Emerson Refrigeration Manual

Part 1 - Fundamentals of Refrigeration

Refrigeration Manual

© 1968 Emerson Climate Technologies, Inc.All rights reserved.

FOREWORD

The practice of refrigeration undoubtedly goes back as far as the history of mankind, but for thousands of years the only cooling mediums were water and ice. Today refrigeration in the home, in the supermarket, and in commercial and industrial usage is so closely woven into our everyday existence it is difficult to imagine life without it. But because of this rapid growth, countless people who must use and work with refrigeration equipment do not fully understand the basic fundamentals of refrigeration system operation.

This manual is designed to fill a need which exists for a concise, elementary text to aid servicemen, salesman, students, and others interested in refrigeration. It is intended to cover only the fundamentals of refrigeration theory and practice. Detailed information as to specific products is available from manufacturers of complete units and accessories. Used to supplement such literature—and to improve general knowledge of refrigeration—this manual should prove to be very helpful.


Section 1 Basic Refrigeration Principles

Thermodynamics 1-1Heat 1-1Temperature 1-1Heat Measurement 1-2Heat Transfer 1-2Change of State 1-3Sensible Heat 1-3Latent Heat of Fusion 1-3Latent Heat of Evaporation 1-4Latent Heat of Sublimation 1-4Saturation Temperature 1-4Superheated Vapor 1-4Subcooled Liquid 1-4Atmospheric Pressure 1-4Absolute Pressure 1-5Gauge Pressure 1-5Pressure-Temperature Relationships, Liquids 1-5Pressure-Temperature Relationships, Gases 1-5Specific Volume 1-6Density 1-6Pressure and Fluid Head 1-6Fluid Flow 1-7Effect of Fluid Flow on Heat Transfer 1-7

Section 2 Refrigerants

Terminology and Examples 2-1Pure Fluid 2-1Mixture and Blend 2-1Azeotropic Refrigerant Mixture 2-1Zeotropic Mixture 2-2Near-Azeotropic Refrigerant Mixture 2-2How are Components Chosen 2-2Mixture Behavior 2-3Azeotrope 2-3Zeotrope 2-3Near-Azeotropic Refrigerant Mixtures 2-3What Happens to Mixture Composition During

System Charging? 2-3Temperature Glide 2-4What Happens to Refrigerant Mixture

Composition During a Leak? 2-5

Types of Refrigerant 2-5Refrigerants 2-8Refrigerant 12 2-8Refrigerant R-401A/B 2-8Refrigerant R-409A 2-8Refrigerant 134a 2-8Refrigerant 22 2-9Refrigerant R-502 2-9Refrigerant R-402A 2-9Refrigerant R-408A 2-9Refrigerant R-404A 2-9Refrigerant R-507 2-10Refrigerant Saturation Temperature 2-10Refrigerant Evaporation 2-10Refrigerant Condensation 2-10Refrigerant-Oil Relationships 2-10Refrigerant Tables 2-11Saturation Properties 2-12Pocket Temperature-Pressure Charts 2-12

Section 3 The Refrigeration Cycle

Simple Compression Refrigeration Cycle 3-1Heat of Compression 3-2Volumetric Efficiency of the Reciprocating

Compressor 3-2Volumetric Efficiency of the Scroll Compressors 3-4Effect of Change in Suction Pressure 3-4Effect of Change in Discharge Pressure 3-4Effect of Subcooling Liquid Refrigerant with

Water or Air 3-4Effect of Subcooling Liquid Refrigerant by

Superheating the Vapor 3-4Effect of Superheating the Vapor Leaving

the Evaporator 3-5Effect of Pressure Drop in the Discharge Line

and Condenser 3-5Effect of Pressure Drop in Liquid Line 3-5Effect of Pressure Drop in the Evaporator 3-5Effect of Pressure Drop in Suction Line 3-6Internally Compound Two-Stage Systems 3-6Externally Compound Systems 3-6Cascade Systems 3-10

Table of Contents


Most users of refrigeration products normally associate refrigeration or air conditioning with cold and cooling, yet the practice of refrigeration engineering deals almost entirely with the transfer of heat. This seeming contradic-tion is one of the most fundamental concepts that must be grasped to understand the workings of a refrigeration or air conditioning system. Cold is really only the ab-sence of heat, just as darkness is the absence of light, and dryness is the absence of moisture.

THERMODYNAMICS

Thermodynamics is that branch of science dealing with the mechanical action of heat. There are certain fundamental principles of nature, often called laws of thermodynamics, which govern our existence here on Earth. Several of these laws are basic to the study of refrigeration.

The first and most important of these laws is the fact that energy can neither be created or destroyed. It can only be converted from one type to another. A study of thermodynamic theory is beyond the scope of this manual, but the examples that follow will illustrate the practical application of the energy law.

HEAT

Heat is a form of energy, primarily created by the trans-formation of other types of energy into heat energy. For example, mechanical energy turning a wheel causes friction and is transformed into heat energy. When a vapor such as air or refrigerant is compressed, the compression process is transformed into heat energy and heat is added to the air or refrigerant.

Heat is often defined as energy in motion, for it is never content to stand still. It is always moving from a warm body to a colder body. Much of the heat on the Earth is derived from radiation from the sun. The heat is being transferred from the hot sun to the colder earth. A spoon in ice water loses its heat to the water and becomes cold. Heat is transferred from the hot spoon to the colder ice water. A spoon in hot coffee absorbs heat from the coffee and becomes warm. The hot coffee transfers heat to the colder spoon. The terms warmer and colder are only comparative. Heat exists at any temperature above absolute zero even though it may be in extremely small quantities.

Absolute zero is the term used by scientists to de-scribe the lowest theoretical temperature possible, the temperature at which no heat exists. This occurs at approximately 460° below zero Fahrenheit, 273° below

Section 1BASIC REFRIGERATION PRINCIPLES

zero Celsius. By comparison with this standard, the coldest weather we might ever experience on Earth is much warmer.

TEMPERATURE

Temperature is the scale used to measure the intensity of heat, the indicator that determines which way the heat energy will move. In the United States, tempera-ture is normally measured in degrees Fahrenheit. The Celsius scale (previously termed Centigrade) is widely used in most other parts of the world. Both scales have several basic points in common, (See Figure 1-1) the freezing point of water, and the boiling point of water at sea level. At sea level, water freezes at 32°F (0°C) and water boils at 212°F (100°C). On the Fahrenheit scale, the temperature difference between these two points is divided into 180 equal increments or degrees F, while on the Celsius scale the temperature difference is divided into 100 equal increments or degrees C. The relation between Fahrenheit and Celsius scales can always be established by the following formulas:

Fahrenheit = 9/5 Celsius + 32° Celsius = 5/9 (Fahrenheit -32°)

COMPARISON OF TEMPERATURE SCALES Figure 1-1

Further observing the two scales, note that at -40°, both the Fahrenheit and Celsius thermometers are at the same point. This is the only point where the two scales are identical. Using this information, the follow-ing formulas can be used to determine the equivalent Fahrenheit or Celsius values.

Fahrenheit = ((Celsius + 40) x 9/5) - 40Celsius = ((Fahrenheit + 40) x 5/9) - 40

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HEAT MEASUREMENT

The measurement of temperature has no relation to the quantity of heat. A match flame may have the same temperature as a bonfire, but obviously the quantity of heat given off is vastly different.

The basic unit of heat measurement used today in the United States is the British Thermal Unit, commonly expressed as a BTU. A BTU is defined as the amount of heat added or removed to change one pound of wa-ter one degree Fahrenheit. For example, to raise the temperature of one gallon of water (approximately 8.3 pounds) from 70°F to 80°F will require 83 BTUs.

1 gallon (8.3 pounds) x (80°F - 70°F)∆T = 83 BTUs heat added

8.3 pounds x 10°∆T = 83 BTUs

In the metric system, the basic unit of heat measure-ment is the Calorie. A Calorie is defined as the amount of heat added or removed to change one gram of water one degree Celsius. For example, to lower one liter of water (1000 grams) from 30°C to 20°C will require 10,000 Calories of heat to be removed.

1000 grams X (30°C - 20°C)∆T = 10,000 Calories of heat removed.

HEAT TRANSFER

The second important law of thermodynamics is that heat always travels from a warm object to a colder one. The rate of heat travel is in direct proportion to the temperature difference between the two bodies.

390°F 400°F

HeatFlow

Figure 1-2

Assume that two steel balls are side by side in a perfectly insulated box. One ball weighs one pound and has a temperature of 400°F, while the second ball weighs 1,000 pounds and has a temperature of 390°F. The heat content of the larger ball is much greater than the small one, but because of the temperature difference, heat will travel from the small ball to the large one (See Figure 1-2) until the temperatures equalize. Heat can

travel in any of three ways; radiation, conduction, or convection.

Radiation is the transfer of heat by waves similar to light waves or radio waves. For example, the sun's energy is transferred to the Earth by radiation.

Figure 1-3

One need only step from the shade into direct sunlight to feel the impact of the heat waves even though the temperature of the surrounding air is identical in both places. Another example of radiation is standing in front of a bonfire. The side of you facing the bon fire is receiving radiant heat and that side is hot. The side away from the fire may feel cool. There is little radiation at low temperatures and at small temperature differ-ences. As a result, radiation is of little importance in the actual refrigeration process. However, radiation to the refrigerated space or product from the outside environ-ment, particularly the sun, may be a major factor in the refrigeration load.

Conduction is the flow of heat through a substance. Actual physical contact is required for heat transfer to take place between two bodies by this means. Conduc-tion is a highly efficient means of heat transfer as any serviceman who has touched a piece of hot metal can testify.

HOT WARM COOL

Figure 1-4

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Figure 1-4 shows a flame heating one end of a metal rod. Heat is conducted to the other end by the process of conduction.

Convection is the flow of heat by means of a fluid me-dium, either vapor or liquid, normally air or water. Air may be heated by a furnace, and then discharged into a room to heat objects in the room by convection.

Figure 1-5

In a typical air conditioning/refrigeration application, heat normally will travel by a combination of processes. The ability of a piece of equipment to transfer heat is referred to as the overall rate of heat transfer. While heat transfer cannot take place without a temperature difference, different materials vary in their ability to conduct heat. Metal is a very good heat conductor. Fiberglass has a lot of resistance to heat flow and is used as insulation.

CHANGE OF STATE

Most common substances can exist as a solid, a liquid, or a vapor, depending on their temperature and the pressure to which they are exposed. Heat can change their temperature, and can also change their state. Heat is absorbed even though no temperature change takes place when a solid changes to a liquid, or when a liquid changes to a vapor. The same amount of heat is given off, rejected, even though there is no temperature change when the vapor changes back to a liquid, and when the liquid is changed back to a solid.

The most common example of this process is water. It generally exists as a liquid, but can exist in solid form as ice, and as a vapor when it becomes steam. As ice it is a usable form for refrigeration, absorbing heat as it melts at a constant temperature of 32°F (0°C). As water, when placed on a hot stove in an open pan, its temperature will rise to the boiling point, 212°F (100°C) at sea level. Regardless of the amount of heat applied, the waters temperature cannot be raised above 212°F (100°C) because the water will vaporize into steam. If

this steam could be enclosed in a container and more heat applied, then the water vapor, steam, temperature could again be raised. Obviously the fluid during the boiling or evaporating process was absorbing heat.

When steam condenses back into water it gives off ex-actly the same amount of heat that it absorbed during evaporation. (The steam radiator is a common usage of this source of heat.) If the water is to be frozen into ice, the same amount of heat that was absorbed in melting must be extracted by some refrigeration process to cause the freezing action.

The question arises, just where did those heat units go? Scientists have found that all matter is made up of mol-ecules, infinitesimally small building blocks which are ar-ranged in certain patterns to form different substances. In a solid or liquid, the molecules are very close together. In a vapor the molecules are much farther apart and move about much more freely. The heat energy that was absorbed by the water became molecular energy, and as a result the molecules rearranged themselves, changing the ice into water, and the water into steam. When the steam condenses back into water, that same molecular energy is again converted into heat energy.

SENSIBLE HEAT

Sensible heat is defined as the heat involved in a change of temperature of a substance. When the temperature of water is raised from 32°F to 212°F, an increase in sensible heat content is taking place. The BTUs required to raise the temperature of one pound of a substance 1°F is termed its specific heat. By definition, the specific heat of water is 1.0 BTU/lb. The amount of heat required to raise the temperature of different substances through a given temperature range will vary. It requires only .64 BTU to raise the temperature of one pound of butter 1°F, and only .22 BTU is required to raise the temperature of one pound of aluminum 1°F. Therefore the specific heats of these two substances are .64 BTU/lb. and .22 BTU/lb. respectively. To raise the temperature of one pound of liquid refrigerant R-22, 1°F from 45° to 46°, requires .29 BTUs, therefore its specific heat is .29 BTU/lb.

LATENT HEAT OF FUSION

A change of state for a substance from a solid to a liquid, or from a liquid to a solid involves the latent heat of fu-sion. It might also be termed the latent heat of melting, or the latent heat of freezing.

When one pound of ice melts, it absorbs 144 BTUs at a constant temperature of 32°F. If one pound of water is to be frozen into ice, 144 BTUs must be removed from the water at a constant temperature of 32°F. In the freezing of food products, it is only the water content

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for which the latent heat of freezing must be taken into account. Normally this is calculated by determining the percentage of water content in a given product.

LATENT HEAT OF EVAPORATION

A change of a substance from a liquid to a vapor, or from a vapor back to a liquid involves the latent heat of evaporation. Since boiling is only a rapid evaporating process, it might also be called the latent heat of boil-ing, the latent heat of vaporization, or for the reverse process, the latent heat of condensation.

When one pound of water boils or evaporates, it absorbs 970 BTUs at a constant temperature of 212°F (at sea level). To condense one pound of steam to water, 970 BTUs must be extracted from the steam.

Because of the large amount of latent heat involved in evaporation and condensation, heat transfer can be very efficient during the process. The same changes of state affecting water applies to any liquid, although at different temperatures and pressures.

The absorption of heat by changing a liquid to a vapor, and the discharge of that heat by condensing the vapor is the keystone to the whole mechanical refrigeration process. The movement of the latent heat involved is the basic means of refrigeration.

When one pound of refrigerant R-22 boils, evaporates, it absorbs 85.9 BTUs at 76 psig. To condense one pound of R-22, 85.9 BTUs must be extracted from the refrigerant vapor.

LATENT HEAT OF SUBLIMATION

A change in state directly from a solid to a vapor without going through the liquid phase can occur with some substances. The most common example is the use of "dry ice" or solid carbon dioxide when used for cooling. The same process can occur with ice below the freez-ing point. This process is utilized in some freeze-drying processes at extremely low temperatures and deep vacuums. The latent heat of sublimation is equal to the sum of the latent heat of fusion and the latent heat of evaporation.

SATURATION TEMPERATURE

The condition of temperature and pressure at which both liquid and vapor can exist simultaneously is termed saturation. A saturated liquid or vapor is one at its boiling point. For water at sea level, the saturation temperature is 212°F. At higher pressures, the saturation temperature increases. With a decrease in pressure, the saturation temperature decreases.

The same condition exists for refrigerants. At the refrig-erants boiling point, both liquid and vapor exist simul-taneously. For example, refrigerant R-22 has a boiling point of 45°F at a pressure of 76 psig. It's boiling point changes only as it pressure changes.

SUPERHEATED VAPOR

After a liquid has changed to a vapor, any further heat added to the vapor raises its temperature. As long as the pressure to which it is exposed remains constant, the resulting vapor is said to be superheated. Since a temperature rise results, sensible heat has been added to the vapor. The term superheated vapor is used to describe a vapor whose temperature is above it's boil-ing or saturation point. The air around us is composed of superheated vapor.

Refrigerant 22 at 76 psig has a boiling point of 45°F. At 76 psig, if the refrigerants temperature is above 45°F, it is said to be superheated.

SUBCOOLED LIQUID

Any liquid that has a temperature lower than the satura-tion temperature corresponding to its saturation pres-sure is said to be subcooled. Water at any temperature less than its boiling temperature (212°F at sea level) is subcooled.

The boiling point of Refrigerant 22 is 45°F at 76 psig. If the actual temperature of the refrigerant is below 45°F at 76 psig, it is said to be subcooled.

ATMOSPHERIC PRESSURE

The atmosphere surrounding the Earth is composed of gases, primarily oxygen and nitrogen, extending many miles above the surface of the Earth. The weight of that atmosphere pressing down on the Earth creates the atmospheric pressure in which we live. At a given point, the atmospheric pressure is relatively constant except for minor changes due to changing weather conditions. For purposes of standardization and as a basic reference for comparison, the atmospheric pres-sure at sea level has been universally accepted. It has been established at 14.7 pounds per square inch, (psi). This is equivalent to the pressure exerted by a column of mercury 29.92 inches high.

At altitudes above sea level, the depth of the atmo-spheric blanket surrounding the Earth is less, therefore the atmospheric pressure is less. At 5,000 feet eleva-tion, the atmospheric pressure is only 12.2 psi., 28.84 inches of mercury.

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unit of measurement since even inches of mercury is too large for accurate reading. The micron, a metric unit of length, is used for this purpose. When we speak of microns in evacuation, we are referring to absolute pressure in units of microns of mercury.

A micron is equal to 1/1000 of a millimeter and there are 25.4 millimeters per inch. One micron, therefore, equals 1/25,400 inch. Evacuation to 500 microns would be evacuating to an absolute pressure of approximately .02 inch of mercury. At standard conditions this is the equivalent of a vacuum reading of 29.90 inches mer-cury.

PRESSURE-TEMPERATURE RELATIONSHIPS, LIQUIDS

The temperature at which a liquid boils is dependent on the pressure being exerted on it. The vapor pressure of the liquid is the pressure being exerted by the tiny molecules seeking to escape the liquid and become vapor. Vapor pressure increases with an increase in temperature until at the point when the vapor pressure equals the external pressure, boiling occurs.

Water at sea level boils at 212°F, but at 5,000 feet eleva-tion it boils at 203°F due to the decreased atmospheric pressure. (See Table 1-1) If some means, a compressor for example, is used to vary the pressure on the sur-face of the water in a closed container, the boiling point can be changed at will. At 100 psig, the boiling point is 337.9°F, and at 1 psig, the boiling point is 215.3°F.

Since all liquids react in the same fashion, although at different temperatures and pressure, pressure provides a means of regulating a refrigerant's temperature. The evaporator is a part of a closed system. A pressure can be maintained in the coil equivalent to the saturation temperature (boiling point) of the liquid at the cooling temperature desired. The liquid will boil at that tempera-ture as long as it is absorbing heat and the pressure does not change.

In a system using refrigerant R-22, if the pressure within the evaporator coil is maintained at 76 psig, the refrigerants boiling point will be 45°F (7.2°C). As long as the temperature surrounding the coil is higher than 45°F (7.2°C), the refrigerant will continue to boil absorbing heat.

PRESSURE-TEMPERATURE RELATIONSHIPS, GASES

One of the basic fundamentals of thermodynamics is the "perfect gas law." This describes the relationship of the three basic factors controlling the behavior of a gas: (1) pressure, (2) volume, and (3) temperature. For all

ABSOLUTE PRESSURE

Absolute pressure, normally expressed in terms of pounds per square inch absolute (psia), is defined as the pressure existing above a perfect vacuum. Therefore in the air around us, absolute pressure and atmospheric pressure are the same.

