is the area on the chart (ph diagram) to the left of the saturated liquid line.

Pgiven or calculated > [Psaturation] at Tgiven or calculated

Tgiven or calculated < [Tsaturation] at Pgiven or calculated

vgiven or calculated < [vf] at (T or P) given or calculated

hgiven or calculated < [hf] at (T or P) given or calculated sgiven or calculated < [sf] at (T or P) given or calculated

2.4.2- Saturation (mixture) Region

Is the section of the chart between the saturated liquid and saturated vapor lines is the mixture region (saturation region) and represents the change in phase of the refrigerant between the liquid and vapor phases.

[Pgiven or calculated] = Psaturation

[Tgiven or calculated] = Tsaturation

[vg ] at (T or P) given or calculated >[vgiven or calculated ] > [vf ] at (T or P) given or calculated

Quality (X) = (v - vf )/(vg - vf)

[hg ] at (T or P) given or calculated >[hgiven or calculated ] > [hf ] at( T or P) given or calculated

Quality (X) = (h - hf )/(hg - hf)

[sg ] at (T or P) given or calculated >[sgiven or calculated ] > [sf ] at (T or P) given or calculated

Quality (X) = (s - sf )/(sg - sf)

2.4.3- Superheated Region

Is the area on the chart (ph diagram) to the right of the saturated vapor line.

Pgiven or calculated < [Psaturation] at Tgiven or calculated

Tgiven or calculated > [Tsaturation] at Pgiven or calculated

vgiven or calculated > [vg] at (T or P) given or calculated

hgiven or calculated > [hg] at (T or P) given or calculated sgiven or calculated > [sg] at (T or P) given or calculated

2.4.4- Saturated liquid

The substance state is located on the saturated liquid line (curve)

Pgiven or calculated = Psaturation

Tgiven or calculated = Tsaturation

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vgiven or calculated = [vf ] at (T or P) given or calculated

hgiven or calculated = [hf ] at (T or P) given or calculated sgiven or calculated = [sf ] at (T or P) given or calculated

2.4.5- Saturated vapor

The substance state is located on the saturated vapor line (curve)

Pgiven or calculated = Psaturation

Tgiven or calculated = Tsaturation

vgiven or calculated = [vg ] at (T or P) given or calculated

hgiven or calculated = [hg ] at (T or P) given or calculated sgiven or calculated = [sg] at (T or P) given or calculated

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Figure 2.1

20

Figure 3.1 ph diagram

2.5- Heat and Work

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Definitions and the nature of heat and work are fully discussed in all modern books of thermodynamics. However, the following definitions are vital and are repeated to assist in the understanding of heat pumps principles.

Heat: is defined as that energy which transferred to (+) or from (-) a system under the sole influence of the temperature difference between the system and its surrounding.

Work: is defined as that energy which transferred to (-) or from (+) a system by the action of a moving force.

2.6- The Ideal Simple Vapor Compression Cycle

An ideal simple vapor compression cycle will be internally and externally reversible. There are a number of such theoretical cycles but the reversed Carnot cycle is the best known.

Reversed Carnot Cycle

Figure 3.2 The ideal vapor

compression refrigeration cycle

The above diagrams illustrate the plant and T-S diagram for a reversed Carnot Cycle refrigerator working on a vapor.

1-2 Isentropic compressio

n of the vapor during which the temperature rises from TL to TH.

2-3 Rejection of heat to the high temperature region as the vapor condenses at the high pressure (Isothermal process).

3-4 Isentropic expansion from the high to the low pressure during which the temperature falls from TH to TL.

4-1 Heat transfer from the low temperature region as evaporation proceeds at the low pressure (Isothermal process)..

The heat transfer to the high temperature region is

QH = TH.(S2-S3)

The heat transfer from the low temperature is

QL = TL.(S2-S3) = TL.(S1-S4)

The net work input is,

QH - QL = (TH- TL) (S2 - S3)

The Carnot Refrigerator Coefficient of Performance C.O.Pc is,

TL

TH

22

COPc

net

L

nevaporatioonCondensati

nevaporatio

LH

L

LH

L

W

Q

TT

T.

TT

T

)S(S)T(T

)S(ST

15273

32

32

2.7. The Simple Vapor Compression Cycle

The Theoretical Vapor Compression Cycle

The cycle is as follows (refer to Figure 3.2)

1-2 Isentropic compression (S = C) of the vapor from the evaporating to the condensing pressure.

2-3 Condensation of the high pressure vapor during which heat is transferred to the high temperature region. (P2 = P3).

3-4 Adiabatic expansion of the condensed vapor from the condensing to the evaporating pressure. (h4 = h3).

4-1 Evaporation of the low pressure liquid during which heat is absorbed from the low temperature source. (P1 = P4).

The energy transfers can readily be determined from p-h diagram per kg.

Net work = h2 –h1

Rejected Heat = h2 – h3

Expansion Process h3 = h4

Refrigerating effect = h1 - h4

Refrigerating efficiency

The Refrigerating efficiency (r)is the ratio of C.O.P. of a cycle or a system to C.O.Pc.

Example 2.1

Determine the C.O.P. and refrigerating efficiency for a theoretical single stage cycle operating with a condenser pressure of 1200 kPa and evaporator pressure of 200 kPa using R22 as a refrigerant. The work supplied to the cycle is 49 kJ/kg of refrigerant.

Solution

From R22 ph diagram h1=395 kJ/kg , h3 = h4 = 238 kJ/kg

C.O.P.= (h1-h4)/Wnet =3.2

C.O.Pc.= (273.15+Tevap.)/ (Tcond. –Tevap.) 248/55 = 4.5

r = C.O.P /C.O.Pc = 3.2/4.5 = 0.71 or 71 %.

Example 2.2

A R 134a refrigerator cycle flow rate of 0.05 kg/sec and the following data are given. Determine, Total heat rejected at condenser, C.O.P., and System Capacity

Data Calculation

State T oC P (Bar) v ( m3/kg) h (kJ/kg) s (kJ/kg K) Quality

23

1 - 20 1.337 0.14649 3861 1.740 1

2 45 8.876 0.0235 427.5 1.740 -

3 35 8.876 0.000857 249.12 1.163 0

4 -20 1.337 0.0525 249.1 1.120 0.355

Solution

From the R 134a thermodynamic properties tables the enthalpy at each point can be calculated and tabulated as shown in the above table.

Refrigerating effect =h1 - h4 = 136.9 kJ/kg

Net Work (Heat of Compression) =h2 - h1 = 41.5 kJ/kg

Total heat rejected at condenser = h2-h3 = 178.4 kJ/kg

C.O.P.= Refrigerating effect/ Net Work = 136.9/41.5 =3.3

System Capacity = m (Refrigerating effect) = 0.05 *136.9 =6.845 kW.

1 =h1 = hg at T= -20

oC

2 h3 =hf at T= 35

oC

24

25

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2.8 The Practical Vapor Compression Cycle

Although real refrigeration systems do not behave completely in the idealized manner just covered the degree of deviation is small enough for analyses based on the ideal model to be adequate. The practical cycle differs from the idealized cycle in the following ways:

i- Compression is neither reversible nor adiabatic, there being both heat transfers and frictional effects in the compressor.

ii- There is a small pressure drop as the fluid passes through the condenser. Also, some sub-cooling of the liquid usually occurs.

iii- The throttling process is unlikely to be adiabatic.

iv- There is usually a small pressure drop through the evaporator and also, for practical reasons, it is normal to ensure that there is a small amount of superheat at compressor suction.

The state points of a practical cycle are as shown in Figure below.

In the following sections we will examine how the real systems do deviate from the model in quantitative terms. Also the effect of suction and condensing temperatures are taken into account and their effect on the cycle C.O.P.is studied in details

Note: Example 2.3 is used in studying the effect of the above mentioned parameters.

2

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i-The effect of evaporating temperature on cycle C.O.P.

A: ( -20 oC) evaporating Temp.) B: ( -10 oC ) evaporating Temp.)

State h (kJ/kg) State h (kJ/kg)

Cycle 1-2-3-4 1 386 Cycle 1‘-2‘-3-4‘ 1‘ 392.285

qe = h1 – h4 2 427.5 qe = h1‘ – h4‘ 2‘ 424.0

w = h2 – h1 3 249.1 w = h2‘ – h1‘ 3 249.1

qc = h2– h3 4 249.1 qc = h2‘ – h3 4‘ 249.1

Mass flow rate per kW of refrigerating capacity

sec/0073.09.136

11)20(

41

kgkW

hh

kWCo

sec/00698.0185.143

11)10(

'4'1

kgkW

hh

kWCo

%536.41000073.0

00698.00073.0rate flow mass in the decrease %

Although the difference in the mass flow rate per kilowatt of refrigerating capacity at the various vaporizing temperature is usually relatively small. The volume of vapor that the compressor must handle per second per kilowatt varies greatly with changes in the vaporizing temperature. This is probably one of the most important factors influencing the capacity and efficiency of a vapor-compression refrigerating system and is the one that is most likely to be overlooked by the trainee when studying cycle diagrams. The difference in the volume of vapor to be displaced per second per kilowatt at the various suction temperatures can be clearly illustrated by a compression of the two cycles in question.

at (-20 oC) , m (v) = 0.0073 * 0.146491 = 0.0010694 m3/sec

at (-10 oC) , m (v) = 0.00698 * 0.0992092 = 0.0006925 m3/sec

It is of interest to note that, whereas the decrease in the mass flow rate at the higher suction temperature is only 4.536 %, the decrease in the volume of vapor handled by the compressor per kilowatt of refrigerating capacity is

%244.351000010694.0

0006925.00010694.0

Refrigerating effect

at (-20 oC) = h1 – h4 = 386-249.1= 136.9 kJ/kg

1 v = vg at T = -20

oC.

2 v = vg at T = -10

oC.