GAUGE PRESSURE

A pressure gauge is calibrated to read 0 psi regardless of elevation when not connected to a pressure producing source. The absolute pressure of a closed system will always be gauge pressure plus atmospheric pressure. At sea level, atmospheric pressure is 14.7 psi, therefore, at sea level, absolute pressure will be gauge pressure plus 14.7. Pressures below 0 psig are actually negative readings on the gauge, and are usually referred to as inches of mercury vacuum. A refrigeration compound gauge is calibrated in the equivalent of inches of mer-cury for negative readings. Since 14.7 psi is equivalent to 29.92 inches of mercury, 1 psi is approximately equal to 2 inches of mercury on the gauge dial. In the vacuum range, below 0 psig, 2 inches of mercury vacuum is approximately equal to a -1 psig.

It is important to remember that gauge pressure is only relative to absolute pressure. Table 1-1 shows rela-tionships existing at various elevations assuming that standard atmospheric conditions prevail.

Table 1-1Pressure Relationships at Varying Altitudes

Altitude (Feet) PSIG PSIA Inches

Hg.Boiling Point of Water

0 0 14.7 29.92 212°F1000 0 14.2 28.85 210°F2000 0 13.7 27.82 208°F3000 0 13.2 26.81 206°F4000 0 12.7 25.84 205°F5000 0 12.2 24.89 203°F

Table 1-1 shows that even though the gauge pressure remains at 0 psig regardless of altitude, the absolute pressure does change. The absolute pressure in inches of mercury indicates the inches of mercury vacuum that a perfect vacuum pump would be able to reach at the stated elevation. At 5,000 feet elevation under standard atmospheric conditions, a perfect vacuum would be 24.89 inches of mercury. This compares to 29.92 inches of mercury at sea level.

At very low pressures, it is necessary to use a smaller

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practical purposes, air and highly superheated refriger-ant vapors may be considered perfect gases, and their behavior follows this relationship:

Pressure One x Volume One = Pressure Two x Volume Two Temperature One Temperature Two

This is most commonly stated, P2V2

T2

P1V1

T1= .

Although the "perfect gas" relationship is not exact, it provides a basis for approximating the effect on a gas with a change in one of the three factors. In this relation-ship, both pressure and temperature must be expressed in absolute values, pressure in psia, and temperature in degrees Rankine or degrees Fahrenheit above abso-lute zero (°F plus 460°). Although not used in practical refrigeration work, the perfect gas relation is valuable for scientific calculations and is helpful in understanding the performance of a refrigerant vapor.

One of the problems of refrigeration is disposing of the heat that has been absorbed during the cooling process. A practical solution is achieved by raising the pressure of the vapor so that its saturation or condensing tem-perature will be sufficiently above the temperature of the available cooling medium (air or water) to assure efficient heat transfer. This will provide the ability of the cooling medium to absorb heat from the refrigerant and cool it below its boiling point (dew point). When the low pressure vapor with its low saturation temperature is drawn into the cylinder of a compressor, the volume of the gas is reduced by the stroke of the compressor piston. The vapor is discharged as a high pressure high temperature vapor and is readily condensed because of its high saturation temperature.

If refrigerant R-22’s pressure is raised to 195 psig, its saturation temperature will be 100°F (37.8°C). If the cooling medium’s temperature is lower than 100°F, heat will be extracted from the R-22 and it will be condensed, converted back to a liquid.

SPECIFIC VOLUME

Specific volume of a substance is defined as the number of cubic feet occupied by one pound (ft3/lb). In the case of liquids and gases, it varies with the temperature and the pressure to which the fluid is subjected. Following the perfect gas law, the volume of a gas varies with both temperature and pressure. The volume of a liquid varies with temperature. Within the limits of practical re-frigeration practice, it is regarded as non-compressible. Specific volume is the reciprocal of density (lb/ft3).

DENSITY

The density of a substance is defined as weight per

unit volume. In the United States, density is normally expressed in pounds per cubic foot (lb./ft3). Since by definition, density is directly related to specific volume, the density of a vapor may vary greatly with changes in pressure and temperature, although it still remains a vapor, invisible to the naked eye. Water vapor or steam at 50 psia pressure and 281°F temperature is over 3 times as heavy as steam at 14.7 psia pressure and 212°F.

Refrigerant 22 vapor at 76 psig and at 45°F has a density of 1.66 lb/ft3. At 150 psig and at 83°F, the refrigerants density is 3.02 lb/ft3 or 1.82 times as heavy.

PRESSURE AND FLUID HEAD

It is frequently necessary to know the pressure created by a column of liquid, or possibly the pressure required to force a column of refrigerant to flow a given vertical distance upwards.

Densities are usually available in terms of pounds per cubic foot, and it is convenient to visualize pressure in terms of a cube of liquid one foot high, one foot wide, and one foot deep. Since the base of this cube is 144 square inches, the average pressure in pounds per square inch is the weight of the liquid per cubic foot di-vided by 144. For example, water weighs approximately 62.4 pounds per cubic foot, the pressure exerted by 1 foot of water is 62.4 ÷ 144 or .433 pounds per square inch. Ten feet of water will exert a pressure of 10 X .433 or 4.33 pounds per square inch. The same relation of height to pressure holds true, no matter what the area of a vertical liquid column. The pressure exerted by other liquids can be calculated in exactly the same manner if the density is known.

The density of liquid refrigerant R-22 at 45°F, 76 psig is 78.8 lb./ft3. The pressure exerted by one foot of liquid R-22 is 78.8 ÷ 144 or .55 psig. A column of liquid R-22 10 feet high would then exert a pressure of 5.5 psig. At 100°F liquid temperature, the density is 71.2 lb./ft3. A one foot column then exerts a pressure of .49 psig. A ten foot column exerts a pressure of 4.9 psig.

Comparing other refrigerants at 45°F, R-404A has a den-sity of 70.1 lb./ft3. It then exerts a pressure of .49 psig per foot of lift. R-134a has a density of 79.3 lb./ft3, therefore it exerts a pressure of .55 psig per foot of lift.

Fluid head is a general term used to designate any kind of pressure exerted by a fluid that can be expressed in terms of the height of a column of the given fluid. Hence a pressure of 1 psi may be expressed as being equivalent to a head of 2.31 feet of water. (1 psi ÷ .433 psi/ft. of water). In air flow through ducts, very small pressures are encountered, and these are commonly

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expressed in inches of water. 1 inch of water = .433 ÷ 12 = .036 psi.

Table 1-2Pressure Equivalents in Fluid Head

Pounds Per Square Inch

AbsoluteInches

MercuryInches Water

Feet Water

.036 .07 1.0 .083

.433 .90 12.0 1.0

.491 1.0 13.6 1.131.0 2.03 27.7 2.31

14.7 29.92 408.0 34.0

FLUID FLOW

For a fluid to flow from one point to another, there must be a difference in pressure between the two points. With no pressure difference, no flow will occur. Fluids may be either liquids or vapors, and the flow of each is important in refrigeration.

Fluid flow through pipes or tubing is governed by the pressure exerted on the fluid, the effect of gravity due to the vertical rise or fall of the pipe, restrictions in the pipe resisting flow, and the resistance of the fluid itself to flow. For example, as a faucet is opened, the flow increases even though the pressure in the water main is constant and the outlet of the faucet has no restriction. Obviously the restriction of the valve is affecting the rate of flow. Water flows more freely than molasses due to a property of fluids called viscosity which describes the fluid's resistance to flow. In oils, the viscosity can be affected by temperatures, and as the temperature decreases the viscosity increases.

As fluid flows through tubing, the contact of the fluid and the walls of the tube create friction, therefore resistance to flow. Valves, fittings, sharp bends in the tubing and other obstructions also create resistance to flow. The basic design of the piping system and its installation will determine the pressure required to obtain a given flow rate.

In a closed system containing tubing through which a fluid is flowing, the pressure difference between two given points will be determined by the velocity, viscosity, and the density of fluid flowing. If the flow is increased, the pressure difference will increase since more friction will be created by the increased velocity of the fluid. This pressure difference is termed pressure loss or pressure drop.

Since control of evaporating and condensing pressures is critical in mechanical refrigeration work, pressure drop through connecting lines can greatly affect the performance of the system. Large pressure drops must be avoided. When designing and installing refrigeration and air conditioning system piping, pressure drop and refrigerant velocity must be given serious consideration. Section 18 in this series of manuals discusses piping and proper sizing and installation.

EFFECT OF FLUID FLOW ON HEAT TRANSFER

Heat transfer from a fluid through a tube wall or through metal fins is greatly affected by the action of the fluid in contact with the metal surface. As a rule, the greater the velocity of flow and the more turbulent the flow, the greater will be the rate of heat transfer. Rapid boiling of an evaporating liquid will also increase the rate of heat transfer. Quiet liquid flow (laminar flow) on the other hand, tends to allow an insulating film to form on the metal surface that resists heat flow, and reduces the rate of heat transfer.

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Large quantities of heat can be absorbed by a substance through an increase in sensible heat involving either a large temperature difference between the cooling media and the product being cooled or a large quantity (weight) of the cooling media. When the cooling media is involved in a change of state, latent heat, a smaller amount of the cooling media is needed to absorb an equivalent large quantity of heat. (Refer to Section 1.)

In mechanical refrigeration, a process is required that can transfer large quantities of heat economically and efficiently on a continuous basis. The processes of evaporation and condensing of a liquid are the logical steps in the refrigeration process.

Many liquids could be used for absorbing heat through the evaporation process. Water is ideal in many re-spects. Unfortunately it boils at temperatures too high for ordinary cooling purposes. It freezes at temperatures too high for low temperature conditions. A refrigerant must satisfy three main requirements:

1. It must readily absorb heat and change state to a vapor at the temperature required by the load.

2. It must readily reject heat and be returned to a liquid at a temperature required by the external cooling media, water or air.

3. For economy and continuous cooling, the system must use the same refrigerant over and over again.

There is no perfect refrigerant for all applications. There are varying opinions as to which refrigerant is best for a specific application.

TERMINOLOGY AND EXAMPLES

The following definitions primarily deal with the way the described materials behave as a working fluid in a thermodynamic system. There may be more specific or technically complete definitions which deal with the chemistry, transport properties, or other aspects of these materials’ composition or behavior which are unimport-ant in the present context.

PURE FLUID

A pure fluid is a single component fluid which does not change composition when boiling or condens-ing. A pure fluid is made up of one type of molecule. Examples: R-11, R-12, R-22, R-134a.

MIXTURE AND BLEND

Technically, there is no difference in the terms mixture and blend. They include any fluids which are composed of more than one component (i.e., more than one type of molecule). Azeotropic Refrigerant Mixtures (ARMs), Near-ARMs, and zeotropes (each is discussed below) are subsets of the larger group consisting of blends and mixtures.

The following definitions apply to dual component mix-tures, but three (“ternary”) or more component mixtures have similar but more complicated characteristics. From a thermodynamic working fluid point of view, the number of components in the fluid has little or no effect.

AZEOTROPIC REFRIGERANT MIXTURE

An azeotropic refrigerant mixture (ARM) is a multi-com-ponent which at the azeotropic point does not change composition when it evaporates or condenses since both components have exactly the same boiling tem-perature at that composition and pressure. It is made up of two or more types of molecules. In actuality, an ARM only exhibits such behavior at one temperature and pressure. Deviations from this behavior at other pres-sures are very slight and essentially undetectable.

ARMs are fairly complex mixtures whose properties depend upon molecular interactions which may result from polarity differences. They can be either minimum or maximum boiling point ARM’s. Even more complicated behavior can occur with ARMs. However, factors such as these are relatively unimportant when considering their performance in a system. The most important fac-tor is that they essentially behave as a pure substance when changing phase.

Examples of how minimum and maximum boiling point ARMs behave at their azeotropic and zeotropic compo-sition ratios at constant pressure are shown in Figures 2-1 and 2-2. These figures show the “Dew Line” (the temperature at which droplets appears as superheated vapor is cooled) and “Bubble Line” (the temperature at which bubbles first appear as subcooled liquid is heated) for mixtures at various fluid temperature and concentra-tions for one pressure value. At concentrations values away from the azeotropic value, the components (“A” and “B”) have different boiling temperature, and the liq-uid and vapor phases change percentage composition as the mixture evaporates or condenses.

Section 2REFRIGERANTS

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Figure 2-1

Figure 2-2

Examples:

R-502, an azeotrope of 48.8% R-22 and 51.2% R-115 at +66°F, has lower discharge temperatures than does R-22 for high compression ratio applications.

R-500, an azeotrope of 73.8% R-12 and 26.2% R-152a at +32°F, has approximately 15% more capacity than pure R-12 and was used to compensate for the capac-ity reduction arising from using a 60 Hz. R-12 system on 50 Hz.

R-507, a non-ozone depleting azeotrope at -40°F is 50% R-125 and 50% R-143a. It is an HFC replacement for R-502.

ZEOTROPIC MIXTURE

A zeotrope is a working fluid with two or more com-ponents of different vapor pressure and boiling points whose liquid and vapor components have different com-

position when the fluid evaporates or condenses. It is made of two or more types of molecules. Under constant pressure, the evaporating and condensing temperatures change with composition. (See Figure 2-3.) This change in temperature during constant pressure phase change is called glide, and varies with the components used and their proportions. (See Temperature Glide.) The amount of glide exhibited by a particular zeotrope is a measure of its deviation from being an azeotrope. By definition, azeotropes have zero glide at their azeotropic point. At other conditions, however, they can exhibit glide.

Figure 2-3

Examples:

R-401A is a mixture of R-22, R-152a, and R-124 which closely approximates the vapor pressure and perfor-mance of R-12. (R-401A is considered to be a Near Azeotropic mixture.)

NEAR-AZEOTROPIC REFRIGERANT MIXTURE

A Near-ARM is a zeotropic fluid whose composition is such that it exhibits a “small” amount of glide. Thus, “near-azeotropic” is a relative term. (See Figure 2-4.) Some researchers use a maximum glide temperature value of 10°F to distinguish Near-ARMs from zeo-tropes.

R-404A is a ternary mixture which closely approximates the vapor pressure and performance characteristics of R-502. R-402A is a mixture of R-22, R-125, and R-290 (propane) which closely approximates the vapor pres-sure and performance characteristics of R-502.

CHARACTERISTICS OF MIXTURES

HOW ARE COMPONENTS CHOSEN?

Components are primarily chosen based on the final characteristics desired in the mixture. These character-

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istics could include vapor pressure, transport properties, lubricant and materials compatibility, thermodynamic performance, cost, flammability, toxicity, stability, and environmental properties. Availability of components to a particular chemical manufacturer may also be a factor in component selection.

Proportions of components are chosen based on the exact characteristics desired in the final product. It is possible to modify the percentage composition and alter such parameters as capacity, efficiency, discharge temperature, vapor pressure, etc. Of course, changing one parameter will likely change others as well. There must also be a need to balance proportions to guarantee that a given mixture cannot become flammable, toxic, or environmentally undesirable under any foreseeable circumstance, such as leakage.

In many cases, there are computer programs which can use the properties of the individual components and cal-culate the resulting mixture properties and performance with a fairly high degree of accuracy.

MIXTURE BEHAVIOR

When an azeotropic or zeotropic mixture is entirely in the vapor state (i.e., no liquid is present in the container) the composition is totally mixed and all properties are uniform throughout.

When that same mixture is entirely in the liquid state (i.e., no vapor present in the container), like the 100% vapor state, the composition is totally mixed and all properties are uniform throughout.

In a partially full sealed container of a refrigerant mix-ture, the composition of the vapor and liquid phases can be different. The degree of difference depends upon whether the mixture is an azeotrope, zeotrope, or near azeotrope.

AZEOTROPE

The percentage of an ARM will be virtually the same in the saturated region where both liquid and vapor are in contact with each other, except at the Azeotropic composition point (See Figures 2-1 and 2-2). At this point its liquid and vapor components will have the same boiling point. At this condition each component has the same boiling point and each vaporizes in proportion to the amount present in the liquid phase. The resulting vapor is the same composition as the liquid. The same is true for the reverse (condensing) process. At other conditions, however, the percentage composition of liquid and vapor phases will be slightly different.

ZEOTROPE

The percentage composition of a zeotropic mixture may be substantially different in the saturated mode when liquid and vapor are in contact with each other. This is because there is no unique boiling point for each component, and they will not vaporize at the rates proportional to their composition in the liquid state. The higher vapor pressure component (with the lower boiling point) will vaporize faster than the lower vapor pressure component (with the higher boiling point), and result in percentage composition changes in both the liquid and vapor phases as vaporization progresses. The higher vapor pressure component will be in higher composi-tion in the vapor phase above the liquid. This process is called “fractionation.”

NEAR-AZEOTROPIC REFRIGERANT MIXTURES

The percentage composition of the liquid and vapor phases of a Near-ARM will be nearly identical, due to the very similar vapor pressure values of each component. Thus, a Near-ARM behaves essentially the same as an ARM from this standpoint.

WHAT HAPPENS TO MIXTURE COMPOSITION DURING SYSTEM CHARGING?

Depending on how system charging is performed (i.e. with vapor or liquid being removed from the cylinder), the refrigerant may change phase in the cylinder. Since pure fluids and ARMs (except as discussed in a previous sec-tion) do not change composition with changes in phase, there is no change in composition with these materials during system charging with vapor or liquid. On the other hand vapor charging with a zeotropic mixture can result in significant composition changes due to fractionation of the components as discussed earlier.

If an entire cylinder of refrigerant is used to charge a system, then the composition change process has no effect since the entire contents of the cylinder will go

Figure 2-4

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into the system. However, if only part of a cylinder of a zeotropic refrigerant is vapor charged into a system, the vapor composition can change substantially during the process. As a result, only liquid charging (i.e., what leaves the cylinder) should be used for zeotropes un-less the entire cylinder is to be used for one system. Of course, proper protection against liquid ingestion by the compressor must be provided. This could be in the form of an accumulator-type device which allows the liquid to boil and enter the compressor as a vapor or meters small amounts of liquid into the suction side of the system. Another choice is the “Dial-a-Charge” type of charging system, which takes a measured amount of liquid from the cylinder and puts all of it into the sys-tem being sure that it goes into the suction side of the compressor as a vapor.

Since Near-ARMs are actually zeotropes, they also result in composition changes during charging, but to a much smaller extent than occurs with zeotropes. When charging Near-ARM refrigerants, liquid (from the cylinder) should be used to avoid composition changes (unless the entire contents is going into the system). The last few percent of the contents of a cylinder should not be used as this is when composition changes can be the greatest.

Guidelines for charging procedures and how much of cylinder’s refrigerant to use during charging will be provided by the refrigerant manufacturers.

Azeotropes, as described above behave as a pure material during boiling and condensing, and do not ap-preciably change percentage composition.

Zeotropes, on the other hand, do not behave as a pure material during boiling and condensing, and the percent-age composition of the liquid and vapor phases can be different. This characteristic can have a significant effect on the composition of the refrigerant left in the system after a leak in the vapor-containing region of a system, and to the subsequent composition after the lost refrigerant had been replaced.

The magnitude of this effect depends strongly upon how much the mixture departs from being an ARM.

Leakage scenarios are discussed later.