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at (-10 oC) = h1‘ – h4‘ = 392.285-249.1= 143.185 kJ/kg

%6.41009.136

9.136185.143%

effectingrefrigeratinincrease

Work of compression

at (-20 oC) = h2 – h1 = 427.5 - 386 = 41.5 kJ/kg

at (-10 oC) = h2‘ – h1‘ = 424 -392.285 = 31.715 kJ/kg

%58.231005.41

715.315.41%

ncompressioofworkindecrease

The theoretical power required per kilowatt of refrigerating capacity for the –20 oC and –10 oC cycles respectively

at (–20 oC ) Pt = m (h2 – h1) = 0.0073 (427.5 – 386)= 0.302 kW

at (–10 oC ) Pt = m (h2‘ – h1‘) = 0.00698 (424 -392.285)= 0.2214 kW

In this instance, increasing the vaporising temperature of the cycle from –20 oC to –10 oC reduces the theoretical power required by

[(0.302 – 0.2214)/0.302 ] x 100 = 26.7 %

The rate of heat rejected at the condenser per kW of refrigerating Capacity

at (-20 oC) = m (h2 – h3) = 0.0073 (427.5 – 249.1) = 1.302 kJ/sec

at (-10 oC) = m (h2‘ – h3) =0.00698 (424 –249.1) = 1.220 kJ/sec

The rate of heat rejection at the condenser is lower for the higher suction temperature because (1) the lower mass flow rate and (2) the smaller heat of compression per unit mass.

It has been shown previously that the heat rejected at the condenser per unit mass of refrigerant circulated is the sum of the heat absorbed in the evaporator per unit mass (Refrigerating effect) and the heat of compression per unit mass (work of compression). Since increasing the vaporising temperature of the cycle brings about an increase in the refrigerating effect as well as a decrease in the heat of compression, the quantity of heat rejected at the condenser per unit mass remains nearly close for both cycles (178.4 kJ/kg at -20 oC. as compared to 174.9 kJ/kg at -10 oC. In general, this is true for all vaporising temperatures because any increase in the heat of compression is usually accompanied by an offsetting increase or decrease in the refrigerating effect.

The Coefficient Of Performance

3.33865.427

1.249386)20(

12

41

hh

hhCo

5.4285.392424

1.249285.392)10(

'1'2

4'1

hh

hhCo

It is evident that the coefficient of performance of the cycle improves considerably as the vaporising temperature increases. In this instance, increasing the vaporising temperature from –20 oC to-10 oC increases the C.O. P. of the cycle by

%36.361003.3

3.35.4...%

POCinincrease

29

ii-The effect of condensing temperature on cycle C.O.P.

A: ( 35 oC) Condensing Temp.) B: ( 45 oC ) Condensing Temp.)

State h (kJ/kg) State h (kJ/kg)

Cycle 1-2-3-4 1 386 Cycle 1‘-2‘-3-4‘ 1 386

qe = h1 – h4 2 427.5 qe = h1 – h4‘ 2‘ 433.0

w = h2 – h1 3 249.1 w = h2‘ – h1 3‘ 264.1

qc = h2– h3 4 249.1 qc = h2‘ – h3‘ 4‘ 264.1

Mass flow rate per kW of refrigerating capacity

sec/0073.09.136

11)35(

41

kgkW

hh

kWCo

sec/0082.0185.143

11)45(

'41

kgkW

hh

kWCo

%325.121000073.0

0073.00082.0rate flow mass in the increase %

It is of interest to note that, whereas the increase in the mass flow rate at the higher condensing temperature is only 12.325 %, the increase in the volume of vapor handled by the compressor per kilowatt of refrigerating capacity is

at 35 oC and -20 oC , m (v) = 0.0073 * 0.14649 = 0.0010694 m3/sec

at 45 oC and -20 oC , m (v) = 0.0082 * 0.14649 = 0.0012012 m3/sec

%325.121000010694.0

0010694.00012012.0

Note that the percent increase in the volume of vapor handled by the compressor is exactly equal to the percent increase in the mass flow rate.

Refrigerating effect

at (35 oC) = h1 – h4 = 386-249.1= 136.9 kJ/kg

at (45 oC) = h1 – h4‘ = 386-264.1= 121.9 kJ/kg

%957.101009.136

9.1219.136%

effectingrefrigeratindecrease

Work of compression

at (35oC) = h2 – h1 = 427.5 - 386 = 41.5 kJ/kg

at (45 oC) = h2‘ – h1 = 433 -386 = 47.0 kJ/kg

30

%25.131005.41

5.4147%

ncompressioofworkinincrease

The theoretical power required per kilowatt of refrigerating capacity for the 35 oC and 45 oC cycles respectively

at (35oC ) Pt = m (h2 – h1) = 0.0073 (427.5 – 386)= 0.302 kW

at (45oC ) Pt = m (h2‘ – h1) = 0.0082 (433 -386)= 0.3854 kW

In this instance, increasing the condensing temperature of the cycle from 35oC to 45oC increases the theoretical power required by

[(0.3854 – 0.302)/0.302 ] x 100 = 27.615 %

This increase is accounted for the increase in mass of refrigerant circulated per unit of capacity.

The rate of heat rejected at the condenser per kW of refrigerating Capacity

at (35oC) = m (h2 – h3) = 0.0073 (427.5 – 249.1) = 1.302 kJ/sec

at (45oC) = m (h2‘ – h3‘) = 0.0082 (433.0 – 264.1) = 1.385 kJ/sec

The rate of heat rejection at the condenser is higher for the higher condensing temperature because (1) the higher mass flow rate and (2) the higher heat of compression per unit mass.

The percent increase is

[(1.385-1.302)/1.302] x 100 = 6.375%

The Coefficient Of Performance

3.33865.427

1.2493862035

12

41

hh

hhCandCat oo

6.2386433

1.2643862045

1'2

'41

hh

hhCandCat oo

It is evident that the coefficient of performance of the cycle decreases considerably as the condensing temperature increases. In this instance, increasing the condensing temperature from 35oC to 45 oC decreases the C.O. P. of the cycle by

%212.211003.3

6.23.3...%

POCindecrease

Comment: Since the capacity and C.O.P of a refrigerating system improve significantly as the evaporator temperature is increased, it should always be designed to operate at the highest practical evaporator temperature. Although the effect of the condensing temperature on the capacity and C.O.P. of the refrigeration cycle is considerably less than that of the evaporator temperature, it nonetheless should be kept as low as is practical.

31

iii- The effect of superheating the suction vapor on Cycle C.O.P.

Superheat definition: Is the temperature difference between the refrigerant temperature at the compressor suction point minus the refrigerant saturation temperature at the evaporator pressure.

Superheat setting: For thermostatic expansion valves, it is generally considered between 5-7 oC of superheat is the

minimum for stability. 4-5% reduction in evaporator rating (UA) per 1 oC increase in superheat. Also superheat

temperature should not be more than the evaporator temperature difference (TD) should, otherwise the evaporator will starve to get the required superheat.

Minimum TD : One of the most important factors to be considered in selecting the proper evaporator for a given application is the evaporator TD which is really depends on RH in the space that desired to be cooled.

Relative Humidity (RH)1 % TD

oC / Forced Convection

95 91 5 6

90 86 6 7

85 81 7 8

80 76 8 9

75 70 9 10

Table 4.1

How to Determine Superheat correctly

In order to determine the superheat correctly the following procedure should be followed;

1. Measure the temperature of the suction line at the point where the remote bulb is clamped.

2. Obtain the suction pressure that exists in the suction line at the remote bulb location by either fitting the pressure gauge in the external equalizer or read the suction pressure at the suction valve of the compressors and add the pressure drop between the remote bulb location and the compressor.

3. Find the saturation temperature correspondent to the pressure that calculated in step (2), which can be obtained from the ph diagram or from the pressure gauge directly.

4. The superheat temp can be calculated by subtracting the two temperatures obtained in 1 and 3.

For thermostatic expansion valves, it is generally considered between 5-7 oC of superheat is the minimum for stability. 4-

5% reduction in evaporator capacity per 1 oC increase in superheat. Also superheat temperature should not be more

than the evaporator TD, otherwise the evaporator will starve to get the required superheat.

A: Saturated Cycle B: 10 oC Superheated Cycle

State h (kJ/kg) State h (kJ/kg)

Cycle 1-2-3-4 1 386 Cycle 1’-2’-3-4 1‘ 394

qe = h1 – h4 2 427.5 qe = h1‘ – h4 2‘ 436

w = h2 – h1 3 249.1 w = h2‘ – h1‘ 3 249.1

1 RH will be considered in the next chapter

32

qc = h2– h3 4 249.1 qc = h2‘ – h3 4 249.1

Mass flow rate per kW of refrigerating capacity

sec/0073.09.136

11

41

kgkW

hh

kWcycleSaturated

sec/0069.09.144

11

4'1

kgkW

hh

kWcycledSuperheate

%48.51000073.0

0069.00073.0rate flow mass in the decrease %

It is of interest to note that, whereas the decrease in the mass flow rate at the higher superheat temperature is only 5.48 %, the decrease in the volume of vapor handled by the compressor per kilowatt of refrigerating capacity is

Saturated cycle, m (v) = 0.0073 * 0.14649 = 0.0010694 m3/sec

Superheat cycle, m (v) = 0.0069 * 0.1535 = 0.00105915 m3/sec

%11000010694.0

00105915.00010694.0

33

Refrigerating effect

Saturated cycle = h1 – h4 = 386-249.1= 136.9 kJ/kg

Superheated cycle = h1‘ – h4‘ = 394-249.1= 144.9 kJ/kg

%844.51009.136

9.1369.144%

effectingrefrigeratinincrease

Work of compression

Saturated cycle = h2 – h1 = 427.5 - 386 = 41.5 kJ/kg

Superheated cycle = h2‘ – h1‘ = 436 -394 = 42.0 kJ/kg

%2.11005.41

5.4142%

ncompressioofworkinincrease

The theoretical power required per kilowatt of refrigerating capacity for the saturated and superheated cycles respectively

Saturated cycles, Pt = m (h2 – h1) = 0.0073 (427.5 – 386)= 0.302 kW

Superheated cycle, Pt = m (h2‘ – h1‘) = 0.0069 (436 -394)= 0.2898 kW

In this instance, the superheat decreases the theoretical power required by

[(0.302 – 0.2898)/0.302 ] x 100 = 4.35 %

Comment: Even through the specific volume of the suction vapor and the compression work per unit mass are both greater for the superheated cycle that for the saturated cycle, the volume of vapor compressed per unit capacity and power required per unit capacity are both less for the superheated cycle than the saturated cycle. This is because of the reduction in the mass flow rate.

The rate of heat rejected at the condenser per kW of refrigerating Capacity

Saturated cycle = m (h2 – h3) = 0.0073 (427.5 – 249.1) = 1.302 kJ/sec

Superheated cycle = m (h2‘ – h3) = 0.0069 (436.0 – 249.1) = 1.2896 kJ/sec

The percent decrease is

[(1.302-1.2896)/1.302] x 100 = 1%

34

The Coefficient Of Performance

3.33865.427

1.249386

12

41

hh

hhcycleSaturated

45.3394436

1.249394

'1'2

'4'1

hh

hhcycledSuperheate

It is evident that the coefficient of performance for the superheated cycle is higher than that of saturated cycle.

%545.41003.3

3.345.3...%

POCinincrease

35

iv- The effect of sub-cooling the liquid on the cycle C.O.P.

A: Saturated Cycle B: 3 oC Sub-Cooled Cycle

State h (kJ/kg) State h (kJ/kg)

Cycle 1-2-3-4 1 386 Cycle 1-2-3‘-4‘ 1 386

qe = h1 – h4 2 427.5 qe = h1 – h4‘ 2 427.5

w = h2 – h1 3 249.1 w = h2 – h1 3‘ 244.7

qc = h2– h3 4 249.1 qc = h2 – h3‘ 4‘ 244.7

-Mass flow rate per kW of refrigerating capacity

sec/0073.09.136

11

41

kgkW

hh

kWcycleSaturated

sec/007077.03.141

11

'41

kgkW

hh

kWcycleCooledSub

The decrease in the volume of vapor handled by the compressor per kilowatt of refrigerating capacity is

%055.31000073.0

007077.00073.0rate flow mass in the decrease %

Saturated cycle, m (v) = 0.0073 * 0.14649 = 0.0010694 m3/sec

Sub-Cooled cycle, m (v) = 0.007077 * 0.14649 = 0.0010367 m3/sec

%058.31000010694.0

0010367.00010694.0

Refrigerating effect

Saturated cycle = h1 – h4 = 386-249.1= 136.9 kJ/kg

Sub-Cooled cycle = h1 – h4‘ = 386-244.7= 141.3 kJ/kg

%214.31009.136

9.1363.141%

effectingrefrigeratinincrease

Work of compression

Saturated cycle = h2 – h1 = 427.5 - 386 = 41.5 kJ/kg

Sub-Cooled cycle h2 – h1 = 427.5 - 386 = 41.5 kJ/kg

The theoretical power required per kilowatt of refrigerating capacity for the saturated and superheated cycles respectively

Saturated cycle or Sub-Cooled cycle, Pt =m(h2 – h1)=0.0073*(427.5–386)= 0.302 kW

The rate of heat rejected at the condenser per kW of refrigerating Capacity

36

Saturated cycle = m (h2 – h3) = 0.0073 (427.5 – 249.1) = 1.302 kJ/sec

Sub-Cooled cycle = m (h2 – h3‘) = 0.007077 (427.5.0 – 244.7) = 1.29367 kJ/sec

The percent decrease is

[(1.302-1.29367)/1.302] x 100 = 0.6%

The Coefficient Of Performance

3.33865.427

1.249386

12

41

hh

hhCycleSaturated

405.33865.427

7.244386

12

'41

hh

hhcyclecooledSub

%182.31003.3

3.3405.3...%

XPOCinincrease

It is evident that the coefficient of performance for the superheated cycle is higher than that of saturated cycle.

Comment: Sub-cooling of the liquid refrigerant can does occur in several places and several ways. Very often the liquid refrigerant becomes sub-cooled while stored in liquid receiver tank or while passing through the liquid by giving off heat to the surroundings air. In some cases, a special liquid sub-cooler is used to sub-cool the liquid. The gain in system capacity and C.O.P. resulting from sub-cooling is very often more than sufficient to offset the additional cost of the sub-cooler, particularly for low temperature applications.

37

v- The effect of pressure drop

Figure 4.2

From the energy viewpoint, the two areas where pressure drops are most important are the suction and discharge lines. The compressor energy use is a function of the difference between the suction and discharge pressure, whereas the condenser and evaporator performances are related to the condensing and evaporation temperatures, which are a function of the condensation and evaporation pressures. As the pressure drop in the pipelines decreases, the compressor energy use is decreased. This is shown on the diagram above, in which for simplicity, heat gains or losses to the pipes are ignored. It can be seen that the size of the pipeline pressure drops affects the enthalpy change in compression, and thus the energy that the compressor must supply. To do a detailed analysis , it is necessary to include the effect of pressure drops, but in the present course, for a simplified analysis it can be assumed that:

Ps = Pe

Pd = Pc

To choose the pipe diameters needed to achieve small pressure drops, the refrigerant mass flow must be calculated.

As liquid sub-cooled and superheating occurs in practice they are often taken into account in the calculations too.

38

2.8- Revision questions

1- Define the following;

a- Refrigeration b- Refrigeration Load c- Refrigeration effect d- System Capacity

2- Using the refrigerant thermodynamic tables, quote system pressure corresponding to :

a- R22 condensing temperature at 40 oC. b- R134a suction temperature of –10 oC. c- Ammonia (NH3) suction temperature of –25 oC.

3- Sketch a ph diagram for an ideal cycle of a refrigerant and show the following;

Sub-cooled, saturation, and superheated regions.

Constant quality, constant volume, constant temperature, and constant entropy lines.

Latent heat of vaporization.

Work of compression.

Refrigerating effect.

Total heat rejected at condenser.

Latent heat rejected at condenser.

Loss in the expansion process.

Sensible heat.

Quality (Dryness fraction factor) formula

4- Consider an ideal refrigeration cycle which uses R22 as the working fluid. The temperature of the refrigerant in the evaporator is –30 oC and in the condenser it is 45 oC. The refrigerant is circulated at the rate of 0.03 kg/s.

(a) Plot the cycle on the ph diagram. (b) Find the discharge temperature and entropy. (c) Determine the C.O.P. and the system capacity

(d) Determine the Refrigerating efficiency (r).

5-A vapor compression refrigeration cycle uses R134a and follows the theoretical single stage cycle. The condensing temperature is 48 oC and the evaporating temperature is -18 oC. Power input to the cycle is 2.5 kW and mass flow rate of refrigerant is 0.05 kg/s. Determine

a- Heat rejected from the condenser b- The Coefficient of Performance c- The enthalpy at the compressor exit, and d- The refrigerating efficiency.

39

6-A Refrigeration System employing R22 is operating on a simple saturated cycle with an evaporator temperature of 0oC and condensing temperature of 45oC. Determine the following

a- The properties of P, T, v , h and s for state points 1, 2,3 and 4.

b- The compressor displacement in m3/sec required per kW of refrigerating capacity.

c- The theoretical power in Watt required per kW of refrigerating capacity.

d- The total heat rejected at the condenser in kW per kW of refrigerating capacity.

e- The C.O.P. of the cycle.

f- The Refrigerating efficiency (r)

7- An air conditioning plant is served by R134a compressor, having a cooling capacity of 10 ton of refrigeration (1 ton = 3.5 kW) at an evaporation temperature –5 oC, and a condensing temp. 35oC.

Determine:

i- the compressor work for the compressor

ii- the cooling capacity and the compressor work, when it operates at an evaporating temp. of 5 oC and a condensing temp. of 40 oC.

iii- The Compressor work, system capacity, the C.O.P. and Refrigeration efficiency, when the refrigerant enters the compressor with 10 oC superheat and leaves the condenser with 5 oC sub-cooling

40

Chapter 3

Refrigeration systems 3.1- Objectives

The objective of this course is to outlines the most relevant refrigeration cycles and to assist graduate engineers to become more effective refrigeration engineers through understanding how the refrigeration cycles are analyzed. These objectives will be pursued through the application of refrigeration cycles theory to practical examples.

3.2- Industrial refrigeration

Industrial refrigeration is often confused with commercial refrigeration because the division between these two areas is not clearly defined. As a general rule, industrial applications are larger in size than commercial application and have the distinguishing feature of requiring an attendant on duty, usually a licensed operation engineer. Typical industrial application are ice plants, large food packing plants (meat, fish, frozen foods, etc.), creameries, and industrial plants, such as oil refineries, chemical plants, rubber plants, etc.

3.3- Commercial Refrigeration

Commercial refrigeration is concerned with the designing, installation, and maintenance of refrigerated fixtures of the type used by retail stores, restaurants, hotels and institutions for the storing, displaying, processing, and dispensing of perishable commodities of all types.

3.4- Food Preservation

The edibility of foodstuffs can be prolonged by lowering the temperature, since this slows chemical reactions and breakdown by bacteria. Some products can be frozen, and when they are in the solid state all movement in the individual cells will cease, inhibiting further reactions. The preservation, of perishable commodities, particularly foodstuffs, is one of the most common uses of mechanical refrigeration. As such, it is a subject which should be given consideration in any comprehensive study of refrigeration.

Food preservation is one of the significant applications of refrigeration, whether it is by way of processing or for storage. Processing is done by chilling freezing or freeze-drying. Storage may be of either a chilled or frozen product. Some of the important products involved in processing are candy, beverages, meat, poultry, fish, bakery and dairy products, fruits and vegetables, fruit-juice concentrates, precooked foods, etc. The common products preserved by storage after chilling are fruits such as apples, pears, grapes, citrus fruits, etc., vegetables such as onions, potatoes, tomatoes, etc, dry fruits, candies, milk, eggs and their products. Storage under frozen conditions is resorted to for preserving the food value as well as perishable products over a long period. The common items of frozen food are fish, meat, poultry, and some vegetables such as pears, beans, carrots, cauliflower, etc.

3.5- Refrigeration systems COP improvement methods A-Liquid/Suction lines heat exchangers

Condenser

P Pc 3 2

3 2

41

B-Multi-stage expansion i- Direct staging 1-VCC with FC

ececi

eci

TTPPP

or

PPP

T should be in K

Analysis A- Energy balance for the FC gives

56

75Lf

56f75L

6f7L5f5L

6f7L5fL

hh

hhmm

hhmhhm

hmhmhmhm

hmhmhmm

B-Refrigeration Load

81Le hhmQ

C- Compression Work

12L69f2c1cc hhmhhmWWW

D-Total heat rejected at condenser

43fLc hhmmQ

fL

9f2L

3

9f2L3fL

mm

hmhmh

hmhmhmm

Condenser

Evaporator

Flash

Chamber

Comp. # 1

Comp. # 2

h

P Pc

Pe

1 8

7

5

4

6

2 9

1

2

7

5

4

9

3

Lm

fm

Lf mm

6 EV #1

EV# 2

3

Liq. Receiver

fm

Lm

Lf mm

8

Pi

42

2-VCC with BPV

Analysis

A- Energy balance for the FI gives

45

74Lf

hh

hhmm

B-Refrigeration Load

81Le hhmQ

C- Compression Work

12Lfc hhmmW

fL

6f9L

1

6f9L1fL

mm

hmhmh

hmhmhmm

Condenser

Evaporator

Flash

Chamber

h

P Pc

Pe

1 8

7

4

3

5

2

1

2

7

4

3

Lm

fm

Lf mm

5 EV #1

EV# 2

Liq. Receiver

fm

Lm

Lf mm

6

9 BPV

8

9

6

Pi

43

3-VCC with FI (Open type)

Analysis A- Energy balance for the FI gives

56

62L

I

62L56I

72L56I56

56

75L56L

56

75Lf

72L56I56fL

72L56IfL

5IfL2L7L6IfL

hh

hhmm

hhmhhm

hhmhhmhhhh

hhmhhm

hh

hhmm

hhmhhmhhmm

hhmhhmmm

hmmmhmhmhmmm

Im : is the required mass for desuperheating the L.S.C. discharge refrigerant

B-Refrigeration Load

81Le hhmQ

C- Compression Work

63IfL12LHSCLSCc hhmmmhhmWWW

Evaporator

Condenser

F.I.