TEMPERATURE GLIDE

Figure 2-5 shows how a zeotropic two-component mixture behaves during change of phase at various concentrations for a constant pressure. As subcooled liquid is heated, the higher vapor pressure component eventually reaches its boiling point and begins to form vapor. This condition is called the “Bubble Point,” and

is the temperature at which bubbles (flash gas) begin to appear. The vapor is rich with the high vapor pressure component. As the temperature increases, more and more of the high vapor pressure component vaporizes, reducing its component in the liquid phase. At the same time, the lower vapor pressure component eventually reaches its boiling point and begins to vaporize. Finally, the high vapor pressure component is fully evaporated. All that is left is the low vapor pressure component and when the last drop evaporates, the “Dew Point” temperature is established. This is the temperature at which liquid begins to appear when the zeotropic vapor is cooled. The difference between the dew point and the bubble point temperature is known as “temperature glide.” It varies with percentage composition of the components as well as pressure.

Figure 2-5

The practical effect of glide in heat exchangers is that as the refrigerant mixture flows through the tubing at con-stant pressure, the evaporating (or boiling) temperature will change as the composition of the liquid and vapor phase change. Thus, a constant evaporating tempera-ture does not occur, even with constant pressure.

The amount of glide varies with the pressure and percentage composition of each component present in the mixture. Glide can vary from an imperceptible amount with a Near-ARM to ten or more degrees F with a zeotrope. Many researchers consider a Near-ARM to have glide less than ten degrees F. The glide of many mixtures is given in Emerson Climate Technologies, Inc. Changeover Guidelines and TIP card/PT chart.

In any heat exchanger, flow of refrigerant through the tubing results in a pressure drop from the entrance to the exit. Consequently, since the pressure at which the phase change is occurring is decreasing along the length of the heat exchanger, the evaporating or con-

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densing temperature will decrease as the saturated re-frigerant moves through the heat exchanger. Of course, the amount of change in evaporating or condensing temperature depends upon the magnitude of the pres-sure drop, but it can be several degrees. The change in evaporating and condensing temperature which occurs with today’s pure fluids in many systems is similar to that which occurs with Near-ARMs due to glide as they pass through the heat exchanger.

Figure 2-6

A schematic example of how glide with a zeotrope af-fects temperatures in an evaporator and condenser in a system is shown in Figure 2-6. As a result of the composition change as the refrigerant flows through the heat exchangers, the evaporating and condensing temperatures decrease. Of course, with Near-ARMs the temperature change is very slight and probably undetectable.

As a practical matter, the pressure drop in the evapora-tor tends to counteract the temperature glide being less than would be expected at constant pressure conditions. The effects are additive in the condenser.

WHAT HAPPENS TO REFRIGERANT COMPOSITION DURING A LEAK?

In a single component refrigerant there is no change in percentage composition of the refrigerant. In an azeo-tropic mixture there is virtually no change in percentage composition of the refrigerant.

If a leak occurs in a zeotropic mixture in a portion of an operating system where only vapor is present (such as at the compressor discharge or suction line), the system’s refrigerant composition will not change since the percentage composition of the vapor is identical to the mixture and each component will leak at the same

rate. If a leak occurs in a portion of an operating system where only liquid (such as in the liquid line) is present, the composition will not change since the percentage composition of the liquid is identical to the mixture and each component will leak at the same rate.

However, if a leak occurs in a portion of an operating system where both liquid and vapor exist simultaneously (such as in the evaporator or condenser), “fractionation” (unequal evaporation or condensing of the refrigerant in a change in percentage composition between liquid and vapor phase as discussed previously) will occur and there can be a change in percentage composition of the refrigerant left in the system. For example, if a leak oc-curs in the two-phase potion of the evaporator and only vapor leaks out, the vapor will be richer in the higher vapor pressure component, resulting in a change in the percentage composition of the remaining refrigerant in the system. If the system is recharged with the original composition, the mixture in the system can never get back to the original composition, and system perfor-mance (such as capacity or efficiency) may change to some degree. Repeated leak and recharge cycles will result in additional change. However, in most operating systems where two phases are present, turbulent mixing occurs and liquid will leak along with the vapor which minimizes the effect the vapor leakage effect

It is important to keep in mind that to have a change in composition in an operating system, the leak must occur in a portion of the system where both liquid and vapor phases exist simultaneously, and only vapor leaks out.

While a system is off, there will be parts of the system where pure vapor exists and parts where pure liquid exists, and these locations can change with varying environmental conditions. Since the composition of the liquid and vapor phases will be different for a zeotrope in liquid-vapor equilibrium (as discussed previously), a leak in the area where vapor alone is present can result in a composition change in the system. Such a leak could be significant during the long wintertime off cycle of an air conditioning system, or during long periods of non-use for any system.

The effect of such leaks with Near-Azeotropes is much less due to the fact that the percentage compo-sition difference between the liquid and vapor phases is very small. Theoretical leakage effect calculations (verified by initial laboratory testing) with Near-ARMs are undetectable, even with several repeated leak and recharge cycles.

TYPES OF REFRIGERANT

There are many different types of refrigerants avail-

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able. Several have been in common use for a number of years. In the United States, because of the Clean Air Act of 1990 amended, the common refrigerants are changing. The Montreal Protocol also effects the com-mon usage refrigerants in many parts of the world.

In early refrigeration applications, ammonia, sulfur diox-ide, methyl chloride, propane, and ethane were widely used. Some of these are still used today. Because these products are either toxic, dangerous, or have other undesirable characteristics, they have been replaced by compounds developed especially for refrigeration use. Specialized refrigerants are used for ultra-low temperature work, and in large centrifugal compressors. For normal commercial refrigeration and air condition-ing applications utilizing reciprocating compressors, refrigerants R-12, R-22, and R-502 have been used almost exclusively. These were developed originally by DuPont as Freon® refrigerants. The ASHRAE numerical designations are now standard with all manufacturers of refrigerants

These refrigerants use chlorine as one of the elements in their composition. The scientific community has de-termined that chlorine reaching the upper atmosphere causes a reduction in the upper atmospheric Ozone (O3). R-12 and R-502 are classified as ChloroFluoro-Carbons (CFCs) and are being phased out in favor of non Ozone depleting refrigerants. R-22 is classified as a HydroChloroFluoroCarbon (HCFC). the hydrogen molecule allows the chlorine to break down lower in the atmosphere reducing its Ozone Depletion Potential.

The replacement refrigerants are referred to as Hy-droFluoroCarbons (HFCs). These refrigerants do not contain chlorine. R-134a is the HFC replacement for medium temperature R-12 applications. It is not rec-ommended for use in application at saturated suction temperatures below -10°F. R-404A and R-507 are the HFC replacement for R-502 for use in medium & low

temperature applications. R-407C and R-410A are the HFC replacements for R-22 for use in high temperature air conditioning applications. R-410A is a high pressure refrigerant and is not a retrofit refrigerant for R-22. These HFC refrigerants are similar to the CFC and HCFC refrigerants, however, they are not identical with respect to pressures, temperatures and enthalpy.

Service Refrigerants

The Clean Air Act of 1990 Amended prohibits the manufacture or import of CFC refrigerants into the United States after December 31, 1995. In order to maintain those systems already in operation using these refrigerants, the chemical companies have developed service replacement refrigerant blends. These blends are an HCFC. They use R-22 as the base refrigerant and blend other refrigerants with it to achieve a desir-able property.

It is desirable to have a service replacement that looks very much like the CFC refrigerant in the system. However, they cannot be mixed with the CFC refriger-ant. CFC refrigerant in the system must be properly recovered before the replacement HCFC refrigerant is put into the system.

R-401A/B and R-409A are service replacements for R-12. R-402A and R-408A are service replacements for R-502.

NOTE: Not all service replacement refrigerants are approved for use in Copeland® brand compres-sors.

Refrigerants Solubility of Water

Table 2-1 lists several comparative properties of refriger-ants including the Solubility of Water in different refriger-ants at two different temperatures. We would consider the 100°F point to be liquid refrigerant entering the TEV

Table 2-1Comparative Properties of Several Refrigerants

R-12 R-401A R-401B R-409A R-134a R-22 R-407C R-410A R-502 R-402A R-408A R-404A R-507

Saturation Pressure, psig at 70°F 70.2 85.8* 91.9* 106.1* 71.2 121.4 114.9* 200.6* 137.6 160.4* 135.1* 147.5* 153.6

Boiling Point, °F, at 14.7 psia (Standard Atmospheric Pressure) -21.6 -27.3** -30.41** -29.6** -14.9 -41.4 -46.4** -69.9** -49.8 -56.5** -46.7** -51.6** -52.1

Liquid Density, lb./ft3 at 70°F 82.7 75.4 75.2 77.06 76.2 75.5 71.8 67.6 78.6 72.2 80.9 66.5 66.5

Solubility of Water, PPM at 100°F 165 NA NA 1600 1900 1800 NA 2850 740 NA 900 970 970

Solubility of Water, PPM at -40°F 1.7 NA NA 190 NA 120 NA 90 12 NA 100 100 100

* Dew Point Pressure ** Bubble Point TemperaturesNA - Not Available

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and the -40°F point to be saturated refrigerant in the evaporator. The concern of the technician should be how much water can the refrigerant hold before it becomes free water and causes problems. Ideally there should not be any moisture in the refrigerant in the system. Proper installation and service techniques should assure that the system is clean and dry. This includes the proper use of a vacuum pump and micron gage.

The table lists the solubility in PPM (parts per million by weight). PPM may be a meaningless number to the average technician and installer so let’s equate it to something more understandable. Filter dryers are rated in drops of water before they become saturated and can no longer hold any additional water. ARI (The Air Conditioning and Refrigeration Institute) standard is that 20 drops of water equals 1 cc or 1 gram by weight.

Table 2-1 lists R-22 at 100°F having a solubility of 1800 PPM. This simply means that R-22 liquid at 100°F can hold up to 11 drops of water per pound before there is free water. In an Air Conditioning system, when the refrigerants temperature is lowered to +40°F, the solu-bility drops to 690 PPM (Not shown in Table 2-1.). This equates to 6 drops of water before there is free water. In an Air Conditioning system, the free water will not freeze but may cause other chemical reaction damage.

When the refrigerant goes through the TEV and its temperature is lowered to -40°F, the solubility drops to 120 PPM. This now equates to one drop of water per pound of refrigerant. Once there is more than one drop of water per pound at -40°F, there is free water in the system and the TEV will freeze closed.

Comparative Refrigeration Effect

Table 2-2 lists comparative data for different refrigerants.

Each refrigerant has different suction and discharge pressures for the same operating conditions. This should be expected in that each refrigerant is made up of different chemicals. It is interesting however that the compression ratio for the refrigerants is not that dis-similar. The highest is R-401B, 13.4:1, and the lowest is R-502, 8.6:1.

The Specific Volume (ft3/lb.) of the return gas varies significantly, with the medium pressure refrigerants having the largest Specific Volume, the lowest Density. This equates to a fewer number of pounds of refrigerant being circulated through the compressor per revolution of the compressor motor. The Refrigeration Effect is the pounds of refrigerant in circulation times the refrigerants Enthalpy.

Table 2-2Comparative Refrigeration Effect

R-12 R-401A R-401B R-409A R-134a R-22 R-407C R-410A R-502 R-402A R-408A R-404A R-507

Evaporating Pressure, psig 0.6 2.9 *0.8 *1.9 *3.7 10.2 6.2 26.3 15.5 18.1 13.4 16.3 17.8

Condensing Pressure psig 136 166.1 176.3 140.6 146.5 226 256.6 365.4 246 289.3 252.1 270.3 280.6

Compression Ratio 9.9 10.3 13.4 11.3 12.5 9.7 13 9.3 8.6 9.3 9.5 9.2 9.1

Specific Volume of Return Gas, ft3/lb 3.03 3.7 4.2 4.1 5.7 2.53 1.6 1.5 1.66 1.4 2.1 1.79 1.69

Refrigeration Effect BTU/lb 53.7 71.7 72.1 68.6 69.5 73.03 70.7 77.3 48.72 51 62 51.8 53.7

* In/Hg

(Data shown at -20°F evaporating temperature, 110°F condensing temperature, 0°F liquid subcooling, 65°F return gas temperature.) Not all of the refrigerants are recommended at this conditions.

Table 2-3Refrigerant/Lubricant Chart

Conventional Service Blends (HCFC) Non-Ozone Refrigerants Depleting (HFC)

Refrigerants

CFC R-12 R-401A R-134a R-401B R-409A

CFC R-502 R-402A R-404A R-048A R-507A

HCFC R-22 R-407C R-410A

Lubricants

MO AB POE POE AB/MO* POE/MO*

*AB or POE must be at least 50% of the system lubricant.

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Again note that each refrigerant has its own Refrigera-tion Effect. The Density of each refrigerant is different as is the Enthalpy of each refrigerant different. Note that the service blends have a higher Refrigeration Effect than their CFC counter parts. It is for this reason that many systems using the service blends do not need as much refrigerant in circulation as they did with the CFCs that were removed.

Each chemical company can provide the information for the refrigerants they manufacture. The form of the information may be in print per the examples shown at the end of this section in Figures 2-4, 2-5 and 2-6 or may be on a computer disc.

Refrigerants and Lubricants

Table 2-3 is a cross reference of the various refrigerants as one goes from the Conventional Refrigerants, to the Service Blends, to the Non Ozone Depleting Refriger-ants. It also lists the types of lubricants recommended for use with each category. A description of each refrig-erant and its application follows.

Because the mineral oil (MO) used for so many years in air conditioning and refrigeration systems has miscibility issues with the service blends and HFC refrigerants, the proper lubricant approved by Emerson Climate Tech-nologies, Inc. must be used with each type of refriger-ant. (Refer to Application Engineering Bulletin 17-1248, Refrigerant Oils.) Later in this section the refrigerant-oil relationship is discussed.

REFRIGERANTS

REFRIGERANT 12

Refrigerant 12 is a pure fluid and is categorized as a ChloroFluoroCarbon (CFC). It has been widely used in household and commercial refrigeration and air con-ditioning. At temperatures below its boiling point it is a clear, almost colorless liquid. It is not toxic or irritating, and is suitable for high, medium, and low temperature applications. Refrigerant 12 has been determined to cause the depletion of the upper atmosphere Ozone (O3) layer when it reaches the upper atmosphere. It has been assigned an Ozone Depletion Potential (ODP) of 1. All other refrigerants ODP is measured against R-12. R-12 primary ingredients are chlorine, fluorine, and carbon. The chlorine has been determined to cause the ozone depletion. New R-12 cannot be produced or brought into the United States after December 31, 1995. The result should cause the use of this refrigerant to diminish quickly after that date.

REFRIGERANT R-401A/B

Refrigerants R-401A and R-401B are zeotropic HCFC blends. They are the service replacements for R-12 These refrigerants are a blend of R-22, R-124 and R-152a. The difference between the two is the percent-age of each refrigerant in the blend. (R-401A, 53% R-22, 13% R-152a, 34% R-124) (R-401B, 61% R-22, 11% R-152a, 28% R-124) R-401A is the high and me-dium temperature service replacement for R-12 and R-401B is the low temperature service replacement.

Because these refrigerants are not as miscible with mineral oil in the vapor state, Emerson Climate Tech-nologies, Inc. recommends an approved Alkyl Benzene (AB) lubricant be used with R-401A/B. The AB lubricant must be at least 50% of the lubricant in the system.

The Ozone Depletion Factor of R-401 A/B is 0.030 and 0.035 respectively. The glide is 9.5°F and 8.8°F respectively. The Enthalpy of these blends is ap-proximately 25% greater than R-12, and as such, the system charge may be as much as 15% less than the R-12 charge. The systems pressure will be higher than the R-12 pressures and the discharge temperature will be lower. Even though the system may require less refrigerant than the original refrigerant charge, systems using a TEV must be recharged to insure a full column of liquid at the TEV.

REFRIGERANT R-409A

Like refrigerants R-401A/B, refrigerant R-409A is a zeo-tropic HCFC blend. It is considered to be a medium/low temperature Service Replacement refrigerant for R-12. It is a blend of R-22, R-124 and R-142b. The glide for R-409A is 14°F. Because it is less miscible in mineral oil, an approved AB lubricant must be used. The AB lubricant must be at least 50% or more of the lubricant in the system. The pressures and temperatures in the system will be different than when using R-12.

REFRIGERANT 134A

Refrigerant R-134a is a pure fluid and is categorized as an HydroFluoroCarbon (HFC). It is the medium and high temperature replacement for R-12. Like R-12, at temperatures below its boiling point, it is a clear color-less liquid. Its basic chemical components are Hydro-gen, Fluorine, and Carbon. With the chlorine element removed, its Ozone depletion factor is 0. Unlike R-12, R-134a is not recommended for use in systems where the saturated suction temperature is below -10°F. The saturated suction pressure of R-134a compared to R-12 is similar. The discharge pressure will be higher there-fore the compressors compression ratio will be greater. Compressor displacements for R-134a will be similar to

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those for R-12. Unlike R-12, it is not miscible in mineral oil. R-134a requires the lubricant in the compressor/sys-tem be an approved Polyol Ester (POE). R-134a is a single element compound. (Refer to Emerson Climate Technologies, Inc. Application Engineering Bulletins for the listing of approved lubricants.)

REFRIGERANT 22

Refrigerant 22 in most of its physical characteristics is similar to R-12. However, it has much higher saturation pressures than R-12 for equivalent temperatures. It has a much greater latent heat of evaporation, and a lower specific volume. For a given volume of saturated refrig-erant vapor, R-22 has greater refrigerating capacity. This allows the use of smaller compressor displacements, resulting in smaller compressors for performance com-parable with R-12. Where size and economy are critical factors, such as package air conditioning units, R-22 is widely used. R-22 is categorized as an HCFC. Its com-ponents are hydrogen, chlorine, fluorine, and carbon. Because it contains chlorine, it has an Ozone Depletion Factor of .05 and it will ultimately be phased out under the rules of the Clean Air Act of 1990 Amended.

Because of its characteristics at low evaporating tem-peratures and it’s high compression ratios, the tempera-ture of the compressed R-22 vapor becomes so high it can cause damage to the compressor. In the past, Emerson Climate Technologies, Inc. recommended R-22 in single stage systems for high and medium tem-perature applications only. It can however be used in low temperature single stage systems only when using Emerson Climate Technologies, Inc. patented Demand Cooling® system or some method of de-superheating before the refrigerant is compressed. It can also be used in low temperature and ultra-low applications in multi-stage systems where the vapor temperature can be adequately controlled. R-22 is a single element compound.

REFRIGERANT 502

Refrigerant 502 is an azeotropic mixture of R-22 and R-115. Its azeotropic rating point is at +66°F. In most physical characteristics, R-502 is similar to R-12 and R-22. While its latent heat of evaporation is not as high as either R-12 or R-22, its vapor is much heavier, or to describe it differently, its specific volume is much less. For a given compressor displacement, its refrigerat-ing capacity is comparable to that of R-22, and at low temperatures is a little greater. As with R-22, a com-pressor with a smaller displacement may be used for performance equivalent to R-12. Because of its excellent low temperature characteristics, R-502 has been well suited for low temperature refrigeration applications. It

has been the refrigerant of choice for all single stage applications where the evaporating temperature is 0°F or below. It has also been very satisfactory for use in two stage systems for ultra low temperature applications. It gained popularity for use in the medium temperature range.

Like refrigerant R 12, R-502 is considered to be a CFC. The R-115 used to make the azeotropic blend is a CFC. It is this component that makes R-502 a CFC. Like R-12, its production and import into the United States is banned after December 31, 1995. Its Ozone Deple-tion Factor is 0.3.