(open type) L.S.C.

H.S.C

1

2

6

7

4

3

5

8

I

fL

m

mm

Lm

IfL mmm

Hm

Pi

Liq. Receiver

h

P Pc

Pe

1 8

7 5

4

6 2

3

44

4-VCC with closed type FI (subcooler intercooler)

Analysis A- Energy balance for the FI gives

56

6274LIs

erheatingsupde

74L

subcooling

62L

erheatingsupde

56I

subcooling

56s

6IsL7L2L4L5Is

hh

hhhhmmm

hhmhhmhhmhhm

hmmmhmhmhmhmm

sm : is the required mass for subcooling the refrigerant before enters the evaporator

B-Refrigeration Load

81Le hhmQ

C- Compression Work

63IsL12LHSCLSCc hhmmmhhmWWW

Evaporator

Condenser

F.I.

(closed type)

L.S.C.

H.S.C

1

2

6

7

4

3

5

8

Is mm

Lm

HIsL mmmm

Pi

Liq. Receiver

h

P Pc

Pe

1 8

7 4 3

5 2

6

4 Lm

45

3.6 Review of Multi-Evaporator Vapor Compression Systems

This section is a review and continuity to what has been studied at ―Refrigeration systems and Equipment‖ subject. But, the student will go a step further in making the use of C.O.P. for comparing different systems for the same application. You may ask why? Because I personally believe that:

The C.O.P. is a function of the plant design and depends on the relative sizes of the various cooling loads at the various suction temperatures, and C.O.P. gives us a hint on how much energy required (Operating Costs) to the proposed plant.

Example 3.1

Assume, you are a refrigeration engineer working with Al-Salem York J.V., one of your customers asked you to give him a proposal for a refrigeration system uses R134a to supply two stores. The capacity of store # 1 evaporators is 50 T.R. and the evaporation temperature is 5 oC, while the capacity and the evaporation temperature of store # 2 evaporators are 50 T.R. and –20 oC respectively. The condensing temperature is 40 oC. What is your proposal and Why?

Solution

PROPOSAL 1. Single Compressor-Individual Expansion Valves

From R134a Ph diagram, we have

h3 = h4 = h5 = 106 kJ/kg

h6 = h7 = 250.5 kJ/kg

h8 = 235 kJ/kg

skghh

m /21.11065.250

175505.3

46

1

Evaporator # 2

Condenser

Evaporator # 1

Comp.

B.P.V.

h

P

2m

1m

1

2

3

4

5

8

6 7

21 mmmt

Pc

Pe1

Pe2

7

6 4

5

3 2

1

8

46

skghh

m /356.1106235

175505.3

58

2

h1 can be calculated from the Energy Balance around point 1.

8271121 hmhmhmm

kJ/kg3.242356.121.1

66.3181.303

356.121.1

235356.15.25021.1

1

h

Then h2 = 287 kJ/kg

kW114.7242.3)(2872.5661221

h-hmmWc

05.37.114

350

7.114

5.3)5050(

C.O.P.

47

PROPOSAL 2. Individual Compressors – Individual Expansion Valves

From R134a Ph diagram, we have

h4 = h5 = h7 = 106 kJ/kg

h1 = 235 kJ/kg , h3 = 272 kJ/kg

h6 = 250.5 kJ/kg , h2 = 278 kJ/kg

skghh

m /21.11065.250

175505.3

56

1

skghh

m /356.1106235

175505.3

71

2

kWh-hmh-hmt

W 84.323235)-1.356(278250.5)1.21(2721263 21

15.4323.84

5.3)5050(...

t

t

W

QPOC

Condenser

Evaporator # 1

Evaporator # 2

Comp. # 1

Comp. # 2

1

2

4

3

5

7

h

P

1

2 3

6 5

7

4 Pc

Pe1

Pe2

6

2m

1m

48

PROPOSAL 3 Individual Compressors-Individual Expansion Valves with Flash Chamber

From R134a Ph diagram, we have

h4 = h5 = h7 = 106 kJ/kg

h1 = 235 kJ/kg , h3 = 272 kJ/kg

h6 = 250.5 kJ/kg , h2 = 255 kJ/kg

skghh

m /21.11065.250

175505.3

56

1

skghh

m /356.1106235

175505.3

71

2

The refrigerant mass that is required to de-superheated the discharge of low stage compressor

skg

hh

hhmmI /04223.0

1065.250

5.250255356.1

56

622

kW

h-hmh-hIt

mmmW

55.84235)-1.356(256250.5)(2720.04223)1.3561.21(

126321 2

14.455.84

5.3)5050(...

t

t

W

QPOC

Condenser

Evaporator #1

Evaporator # 2

Flash

Chamber

HSC

LSC

h

P

1

2

3

6 5

7

4 Pc

Pe1

Pe2

1

2

6

3

4

5

5

7 2m

Im

1m

It mmmm 21

49

PROPOSAL 4. Single Compressor – Multiple Expansion Valves with F.C.

From R134a Ph diagram, we have

h3 = h4 = 106 kJ/kg

h7 = h8 = 56.5 kJ/kg

h5 = h6 = 250.5 kJ/kg , h9 = 235 kJ/kg

skghh

m /21.11065.250

175505.3

45

1

skghh

m /98.05.56235

175505.3

89

2

Energy balance at Point 4

skg

hh

hhmm f /3357.0

1065.250

5.5610698.0

45

742

Energy Balance at point 1.

1216192 hmmmhmmhm ff

h1 = 244.5 kJ/kg

Then h2 = 289.5 kJ/kg

kWhhmmmW fc 65.1135.2445.2895257.21212

08.365.113

350... POC

Condenser

Evaporator #1

Evaporator # 2

Flash

Chamber

B.P.V.

Comp.

9

1

6

5 4

4

3 2

8

7

h

P Pc

Pe1

Pe2

1 6

9

8

7 4

3

5

2

1m

2m

fm

2mm f

50

PROPOSAL 5 Individual Compressors – Multiple Expansion Valves with F.C.

From R134a Ph diagram, we have

h9 = 273 kJ/kg

h2 = 279 kJ/kg

h1 = 235 kJ/kg

h6 = 250.5 kJ/kg

691122 hhmmhhmW fc

9.77250.5-2733357.021.123527998.0

493.49.77

350... POC

Condenser

Evaporator #1

Evaporator # 2

Flash

Chamber

Comp. # 1

Comp. # 2

1m

h

P Pc

Pe1

Pe2

1 8

7 5

4

6

2 9

1

2

7

8

5

5

4

6

9

3

2m

fm

2mm f

51

PROPOSAL 6. Individual Compressors – with Compound Compression

From R134a Ph diagram, we have

h1 = 235 kJ/kg

h3 = 272 kJ/kg

h4 = h5 = 106 kJ/kg

h2 = 255 kJ/kg

h6 = 250.5 kJ/kg

h7 = h8 = 56.5 kJ/kg

skghh

mm L /98.05.56235

175505.3

81

2

Considering the Evaporator # 1 and the flash chamber as the control volume. Then the Energy balance,

skgmhhmhhm HH /557.25.350 56722

kWhhmhhmW Hc 575.745.250272557.223525598.063122

693.4575.74

350... POC

Hm can be evaluated by using another method

skg

hh

hhmmI /03052.0

1065.250

5.25025598.0

56

622

skgmmmmm IfH /556.203052.03357.098.021.121

Evaporator # 2

Condenser

Evaporator #1

Flash

Chamber L.S.C.

H.S.C

h

P Pc

Pe1

Pe2

1 8

7 5

4

6 2

3

1

2

6

7

5

4

3

5

8

I

f

m

mm

2

1m

Lmm 2

If mmm 2

Hm

52

Comparison Table

Proposed System # C.O.P kW/T.R.

1 3.05 1.1475

2 4.15 0.8434

3 4.14 0.8454

4 3.08 1.1364

5 4.493 0.7790

6 4.693 0.7458

As can be seen from the above comparison table, the operating cost of system # 6 is the cheapest‘ whilst the capital cost is more expensive because of the flash chamber and two compressors. But the difference in capital cost compare to the systems without a flash chamber and one compressor can be paid back after a certain period from the saving that incurred in operating cost.

53

Example (1)

54

A two stage ammonia food-freezing plant as shown in Figure 1. The desired cooling capacity is 561600 kJ/h at -40 oC evaporating temperature and 35 oC condensing temperature. The system has a flash intercooling with liquid subcooler where the refrigerant temperature subcooled by 30 oC. The vapor leaving the evaporator at -30 oC and entering the first stage compressor is at -10 oC. The vapor leaving the flash intercooler is superheated by 10 oC in the suction line to the second stage compressor. Adiabatic efficiencies of the compressors are 0.8. Find the discharge temperatures, the vapor volume

(V1) of the LSC, power requirements of the two compressors and the COP.

Figure 1

Evaporator

Condenser

F.I.

(closed type)

L.S.C.

H.S.C

1

2

6

7

4

3

5

8

Is mm

Lm

HIsL mmmm

Pi

Liq. Receiver

4 Lm

55

Example (2) A two stage ammonia system uses centrifugal compressors, flash intercooler and individual expansion valves as shown in Figure 2. The system operating conditions are as follows

Refrigeration capacity evapQ = 100 T.R.

Evaporating pressure = 0.1MPa.

The vapor leaving the evaporator at -30 oC.

Condensing temperature =30 oC.

The gas leaving the second stage compressor at 95 oC.

Flash intercooler pressure = 0.35 MPa.

The gas leaving the first stage compressor at 70 oC.