REFRIGERANT R-402A

Like refrigerants R-401A/B, R-402A is a Zeotropic HCFC blend. It is considered to be a Service Replacement refrigerant for R-502. It is a blend of R-22, R-125, and R-290. Its enthalpy is similar to R-502 therefore the system charge will be the same. The glide for R-402A is 2.8°F. AB lubricant is recommended for use with this refrigerant and must be at least 50% or more of the lubricant in the system. The system pressures will be higher than with R-502, however the discharge tem-perature will be lower.

REFRIGERANT R-408A

Refrigerant R-408A is a zeotropic Service Replace-ment HCFC blend alternative for R-502A. It is a blend of R-22, R-125 and R-143A. Its enthalpy is similar to R-502 therefore the systems charge will be the same. The glide for R-408A is 1.0°F. Because it is less miscible in mineral oil, it is recommended that an AB lubricant be used. The AB lubricant must be at least 50% or more of the lubricant in the system. The systems pressures and temperatures will be different.

REFRIGERANT 404A

Refrigerant R-404A is a zeotropic blend of three refrig-erants and is one of two refrigerants considered to be the HFC replacement for R-502. The three refrigerants are R-125, R-143a, and R-134a. R-404A has an Ozone Depletion Factor of 0. Like R-502, it is an excellent re-frigerant for low and medium temperature applications. It has performance characteristics similar to R-502 except that it is not miscible with mineral oil. Like R-134a, an ap-proved POE lubricants must be used with R-404A. The system must have less than 5% residual mineral oil.

The saturated suction pressures will be similar to R-502. The discharge pressures will however be higher than R-502. The compressors discharge temperature will be lower when using R-404A as compared to R-502. Because R-404A is a zeotrope, the three refrigerants

2-9


that make up it composition do not boil at the same temperature at a specific pressure. Its glide is 0.8°F.

REFRIGERANT 507

Refrigerant R-507 is another HFC replacement for R-502. Like R-404A, it has an Ozone Depletion Factor of 0. This refrigerant is an azeotropic blend made up of R-125 and R-143a. At its azeotropic rating point it is a true azeotrope. Like R-404A, R-507 is not miscible in mineral oil and an approved POE lubricant must be used when the system is using R-507.

R-507 saturated suction pressures will be similar to R-502, however its discharge pressures will be higher than R-502. The discharge temperature will be a little lower than R-502.

REFRIGERANT SATURATION TEMPERATURE

At normal room temperatures, the above refrigerants can exist only as a vapor unless they are pressurized. Their boiling points at atmospheric pressure are below 0°F (See Table 2-1). Therefore, refrigerants are always stored and transported in special pressure resistant drums. As long as both liquid and vapor are present in a closed system, and there is no external pressure influence, the refrigerant will evaporate or condense as a function of the surrounding temperature. Evaporation or condensation will continue until the saturation pres-sure and temperature corresponding to the surrounding temperature is reached. When this occurs, heat transfer will no longer take place. A decrease in the surrounding temperature will allow heat to flow out of the refrigerant. This will cause the refrigerant to condense and lower the pressure. An increase in the surrounding temperature will cause heat to flow into the refrigerant. This will cause the refrigerant to evaporate, and raise the pressure.

Understanding this principle, and by knowing the surrounding temperature, the refrigerants saturation pressure is known. Conversely, knowing the refriger-ants saturation pressure, the refrigerants temperature is known.

REFRIGERANT EVAPORATION

Presume the refrigerant is enclosed in a refrigeration system and it’s temperature is equalized with the sur-rounding temperature. If the pressure in the refrigera-tion system is lowered, the saturation temperature (the boiling point) will be lowered. The temperature of the liquid refrigerant is now above its boiling point. It will immediately start to boil absorbing heat. In the process, the temperature of the remaining liquid will be reduced. Vapor (flash gas) will occur as the change of state takes place. Heat will now flow into the refrigeration

system from outside of the refrigerant system. This oc-curs because the refrigerants temperature decreased. Boiling will continue until the surrounding temperature is reduced to the saturation temperature of the refriger-ant, or until the pressure in the system again rises to the equivalent saturation pressure of the surrounding temperature. If a means, a compressor, is provided to re-move the refrigerant vapor so that the system pressure will not increase, and at the same time liquid refrigerant is fed into the system, continuous refrigeration will take place. This is the process that occurs in a refrigeration or air conditioning system evaporator.

REFRIGERANT CONDENSATION

Presume the refrigerant is enclosed in a refrigeration system and its temperature is equalized with the sur-rounding temperature. If hot refrigerant vapor is pumped into the system, the pressure in the refrigeration system will be increased and its saturation temperature, boiling point, will be raised.

Heat will be transferred from the incoming hot vapor to the refrigerant liquid and the walls of the system. The temperature of the refrigerant vapor will fall to its condensing, saturation temperature, and condensa-tion will begin. Heat from the refrigerants latent heat of condensation flows from the system to the surrounding temperature until the pressure in the system is lowered to the equivalent of the saturation pressure of the sur-rounding temperature. If a means, the compressor, is provided to maintain a supply of hot, high pressure re-frigerant vapor, while at the same time liquid refrigerant is drawn off, continuous condensation will take place. This is the process taking place in a refrigeration and air conditioning system condenser.

REFRIGERANT-OIL RELATIONSHIPS

In reciprocating and scroll compressors, refrigerant and lubricant mix continuously. Refrigerant gases are soluble in the lubricant at most temperatures and pressures. The liquid refrigerant and the lubricant can be completely miscible, existing as a single phase mixture. Separa-tion of the lubricant and liquid refrigerant into separate layers, two phases, can occur. This generally occurs over a specific range of temperature and composition. This separation occurs at low temperatures, during off-cycles. It occurs in the compressor sump and other places such as accumulators, receivers and oil separa-tors. In the two phase state, the denser liquid refrigerant is underneath the less dense lubricant. This separation does not necessarily affect the lubricating ability of the lubricant but it may create problems in properly supply-ing lubricant to the working parts. Those compressors with oil pumps have their pickup low in the crankcase.

2-10


This will not allow the less dense lubricant to be picked up when the mixture is in the two phase state. Accumu-lators, receivers and oil separators also pick up liquid low in the component. The location and size of the oil pick up hole is critical. It is imperative that the lubricant return to the compressor is not delayed. The industry has had a successful experience with R-12, R-22 and R-502 refrigerants and mineral oils. The experience with POEs and HFCs has not been as extensive but has been very successful.

The new chlorine-free HFC refrigerants are more polar than the current CFC/HCFC refrigerants. The result is that mineral oils are not miscible with the HFC refriger-ants. Polyol Ester (POE) lubricants are more polar than the mineral oils. This polarity of the POE and HFCs make the two miscible and as such, POE is the lubricant to be used with HFC refrigerants. POE lubricants are synthetic, formed by mixing a specific organic acid with a specific alcohol and subjecting them to a reaction. The result is a POE base material and water. The water is driven off and an additive package is mixed with the POE to arrive at a unique approved lubricant. Emerson Climate Technologies, Inc. has approved specific POE lubricants after extensive laboratory and field testing.

POE lubricants are hygroscopic and want to re-absorb water. If care is not taken to keep moisture out of the system, a chemical reaction can occur and produce some weak organic acids. The recommended maximum moisture content in POE is 50 Parts Per Million (PPM). There should be concern when the moisture content is in the 50 to 100 PPM range. Should the moisture content rise above 100 PPM, action must be taken to remove the moisture. It is important that the engineer and the service technician understand the need for “clean and dry” hermetically sealed air-conditioning and refrigera-tion systems. Proper evacuation techniques and the use of approved filter-driers with adequate moisture removal capacity is crucial to avoid system problems.

HCFC Service Blend refrigerants are a blend of HCFC R-22 and other refrigerants. The other refrigerants are of different types and can be an HFC. This blending is done to achieve a service refrigerant that reacts similarly with respect to temperature, pressure and enthalpy as the CFC refrigerant being removed. These blends are not as miscible/soluble in mineral oil as is the CFC.

Alkyl Benzene (AB) lubricant is a synthetic hydrocar-bon. Its composition is more polar than mineral oil. This polar property makes the HCFC service blends more soluble/miscible in the AB lubricant. Like POE, specific AB and Alkyl Benzene Mineral Oil (ABMO) blends have been approved by Emerson Climate Technologies, Inc. after extensive laboratory and field testing.

Since oil must pass through the compressor cylinders to provide lubrication, a small amount of lubricant is always in circulation with the refrigerant. Lubricant and refrigerant vapor do not mix readily. The lubricant can be properly circulated through the system only if vapor velocities are high enough to carry the lubricant along. If velocities are not sufficiently high, lubricant will tend to lie on the bottom of refrigeration tubing, decreasing heat transfer and possibly causing a shortage of lubricant in the compressor. As evaporating temperatures are low-ered, this problem increases. For these reasons, proper design of piping is essential for satisfactory lubricant return. (See Section 18, AE-104)

One of the basic characteristics of a refrigerant and lubricant mixture in a sealed system is the fact that refrigerant is attracted to the lubricant. The refrigerant will vaporize and migrate through the system to the com-pressor crankcase even though no pressure difference exists. On reaching the crankcase the refrigerant will condense into the lubricant. This migration will continue until the lubricant is saturated with liquid refrigerant. Further migration will cause the liquid refrigerant to settle beneath the lubricant.

Excess refrigerant in the compressor crankcase can result in violent foaming and boiling action, driving all the lubricant from the crankcase causing lubrication problems. It can also cause slugging of the compres-sor at start up. Provisions must be made to prevent the accumulation of excess liquid refrigerant in the compressor.

Proper piping and system design for the refrigerants and lubricants is critical for the lubricant return. The new HFC refrigerants are relatively more soluble in POE lubricants than CFC/HCFC refrigerants and mineral oil. It is important that the engineer and the service techni-cian understand that mineral oil can not be used with the HFC refrigerants and only POE lubricants approved by Emerson Climate Technologies, Inc. are to be used in Copeland® brand compressors.

Refer to Emerson Climate Technologies, Inc. Application Engineering Bulletins for a listing of approved lubricants and refrigerants

REFRIGERANT TABLES

To accurately determine the operating performance of a refrigeration system, precise and accurate information is required. This includes various properties of refriger-ants at any temperature and pressure to be considered. Refrigerant manufacturers have calculated and com-piled this data in the form of tables of thermodynamic properties. These tables are made available to design and application engineers and others who have a need

2-11


for this information.

Table 2-4 is an excerpt from an R-134a saturation table. It lists the five major saturation properties of R-134a, both liquid and vapor, at various temperatures. Pres-sure, volume, and density have been discussed previ-ously. Table 2-5 is an excerpt from an R-22 saturation table.

Enthalpy is a term used in thermodynamics to describe the heat content of a substance. In refrigeration practice, enthalpy is expressed in terms of BTU per pound. An arbitrary base of saturated liquid at -40°F. has been accepted as the standard zero value. In other words, the enthalpy of any refrigerant is zero for liquid at -40°F. Liquid refrigerant at temperatures below -40°F. is considered to have a negative enthalpy. Refrigerant at all temperatures above -40°F. has a positive enthalpy value.

The difference in enthalpy values at different parts of the system are commonly used to determine the perfor-mance of a refrigeration unit. When the heat content per pound of the refrigerant entering and leaving a cooling coil is determined, the cooling ability of that coil can be calculated provided the refrigerant flow rate is known.

Entropy can best be described as a mathematical ratio used in thermodynamics. It is used in solving complex refrigeration engineering problems. It is not easily defined or explained. It is seldom used in com-mercial refrigeration applications and a discussion of it is beyond the scope of this manual. For our purpose, the compression process within the compressor is an Isentropic process.

Figure 2-6 is an excerpt from an R-404A superheat table. Superheat tables list saturation evaporating temperature and pressure in increments of 1 psi, and tabulate changes in specific volume, enthalpy, and entropy for various increases in temperature of the refrigerant vapor or superheat. Since superheat tables are quite lengthy and are available separately in bound volumes, complete superheat tables have not been included in this manual.

SATURATION PROPERTIES

Temperature/Pressure tables are specific to a refriger-ant. The temperature and pressure columns of these tables are most usable to a service technician/engineer. These tables are cumbersome for the average service person because they generally are multiple pages per refrigerant. Table 2-7 is an example of the basic pres-sures and saturated temperatures for refrigerant R-507 taken from the basic tables. This consolidates the data into one single table.

POCKET TEMPERATURE-PRESSURE CHARTS

Small pocket sized folders listing the saturation tem-peratures and pressures of common refrigerants are readily available from expansion valve and refrigerant manufacturers. Table 2-8 is a typical example of a pocket sized chart for refrigerants approved for use in a Copeland® brand compressor.

A saturation chart for ready reference is an invaluable tool for the refrigeration and air conditioning technician or for anyone checking the performance of a refrigera-tion or air conditioning system. Suction and discharge pressures can be readily measured by means of gauges. From these pressures, the saturated evaporat-ing and condensing temperatures can be determined. Knowing the saturated temperatures makes it easy for the technician to determine the amount of superheat or sub-cooling.

2-12

© 1968 Emerson Climate Technologies, Inc.All rights reserved.2-13

Table 2-4 R-134a

Saturation Properties – Temperature Table

TEMP. °F

PRESSURE VOLUME lb/ft3

DENSITY lb/ft3

ENTHALPY Btu/lb

ENTROPY Btu/(lb)(°R) TEMP.

°Fpsia LIQUID

Vf VAPOR

VgLIQUID

1/Vf VAPOR

1/Vg LIQUID

hf LATENT

hfg VAPOR

hgLIQUID

SfVAPOR

Sg

90 119.138 0.0136 0.3999 73.54 2.5009 41.6 73.6 115.2 0.0855 0.2194 90 91 121.024 0.0136 0.3935 73.40 2.5416 41.9 73.4 115.3 0.0861 0.2194 91 92 122.930 0.0137 0.3872 73.26 2.5829 42.3 73.1 115.4 0.0868 0.2193 92 93 124.858 0.0137 0.3810 73.12 2.6247 42.6 72.9 115.5 0.0874 0.2193 93 94 126.809 0.0137 0.3749 72.98 2.6672 43.0 72.7 115.7 0.0880 0.2193 94 95 128.782 0.0137 0.3690 72.84 2.7102 43.4 72.4 115.8 0.0886 0.2192 95 96 130.778 0.0138 0.3631 72.70 2.7539 43.7 72.2 115.9 0.0893 0.2192 96 97 132.798 0.0138 0.3574 72.56 2.7981 44.1 71.9 116.0 0.0899 0.2191 97 98 134.840 0.0138 0.3517 72.42 2.8430 44.4 71.7 116.1 0.0905 0.2191 98 99 136.906 0.0138 0.3462 72.27 2.8885 44.8 71.4 116.2 0.0912 0.2190 99

100 138.996 0.0139 0.3408 72.13 2.9347 45.1 71.2 116.3 0.0918 0.2190 100 101 141.109 0.0139 0.3354 71.99 2.9815 45.5 70.9 116.4 0.0924 0.2190 101 102 143.247 0.0139 0.3302 71.84 3.0289 45.8 70.7 116.5 0.0930 0.2189 102 103 145.408 0.0139 0.3250 71.70 3.0771 46.2 70.4 116.6 0.0937 0.2189 103 104 147.594 0.0140 0.3199 71.55 3.1259 46.6 70.2 116.7 0.0943 0.2188 104 105 149.804 0.0140 0.3149 71.40 3.1754 46.9 69.9 116.9 0.0949 0.2188 105 106 152.039 0.0140 0.3100 71.25 3.2256 47.3 69.7 117.0 0.0955 0.2187 106 107 154.298 0.0141 0.3052 71.11 3.2765 47.6 69.4 117.1 0.0962 0.2187 107 108 156.583 0.0141 0.3005 70.96 3.3282 48.0 69.2 117.2 0.0968 0.2186 108 109 158.893 0.0141 0.2958 70.81 3.3806 48.4 68.9 117.3 0.0974 0.2186 109 110 161.227 0.0142 0.2912 70.66 3.4337 48.7 68.6 117.4 0.0981 0.2185 110 111 163.588 0.0142 0.2867 70.51 3.4876 49.1 68.4 117.5 0.0987 0.2185 111 112 165.974 0.0142 0.2823 70.35 3.5423 49.5 68.1 117.6 0.0993 0.2185 112 113 168.393 0.0142 0.2780 70.20 3.5977 49.8 67.8 117.7 0.0999 0.2184 113 114 170.833 0.0143 0.2737 70.05 3.6539 50.2 67.6 117.8 0.1006 0.2184 114 115 173.298 0.0143 0.2695 69.89 3.7110 50.5 67.3 117.9 0.1012 0.2183 115 116 175.790 0.0143 0.2653 69.74 3.7689 50.9 67.0 117.9 0.1018 0.2183 116 117 178.297 0.0144 0.2613 69.58 3.8276 51.3 66.8 118.0 0.1024 0.2182 117 118 180.846 0.0144 0.2573 69.42 3.8872 51.7 66.5 118.1 0.1031 0.2182 118 119 183.421 0.0144 0.2533 69.26 3.9476 52.0 66.2 118.2 0.1037 0.2181 119 120 186.023 0.0145 0.2494 69.10 4.0089 52.4 65.9 118.3 0.1043 0.2181 120 121 188.652 0.0145 0.2456 68.94 4.0712 52.8 65.6 118.4 0.1050 0.2180 121 122 191.308 0.0145 0.2419 68.78 4.1343 53.1 65.4 118.5 0.1056 0.2180 122 123 193.992 0.0146 0.2382 68.62 4.1984 53.5 65.1 118.6 0.1062 0.2179 123 124 196.703 0.0146 0.2346 68.46 4.2634 53.9 64.8 118.7 0.1068 0.2178 124 125 199.443 0.0146 0.2310 68.29 4.3294 54.3 64.5 118.8 0.1075 0.2178 125 126 202.211 0.0147 0.2275 68.13 4.3964 54.6 64.2 118.8 0.1081 0.2177 126 127 205.008 0.0147 0.2240 67.96 4.4644 55.0 63.9 118.9 0.1087 0.2177 127 128 207.834 0.0147 0.2206 67.80 4.5334 55.4 63.6 119.0 0.1094 0.2176 128 129 210.688 0.0148 0.2172 67.63 4.6034 55.8 63.3 119.1 0.1100 0.2176 129 130 213.572 0.0148 0.2139 67.46 4.6745 56.2 63.0 119.2 0.1106 0.2175 130 131 216.485 0.0149 0.2107 67.29 4.7467 56.5 62.7 119.2 0.1113 0.2174 131 132 219.429 0.0149 0.2075 67.12 4.8200 56.9 62.4 119.3 0.1119 0.2174 132 133 222.402 0.0149 0.2043 66.95 4.8945 57.3 62.1 119.4 0.1125 0.2173 133 134 225.405 0.0150 0.2012 66.77 4.9700 57.7 61.8 119.5 0.1132 0.2173 134 135 228.438 0.0150 0.1981 66.60 5.0468 58.1 61.5 119.6 0.1138 0.2172 135 136 231.502 0.0151 0.1951 66.42 5.1248 58.5 61.2 119.6 0.1144 0.2171 136 137 234.597 0.0151 0.1922 66.24 5.2040 58.8 60.8 119.7 0.1151 0.2171 137 138 237.723 0.0151 0.1892 66.06 5.2844 59.2 60.5 119.8 0.1157 0.2170 138 139 240.880 0.0152 0.1864 65.88 5.3661 59.6 60.2 119.8 0.1163 0.2169 139 140 244.068 0.0152 0.1835 65.70 5.4491 60.0 59.9 119.9 0.1170 0.2168 140 141 247.288 0.0153 0.1807 65.52 5.5335 60.4 59.6 120.0 0.1176 0.2168 141 142 250.540 0.0153 0.1780 65.34 5.6192 60.8 59.2 120.0 0.1183 0.2167 142 143 253.824 0.0153 0.1752 65.15 5.7064 61.2 58.9 120.1 0.1189 0.2166 143 144 257.140 0.0154 0.1726 64.96 5.7949 61.6 58.6 120.1 0.1195 0.2165 144 145 260.489 0.0154 0.1699 64.78 5.8849 62.0 58.2 120.2 0.1202 0.2165 145 146 263.871 0.0155 0.1673 64.59 5.9765 62.4 57.9 120.3 0.1208 0.2164 146 147 267.270 0.0155 0.1648 64.39 6.0695 62.8 57.5 120.3 0.1215 0.2163 147 148 270.721 0.0156 0.1622 64.20 6.1642 63.2 57.2 120.4 0.1221 0.2162 148 149 274.204 0.0156 0.1597 64.01 6.2604 63.6 56.8 120.4 0.1228 0.2161 149

Reprinted with permission from E.I. DuPont

© 1968 Emerson Climate Technologies, Inc.All rights reserved. 2-14

Table 2-5 "FREON" 22

Saturation Properties – Temperature Table

TEMP. °F

PRESSURE VOLUME cu ft/lb

DENSITY lb/cu ft

ENTHALPY Btu/lb

ENTROPY Btu/(lb)(°R) TEMP.

°Fpsia psig LIQUID

Vf VAPOR

VgLIQUID

1/Vf VAPOR

1/Vg LIQUID

hf LATENT

hfg VAPOR

hgLIQUID

SfVAPOR

Sg

10 47.464 32.768 0.012088 1.1290 82.724 0.88571 13.104 92.338 105.442 0.02932 0.22592 10 11 48.423 33.727 0.012105 1.1077 82.612 0.90275 13.376 92.162 105.538 0.02990 0.22570 11 12 49.396 34.700 0.012121 1.0869 82.501 0.92005 13.648 91.986 105.633 0.03047 0.22548 12 13 50.384 35.688 0.012138 1.0665 82.389 0.93761 13.920 91.808 105.728 0.03104 0.22527 13 14 51.387 36.691 0.012154 1.0466 82.276 0.95544 14.193 91.630 105.823 0.03161 0.22505 14

15 52.405 37.709 0.012171 1.0272 82.164 0.97352 14.466 91.451 105.917 0.03218 0.22484 15 16 53.438 38.742 0.012188 1.0082 82.051 0.99188 14.739 91.272 106.011 0.03275 0.22463 16 17 54.487 39.791 0.012204 0.98961 81.938 1.0105 15.013 91.091 106.105 0.03332 0.22442 17 18 55.551 40.855 0.012221 0.97144 81.825 1.0294 15.288 90.910 106.198 0.03389 0.22421 18 19 56.631 41.935 0.012238 0.95368 81.711 1.0486 15.562 90.728 106.290 0.03446 0.22400 19

20 57.727 43.031 0.012255 0.93631 81.597 1.0680 15.837 90.545 106.383 0.03503 0.22379 20 21 58.839 44.143 0.012273 0.91932 81.483 1.0878 16.113 90.362 106.475 0.03560 0.22358 21 22 59.967 45.271 0.012290 0.90270 81.368 1.1078 16.389 90.178 106.566 0.03617 0.22338 22 23 61.111 46.415 0.012307 0.88645 81.253 1.1281 16.665 89.993 106.657 0.03674 0.22318 23 24 62.272 47.576 0.012325 0.87055 81.138 1.1487 16.942 89.807 106.748 0.03730 0.22297 24

25 63.450 48.754 0.012342 0.85500 81.023 1.1696 17.219 89.620 106.839 0.03787 0.22277 25 26 64.644 49.948 0.012360 0.83978 80.907 1.1908 17.496 89.433 106.928 0.03844 0.22257 26 27 65.855 51.159 0.012378 0.82488 80.791 1.2123 17.774 89.244 107.018 0.03900 0.22237 27 28 67.083 52.387 0.012395 0.81031 80.675 1.2341 18.052 89.055 107.107 0.03958 0.22217 28 29 68.328 53.632 0.012413 0.79604 80.558 1.2562 18.330 88.865 107.196 0.04013 0.22198 29

30 69.591 54.895 0.012431 0.78208 80.441 1.2786 18.609 88.674 107.284 0.04070 0.22178 30 31 70.871 56.175 0.012450 0.76842 80.324 1.3014 18.889 88.483 107.372 0.04126 0.22158 31 32 72.169 57.473 0.012468 0.75503 80.207 1.3244 19.169 88.290 107.459 0.04182 0.22139 32 33 73.485 58.789 0.012486 0.74194 80.089 1.3478 19.449 88.097 107.546 0.04239 0.22119 33 34 74.818 60.122 0.012505 0.72911 79.971 1.3715 19.729 87.903 107.632 0.04295 0.22100 34

35 76.170 61.474 0.012523 0.71655 79.852 1.3956 20.010 87.708 107.719 0.04351 0.22081 35 36 77.540 62.844 0.012542 0.70425 79.733 1.4199 20.292 87.512 107.804 0.04407 0.22062 36 37 78.929 64.233 0.012561 0.69221 79.614 1.4447 20.574 87.316 107.889 0.04464 0.22043 37 38 80.336 65.640 0.012579 0.68041 79.495 1.4697 20.856 87.118 107.974 0.04520 0.22024 38 39 81.761 67.065 0.012598 0.66885 79.375 1.4951 21.138 86.920 108.058 0.04576 0.22005 39

40 83.206 68.510 0.012618 0.65753 79.255 1.5208 21.422 86.720 108.142 0.04632 0.21986 40 41 84.670 69.974 0.012637 0.64643 79.134 1.5469 21.705 86.520 108.225 0.04688 0.21968 41 42 86.153 71.457 0.012656 0.63557 79.013 1.5734 21.989 86.319 108.308 0.04744 0.21949 42 43 87.655 72.959 0.012676 0.62492 78.892 1.6002 22.273 86.117 108.390 0.04800 0.21931 43 44 89.177 74.481 0.012695 0.61448 78.770 1.6274 22.558 85.914 108.472 0.04855 0.21912 44

45 90.719 76.023 0.012715 0.60425 78.648 1.6549 22.843 85.710 108.553 0.04911 0.21894 45 46 92.280 77.584 0.012735 0.59422 78.526 1.6829 23.129 85.506 108.634 0.04967 0.21876 46 47 93.861 79.165 0.012755 0.58440 78.403 1.7112 23.415 85.300 108.715 0.05023 0.21858 47 48 95.463 80.767 0.012775 0.57476 78.280 1.7398 23.701 85.094 108.795 0.05079 0.21839 48 49 97.085 82.389 0.012795 0.56532 78.157 1.7689 23.988 84.886 108.874 0.05134 0.21821 49

50 98.727 84.031 0.012815 0.55606 78.033 1.7984 24.275 84.678 108.953 0.05190 0.21803 50 51 100.39 85.69 0.012836 0.54698 77.909 1.8282 24.563 84.468 109.031 0.05245 0.21785 51 52 102.07 87.38 0.012856 0.53808 77.784 1.8585 24.851 84.258 109.109 0.05301 0.21768 52 53 103.78 89.08 0.012877 0.52934 77.659 1.8891 25.139 84.047 109.186 0.05357 0.21750 53 54 105.50 90.81 0.012898 0.52078 77.534 1.9202 25.429 83.834 109.263 0.05412 0.21732 54

55 107.25 92.56 0.012919 0.51238 77.408 1.9517 25.718 83.621 109.339 0.05468 0.21714 55 56 109.02 94.32 0.012940 0.50414 77.282 1.9836 26.008 83.407 109.415 0.05523 0.21697 56 57 110.81 96.11 0.012961 0.49606 77.155 2.0159 26.298 83.191 109.490 0.05579 0.21679 57 58 112.62 97.93 0.012982 0.48813 77.028 2.0486 26.589 82.975 109.564 0.05634 0.21662 58 59 114.46 99.76 0.013004 0.48035 76.900 2.0818 26.880 82.758 109.638 0.05689 0.21644 59

60 116.31 101.62 0.013025 0.47272 76.773 2.1154 27.172 82.540 109.712 0.05745 0.21627 60 61 118.19 103.49 0.013047 0.46523 76.644 2.1495 27.464 82.320 109.785 0.05800 0.21610 61 62 120.09 105.39 0.013069 0.45788 76.515 2.1840 27.757 82.100 109.857 0.05855 0.21592 62 63 122.01 107.32 0.013091 0.45066 76.386 2.2190 28.050 81.878 109.929 0.05910 0.21575 63 64 123.96 109.26 0.013114 0.44358 76.257 2.2544 28.344 81.656 110.000 0.05966 0.21558 64



Table 2-6 R-404A

SUVA® HP62 Superheated Vapor – Constant Pressure TablesV = Volume in ft3/lb H = Enthalpy in Btu/lb S = Entropy in Btu/(lb)(°R) (Saturation properties in parentheses)

ABSOLUTE PRESSURE, psia

TEMP. °F

16.00 (-47.11ºF)

17.00 (-44.78ºF)

18.00 (-42.57ºF)

19.00 (-40.44ºF) TEMP.

°FV H S V H S V H S V H S(2.7271) (84.2) (0.2045) (2.5763) (84.5) (0.2042) (2.4416) (84.9) (0.2039) (2.3206) (85.2) (0.2036)

-40 2.7793 85.5 0.2076 2.6094 85.4 0.2063 2.4584 85.3 0.2050 2.3233 85.3 0.2038 -40-30 2.8524 87.3 0.2119 2.6786 87.2 0.2106 2.5240 87.2 0.2093 2.3857 87.1 0.2081 -30-20 2.9253 89.1 0.2161 2.7474 89.1 0.2148 2.5893 89.0 0.2136 2.4478 89.0 0.2124 -20-10 2.9979 91.0 0.2203 2.8160 90.9 0.2190 2.6543 90.9 0.2178 2.5096 90.9 0.2166 -100 3.0703 92.9 0.2245 2.8843 92.8 0.2232 2.7191 92.8 0.2220 2.5712 92.8 0.2208 0

10 3.1425 94.8 0.2285 2.9525 94.8 0.2273 2.7837 94.7 0.2261 2.6326 94.7 0.2250 1020 3.2145 95.8 0.2327 3.0205 96.7 0.2315 2.8481 96.7 0.2302 2.6938 96.6 0.2291 2030 3.2864 98.7 0.2368 3.0883 98.7 0.2355 2.9123 98.6 0.2343 2.7548 98.6 0.2331 3040 3.3581 100.7 0.2408 3.1560 100.7 0.2396 2.9764 100.6 0.2383 2.8157 100.6 0.2372 4050 3.4296 102.7 0.2448 3.2235 102.7 0.2436 3.0403 102.7 0.2423 2.8764 102.6 0.2412 5060 3.5010 104.8 0.2488 3.2909 104.8 0.2475 3.1041 104.7 0.2463 2.9370 104.7 0.2452 6070 3.5723 106.9 0.2528 3.3582 106.8 0.2515 3.1678 106.8 0.2503 2.9975 106.7 0.2491 7080 3.6435 109.0 0.2567 3.4253 108.9 0.2554 3.2313 108.9 0.2542 3.0578 108.8 0.2530 8090 3.7146 111.1 0.2606 3.4923 111.0 0.2593 3.2948 111.0 0.2581 3.1180 111.0 0.2569 90

100 3.7855 113.2 0.2644 3.5593 113.2 0.2631 3.3581 113.1 0.2619 3.1782 113.1 0.2608 100110 3.8564 115.4 0.2682 3.6261 115.3 0.2670 3.4214 115.3 0.2658 3.2382 115.3 0.2646 110120 3.9272 117.6 0.2721 3.6928 117.5 0.2708 3.4845 117.5 0.2696 3.2981 117.5 0.2684 120130 3.9978 119.8 0.2758 3.7595 119.7 0.2746 3.5476 119.7 0.2734 3.3580 119.7 0.2722 130140 4.0685 122.0 0.2796 3.8260 122.0 0.2783 3.6105 121.9 0.2771 3.4177 121.9 0.2760 140150 4.1390 124.3 0.2833 3.8925 124.2 0.2821 3.6734 124.2 0.2809 3.4774 124.2 0.2797 150160 4.2094 126.5 0.2870 3.9589 126.5 0.2858 3.7363 126.5 0.2846 3.5370 126.4 0.2834 160170 4.2798 128.8 0.2907 4.0253 128.8 0.2894 3.7990 128.8 0.2883 3.5966 128.8 0.2871 170180 4.3502 131.2 0.2944 4.0916 131.1 0.2931 3.8617 131.1 0.2919 3.6561 131.1 0.2908 180190 4.4204 133.5 0.2980 4.1578 133.5 0.2967 3.9244 133.5 0.2956 3.7155 133.4 0.2944 190200 4.4906 135.9 0.3016 4.2240 135.8 0.3004 3.9869 135.8 0.2992 3.7748 135.8 0.2980 200210 4.5608 138.3 0.3052 4.2901 138.2 0.3040 4.0495 138.2 0.3028 3.8342 138.2 0.3016 210220 4.6309 140.7 0.3088 4.3561 140.6 0.3075 4.1119 140.6 0.3063 3.8934 140.6 0.3052 220230 4.7009 143.1 0.3123 4.4221 143.1 0.3111 4.1743 143.1 0.3099 3.9526 143.0 0.3088 230240 4.7709 145.6 0.3159 4.4881 145.5 0.3146 4.2367 145.5 0.3134 4.0118 145.5 0.3123 240250 4.8409 148.0 0.3194 4.5540 148.0 0.3181 4.2990 148.0 0.3169 4.0709 148.0 0.3158 250260 4.9108 150.5 0.3229 4.6199 150.5 0.3216 4.3613 150.5 0.3204 4.1299 150.5 0.3193 260

TEMP. °F

20.00 (-38.40ºF)

21.00 (-36.44ºF)

22.00 (-34.55ºF)

23.00 (-32.73ºF) TEMP.

°FV H S V H S V H S V H S(2.2112) (85.5) (0.2034) (2.1119) (85.8) (0.2032) (2.0213) (86.1) (0.2029) (1.9383) (86.4) (0.2027)

-30 2.2612 87.1 0.2070 2.1485 87.0 0.2059 2.0461 87.0 0.2049 1.9525 86.9 0.2039 -30-20 2.3204 88.9 0.2113 2.2051 88.9 0.2102 2.1003 88.8 0.2092 2.0046 88.8 0.2082 -20-10 2.3793 90.8 0.2155 2.2615 90.8 0.2145 2.1543 90.7 0.2134 2.0565 90.7 0.2125 -100 2.4380 92.7 0.2197 2.3176 92.7 0.2186 2.2081 92.6 0.2176 2.1081 92.6 0.2167 0

10 2.4966 94.6 0.2239 2.3735 94.6 0.2228 2.2616 94.5 0.2218 2.1594 94.5 0.2208 1020 2.5549 96.6 0.2280 2.4292 96.5 0.2269 2.3150 96.5 0.2259 2.2106 96.5 0.2249 2030 2.6130 98.6 0.2320 2.4848 98.5 0.2310 2.3681 98.5 0.2300 2.2617 98.4 0.2290 3040 2.6710 100.6 0.2361 2.5401 100.5 0.2350 2.4212 100.5 0.2340 2.3125 100.4 0.2331 4050 2.7289 102.6 0.2401 2.5954 102.5 0.2391 2.4740 102.5 0.2381 2.3632 102.5 0.2371 5060 2.7866 104.6 0.2441 2.6505 104.6 0.2430 2.5267 104.6 0.2420 2.4138 104.5 0.2411 6070 2.8442 106.7 0.2480 2.7054 106.7 0.2470 2.5793 106.6 0.2460 2.4642 106.6 0.2450 7080 2.9016 108.8 0.2520 2.7603 108.8 0.2509 2.6318 108.7 0.2499 2.5145 108.7 0.2490 8090 2.9590 110.9 0.2558 2.8150 110.9 0.2548 2.6842 110.9 0.2538 2.5647 110.8 0.2529 90

100 3.0162 113.1 0.2597 2.8696 113.0 0.2587 2.7364 113.0 0.2577 2.6148 113.0 0.2567 100110 3.0733 115.2 0.2635 2.9242 115.2 0.2625 2.7886 115.2 0.2615 2.6647 115.1 0.2606 110120 3.1304 117.4 0.2674 2.9786 117.4 0.2663 2.8406 117.4 0.2653 2.7146 117.3 0.2644 120130 3.1873 119.6 0.2711 3.0329 119.6 0.2701 2.8926 119.6 0.2691 2.7644 119.5 0.2682 130140 3.2442 121.9 0.2749 3.0872 121.8 0.2739 2.9445 121.8 0.2729 2.8141 121.8 0.2720 140150 3.3010 124.1 0.2786 3.1414 124.1 0.2776 2.9963 124.1 0.2766 2.8638 124.0 0.2757 150160 3.3577 126.4 0.2824 3.1955 126.4 0.2813 3.0480 126.4 0.2804 2.9133 126.3 0.2794 160170 3.4144 128.7 0.2860 3.2495 128.7 0.2850 3.0997 128.7 0.2840 2.9628 128.6 0.2831 170180 3.4710 131.1 0.2897 3.3035 131.0 0.2887 3.1513 131.0 0.2877 3.0123 131.0 0.2868 180190 3.5275 133.4 0.2934 3.3574 133.4 0.2923 3.2028 133.3 0.2914 3.0616 133.3 0.2904 190200 3.5840 135.8 0.2970 3.4113 135.7 0.2960 3.2543 135.7 0.2950 3.1109 135.7 0.2940 200210 3.6404 138.2 0.3006 3.4651 138.1 0.2996 3.3057 138.1 0.2986 3.1602 138.1 0.2976 210220 3.6967 140.6 0.3041 3.5188 140.5 0.3031 3.3571 140.5 0.3022 3.2094 140.5 0.3012 220230 3.7531 143.0 0.3077 3.5725 143.0 0.3067 3.4084 143.0 0.3057 3.2585 142.9 0.3048 230240 3.8093 145.5 0.3112 3.6262 145.4 0.3102 3.4596 145.4 0.3092 3.3076 145.4 0.3083 240250 3.8655 147.9 0.3147 3.6798 147.9 0.3137 3.5109 147.9 0.3128 3.3567 147.9 0.3118 250260 3.9217 150.4 0.3182 3.7333 150.4 0.3172 3.5620 150.4 0.3163 3.4057 150.4 0.3153 260270 3.9779 152.9 0.3217 3.7868 152.9 0.3207 3.6132 152.9 0.3197 3.4546 152.9 0.3188 270



R-507 Pressure/Temperature Chart

Figure 2-7

Vapor Vapor Pressure Temperature PSIG °F




0 -52.1 1 -49.7 2 -47.3 3 -46 4 -43 5 -41 6 -39 7 -37 8 -35.3 9 -33.5

10 -31.8 11 -30.2 12 -28.6 13 -27 14 -25.5

15 -24 16 -22.6 17 -21.2 18 -19.8 19 -18.4

20 -17.1 21 -15.8 22 -14.6 23 -13.4 24 -12.2

25 -11 26 -9.8 27 -8.7 28 -7.5 29 -6.4

30 -5.4 31 -4.3 32 -3.3 33 -2.2 34 -1.2

35 -0.2 36 0.8 37 1.7 38 2.7 39 3.6 40 4.6 41 5.5 42 6.4 43 7.3 44 8.2

45 9 46 9.9 47 10.7 48 11.6 49 12.3

50 13.1 51 14 52 14.8 53 15.6 54 16.4

55 17.2 56 17.9 57 18.7 58 19.4 59 20.2

60 20.9 61 21.6 62 22.3 63 23.1 64 23.8

65 24.5 66 25.2 67 25.8 68 26.5 69 27.2

70 27.9 71 28.5 72 29.2 73 29.8 74 30.5 75 31.5 76 31.7 77 32.4 78 33 79 33.6

80 34.2 81 34.8 82 35.4 83 36 84 36.6

85 37.2 86 37.8 87 38.4 88 39 89 39.5

90 40.1 91 40.7 92 41.2 93 41.8 94 42.4

95 42.9 96 43.4 97 44 98 44.5 99 45.1

100 45.6 101 46.1 102 46.7 103 47.2 104 47.7

105 48.2 106 48.7 107 49.2 108 49.7 109 50.2

110 50.7 115 53.2 120 55.6 125 57.9 130 60.1

135 62.3 140 64.4 145 66.5 150 68.6 155 70.6 160 72.5 165 74.4 170 76.3 175 78.1 180 79.9

185 81.7 190 83.4 195 85.1 200 86.7 205 88.3

210 89.9 215 91.5 220 93.1 225 94.6 230 96.1

235 97.5 240 99 245 100.4 250 101.8 255 103.2

260 104.6 265 105.9 270 107.2 275 108.6 280 109.8

285 111.1 290 112.4 295 113.6 300 114.8 305 116

310 117.2 315 118.4 320 119.6 325 120.7 330 121.8

335 123 340 124.1 345 125.2 350 126.3 355 127.3

360 128.4 365 129.4 370 130.5 375 131.5 380 132.5

385 133.5 390 134.5 395 135.5 400 136.4

2-16


Typical Pocket Pressure/Temperature Chart Table 2-8

2-17


Continuous refrigeration can be accomplished by sev-eral different processes. In the great majority of applica-tions, and almost exclusively in the smaller horsepower range, the vapor compression system is used for the refrigeration process. However, absorption systems are being successfully used in many applications. In larger equipment, centrifugal systems are used which basically is an adaptation of the compression cycle.