The liqiud refrigerant leaving the condenser at 25 oC. Pressure losses due to friction and heat losses are negligible and assume the refrigerant enters the second

stage as saturated vapor. Find the vapor volume (V1) of the LSC and HSC, power requirements of the two

compressors and the COP.

Figure 2

Evaporator

Condenser

F.I.

(open type) L.S.C.

H.S.C

1

2

6

7

4

3

5

8

Pi

Liq. Receiver

56

Example (3)

A two stage ammonia system as shown in Figure 3 (open flash intercooler, load at intermediate pressure). This refrigeration system supplys two cold stores. The capacity of store # 1 evaporators is 20 T.R. and the evaporating pressure is 300 kPa, while the capacity and the evaporating pressure of store # 2 evaporators are

30 T.R. and100 kPa respectively. The condensing temperature is 30 oC, and the refrigerant leaves the

evaporators as saturated vapor and the condebser as saturated liquid. Find the power requirements of the two compressors and the COP.

Figure 3

Evaporator # 2

Condenser

Evaporator #1

Flash

Intercooler L.S.C.

H.S.C

1

2

6

7

5

4

3

5

8

1m

2m

57

ii- Cascade staging

The use of a single refrigerant in a single VCC for the production of low temperatures is limited by the following reasons: (i) Solidification temperature of the refrigerant. (ii) Extremely low pressures in the evaporator and large suction volumes if a high-boiling refrigerant is

selected. (iii) Extremely high pressures in the condenser if a low-boiling refrigerant is selected. (iv) Very high pressure ratio and, therefore, a low coefficient of performance. (v) Difficulties encountered in the operation of any mechanical equipment at very low temperatures. We know that multistage compression is employed when low evaporator temperatures are required and when the pressure ratio is high. Refrigerant 22 is used in a two-stage system up to -50°C and in a three-stage system up to about -65 °C. If vapor compression systems are to be used for the production of low temperatures, the common alternative to stage compression is the cascade system in which a series of refrigerants, with progressively lower boiling points, are used in a series of single-stage units. The system provides a solution to all the problems mentioned above except the last one.

Condenser

Evaporator

Condenser

Evaporator

HSC

LSC

1

6

4 3

5

8

h

P

1

2

3

6 5 7

4 Pc

Pe1

Pe2

7

1m

2m

8 2

Upper Cascade

Lower Cascade

58

Chapter 4

Refrigeration Equipments

Introduction

A basic refrigeration system is mainly composed of four components in addition to some auxiliary elements, which are connected together by means of pipes. These main four components are: a compressor, a condenser, an expansion device, and an evaporator. Many auxiliary elements may be used according to the applications such as: filter drier, sight glass, solenoid valve, liquid receiver, suction accumulator, non-return valve, regulator, heat exchanger, separator,. etc.).

I- Compressors

The compressor in a refrigeration system is the same as the heart in human body. It withdraws the refrigerant vapor from the evaporator and compresses it to increase its pressure and consequently its temperature to be able to loose it‘s heat to the ambient and condensate. Also, the compressor forces the refrigerant through the expansion device in which it‘s pressure and temperature are both decreased to able to absorb heat from the refrigerated space through the evaporator.

Generally, the compressors are categorized into two categories: External-Drive (Open) and Hermetic Compressors as shown in Figures (4.1 – 4.3).

Figure (4.1a) Open type reciprocating compressors

59

Figure (4.1 b) Four cylinder Open, V type reciprocating compressors

Figure (4.2) Semi-hermetic compressor

Cylinder

Connecting Rod

Piston

Crankshaft

60

Figure (4.3) Hermetic Scroll compressor

There are five basic types of compressors in use in refrigeration and air conditioning industry:

Reciprocating (Piston-cylinder)

Rotary

Scroll

Screw

Centrifugal

4.2.1- Reciprocating compressors

Reciprocating compressor is a positive displacement compressor. It consists of a piston moving back and forth in a cylinder with suction and discharge valves that arranged to allow pumping to take place. This piston is connected to a connecting rod, which is connected to a crankshaft. This crankshaft is rotating and the connecting rod changes this motion to reciprocating one. So, the piston is rotating back and forth in the cylinder.

Compressor capacity control

Because of the thermal load variation, the compressor capacity must be varied accordingly. Many methods are available for compressor capacity control:

ON – OFF of the compressor (small units), in which the thermostat stops and starts the compressor motor according to refrigerated space temperature.

Motor

winding

Crankshaft

Connecting Rod Discharge Line

connection

Body

Discharge Muffler

Electrical terminals

Suction Line

connection

Pistons

Motor

61

Loading and unloading of compressor cylinders as shown in Figure 4.6, in which the thermostat energizes the solenoid coil to pull the bypass piston for opening the bypass passage for unloading and vice versa.

Loaded Operation. When the suction pressure is above the control point the poppet valve will close. The discharge gas bleeds into the valve chamber, and the pressure closes the bypass piston and the cylinder bank loads up. Discharge gas pressure forces the check valve open permitting gas to enter the discharge manifold (Figure 4.6 a)

Unloaded Operation. When the suction pressure falls below the valve control point, the poppet valve will open. The discharge gas now bleeds from behind the bypass piston to the suction manifold. The bypass piston opens and the discharge gas is re-circulated back into the suction manifold and the cylinder bank is unloaded. A reduction in the discharge pressure will cause the check valve to close, isolating the cylinder bank from the discharge manifold (Figure 4.6 b)

Hot gas bypass from discharge to suction as shown in Figure 4.7, in which the thermostat open the hot gas solenoid to bypass it back to the compressor suction. This eliminates refrigerant to the cycle to decrease cooling capacity.

Compressor speed variation (car AC, variable speed motor), also, the new compressors have this facility

Figure (4.6 a) Loaded cylinder

62

Figure (4.6 b) Unloaded cylinder

Figure (4.7) Typical Piping arrangement for hot gas bypass circuit using hot gas pressure regulator

63

4.2.2- Rotary Compressors

Rotary compressor is a positive displacement type compressor. The displacement is obtained by rotary motion. It consists of eccentric rotor, cylinder, slide blade (vane), exhaust port as shown in Figure 8.8. Rotary compressors are commonly used to power small refrigerated appliances such as window air conditioners, packaged terminal air conditioners, and heat pumps up to five tons. There are two basic types of rotary compressors: stationary blade and rotating blade. The blades (vanes) on a rotating blade rotary compressor rotate with the shaft. The stationary blade has a blade that remains stationary and is part of the housing assembly. In both types, the blade provides a continuous seal for the refrigerant vapor. Figure (4.8) shows a typical rotating two-blade compressor. The low-pressure vapor from the suction line is drawn into the opening. The vapor fills the space behind the blade as it revolves. As the blades revolve, trapped vapor in the space ahead of the blade is compressed until it can be pushed into the exhaust line to the condenser.

A commercial rotary blade compressor, using eight blades, is pictured in Figure (4.9). The basic operation of the eight-blade compressor is the same as the two-blade.

Figure (4.10) represents a stationary blade (often called a divider block) rotary compressor. An eccentric shaft rotates an impeller in a cylinder. This impeller constantly rubs against the outer wall of the cylinder.

As the impeller (or roller) revolves, the blade traps quantities of vapor. The vapor is compressed into a smaller and smaller space. The pressure and temperature build up. Finally the vapor is forced through the exhaust port. It enters the high-pressure side of the system (condenser).

The compression action on one quantity of vapor takes place at the same time another quantity of vapor is filling the cylinder on the intake stroke. All of the parts must be fitted to extremely close tolerances and clearances. The dimensions are very accurate and the surfaces quite smooth. Therefore, no gaskets are needed in the compressor assembly.

Figure (4.8) A rotary blade compressor. Black arrows indicate direction of rotation of rotor

64

Figure (4.9) Eight-blade rotary compressor

65

Figure (4.10) A stationary blade rotary compressor

66

4.2.3- Scroll Compressors

The scroll compressor is commonly used in residential air conditioning and heat pump applications. Benefits of the scroll include fewer moving parts, less internal friction, smooth compression cycle with low torque, low noise levels, and low vibration levels. A scroll compressor generates a series of crescent-shaped gas pockets between two scrolls. Figure (4.11). One scroll—the fixed scroll—remains stationary. The other scroll—the orbiting scroll—rotates through the use of the swing link. As the motion occurs, the pockets between the two forms are slowly pushed to the centre of the two scrolls. This reduces the gas volume. When the pocket reaches the centre of the scroll, the gas is at a high pressure. It is discharged out of the centre port. During this compression process, several pockets are being formed at the same time. The suction process from the outer portion of the scroll and the discharge from the inner portion are continuous. This continuous process gives the compressor very smooth action.

Scroll compressor design is shown in Figure (4.12). A scroll compressor used on domestic room air conditioners is shown in Figure (4.3).

The scroll compressor has fewer moving parts and less torque variation than reciprocating compressors. This results in very smooth and quite operation.

Figure (4.11) Compression in the scroll is caused by the interaction of an orbiting scroll mated within a stationary scroll. 1— Gas is drawn into an outer opening as one of the scrolls orbits. 2— As the orbiting motion continues, the open passage is sealed off and the gas is forced to the centre of the scroll. 3— The pocket becomes progressively small in volume. This creates increasingly higher gas pressures. 4— Discharge pressure is reached at the centre of the pocket. Gas is released from the port of the stationary scroll member. 5— In actual operation, six gas passages are in various stages of compression at all times. This creates nearly continuous suction and discharge.

67

Figure (4.12) Scroll compressor design. A—two scrolls are used to produce a vapor compression. The upper scroll is stationary and the lower scroll is driven. Note intake and discharge ports. B—Note how the rotation of the motor shaft

causes the orbiting scroll to orbit—not rotate about the shaft centre.

4.2.4- Screw Compressors

Screw compressors are often used in large-capacity systems ranging from 20 to 300 tons. They are offered as open, externally-driven compressors, or hermetic, internally-driven compressors. Open screw compressors are most often used with ammonia systems. Hermetic screw compressors are used with halocarbon refrigerants.

The screw compressor uses a pair of special helical rotors. These trap and compress refrigerant as they revolve in an accurately machined compressor cylinder.

Figure (4.13) illustrates a cross section of a screw compressor. The two rotors are not the same shape. One is male, the other female. The male rotor, A, is driven by the motor. It has four lobes. The female rotor, B, meshes with and is driven by the male rotor. It has six interlobe (grooves) spaces. The cylinder, C, encloses both rotors.

In operation, the refrigerant vapor is drawn in as shown in Figure (4.14). The intake (low-pressure vapor) enters at one end of the compressor and is discharged (compressed vapor) at the opposite end.

The male rotor revolves more rapidly than the female rotor. (There are four lobes on the male rotor and six on the female rotor.) The rotors are helixes. They provide a continuous pumping action rather than pulsating as with a reciprocating compressor. With this pumping action, there is very little vibration during operation.

Orbiting Scroll

Suction gas

68

Figure (4.13) Cross section of screw compressor. A—Male rotor. B—Female rotor. C—Cylinder. Vaporized refrigerant enters at one end and exhausts at other end.

Figure (4.14) Basic operation of screw compressor. Revolving rotor compresses vapor. A—Compressor interlobe spaces being filled. B—Beginning of compression. C—Full compression of trapped vapor. D—Beginning of discharge of compressed vapor. E—Compressed vapor fully discharged from interlobe spaces.

Figure (4.15) illustrates a 3600-rpm, single screw compressor. It utilizes one main rotor that meshes with two diametrically opposed star-shaped gate rotors. The main rotor contains six grooves. It has straight roller bearings at the shaft ends. Two capacity control slide valves, one on each side, help to determine the capacity control.

Af

Am

69

Figure (4.15) Single screw compressor. Note the location of the main rotor in relation to the two gate rotors.

70

4.2.5- Centrifugal Compressors

Centrifugal compressors are designed for use with large-capacity systems ranging in size from 50 to 5,000 tons. In this type of compressor, vapor moves outward as it is moved rapidly in a circular path. This action is called centrifugal force. (However, the correct term is ―centripetal force.‖).

The vapor is fed into a housing near the centre of the compressor. A disk with radial blades (impellers) spins rapidly in this housing. This forces vapor against the outer diameter.

The pressure gained is small, so several of these compressor wheels or impellers are put in series. This creates greater pressure difference and pumps a sufficient volume of vapor. A centrifugal compressor looks like a steam turbine or axial flow air compressor for a gas turbine engine.

The centrifugal compressor has the advantage of simplicity. There are no valves or pistons and cylinders. The only wearing parts are the main bearings. Pumping efficiency increases with speed, so the compressors are designed to operate at high speeds.

Figure (4.16) is a cross section through a two-stage centrifugal compressor. The driving motor is mounted between stages. The inlet is at the left on the illustration. The discharge is in the back at the right end of the illustration and is not shown.

Figure (4.17), right view, shows a section through a hermetic centrifugal compressor. These compressors operate at a high speed. They are usually driven by an electric motor or steam turbine.

71

Figure (4.16) Two-stage centrifugal compressor. I—Second-stage variable inlet guide vane. 2—First-stage impeller. 3—Second-stage impeller. 4—Water-cooled motor. 5—Base, oil tank, and lubricating oil pump assembly. 6—First stage guide vanes and capacity control. 7—Labyrinth seal. 8—Cross-over connection. 9—Guide vane actuator. 10—Volute

casing. 11—Pressure-lubricated sleeve bearing. Note that discharge opening is not shown.

Figure (4.17). Hermetic centrifugal compressor. The impeller is shown at left above. Major components of the compressor are: A—Intake. B—First-stage impeller. C—Second-stage impeller. D—Hermetic motor. E—Exhaust.

II- Condensers

Introduction

72

Condenser is a heat exchanger withdraws the heat from the refrigerant gas of high pressure and temperature and releases it to the ambient to condense it. The compressor must therefore raise the pressure of the gases to the necessary saturation pressure at which the transformation from gas to liquid can take place by heat transfer to the available cooling medium, usually air or water or both, taking into account the temperature of the cooling medium.

When the gas enters the condenser, the superheat is first removed. Next, condensation takes place and, finally, the liquid may be subcooled. In all stages the beat-transfer coefficient of course is different. In the first stage it is very low, because on the refrigerant side there is only gas; in fact, in air-cooled condensers there is gas on both sides – refrigerant and air.

This is also the case in some evaporative condensers, where water consumption is economized by desuperheating using air only. In the second stage, the heat transfer is very good because of the condensation process. In the third stage of ubcooling, the heat-transfer coefficients are higher as there is liquid on both sides; except of course in the case of air-cooled condensers.

The condenser capacity

The condenser capacity that is related to the refrigerant mass flow rate can be expressed as,

eirc hhmQ (1)

where hi and he are the enthalpies of the refrigerant at the condenser inlet and exit states respectively.

Also, the condenser capacity can be calculated from the cooling medium side, which is usually water or air or both.

Additionally, the condenser capacity is calculated as,

cec WQQ (2)

or

cQ Compressor Capacity x Heat Rejection Factor (HRF) (3)

The HRF depends on the type of the compressor (open or hermetic), condensing and evaporating temperatures. Some compressors manufacturers publish total heat rejection data as part of their compressor ratings as shown in Table (1). These data should be used as a basis for condenser selection.

73

Table (1)

Example 1

Determine the condenser capacity for an open type compressor having a capacity of 250 Kw when operating with evaporating temperature of -20 oC and condensing temperature of 40 oC.

Example 2

Determine the condenser capacity for the conditions given in example (1) if a suction cooled hermetic compressor is used.

Since the heat transfer through the walls of the condenser is by conduction, condenser capacity is a function of the fundamental heat transfer equation:

mc TUAQ (4)

where

A: the heat transfer area of the condenser (m2)

U: overall heat transfer coefficient (Kw/m2 oC)

Tm: the log mean temperature difference between the condensing refrigerant and the condensing medium (water or air) in Celsius, and is expressed as (see the following Figure )

74

off

on

offon

m

T

Tn

TTT

(5)

1-Water-cooled Condenser

The simplest condenser is the water-cooled condenser, of which several types have been used in the past. The first was the atmospheric condenser. A spray system divided the water flow over a number of vertical parallel coils in which the refrigerant was condensed.

The cooling effect relied partly on the water and partly on the evaporation of some of the circulating water – the cooling tower effect which depended on the atmospheric conditions.

The disadvantages of these condensers were that they were expensive and occupied too much space, so they had to be placed outside the building where they suffered from corrosion and the growth of algae.

Another older type. Which is still in use in very small installations, is the double pipe condenser. It is very compact with a high Ii-value. The refrigerant condenses in the space between inner tube and outer tube, while the cooling water flows in the inner tube. The complicated construction makes them too expensive for normal applications.

Nowadays horizontal shell and tube condensers, constructed like liquid chillers, are normally used (see figure). The vertical shell and lube condenser shown in figure has almost disappeared from the market; however, for large capacities they arc not so much more expensive than the horizontal ones. They have the advantage of taking up little ground space, are easy to maintain even during operation and have a high U-value. Partly because of the chimney effect.

Refrigerant gas in

Baffles Water out

onT

offT

Condensing temperature

Condensing medium entering temperature

Condensing medium leveing temperature

75

Horizontal shell and tube water-cooled condenser

Vertical shell and tube water cooled condenser

The horizontal shell and tube condenser is still in use today, although mainly in combination with a cooling tower owing to the shortage of clean water. On board ship or near the coast, water is available in sufficient quantity; in this environment the tubes must be coaled with plastic or be made of special alloys to prevent corrosion from sea-water. In marine plant, galvanic plugs are placed in the water covers to protect against corrosion.

Finned tubes are used to increase the surface, both inside and outside the tubes.

Water in Liquid refrigerant

out

76

Water cooled condenser calculations

Additionally to equation (9.4), the condenser capacity that is related to the cooling water side can be expressed as,

wwpwwiewc TcmhhmQ , (6)

Example 3

In a refrigeration system, in which R134a is the refrigerant, R134a enters a water cooled condenser at 1 Mpa and 70 oC and leaves as a liquid at 0.95 Mpa and 35 oC. The inlet and outer water temperature difference is 7 oC. Determine the rate at which cooling water flows through the condenser.

2-Evaporative Condensers

The evaporative condenser is a combination of condenser and cooling tower. It is an atmospheric condenser with a forced water and air flow as shown in figure . As in the case of an atmospheric condenser, water is sprayed over the rows of coils, which are placed parallel to each other and through which the refrigerant flows. The water is collected in a tank to which the circulation pump is connected, the coils are placed inside a casing, and air is blown or sucked in counterflow with the water stream by means of a fan.

Today, centrifugal fans arc normally used, placed in the bottom of the tower. This is the best way to avoid maintenance and noise-level problems. These condensers are normally placed in the open air and must therefore be protected against freezing. To do tills a heater is placed in the water tank. Sometimes the water is drained to a water receiver placed in a frost-free place. The circulation pump is then placed inside the building.

Such condensers are used when water is difficult to obtain or is of inferior quality. The water is used in a closed circuit and a small part of the water evaporates, cooling down the rest of the water in the same way as is done in a cooling tower. The evaporated water is replaced by adding water using a float-valve built into the water tank. Water consumption is very low: about 5% of a water-cooled condenser. It can be calculated by reading the value of dx from the Mollier graph, dx being the difference in water content of the air between the inlet and the outlet, and multiplying this value by the mass of the air stream. The resulting value is then multiplied by 3 to allow for regular replenishment of water in the storage tank.

Compared with air-cooled condensers they have the advantage of working with lower condensing temperatures. Normally the condensing temperatures lie between II and 13K above the wet-bulb temperature, which means for European conditions +295K + 13K = +308K. In addition to the type with an open water tank, called open circuit, there is a type with a completely sealed circuit, which is used when pollution is expected.

When two or more condensers are connected to one refrigeration circuit, care must be taken to follow exactly the erection instructions of the manufacturers in order to prevent unequal pressure drops which will block up the refrigerant in one of the condensers, which could then lead to a lower condenser efficiency. See figure 9.5 for a typical parallel pipework arrangement for two evaporative condensers and a liquid receiver. The ratio air volume: refrigerant capacity has been determined by experiments at between 100 and 150 m3/h per Kw,

77

Evaporative condenser calculations

The evaporative condenser capacity can be calculated from the energy balance as,

wwiaoaaec hmhhmQ ,, (7)

where

am : the air flow rate

iaoa hh ,, , : the air specific enthalpies at the outlet and inlet of the evaporative condenser.

wh : the water specific enthalpy = (hf) at water temperature

wm : the mass flow rate of water that is carried out by the air, and is expressed as,

ioaw WWmm (8)

Example 9.4

Calculate the evaporative condenser capacity and the make up water if the water bleeding rate is about 30% of the water that is carried out by the air. The volumetric flow rate of air is 6 m3/s, and the entering air wet and dry bulb temperatures are 22 oC and 30 oC respectively. The leaving air conditions are 30 oC wet bulb temperature and 36 oC dry bulb temperature.