Copeland® brand compressors, as their name implies, are designed for use with the compression cycle. This section of this manual will cover only that form of re-frigeration.

SIMPLE COMPRESSION REFRIGERATION CYCLE

There are two pressures existing in a compression system, the evaporating or low pressure, and the con-densing or high pressure.

The refrigerant acts as a transportation medium to move heat absorbed in the evaporator to the condenser where

it is rejected. The heat rejected may be given off to the ambient air, or in a water cooled system, to the cooling water. A change of state from liquid to vapor and back to liquid allows the refrigerant to absorb and reject large quantities of heat efficiently and repeatedly.

The basic cycle operates as follows:

High pressure liquid refrigerant is fed from the receiver or condenser through the liquid line, and through the filter-drier to the metering device. It is at this point that the high pressure side of the system is separated from the low pressure side. Various types of control devices may be used, but for purposes of this illustration, only the thermostatic expansion valve (TEV) will be con-sidered.

The TEV controls the quantity of liquid refrigerant be-ing fed into the evaporator. The TEV’s internal orifice causes the pressure of the refrigerant to the evaporat-ing or low side pressure to be reduced. This reduction of the refrigerant pressure, therefore its boiling point,

Section 3THE REFRIGERATION CYCLE

TYPICAL COMPRESSION REFRIGERATION SYSTEM Figure 3-1

Condenser

ReceiverFilter Dryer

Thermostatic Expansion Valve

Evaporator

Compressor

3-1


causes it to boil or vaporize, absorbing heat until the refrigerant is at the saturation temperature correspond-ing to its pressure. As the low temperature refrigerant passes through the evaporator coil, heat flows through the walls of the evaporator tubing into the refrigerant. The boiling action continues until the refrigerant is completely vaporized.

The TEV regulates the quantity of refrigerant, (lb/min) through the evaporator to maintain a preset tempera-ture difference or superheat between the evaporating refrigerant and the vapor leaving the evaporator. As the temperature of the gas leaving the evaporator varies, the expansion valve power element bulb senses this temperature, and acts to modulate the feed of refriger-ant through the expansion valve.

The superheated refrigerant vapor leaving the evapora-tor travels through the suction line to the compressor inlet. The compressor takes the low pressure vapor and compresses it, increasing it’s pressure and temperature. The hot, high pressure vapor is forced out of the com-pressor discharge valve(s), and into the condenser.

As the high pressure high temperature vapor passes through the condenser, it is cooled by an external means. In air cooled systems, a fan, and fin-type condenser surface is normally used. In water cooled systems, a refrigerant-to-water heat exchanger is employed. As the temperature of the refrigerant vapor is lowered to the saturation temperature corresponding to the high pres-sure in the condenser, the vapor condenses into a liquid and flows back to the receiver or directly to the TEV to repeat the cycle. The refrigerating process is continuous as long as the compressor is operating.

HEAT OF COMPRESSION

Heat of compression is defined as the heat added to the refrigerant vapor as a result of the work energy used in the compression process. When the refrigerant vapor is compressed in a compressors cylinder, its pressure is increased and the volume is decreased. The change in pressure and volume tend to maintain equilibrium in the perfect gas law equation, so this change alone would not greatly affect the temperature of the refriger-ant vapor. In order to compress the refrigerant vapor, work or energy is required. Following the first law of thermodynamics, this energy cannot be destroyed, and all of the mechanical energy necessary to compress the vapor is transformed into heat energy. With the excep-tion of a small fraction of the total heat given off to the compressor body, all of this heat energy is transferred to the refrigerant vapor. This causes a sharp increase in the temperature of the compressed gas, therefore, in a reciprocating compressor, the discharge valves are

always subjected to the highest temperature existing in the refrigerating system. In the Copeland Compli-ant Scroll®, the discharge port or dynamic discharge valve will be subjected to the highest temperature in the system.

The heat which must be discharged by the condenser, termed the heat of rejection, is the total of the heat absorbed by the refrigerant in the evaporator, the heat of compression, and any heat added to the system due to motor inefficiency. Any heat absorbed in the suction and/or discharge lines must also be rejected by the condenser. For hermetic and accessible hermetic mo-tor-compressors, the heat which must be rejected in addition to the refrigeration load can be approximated by the heat equivalent of the electrical power input to the compressor expressed in BTU/hr. (Motor watts X 3.1416 = BTU/hr. of heat to be rejected)

VOLUMETRIC EFFICIENCY OF THE RECIPROCATING COMPRESSOR

Volumetric efficiency is defined as the ratio of the actual volume of refrigerant vapor pumped by the compressor to the volume displaced by the compressor pistons. The volumetric efficiency of a piston compressor will vary over a wide range, depending on the compressor design and the compression ratio.

The compression ratio of a compressor is the ratio of the absolute discharge pressure (psia) to the absolute suction pressure (psia). (Discharge Pressure Absolute ÷ Suction Pressure Absolute)

Several design factors can influence compressor effi-ciency including the clearance volume above the piston, the clearance between the piston and the cylinder wall, valve spring tension, valve leakage and the volume of the valve plate discharge ports. Reed compressors have from one to three discharge ports per cylinder to allow the compressors refrigerant to exit the cylinder/piston area with a minimum pressure drop. These discharge ports however hold high pressure vapor in them that cannot be sent out to the system. (See Figure 3-2) To improve the volumetric efficiency of low temperature compressors, the number and size of the discharge ports are reduced. (Figure 3-3) Because the volume of refrigerant is less in a low temperature compressor, this can be done with little effect on internal pressure drop but with positive results in increased volumetric efficiency and compressor capacity. Compressor effi-ciency, because of design, is fairly constant for a given compressor. Volumetric efficiency will vary inversely with the compression ratio.

Two factors cause a loss of volumetric efficiency with an increase in compression ratio. As the vapor is subjected

3-2


to greater compression, the residual vapor remaining in the cylinder clearance space and in the valve plate discharge ports becomes more dense. Since it does not leave the cylinder on the discharge stroke, it re-ex-pands on the suction stroke, preventing the intake of a full cylinder of vapor from the suction line. The higher the pressure exerted on the residual vapor, the more dense it becomes, and the greater volume it occupies on re-expansion.

Discus® compressors have less clearance volume and almost no trapped high pressure refrigerant in the discharge ports. The Discus® discharge valve seats at the bottom of the valve plate basically eliminating the trapped high pressure vapor in the valve plate. This reduction in trapped high pressure refrigerant reduces the amount of re-expansion and increases the com-pressors capacity and efficiency (See Figure 3-4). The

Discus® compressor is more volumetric efficient than the same displacement reed compressor and as such circulates more pounds of refrigerant therefore delivers more BTUs of refrigeration.

The second factor is the high temperature of the cylinder walls resulting from the heat of compression. As the compression ratio increases, the heat of compression increases, and the cylinders and head of the compressor become very hot. Suction vapor entering the cylinder on the intake stroke is heated by the cylinder walls, and expands, resulting in a reduced weight of vapor entering the compressor.

Typical volumetric efficiency curves are shown in Figure 3-5. Air Conditioning and refrigeration compressors are designed with a minimum of clearance volume. As previously stated, clearance volume is a loss in ac-tual capacity versus theoretical capacity. The Discus®

Figure 3-3Figure 3-2

High Temperature Valve Plate

Low Temperature Valve Plate

3-3

Figure 3-5

TYPICAL COMPRESSOR VOLUMETRIC EFFICIENCY CURVES

Figure 3-4

Typical lowtemperatureDiscus® compressor


compressor because of its reduced clearance volume is more volumetric efficient than an equivalent displace-ment, horse power, reed compressor

While the volumetric efficiency of each stage of a two stage compressor would resemble the typical single stage curves, the overall volumetric efficiency is rela-tively constant over a wide compression ratio range. Since the use of a liquid subcooler with the two stage compressor can increase the capacity so dramatically. a dotted curve has been added for comparison.

VOLUMETRIC EFFICIENCY OF SCROLL COMPRESSORS

The volumetric efficiency of the Copeland Compliant Scroll® is 100%. When the first pocket of the Scroll closes and captures a volume of refrigerant, all of it will be swept along in the compression process and discharged out of the Scroll members to the system. Unlike the piston compressor, there is no clearance area, clearance volume, to create losses.

EFFECT OF CHANGE IN SUCTION PRESSURE

Other factors remaining equal, as the suction pres-sure is reduced, the specific volume(ft3/#) of the vapor returning to the compressor increases. Density (#/ft3) and specific volume are inversely proportional, there-fore the refrigerant density decreases. Since a given compressor's pumping capacity (CFH) is fixed by its speed and displacement, the reduction in density of the suction vapor decreases the weight (#/hr.) of the refriger-ant pumped, resulting in a reduction in the compressors capacity (BTU/hr). The loss of capacity with a reduction in suction pressure is extremely rapid. Since the energy input required by the compressor to perform its work does not decrease at the same rate, the BTU/watt ratio decreases rapidly with a drop in suction pressure. This reflects in the performance of the compressors per unit of electrical energy consumed, the Energy Efficiency Ratio (EER).

In addition to the specific volume of the refrigerant be-ing reduced when the suction pressure is reduced, the compression ratio is increased. As stated before, as the compression ratio is increased, the compressors discharge temperature will also be increased. For best capacity performance, operating economy and lowered discharge temperature, it is most important that refrig-eration and air-conditioning systems operate at the highest suction pressure possible for the application.

EFFECT OF CHANGE IN DISCHARGE PRESSURE

An increase in the condensing pressure, commonly termed the discharge pressure or head pressure, results

in an increase in the compression ratio. This results in a consequent loss of volumetric efficiency except for the Scroll compressor. While the loss of capacity is not as great as that caused by an equivalent decrease in suction pressure, it still is severe.

For operating economy and maximum capacity, the discharge pressure should be kept as low as practical but should not be lower than the equivalent of 70°F saturated discharge pressure.

EFFECT OF SUBCOOLING LIQUID REFRIGERANT WITH WATER OR AIR

When the hot high pressure liquid refrigerant is fed into the evaporator through the TXV, the refrigerants temperature must first be reduced to the evaporating temperature in the evaporator before it can start absorb-ing heat. This is accomplished by almost instantaneous boiling or “flashing” of a portion of the liquid into vapor. The latent heat of vaporization involved in the change of state absorbs heat from the remaining liquid refrigerant lowering its temperature.

The resulting flash gas will produce little to no further cooling. In effect the refrigerating capacity of the re-frigerant has been reduced by the heat absorbed in lowering the liquid temperature. If a portion of this heat could be extracted from the liquid prior to its entry into the evaporator, the effective capacity of the system will be increased. This happens because not as much liquid will flash off to cool the remaining liquid to its desired temperature.

This can be accomplished by subcooling the liquid re-frigerant after condensing by means of water or air. If condensing temperatures are relatively high, capacity increases of 5% to 15% are easily obtainable. Since no power is required other than that involved in moving the cooling medium, subcooling the liquid can result in substantial savings in operating cost.

EFFECT OF SUBCOOLING LIQUID REFRIGERANT BY SUPERHEATING THE VAPOR

A suction gas to liquid refrigerant heat exchanger is frequently used for the following reasons:

1. To subcool the liquid refrigerant sufficiently to offset any pressure drop that might occur in the liquid line; to compensate for any heat picked up in the liquid line preventing the formation of flash gas in the liquid line.

2. To provide a source of heat to evaporate any liquid refrigerant which might have flooded through the

3-4


EFFECT OF PRESSURE DROP IN LIQUID LINE

If the pressure of the liquid refrigerant falls below its saturation temperature, a portion of the liquid will flash into vapor. This will cool the liquid refrigerant to a new saturation temperature. This will occur in a liquid line if the pressure drops significantly because of friction or be-cause of vertical lift. If flashing occurs, the feed through the expansion valve will be erratic and inadequate for the evaporator demand.

Subcooling of the liquid refrigerant after condensing by an amount sufficient to offset the pressure drop will insure a solid column of liquid refrigerant at the inlet to the expansion valve. At 120°F saturated condens-ing temperature, 10°F of liquid subcooling will protect against flash gas forming in the liquid line for pressure drops up to those shown in Table 3-1. These are the maximum allowable that can be tolerated to prevent flashing of the refrigerant in the liquid line.

Table 3-1Liquid Line Pressure Drop

Refrig.Press Drop (psig)



R-12 21.3 R-22 34.5 R-502 33.9R-401A 25.9 R-

407C38.5 R-402A 41.1

R-401B 27.2 R-410A 57.7 R-408A 28.5R-134a 24.8 R-404A 39.4R-409A 25.4 R-507 41.3

All of the refrigerants listed in Table 3-1 are slightly heavier than water. A head of two feet of liquid refriger-ant is approximately equivalent to 1 psi. Therefore if a condenser or receiver in the basement of a building 20 feet tall is to supply liquid refrigerant to an evaporator on the roof, a pressure drop of approximately 10 psi for the vertical head will occur. This must be provided for in system design. (Refer to Section 1 - Pressure & Fluid Head.)

EFFECT OF PRESSURE DROP IN THE EVAPORATOR

Pressure drop occurring in the evaporator due to fric-tional resistance to flow results in the leaving evaporator pressure being less than the pressure of the refrigerant at the entrance of the evaporator. For a given load and coil, the required average refrigerant temperature is fixed. The greater the pressure drop, the greater the difference between the average evaporator refrigerant pressure and the leaving evaporator refrigerant pres-sure.

evaporator, thus preventing the return of liquid re-frigerant to the crankcase.

As pointed out in the previous section, subcooling the liquid refrigerant increases the refrigerating capacity per pound of the refrigerant circulated. In a perfectly insulated system with negligible heat transfer into the suction line outside the refrigerated space, a liquid to suction heat exchanger theoretically will increase sys-tem capacity slightly since the heat transferred from the liquid refrigerant to the refrigerant vapor is greater than the capacity reduction at the compressor resulting from the increase in specific volume of the vapor.

EFFECT OF SUPERHEATING THE VAPOR LEAVING THE EVAPORATOR

It is essential that the temperature of the vapor return-ing to the compressor be superheated to avoid carrying liquid refrigerant back to the compressor. It is gener-ally recommended that the minimum superheat value be 20°F when the system is at low load. If this heat is added to the vapor inside the refrigerated space, the heat absorbed increases the refrigeration capacity, while the increase in specific volume of the gas decreases the compressor capacity. These two factors tend to offset one another, with a negligible effect on capacity.

Heat entering the refrigerant through the suction line from the ambient air outside the refrigerated space results in a net loss of system capacity. These losses may be as high as 10% to 15%. Insulation of the suction line is a worthwhile investment, and may be necessary to prevent the return gas temperature from rising too high. This will also prevent the compressors discharge temperature from rising too high.

EFFECT OF PRESSURE DROP IN THE DISCHARGE LINE AND CONDENSER

Pressure drop due to friction as the refrigerant vapor flows through the discharge line and condenser reduces compressor capacity. This results in higher compressor discharge pressure and lower volumetric efficiency. Since the condensing temperature is not greatly af-fected, pressure drops of less than 5 psig have very little effect on system capacity.

However, compressor power consumption will increase because of the higher compressor discharge pressure. For best operating economy, excessively high pres-sure drops in the discharge line should be avoided. A pressure drop in the discharge line between five and ten psig. should be considered normal. Pressure drops over ten psig. should be avoided.

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As the suction pressure leaving the evaporator is de-creased, the specific volume of the gas returning to the compressor increases, and the weight of the refriger-ant pumped by the compressor decreases. Therefore pressure drop in the evaporator causes a decrease in system capacity, and it is important that the evaporator be sized so that abnormally high pressure drops do not occur.

EFFECT OF PRESSURE DROP IN SUCTION LINE

The effect of pressure loss in the suction line is similar to pressure drop in the evaporator. Since pressure drop in the suction line does not result in a corresponding decrease in the refrigerant evaporating temperature in the evaporator. Pressure drop in the suction line can be extremely detrimental to system capacity. Suction lines must be sized to prevent excessive pressure losses.

Table 3-2 shows the capacity loss for a typical 7-1/2 HP compressor as the result of suction line pressure drop. The table lists the losses for both R-12 and R-507 refrigerants at a specific saturated suction tempera-ture. The loss in capacity for an R-12 compressor for a change of 3 psig., 1 psig. to 4 psig., is 20%. The loss in the R-507 compressor for the same additional pres-sure drop is 12%.

Table 3-2Suction Line Pressure Drop

R-12

Evap. Temp.

Suction Line Pressure

DropPressure at Comp.

BTU/hr. Capacity

-10°F 1 psi 3.5 psig 27,490-10°F 2 psi 2.5 psig 25,950-10°F 3 psi 1.5 psig 24,410-10°F 4 psi 0.5 psig 22,100

R-507

Evap. Temp.

Suction Line Pressure

DropPressure at Comp.

BTU/hr. Capacity

-10°F 1 psi 24 psig 40,400-10°F 2 psi 23 psig 39,400-10°F 3 psi 22 psig 37,400-10°F 4 psi 21 psig 35,500

INTERNALLY COMPOUND TWO-STAGE SYSTEMS

As the compression ratio increases, the volumetric ef-ficiency of the compressor decreases and the heat of compression increases. For low temperature applica-

tions, the decreasing efficiency and excessively high discharge temperatures become increasingly critical. The lowest recommended evaporating temperature for compressors operating on the simple single stage compression cycle, is -40°F.

In order to increase operating efficiency at low tem-peratures the compression can be done in two steps or stages. For internally compound two stage operation with equal compression ratios, the compression ratio of each stage will be equal to the square root of the total compression ratio (approximately 1/4 of the total compression ratio for the normal two-stage operating range.) Since each stage of compression then is at a much lower compression ratio, the compressor ef-ficiency is greatly increased. The temperature of the refrigerant vapor leaving the first stage and entering the second stage may be high due to the heat of compres-sion. This can result in overheating the second stage cylinders and valves. To prevent compressor damage, saturated refrigerant must be injected between stages to properly cool the compressor.

A two-stage compressor is designed so that suction gas is drawn directly into the low stage cylinders and then discharged into the high stage cylinder or cylinders. On Copelametic® two-stage compressors the ratio of low stage to high stage displacement is 2 to 1. The greater volume of the low stage cylinders is necessary because of the difference in specific volume of the gas at the low and interstage pressures.

Figures 3-6 and 3-7 illustrate typical two-stage com-pressors as applied to low temperature systems. Two-stage refrigeration is effective down to evaporator temperatures of -80°F. Below that level, efficiency drops off rapidly.

For additional application and service information on internally compound compressors, refer to Application Engineering Bulletin AE 19-1132.

EXTERNALLY COMPOUND SYSTEMS

Two stage compression can be accomplished with the use of two compressors. The discharge of the first compressor becomes the suction of the second com-pressor. (See Figure 3-8) Like the internally compound compressor, ideally the first compressor will have twice the displacement of the second. However, in the exter-nally compound system, it is not critical.

In the externally compound system, the ideal interstage pressure absolute, can be calculated. It is the square root of the absolute suction pressure times the absolute discharge pressure.

3-6


The externally compound system can have compres-sors in parallel in either or both stages of the system. Compressors can have unloaders. Parallel compres-sors in both stages can be turned on and off to meet the demands of the low temperature and interstage pressures.

System With 6-Cylinder Compressor (Without Liquid Sub-Cooler) Figure 3-6

3-7


System With 6-Cylinder Compressor (With Sub-Cooler) Figure 3-7

3-8


Externally Compound System Figure 3-8

3-9

Evaporator

Condenser-Receiver

Assembly


CASCADE SYSTEMS

In order to operate satisfactorily at even lower evap-orating temperatures, and to increase the flexibility of system design, multiple stage refrigeration can also be accomplished by using separate systems with the evaporator of one serving as the condenser of the second by means of a heat exchanger. (See Figure 3-9) This type of design is termed a cascade system, and allows the use of different refriger-

ants in the separate systems. Refrigerants with characteristics and pressures suitable for ultra-low temperature refrigeration can be used in the low stage system. Cascade systems in multiples of two, three, or even more separate stages make possible refrigeration at almost any desired evaporating temperature. Cascade systems composed of both single and two-stage compressors can be used very effectively.

Cascade System Figure 3-9

3-10

HighStageComp.

HighStageCond .

HighStageEvap .

LowStageCond .

LowStageComp.

LowStageEvap .

1675 W. Campbell Rd.Sidney, OH 45365

EmersonClimate.com

Form No. AE 101 R2 (10/06)Emerson®, Emerson. Consider It Solved™, Emerson Climate Technologies™ and the Emerson Climate Technologies™ logo are the trademarks and service marks of Emerson Electric Co. and are used with the permission of Emerson Electric Co.Discus®, Copeland Compliant Scroll®, Copelametic®, Copeland®, and the Copeland® brand products logo are the trademarks and service marks of Emerson Climate Technologies, Inc.All other trademarks are the property of their respective owners.Printed in the USA. © 1968 Emerson Climate Technologies, Inc. All rights reserved.

Part 2 - Refrigeration System Components



This is the second of a series of publications comprising the Emerson Climate Technologies, Inc. Refrigeration Manual, and follows Part 1, “Fundamentals of Refrigeration.”

The information included on refrigeration components is general in nature and is intended only to give a brief description of their operation. Detailed information as to specific products is available from manufacturers of components and accessories.


1

Part IIREFRIGERATION SYSTEM COMPONENTS

Section 4. COMPRESSORS Reciprocating Compressors ............................. 4-1 Open Type Compressors .................................. 4-2 Accessible-Hermetic Motor-Compressors ........ 4-2 Welded Hermetic Motor-Compressors ............. 4-2 Compressor Speed........................................... 4-2 Basic Compressor Operation ........................... 4-4 Suction and Discharge Valves .......................... 4-4 Compressor Displacement ............................... 4-4 Clearance Volume ............................................ 4-4 Lubrication ........................................................ 4-5 Dry Air Holding Charge ..................................... 4-6 Compressor Cooling ......................................... 4-6 Compressor Capacity ....................................... 4-6 Two Stage Compressors .................................. 4-6 Compressors with Unloaders ......................... 4-7 Tandem Compressors ...................................... 4-7

Section 5. CONDENSERS

Air Cooled Condensers .................................... 5-1 Water Cooled Condensers ............................... 5-2 Evaporative Condensers .................................. 5-4 Condenser Capacity ......................................... 5-5 Condensing Temperature ................................. 5-5 Non-Condensable Gases ................................. 5-5 Condensing Temperature Difference ................ 5-6

Section 6. EVAPORATORS

Types of Evaporators........................................ 6-1 Blower Coil Construction .................................. 6-1 Pressure Drop and Other Factors in Evaporator Design ................................. 6-2 Evaporator Capacity ......................................... 6-2 Temperature Difference and Dehumidification ........................................ 6-2 Defrosting of Blower Coils ................................ 6-3

Section 7. CONTROL DEVICES, REFRIGERANT

Thermostatic Expansion Valves ....................... 7- 1 Other Types of Expansion Valves ..................... 7- 2 Distributors ....................................................... 7- 2 Capillary Tubes ................................................. 7- 2 Float Valves ...................................................... 7- 8Solenoid Valves ................................................ 7- 8 Crankcase Pressure Regulating Valves ........... 7- 9 Evaporator Pressure Regulating Valve ............. 7- 9 Hot Gas Bypass Valves .................................... 7- 9 Reversing Valves .............................................. 7-10 Check Valves .................................................... 7-10 Manual Shut-Off Valves .................................... 7-11 Compressor Service Valves ............................. 7-11 Schrader Type Valve ........................................ 7-11 Pressure Relief Valves ..................................... 7-12 Fusible Plugs .................................................... 7-12 Water Regulating Valves .................................. 7-12

Section 8. CONTROL DEVICES, ELECTRICAL Control Differential ............................................ 8-1 Line Voltage and Low Voltage Controls ............ 8-1 Low Pressure and High Pressure Controls ...... 8-1 Condenser Fan Cycling Control ....................... 8-2 Thermostats...................................................... 8-2 Oil Pressure Safety Control .............................. 8-2 Time Clocks ...................................................... 8-2Relays............................................................... 8-3 Time Delay Relay ............................................. 8-3 Transformers .................................................... 8-3


Section 9. MOTORS

Motor Temperature ........................................... 9-1 Open Type Motors and Belt Drives ................... 9-1 Hermetic Motors ............................................... 9-2 Nameplate Amperage ....................................... 9-2 Voltage and Frequency..................................... 9-3 Three Phase Motors ......................................... 9-3 Single Phase Motors ........................................ 9-3 Split Phase Motors ........................................... 9-3 Capacitor Start-Induction Run Motors (CSIR) ....................................................... 9-4 Capacitor Start-Capacitor Run Motors (CSR) ........................................................ 9-4 Permanent Split Capacitor Motors (PSC) ......... 9-5 Dual Voltage Motors ......................................... 9-5 Two Phase Motors ............................................ 9-6

Section 10. STARTING EQUIPMENT AND MOTOR PROTECTORS

Contactors and Starters.................................... 10-1 Capacitors ........................................................ 10-1 Start Capacitors ................................................ 10-2Run Capacitors ................................................. 10-2Reduced Voltage Starting ................................. 10-3

Motor Protection ............................................... 10-8 Internal Inherent Line Break Protector.............. 10-8 External Inherent Protector............................... 10-9Internal Thermostats ......................................... 10-9 External Thermostats ....................................... 10-9Current Sensitive Protectors............................. 10-9Thermotector .................................................... 10-9Solid State Protectors ....................................... 10-9Fuses and Circuit Breakers .............................. 10-9 Effect of Unbalanced Voltage and Current on Three Phase Motor Protection ............. 10-10

Section 11. ACCESSORIES

Receivers.......................................................... 11-1 Heat Exchangers .............................................. 11-1 Suction Accumulators ....................................... 11-1 Oil Separators................................................... 11-2 Dehydrators ...................................................... 11-2 Suction Line Filters ........................................... 11-2 Vibration Eliminators ......................................... 11-2 Strainers ........................................................... 11-3 Sight Glass and Moisture Indicators ................. 11-3 Discharge Mufflers............................................ 11-3 Crankcase Heaters ........................................... 11-3Refrigeration Gauges ....................................... 11-4


SECTION 4 COMPRESSORS

RECIPROCATING COMPRESSORS The design of the reciprocating compressor is somewhat similar to a modern automotive engine, with a piston driven from a crankshaft making alternate suction and compression strokes in a cylinder equipped with suction and discharge valves. Since the reciprocating compressor is a positive displacement pump, it is suitable for small displacement volumes, and is quite efficient at high condensing pressures and high compression ratios. Other advantages are its adaptability to a number of different refrigerants, the fact that liquid refrigerant may be easily run through connecting piping because of the high pressure created by the compressor, its durability, basic simplicity of design, and relatively low cost.

An exploded view of a typical Copelametic® accessible-hermetic motor-compressor is shown in Figure 10.

The compressor has two functions in the compression refrigeration cycle. First it removes the refrigerant vapor from the evaporator and reduces the pressure in the evaporator to a point where the desired evaporating temperature can be maintained. Second, the compressor raises the pressure of the refrigerant vapor to a level high enough so that the saturation temperature is higher than the temperature of the cooling medium available for condensing the refrigerant vapor.

There are three basic types of compressors; reciprocating, rotary, and centrifugal. Centrifugal compressors are widely used in large central air conditioning systems, and rotary compressors are used in the domestic refrigerator field, but the overwhelming majority of compressors used in the smaller horsepower sizes for commercial, domestic, and industrial applications are reciprocating, and this manual will cover only reciprocating compressors.

4-1


OPEN TYPE COMPRESSORS

Early models of refrigeration compressors were of the so-called open type, with the pistons and cylinders sealed within a crankcase, and a crankshaft extending through the body for an external power source. A shaft seal around the crankshaft prevented the loss of refrigerant and oil from the body.

Although at one time open type compressors were widely used, they have many inherent disadvantages such as greater weight, higher cost, larger size, vulnerability to seal failures, difficult shaft alignment, excessive noise, and short life of belts or direct drive components. As a result, the open type compressor has been largely replaced with the accessible-hermetic and hermetic type motor-compressor in most applications, and the use of open type compressors continues to decline except for specialized applications such as automobile air conditioning.

ACCESSIBLE-HERMETIC MOTOR-COMPRESSORS

The accessible-hermetic motor-compressor design was pioneered by Emerson Climate Technologies, Inc. and is widely used in the popular Copelametic® models. The compressor is driven by an electric motor mounted directly on the compressor crankshaft, with both the motor and the compressor working parts hermetically sealed within a common enclosure. The troublesome shaft seal is eliminated, motors can be sized specifically for the load to be handled, and the resulting design is compact, economical, efficient, and basically maintenance free.

Removable heads, stator covers, bottom plates, and housing covers allow access for easy field repairs in the event of compressor damage.

WELDED HERMETIC MOTOR-COMPRESSORS

In an effort to further decrease size and cost, the welded hermetic motor-compressor has been developed, and is widely used in small horsepower unitary equipment. As in the case of the accessible-hermetic motor-compressor an electric motor is mounted directly on the compressor crankshaft, but the body is a formed metal shell hermetically sealed by welding. No internal field repairs can be performed on this type of compressor since the only means of access is by cutting open the compressor shell.

COMPRESSOR SPEED Early models of compressors were designed for relatively slow speed operation, well below 1,000 RPM. In order to utilize standard 4 pole electric motors, accessible-hermetic and hermetic motor-compressors introduced operation at 1,750 RPM (1,450 RPM on 50 cycle). The increasing demand for lighter weight and more compact air conditioning equipment has been instrumental in the development of hermetic motor-compressors equipped with 2 pole motors operating at 3,500 RPM (2,900 RPM on 50 cycle).

Specialized applications such as aircraft, automotive, or military air conditioning equipment utilize even higher speed compressors, but for the normal commercial and domestic application, the existing 60 cycle electric power supply will generally limit compressor speeds to the presently available 1,750 and 3,500 RPM.

Higher compressor speeds introduce lubrication and life problems, and these factors as well as cost, size and weight must be considered in compressor design and application.

4-2

(continued on p. 4-4)


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BASIC COMPRESSOR OPERATION

A cross-sectional view of a typical Copelametic® motor-compressor is shown in Figure 13. Following is a brief description of its operation.

As the piston moves downward on the suction stroke, pressure is reduced in the cylinder. When the pressure falls below that in the compressor suction line, the pressure differential causes the suction valves to open and forces the refrigerant vapor to flow into the cylinder.

As the piston reaches the bottom of its stroke and starts upward on the compression stroke, pressure is developed in the cylinder, forcing the suction valves closed. The pressure in the cylinder continues to rise as the piston moves upward, compressing the vapor trapped in the cylinder. When the pressure in the cylinder exceeds the pressure existing in the compressor discharge line, the discharge valves are forced open, and the compressed gas flows into the discharge line and on into the condenser.

When the piston starts downward, the reduction in pressure allows the discharge valves to close because of the higher pressure in the condenser and discharge line, and the cycle is repeated.

For every revolution of the crankshaft, there is both a suction and compression stroke of each piston, so in 1,750 RPM motor-compressors there are 1,750 complete compression and suction cycles in each cylinder each minute, and in 3,500 RPM motor-compressors, 3,500 complete cycles each minute.

SUCTION AND DISCHARGE VALVES

Since the parts of the compressor most apt to require service are the suction and discharge valves, on Copelametic® compressors these valves are mounted on a valve plate which can be removed for easy service or replacement. A typical valve plate is shown in Figure 10, part number 11.

Most reciprocating compressor valves are of the reed type, and must seat properly to avoid leakage. The least bit of foreign material or corrosion under the valve will cause leakage and the utmost care must be used in protecting the compressor against contamination.

COMPRESSOR DISPLACEMENT

The displacement of a reciprocating compressor is the volume displaced by the pistons. Emerson Climate

Technologies, Inc. publishes the displacement of a compressor in terms of cubic feet per hour, but some manufacturers rate their compressors in terms of cubic inch displacement per revolution, or in cubic feet per minute. For comparative purposes, compressor displacement may be calculated by the following formulas:

DISPLACEMENT

CFM = π x D² x L x RPM x N 4 x 1728

CFH = π x D² x L x RPM x N x 60 4 x 1728

Cu. In./Rev. = π x D² x L x N 4

CONVERSION FACTORS

1750 RPM 3500 RPM CFH = 60 x CFM 60 x CFM CFH = 60.78 x 121.5 x Cu. In./Rev. Cu. In./Rev. CFM = 1.013 x 2.025 x Cu. In./Rev. Cu. In./Rev. Cu. In./Rev. = .01645 x CFH .00823 x CFH CFM = Cubic feet per minute CFH = Cubic feet per hour Cu. In./Rev. = Cubic inch d isp lacement per revolutionπ = 3.1416 D = Cylinder bore, inches L = Length of stroke, inches N = Number of cylinders RPM = Revolutions per minute 1728 = Cubic inches per cubic foot πD² = Area of a circle 4

CLEARANCE VOLUME

As mentioned previously, the volumetric efficiency of a compressor will vary with compressor design. If the valves seat properly, the most important factor affecting compressor efficiency is clearance volume.

At the completion of the compression stroke, there still remains some clearance space which is essential if the piston is not to hit the valve plate. There is also a great deal more space in the discharge valve ports in the valve plate, since the discharge valves are on top of the valve plate. This residual space which is unswept by the piston at the end of the stroke is termed clearance volume, and remains filled with hot, compressed gas at the end of the compression stroke.

4-4


When the piston starts down on the suction stroke, the residual high pressure gas expands and its pressure is reduced. No vapor from the suction line can enter the cylinder until the pressure in the cylinder has been reduced below the suction line pressure. Thus, the first part of the suction stroke is actually lost from a capacity standpoint, and as the compression ratio increases, a greater percentage of the suction stroke is occupied by the residual gas.

With high suction pressures, the compression ratio is low and clearance volume is not critical from a capacity standpoint. Additional clearance volume is also helpful in reducing the compressor noise level. Since lower gas velocities through the discharge ports reduce both wear and operating power requirements, on Copeland® brand air conditioning compressors, valve plates are designed with greater clearance volume by increasing the diameter of the discharge ports.

On low temperature applications, it is often necessary to reduce the clearance volume to obtain the desired capacity. Low temperature valve plates having smaller discharge port sizes to reduce the clearance volume are used on low temperature Copelametic® compressors.

LUBRICATION

An adequate supply of oil must be maintained in the crankcase at all times to insure continuous lubrication. The normal oil level should be maintained at or slightly above the center of the sight class.

On all Copelametic® compressors 5 H.P. and larger in size, and on 3 H.P. “NR” models, compressor lubrication is provided by means of a positive displacement oil pump. The pump is mounted on the bearing housing, and is driven from a slot in the crankshaft into which the flat end of the oil pump drive shaft is fitted.

Oil is forced through a hole in the crankshaft to the compressor bearings and connecting rods. A spring loaded ball check valve serves as a pressure relief device, allowing oil to bypass directly to the compressor crankcase if the oil pressure rises above its setting.

Since the oil pump intake is connected directly to the compressor crankcase, the oil pump inlet pressure will always be crankcase pressure, and the oil pump outlet pressure will be the sum of crankcase pressure plus oil pump pressure. Therefore, the net oil pump pressure is always the pump outlet pressure minus the crankcase pressure. When the compressor is operating with the suction pressure in a vacuum, the crankcase pressure is negative and must be added to the pump outlet pressure to determine the net oil pump pressure.

A typical compound gauge is calibrated in inches of mercury for vacuum readings, and 2 inches of mercury are approximately equal to 1 psi.

For example: Pump Net Oil Crankcase Outlet Pump Pressure Pressure Pressure 50 psig 90 psig 40 psi 8” vacuum 36 psig 40 psi (equivalent to a reading of minus 4 psig)

In normal operation, the net oil pressure will vary depending on the size of the compressor, the temperature and viscosity of the oil, and the amount of clearance in the compressor bearings. Net oil pressures of 30 to 40 psi are normal, but adequate lubrication will be maintained at pressures down to 10 psi. The bypass valve is set at the factory to prevent the net pump pressure from exceeding 60 psi.

The oil pump may be operated in either direction, the reversing action being accomplished by a friction plate which shifts the inlet and outlet ports. After prolonged operation in one direction, wear, corrosion, varnish formation, or burrs may develop on the reversing plate, and this can prevent the pump from reversing. Therefore, on installations where compressors have been in service for some time, care must be taken to maintain the original phasing of the motor if for any reason the electrical connections are disturbed.