.

Figure Schematic diagram of an evaporative condenser

78

Figure An evaporative condenser

Figure parallel evaporative condensers with common liquid heater above receiver level

3- Air Cooled Condensers

These condensers are the most commonly used type today owing to the difficulty in obtaining water of good quality, in sufficient quantity and at reasonable price. Originally, they were only used as natural draught condensers for very small capacities, such as household refrigerators. Some condensers with forced draught air flow using fans were developed for commercial refrigeration up to capacities of 30 kW. A typical model is shown in figure. Today, air-cooled condensers are in use in industrial installations, even for those plants using ammonia.

Their disadvantages can be listed as follows:

High condensing temperatures and high energy consumption of the compressors

79

Power consumption of the condenser fans

High condensing temperatures also mean high end-compression temperatures and high oil temperatures

In some locations, noise can be a problem.

Normally, the condensing temperature is between 10 and 15K above the Maximum ambient temperature.

Compared with water-cooled or evaporative condensers, the U-value is low because the heat-exchange surface has gas on both sides for much of its effective area. The surface is increased by adding fins on the tubes. Since the fin spacing is small, 1-2 mm, and becomes quickly choked, the result is a higher energy consumption and the need for frequent cleaning of the condenser fins.

The condenser must always be located outside, or alternatively, be in contact with the outside air by means of air ducts. Small condenser types can only be positioned inside in a place where there is a large free volume of air available. Tables are available to guide such a choice. Summer and winter working conditions are very different because operation is directly related to the climatic conditions, in particular the dry-bulb temperature of the air.

Air cooled condenser calculations

Additionally to equation (9.4), the condenser capacity that is related to the cooling air side can be expressed as,

aapaaieac TcmhhmQ , (9)

Example 5

An ambient air at the rate of 15 kg/s and 35 oC is pushed through a 60 Kw condenser. The overall heat transfer coefficient is 30 W/m2 K and the refrigerant condensing temperature is 48 oC. Calculate the surface area of the condenser at the air side.

Figure Horizontal air-cooled condenser

Air Flow

Air Flow Fans

80

81

III- Evaporators

Introduction

Three major vapor .compression system components namely compressor, condenser and expansion device have already been discussed. The fourth component to complete the cycle is the evaporator. The evaporator in the vapor compression cycle, is a heat exchanger which absorbs heat from the substance to be cooled and transfers it to a boiling refrigerant.

Types of Evaporators

The evaporators may be classified as Forced convection type or Free convection type depending upon whether the substance to be cooled is forced by pump or fan through the heat transfer surfaces of the evaporator, or it flows naturally by density difference of warmer and colder fluid. Some evaporators have refrigerant in the tubes and substance to be cooled surrounding the tubes, but, in other cases, the refrigerant is in the shell with substance to fee cooled passing through the tubes.

Evaporators are also classified as flooded type and Dry type depending upon whether liquid refrigerant covers all heat transfer surfaces or some portion is having gas being superheated. The ‗evaporators with thermostatic expansion valve will have some portion of heat transfer surface where superheating is taking place and can be designated as dry evaporator; whereas evaporators with float valve will be flooded type.

Flooded Evaporator

A typical flooded evaporator with float control is shown in Figure. The liquid on its flow passage upwards through the tubes, boils due to absorption of heat from the warmer substance which is to be cooled. The vapor so formed on boiling bubbles up in flash chamber. The flash chamber separates vapor from liquid, which flows back to the .evaporator whereas vapors are sucked by the compressor. The flash chamber collects the flash or vapor obtained in the expansion device, plus the vapor formed by refrigerant liquid boiling in the evaporator. In a flooded type evaporator refrigerant liquid level is maintained. Float valve is used as throttling device. The heat-transfer efficiency increases because the entire surface is in contact with the liquid refrigerant. But the refrigerant charge is relatively large as compared to dry expansion type. To prevent liquid carry over to compressor, accumulator or flash chamber is used. The evaporator coil is connected to accumulator and the liquid flow from the accumulator to the evaporator coil is generally by gravity. The vapor formed by the vaporizing of the liquid in the coil being lighter rises up and passes on to the top of the accumulator from where it enters the suction line. In some cases liquid eliminators are provided in the accumulator top to prevent the possible carryover of liquid to suction line. Further a liquid suction heat exchanger is used on the suction line to superheat the suction vapor.

Liquid chiller

Refer to the following Figures for two types of liquid chillers. The former has refrigerant in the shell and liquid to be chilled in the tubes whereas the latter has refrigerant in the tubes and liquid to be chilled in the shell. When the refrigerant is in the shell, the refrigerant liquid level is so kept that there is enough space on the top portion of the shell for the liquid and vapor to separate Vapors are drawn from the top portion by the compressor. Liquid level must be- maintained constant as the chilled tubes are also, immersed in the refrigerant liquid. Thus float control is preferred. But when the liquid to be chilled is in the shell and the refrigerant is in the tubes as shown in Figure, thermostatic expansion

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valve is preferred. The refrigerant gets superheated in the last ‗portion of the set of tubes and is collected in. the end chamber from where it is sucked by the compressor. In order to facilitate proper contact of water with the refrigerant tubes, baffles are provided to ensure larger circuit up and down for the water, resulting in increased turbulence and hence better over heat transfer coefficient. Thus evaporator shown in figure can be termed as flooded whereas evaporator shown in Figure is dry.

Figure Flooded evaporator

Figure Dry evaporator

Direct Expansion Coil Evaporator

In the liquid chiller, the chilled liquid is fed to the coils which are used for cooling air. But, if the coils of the evaporator with refrigerant passing through them are used directly to cool air by natural or forced convection, the evaporator is called direct expansion evaporator. (Figure). The refrigerant feed comes through the thermostatic expansion valve more often located at the top particularly for Freon-12 and Freon-22 to improve the lubricating oil return to the compressor. Air is blown over the –outside of the finned tubes. For air conditioning purposes, the direct expansion coil is preferred where the evaporator is very near to the compressors. It is direct method of cooling the substance and, therefore, quite efficient.

But when the coil has to be located very far away from the compressor, it is preferred to chill the water and pump it to the air cooling coil. For long distances, there is possibility of refrigerant leakage and the cost of the refrigerant would be also high. Besides, the pressure drop in the line would impair

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evaporator efficiency and coefficient of performance. The pressure effect becomes very significant if the direct expansion coil is located at a great height from the compressor and condenser unit.

The expansion valve controls the rate of flow of the refrigerant to evaporator in such a way that all the liquid is vaporized and the vapor is also superheated to a limited extent. The inside of evaporator is far from dry but wetted with liquid. All the same this type is called dry expansion to distinguish it from flooded system and also probably because by the time the refrigerant reaches the evaporator outlet it no more wet but dry superheated vapor

Figure Direct Expansion Coil Evaporator

The air passes through the evaporator coil either naturally or by force. When the moist air is cooled to a temperature below its dew point temperature some of the water vapor will condensate and leave the stream.

The thermal evaporator load can be expressed as,

vwoiaoipaaoiae hWWmTTcmhhmQ ,)(

Where

hw,v = is the specific enthalpy of water vapor in the air = hg+cpv xTdb

hg: =2501.3 kJ/kg

Tdb is the air dry bulb temperature, oC

cpv specific heat of water vapor =1.8723 kJ/kg K

Example

An evaporator is used to cool down air from Tdbi = 32 oC and Wi=0.015 kg/kg da to Tdbo = 14 oC and Wo=0.009 kg/kg da. Calculate the evaporator thermal load if the air volumetric flow rate is 2 m3/s.

Solution:

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IV Expansion devices

Introduction

The expansion device also forms a very important component of the vapor compression refrigerant system. Compressors and condensers have already been discussed. After the condenser comes the expansion device. Common expansion devices are :

(1) Capillary tube

(2) Thermostatic expansion valves

(3) Low and high side float valves

(4) Constant pressure expansion valve.

Capillary Tube

Almost all fractional horsepower, vapor compression refrigeration units employ capillary tube. The use is being extended presently to larger units up to about 5 hp. The capillary tube when used as a liquid refrigerant expanding device usually consists of an extremely small bore tube from 0.5 mm to 2.5 mm of about 0.5 m to 5 m long as shown in Figure. Numerous combinations of bore and length arc available to get the desired restriction. Its extreme simplicity and very low cost make it very popular. In its operation, liquid refrigerant enters the capillary tube and due to flow, there is pressure drop due to friction. Some of the liquid flashes into vapor as the refrigerant flows through the tube. Once the sizing and length of tube is selected, no modifications are possible to adjust itself to variation in discharge pressure, suction pressure and load. Care must be taken to prevent plugging of the tube by any dirt. Ice or any other decomposed material. The capillary tube is substituted for the convectional liquid line from the condenser and soldered to a length of the suction line form a simple heat exchanger.

Thermostatic Expansion Valve

This is the most popular and very efficient type of expansion device in use at present. The operation of thermostatic expansion valve is based on the principle of constant degree of superheat for the evaporator exit. This ensures the evaporator completely filled with refrigerant irrespective of the load and also no liquid can spill over to the suction line to the compressor. Because of its adaptability to load changes, it is specially suitable for variable load systems. Figure gives a schematic diagram of a thermostatic expansion valve to explain its working.

Capilart

y Tube

Strainer

85

Operation

The remote bulb charged with fluid which is open on one side of the diaphragm through a capillary tube is clamped firmly to the evaporator outlet. The temperature of the saturated liquid vapor mixture is the same as the temperature of the superheated gas leaving the evaporator at this location. The pressure of the fluid in the bulb tends to open the valve. This pressure is balanced by pressure due to spring plus pressure in the evaporator. There is thus, interaction of three independent forces namely force due to evaporator pressure, force due to spring compression and the force due to saturated liquid-vapor in the bulb.

Assuming that evaporator contains Feron-12 and that pressure in the evaporator corresponding to - 4 oC saturation temperature, i.e. P1= 250 kPa. Let us also assume that adjusting screw is set to exert a pressure of P2=50 kPa. Therefore, the total pressure tending to close the valve is P1+P2 =300 kPa (see Figure 11.2 a). The pressure due to friction in the evaporator may be neglected. Therefore, the saturation temperature and corresponding pressure has been shown as same the evaporator except for a portion where superheating is done. Superheat of 7°C has been shown in Figure1.1 and P3= P1+P2 =300 kPa. This corresponds to the saturation temperature of 0°C, i.e. equal to the temperature of superheated evaporator gas. Thus the diaphragm is on balance point and in equilibrium condition. Any change in degree of superheat of the suction gas will alter the valve position. If degree of superheat becomes less than 7°C, the sum of evaporator and spring pressure will exceed the pressure exerted by the bulb. This will tend to close the valve and throttle the flow to the evaporator till again the same degree superheat is obtained. Reverse action will take place for increase of the superheat. It may however be noted that the equilibrium condition degree of superheat can be changed by the adjusting screw. Thus spring adjustment is also called superheat adjustment. A typical internally equalized thermostatic expansion valve is shown in Figure.

Automatic Expansion Valve

An explanatory sketch of an automatic expansion valve is shown in Figure 11.3 This valve simply works on the principle of maintaining a set pressure in the evaporator by adjusting the spring. The constant pressure in the evaporator is maintained by two opposing forces namely evaporator pressure and the spring pressure. The evaporator pressure tries to close the valve whereas the spring pressure tries to open the valve.

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Suppose the evaporator pressure falls below the set value for which spring is set. The valve will open more and increase the liquid flow to the evaporator resulting in flooding more evaporator surface. When more evaporator surface becomes effective, the vaporization rate increase resulting in rise of evaporator pressure till again it balances the spring pressure. If the evaporator pressure rises above the set valve, the valve tends to close and therefore less refrigerant liquid flows to the evaporator resulting in reducing the effective surface area of the evaporator. This leads to reduction of pressure in the evaporator till again it is equal to the spring pressure.

When the plant is shut off, the evaporator pressure will build and keep the valve also firmly shut. On starting the plant, the suction from the evaporator starts resulting in the fall in pressure in the evaporator below the spring pressure to resume the flow of refrigerant liquid through the valve. The main disadvantage of automatic expansion valve is relatively low efficiency as compared to thermostatic expansion valve. In view of the matching of evaporator and compressor, it is imperative that rate of vaporization should be kept constant to maintain the evaporator pressure unaltered. Thus varying loads cannot be adequately taken care of without sacrificing efficiency. When the load on the evaporator is very heavy and heat transfer capacity per unit surface area of the evaporator is high, there will be severe throttling to limit the effective surface area as shown in Figure 11 4. The decrease of load and consequently less heat transfer capacity per unit area is achieved by flooding more surface of the evaporator as shown in Figure.

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Due to low efficiency at heavy load, the automatic expansion valve is used for small equipment having relatively constant load.

For varying load operation thermostatic expansion valve has completely replaced constant pressure valve.

Thermal-electric expansion valve

The thermoelectric expansion valve depends upon the use of thermistors directly exposed to the refrigerant in the suction line, to control the expansion valve needle opening. It does not use a pressure element as in the thermostatic expansion valve.

The resistance to electrical flow in the thermistor changes with its temperature. Increasing temperature reduces resistance. Therefore, with a given voltage, it increases the current rate of flow. This increased current flow heats the bimetal in the valve body and makes the bimetal bend, opening the valve.

Fig. illustrates a typical thermal-electric expansion valve installation. The thermistor, C, is placed in immediate contact with the refrigerant vapor inside the suction line from the evaporator.

A low voltage transformer is the power source. This connected to the expansion valve control mechanism at B. The transformer is in series with the thermistor and electric device at B so that increasing current flow through the thermistor increases the opening of the expansion valve and, therefore, increases the rate of flow of the refrigerant into I the evaporator.

Increasing the current causes the valve needle to open, while a decrease closes the valve. Thereby the refrigerant flow is controlled. The thermal-electric expansion valve is not dependent on the pressure in the evaporator. It restricts the flow of refrigerant and controls the suction line superheat in order to prevent flooding of the compressor.

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Electronic expansion valve

As shown in Figure it consists of stepper motor, lead screw, piston, orifice assembly. The stepper motor moves by an electronic signal from electronic device in steps to left the piston up and down to open or close the orifice slots to control the refrigerant flow rate. The electronic device receives many signals from the most refrigeration elements to adjust the electronic valve. Some of these signals are from: overload, high pressure and temperature, low pressure and temperature, oil pressure, . . . . . etc.

Float Valves

Low pressure float valve. The low pressure float control maintains the liquid at constant level in the evaporator by regulating the flow into the evaporator in accordance with the supply from the evaporator to the compressor or the rate of vaporisation in the evaporator. If the refrigeration load increases, the evaporator temperature and pressure rises, which temporarily allows the compressor to pump a greater mass flow rate than the valve is feeding. The valve reacts to keep the level constant by opening more. If the refrigeration load decreases, the evaporator pressure falls and the

89

compressor now pumps less mass flow rate temporarily and the level in float chamber rises resulting in the tendency to close the valve.

It may be noted that this device is responding to the level of liquid in the evaporator and keeps the evaporator always filled to the desired level without regard for evaporator temperature or pressure. This device may be incorporated directly in the evaporator or accumulator in which it has to control the level as shown in Figure 11.8. It can also be installed external to the unit in separate float chamber.

High Pressure Valve

The high pressure float valve also maintains the flow to the evaporator by actuating the level in the float chamber in the same manner as the low pressure float valve except that the high pressure float valve is located on the high pressure side and controls the amount of liquid by maintaining level in the float chamber. The condensation rate and the evaporation rate are matched by the high pressure float valve by actuating the level and thus altering the opening and closing of the needle valve. To have further control on the expansion of liquid refrigerant, a pressure reducing valve is also used in the circuit as shown in Figure.

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Defrosting Systems Introduction Defrosting is a process of removing frost accumulation from the evaporator which cools moist air in frozen applications. When the moist air passes through an evaporator at temperature lower than zero, the moisture in the air condenses and freezes on the evaporator surface. This frost clogs the air passes through the evaporator and decreases the heat transfer between the refrigerant and the air. For these reasons, the frost has to be removed periodically. There are some methods to defrost the evaporator such as:

Electric heater Hot gas Water Warm air

Electric Heater Defrosting This method is used in a small capacity unit, it uses resistance wire heating elements mounted inside the evaporator, in the drain pan and along the drain pipe (see Figure ) A timer stops the refrigerating unit and operates the heaters. If the evaporator is warm enough to insure that all frost is gone, a thermostat sensor on the evaporator surface stops the heaters. When the defrosting time finished, the timer returns back the system to normal cooling operation. Figure shows the control diagram of electric heater defrosting.

Figure Control diagram of electric heater defrosting

T OL

Def. H Def. Th.

EF

C

Def. T

N

Electric heaters

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Figure Control diagram of electric heater defrosting For relatively high capacity units, the timer closes the liquid line solenoid valve (LLSV), while the compressor still in operation till the evaporator becomes nearly empty of refrigerant and the compressor stops and the heater on. This is to eliminate the pressure in the evaporator during the defrosting. This is denoted as pump down control system. Hot gas defrosting In this type, hot refrigerant (gas) from discharge line is pumped directly into the evaporator through a special line and timed hot gas solenoid valve (HGSV). The hot gas rushes through the evaporator, warms it and then returns back to the compressor along the suction line. Because some of the hot gas may condensate into the evaporator, an accumulator is required and additional source of heat is needed to re-evaporate the condensation. Some methods are used such as:

Figure

Using of the cooling evaporator heat in multi evaporators at different temperatures system as shown in Figure. In cooling and freezing processes, the refrigerant goes through the two evaporators as

Legends

T: Thermostat

OL: Over load

C: Comp. Motor

Def. H.: Defrost heater

ef.Th: Def. Thermostat.

Def. T: Defrosting Timer

CF: Cond. Fan

EF: Evapor. Fan

DH: Door Heater

DS: Door Switch

L: Lamp

T

C

OL

110-1-60

N

L1

EF

CF

L

DS

1 3 2 4 6

2 1

3

4

Def.T

DH

5 7

Def

.H

Def.Th

Cond

Evap. # 2

Evap. # 1 s

s

TX

V

#

1

TX V # 2

TX V # 1

LLSV 1

LLSV 2

Closed

Open

s

s

Open

Closed

HGSV 1

HGSV 2

Suction

Acc.

Comp

EPR

NR

BPR

92

usual. Hot gas solenoid valve HGSV is closed, evaporator regulator valve keeps nonfrost evaporator at higher pressure and temperatures than frost one. Where as in defrosting processes for the frost evaporator, liquid line solenoid valve LLSV closes and hot gas solenoid opens. So, the hot gas enters the frost evaporator to defrost it. The by pass regulator valve BPRV injects some gas at the evaporator outlet to evaporate any gas condensation. The compressor still running to absorb heat load from the non frost evaporator and provide the required heat to the frost evaporator. In large refrigeration systems where the defrosting is very important, the gas condensation is sent back with the suction line to the Low Pressure vessel (LP) or the intermediate pressure IP as shown in Figure . As can be seen from Figure there are two ways of feeding the hot gas to the evaporator, either from the liquid line or from the suction line.

Figure

Figure Also, hot gas can be accomplished by reversing the flow of the refrigerant in the refrigeration unit itself as shown in Figure. During the cooling process, the discharge gas goes through the reversed valve RV to the condenser, Non-Return valve NRV, TEX, evaporator, reversed valve RV and then back to the compressor. Whereas, during the defrosting process, the hot gas goes through RV, evaporator (Condenser), NRV,TEX , condenser (evaporator) ,reversed valve RV, and then back to the compressor 7.4- Water defrosting Water defrosting may be manually or automatically runs top water over the evaporator while system is not in operation. Water is sprayed over the evaporator and the lovers are closed as in Figure.

Evap

NR

PRV

Suction line

Liquid line

Suction Stop Valve

LLSV

Closed

Open

Hot gas line HGSV

Evap.

NR

PRV

Suction line

Hot gas line

Liquid line

Suction Stop Valve

LLSV

HGSV

93

The timer stops the compressor and the evaporator fans and opens the solenoids valve S1 to spray the water on the frost evaporator. Water falls down to the drain. Safety float controls level to prevent flooding .

7.5- Warm air defrosting

Cabinet air at right temperature can be used to defrost low temperature evaporator. Some designs use outside air for defrosting using controlled duct and fan.

Evap. Fans

S2

S1

Timer Comp

Liqid Line

Suction line

drain

Water line

Water Spray

Comp

Cond

evap

TX V

Sucti

on

Accu

mulat

or

RV

Cooling

Defrosting

NR

(evap)

(cond)

(4 way

Valve)

TX V

NR

RV

(B)

RV

( C )

(A)