The presence of liquid refrigerant in the crankcase can materially affect the operation of the oil pump. Violent foaming on start up can result in the loss of oil from the crankcase, and a resulting loss of oil pressure until oil returns to the crankcase. If liquid refrigerant or a refrigerant rich mixture of oil and refrigerant is drawn into the oil pump, the resulting flash gas may result in large variations and possibly a loss of oil pressure. Crankcase pressure may vary from suction pressure since liquid refrigerant in the crankcase can pressurize the crankcase for short intervals, and the oil pressure safety switch low pressure connection should always be connected to the crankcase.

During a rapid pull-down of the refrigerant evaporating temperature, the amount of refrigerant in solution in the crankcase oil will be reduced, and may cause flash gas at the oil pump. During this period the oil pump must pump both the flash gas and oil, and as a result the oil pressure may decrease temporarily. This will merely cause the oil pump to bypass less oil, and so long as the oil pressure remains above 9 psi, adequate lubrication

4-5


will be maintained. As soon as a stabilized condition is reached, and liquid refrigerant is no longer reaching the pump, the oil pressure will return to normal.

DRY AIR HOLDING CHARGE

All Copeland® brand compressors are thoroughly dehydrated at the factory, and are shipped with a dry air holding charge. The pressure inside a factory processed compressor is a guarantee that the compressor is leak tight, and the interior is absolutely dry. When installed, the compressor must be evacuated to remove the air from the system.

COMPRESSOR COOLING

Air cooled compressors require an adequate flow of cooling air over the compressor body to prevent the compressor from overheating. The air flow from the fan must be discharged directly on the motor-compressor. Air drawn through a compartment in which the compressor is located usually will not cool the compressor adequately.

Water cooled compressors are provided with a water jacket or wrapped with a copper water coil, and water must be circulated through the cooling circuit when the compressor is in operation.

Refrigerant cooled motor-compressors are designed so that suction gas flows around and through the motor for cooling. At evaporating temperatures below 0° F. additional motor cooling by means of air flow is necessary since the decreasing density of the refrigerant gas reduces its cooling ability.

COMPRESSOR CAPACITY

Capacity data is available from the manufacturer on each model of compressor for the refrigerants with which the compressor can be used. This data may be in the form of curves or in tabular form, and lists the BTU/hr. capacity at various saturated suction and discharge temperatures.

It is difficult to estimate compressor capacities accurately on the basis of displacement and compression ratio because of design differences between different models, but occasionally these factors can be valuable in estimating the comparative performance of compressors on the same application.

TWO STAGE COMPRESSORS Because of the high compression ratios encountered in ultra-low temperature applications, two stage

compressors have been developed for increased efficiency when evaporating temperatures are in the -30° F. to -80° F. range.

Two stage compressors are divided internally into low (or first) and high (or second) stages. On Copelametic® two stage compressors now in production, the ratio of low stage to high stage displacement is 2 to 1. The three cylinder models have two cylinders on the low stage and one on the high, while the six cylinder models have four cylinders on the low and two on the high.

The suction gas enters the low stage cylinders directly from the suction line, and is discharged into the interstage manifold at interstage pressure. Since the interstage discharge vapor has a relatively high temperature, liquid refrigerant must be metered into the interstage manifold by the desuperheating expansion valve to provide adequate motor cooling and prevent excessive temperatures during second stage compression. The discharge of the low stage enters the motor chamber and crankcase, so the crankcase is at interstage pressure.

Desuperheated refrigerant vapor at interstage pressure enters the suction ports of the high stage cylinders, and is then discharged to the condenser at the condensing pressure.

See Figures 6 and 7 on pages 3-6 and 3-7 of Part I for typical two stage systems.

4-6


TANDEM COMPRESSORS It is often desirable to interconnect two compressors on a single refrigeration system as a means of varying capacity according to the system requirement. This immediately introduces lubrication problems, for unless the pressures in the two crankcases are equalized, the oil will leave the crankcase having the highest pressure.

In order to solve the troublesome problems of oil equalization and vibration of connecting oil lines while obtaining the advantage of interconnected compressors, the tandem compressor was developed.

Basically this consists of two individual compressors with an interconnecting housing replacing the individual stator covers. Since each compressor may be operated individually, the tandem provides simple, foolproof capacity reduction with maximum power savings, and greatly simplifies system control.

The tandem offers a much greater factor of safety than a single compressor, and allows staggered starting to reduce inrush current requirements. In the event of failure of one of the compressors, emergency operation of the remaining compressor may be continued until replacement of the inoperative motor-compressor. In order to provide maximum protection for the system in the event of the failure of one compressor, a suction line filter should always be provided in the suction line of a tandem compressor, and an adequately sized liquid line filter-drier should be provided in the liquid line.

COMPRESSORS WITH UNLOADERS

In order to provide a means of changing compressor capacity under fluctuating load conditions, larger compressors are frequently equipped with unloaders. Unloaders on reciprocating compressors are of two general types. In the first, suction valves on one or more cylinders are held open by some mechanical means in response to a pressure control device. With the suction valves open, refrigerant vapor is forced back into the suction chamber during the compression stroke, and the cylinder performs no pumping action.

A second means of unloading is to bypass a portion of the discharge gas into the compressor suction chamber. Care must be taken to avoid excessive discharge temperatures when this is done.

Copelametic® compressors with unloaders have a bypass valve so arranged that discharge gas from an unloaded cylinder is returned to the suction chamber. During the unloaded operation, the unloaded cylinder is sealed from the discharge pressure created by the loaded cylinders. Since both suction and discharge pressures on the unloaded cylinder are approximately the same, the piston and cylinder do no work other than pumping vapor through the bypass circuit, and the problem of cylinder overheating while unloaded is practically eliminated. Because of the decreased volume of suction vapor returning to the compressor from the system and available for motor cooling, the operating range of unloaded compressors must be restricted, and operation beyond established limits can cause compressor overheating.

4-7


The condenser is basically a heat exchanger where the heat absorbed by the refrigerant during the evaporating process is given off to the condensing medium. As mentioned previously, the heat given off by the condenser is always greater than the heat absorbed during the evaporating process because of the heat of compression. As heat is given off by the high temperature high pressure vapor, its temperature falls to the saturation point and the vapor condenses to a liquid, hence the name condenser.

AIR COOLED CONDENSERS

The most commonly used condenser is of tube and external fin construction, which dissipates heat to the ambient air. Except for very small domestic units, which depend on gravity air circulation, heat transfer is efficiently accomplished by forcing large quantities of air through a compact condenser assembly. A typical refrigeration condensing unit equipped with an air cooled condenser is shown in Figure 16.

Air cooled condensers are easy to install, inexpensive to maintain, require no water, and there is no danger of freezing in cold weather. However, an adequate supply of fresh air is necessary, and the fan may create noise problems in large installations. In very hot regions, the relatively high temperature of the ambient air may result in high condensing pressures, but if the condenser

surface is amply sized, air cooled condensers can be used satisfactorily in all climatic regions. They have been used very successfully for many years in hot and dry areas where water is scarce. Because of the increasing scarcity of water in densely populated areas, the use of air cooled condensers will undoubtedly increase in the future.

When space permits, condensers may be made with a single row of tubing, but in order to achieve compact size, condensers are normally constructed with a relatively small face area and several rows of tubing in depth. As the air is forced through the condenser, it absorbs heat and the air temperature rises. Therefore, the efficiency of each succeeding row in the coil decreases, although coils up to eight rows in depth are frequently used.

Draw-through fans, which pull the air through the condenser, result in a more uniform air flow through the condenser than the blow-through type. Since even air distribution will increase the condenser efficiency, draw-through type fans are normally preferred.

Most air cooled refrigeration systems which are operated in low ambient temperatures are susceptible to damage due to abnormally low head pressure, unless adequate means of maintaining normal head pressure are provided. This is true, especially with refrigerated truck units parked outdoors or in unheated garages, roof mounted refrigeration or air conditioning systems, or any system exposed to low outside ambient temperatures. The capacity of refrigerant control devices (expansion valves, capillary tubes, etc.) is dependent upon the pressure difference across the device. Since they are selected for the desired capacity with normal operating pressures, abnormally low head pressure reducing the pressure difference across the expansion valve or capillary tube, may result in insufficient refrigerant flow. This can cause erratic refrigerant feed to the evaporator, and may result in frosting of the evaporator coil on air conditioning applications. The lower refrigerant velocity, and possibly lower evaporator pressure, permits oil to settle out and trap in the evaporator, sometimes causing shortage of oil in the compressor crankcase.

Several proprietary systems are available employing the principle of partially flooding the condenser with liquid refrigerant to reduce condensing capacity. Some of these systems result in very stable condensing pressures, but usually they require a large increase in the refrigerant charge which may cause problems in system performance. Controlling the condenser air

SECTION 5 CONDENSERS

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flow by means of louvers is also an effective means of condensing pressure control. Cycling the condenser fan is a simple but less effective means of control.

WATER COOLED CONDENSERS

When adequate low cost condensing water is available, water cooled condensers are often desirable because of the lower condensing pressures and better head pressure control is possible. Water, particularly from underground sources, is frequently much colder than daytime air temperatures. If evaporative cooling towers are used, the condensing water can be cooled to a point closely approaching the ambient wet bulb temperature. This allows the continuous recirculation of condensing water and reduces water consumption to a minimum.

Because of water’s excellent heat transfer characteristics, water cooled condensers can be quite compact. Several different types of construction are used including shell and coil, shell and tube, and tube within a tube styles. Normally the cooling water is run through tubing or coils within a sealed shell into which the hot gas is discharged from the compressor. As the refrigerant condenses it can be fed out the refrigerant liquid line, thus making the use of a separate receiver unnecessary. A water cooled condensing unit equipped with a shell and tube condenser is shown in Figure 17.

A pressure or temperature sensitive modulating water control valve can be used to maintain condensing pressures within the desired range by increasing or decreasing the rate of water flow as necessary.

Cooling water circuits in compressors with water jackets and in water cooled condensers may be either series or parallel as required by the particular application. The use of parallel circuits results in a lower pressure drop through the circuit, and may be necessary when the temperature of the cooling water is such that the water temperature rise must be held to a minimum. Occasionally condensers may be damaged by excessive water velocities or cavitation on the water side of the condenser tubes. In order to prevent operating difficulties, care should be taken to follow the installation recommendations as outlined below:

1. Water velocities through the condenser should not exceed 7 feet per second. Higher velocities can result in “impingement corrosion”. This is a condition in which progressive erosion of the tube can occur due to the high water velocity washing away the inner oxidized surface of the tube at points where excessive turbulence may occur. This can originate with a minute imperfection on the tube inner surface, but it becomes progressively worse as the pitting increases.

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Figure No.18 illustrates the type of circuiting normally used on all standard condensing units using city water supply. All water cooled condensing units are shipped from the factory with the connections as shown above, and water connections must be modified in the field if parallel circuits are desired.

Figure No.19 illustrates a condenser with parallel circuits connected to a motor-compressor with a straight-through circuit. This type of circuiting is frequently used when the condensing water is cooled by a water tower . The straight-through compressor circuit would be used when connecting a motor-compressor wrapped with an external water coil.

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discharge connection with a high vertical drop could result in cavitation in a manner similar to a pump on the outlet of the condenser.

EVAPORATIVE CONDENSERS

Evaporative condensers are frequently used where lower condensing temperatures are desired than are obtainable with air cooled condensers, and where the available water supply may not be adequate for heavy water usage. The hot refrigerant vapor is piped through a spray chamber where it is cooled by evaporation of the water coming in contact with the refrigerant tubing.

Water which is exposed to air flow in a spray chamber will evaporate rapidly. Latent heat required for the evaporating process is obtained by a reduction in sensible heat and, therefore, a reduction in the temperature of the water remaining. An evaporative spray chamber can reduce the water temperature to a point closely approaching the wet bulb temperature of the air.

Wet bulb temperature is a term used in air conditioning to describe the lowest temperature that can be obtained by the evaporating process. The term wet bulb temperature is derived from the fact that a common mercury bulb thermometer exposed to the ambient air

Figure No.20 shows parallel circuits in both water cooled condenser and the motor-compressor water jacket. Each water jacket circuit is connected in series with one circuit of the split condenser. This type of water circuiting is used when a minimum of water pressure drop is required.

In order to maintain water velocities at an acceptable level, parallel circuiting of the condenser may be necessary when high water flow is required. 2. If a water circulating pump is used, install so that the condenser is fed from the discharge side of the pump. If the pump were on the discharge side of the condenser, the condenser would have a slight vacuum in the water system, and therefore the water would be much nearer its boiling point. A combination of a localized hot spot in the condenser together with a localized velocity increase that might reduce pressures even lower, could result in triggering a cavitation condition.

Cavitation is basically a condition where a fluctuating combination of pressure and temperature can cause instantaneous boiling or flashing of water into vapor, with the subsequent collapse of the bubbles as the conditions vary. This can result in very rapid erosion and destruction of the water tube. Maintaining a positive pressure in the condenser will prevent this condition.

3. If the condenser is installed more than 5 feet higher than the outlet drain point of the condenser, a vacuum breaker or open vent line should be provided to prevent the discharge line from creating a partial vacuum condition in the condenser water system. An unvented

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indicates the dry bulb or ambient temperature, while if a wick wetted with water is placed around the mercury bulb and the thermometer is exposed to rapid air movement, the temperature indicated by the thermometer will be the wet bulb temperature. The difference between the dry bulb and wet bulb readings is determined by the rate of evaporation from the wet surface of the wick, and this in turn is proportional to the moisture content or vapor pressure of the air. The wet bulb temperature is always lower than the dry bulb temperature, and for a given dry bulb, the less the moisture content of the air, the lower the wet bulb temperature will be.

Since the cooling is accomplished by evaporation of the water, water consumption is only a fraction of that used in conventional water cooled applications in which the water once used is discharged to a drain. Evaporative condensing is therefore widely used in hot, arid regions of the world.

Corrosion, scale formation, and the danger of freezing are problems that must be solved with both evaporative and water cooled condensers. With both cooling towers and evaporative condensers, a bleed to a drain must be provided to prevent the concentration of contaminants in the cooling water.

CONDENSER CAPACITY

The heat transfer capacity of a condenser depends upon several factors:

1. Surface area of the condenser .

2. Temperature difference between the cooling medium and the refrigerant gas.

3. Velocity of the refrigerant gas in the condenser tubes. Within the normal commercial operating range, the greater the velocity, the better the heat transfer factor, and the greater the capacity.

4. Rate of flow of the cooling medium over or through the condenser. Heat transfer increases with velocity for both air and water, and in the case of air, it also increases with density.

5. Material of which the condenser is made. Since heat transfer differs with different materials, more efficient metals will increase the capacity.

6. Cleanliness of the heat transfer surface. Dirt, scale, or corrosion can reduce the heat transfer rate.

For a given condenser, the physical characteristics are fixed, and the primary variable is the temperature

difference between the refrigerant gas and the condensing medium.

CONDENSING TEMPERATURE

The condensing temperature is the temperature at which the refrigerant gas is condensing from a vapor to a liquid. This should not be confused with the temperature of the cooling medium, since the condensing temperature must always be higher in order for heat transfer to take place.

In order to condense the refrigerant vapor flowing into the condenser, heat must flow from the condenser at the same rate at which heat is introduced by the refrigerant gas entering the condenser. As mentioned previously, the only way in which the capacity of the condenser can be increased under a given set of conditions is by an increase in the temperature difference through the condenser walls.

Since a reciprocating compressor is a positive displacement machine, the pressure in the condenser will continue to increase until such time as the temperature difference between the cooling medium and the refrigerant condensing temperature is sufficiently great to transfer the necessary amount of heat. With a large condenser, this temperature difference may be very small. With a small condenser or in the event air or water flow to the condenser has been blocked, the necessary temperature difference may be very large. This can result in dangerously high pressures, and it is essential that the condenser is operating properly any time a refrigeration unit is in operation.

The condensing temperature and therefore the condensing pressure is determined by the capacity of the condenser, the temperature of the cooling medium, and the heat content of the refrigerant gas being discharged from the compressor, which in turn is determined by the volume, density and temperature of the gas discharged.

NON-CONDENSABLE GASES

Air is primarily composed of nitrogen and oxygen, and both elements remain in gaseous form at all temperatures and pressures encountered in commercial refrigeration and air conditioning systems. Therefore, although these gases can be liquefied under extremely high pressures and extremely low temperatures, they may be considered as non-condensable in a refrigeration system.

Scientists have discovered that one of the basic laws of nature is the fact that in a combination of gases, each gas exerts its own pressure independently of others,

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and the total pressure existing in a system is the total of all the gaseous pressures present. A second basic characteristic of a gas is that if the space in which it is enclosed remains constant, so that it cannot expand, its pressure will vary directly with the temperature. Therefore, if air is sealed in a system with refrigerant, the nitrogen and oxygen will each add their pressure to the system pressure, and this will increase as the temperature rises.

Since the air is non-condensable, it will usually trap in the top of the condenser and the receiver. During operation the compressor discharge pressure will be a combination of the refrigerant condensing pressure plus the pressure exerted by the nitrogen and oxygen. The amount of pressure above normal condensing pressure that may result will depend on the amount of trapped air, but it can easily reach 40 to 50 psig or more. Any time a system is running with abnormally high head pressure, air in the system is a prime suspect .

CONDENSING TEMPERATURE DIFFERENCE

A condenser is normally selected for a system by sizing it to handle the compressor load at a desired temperature difference between the condensing temperature and the expected temperature of the cooling medium. Most air cooled condensers are selected to operate on temperature differences (commonly called TD) of 20° F. to 30° F. at design conditions, but higher and lower TDs are sometimes used on specialized applications. Standard production air cooled condensing units are often designed with one condenser for a wide range of applications. In order to cover as wide a range as possible, the TD at high suction pressures may be from 30° F. to 40° F., while at low evaporating temperatures the TD often is no more than 4° F. to 10° F. The design condensing temperature on water cooled units is normally determined by the temperature of the water supply and the water flow rate available, and may vary from 90° F. to 120° F.

Since the condenser capacity must be greater than the evaporator capacity by the heat of compression and the motor efficiency loss, the condenser manufacturer may rate condensers in terms of evaporator capacity, or may recommend a factor to allow for the heat of compression in selecting the proper condenser size.

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SECTION 6 EVAPORATORS

The evaporator is that part of the low pressure side of the refrigeration system in which the liquid refrigerant boils or evaporates, absorbing heat as it changes into a vapor. It accomplishes the actual purpose of the system, refrigeration.

TYPES OF EVAPORATORS

Evaporators are made in many different shapes and styles to fill specific needs. The most common style is the blower coil or forced convection evaporator in which the refrigerant evaporates inside of finned tubes, extracting heat from air blown through the coil by a fan. However, specific applications may use bare coils with no fins, gravity coils with natural convection air flow, flat plate surface, or other specialized types of heat transfer surface.

Direct expansion evaporators are those in which the refrigerant is fed directly into the cooling coil through a metering device such as an expansion valve or capillary tube, absorbing the heat directly through the walls of the evaporator from the medium to be cooled. Figure 21 shows a direct expansion coil of one manufacturer prior to assembly in a blower unit.

In other types of systems, secondary refrigerants such as chilled water or brine may be used for the actual space or product refrigeration while the evaporator is the water or brine chiller. A complete packaged water chiller, designed to furnish chilled water for air conditioning or other cooling applications is shown in Figure 22.

BLOWER COIL CONSTRUCTION

A typical blower coil is made up of a direct expansion coil, mounted in a metal housing complete with a fan for forced air circulation. The coil is normally constructed of copper tubing supported in metal tube sheets, with aluminum fins on the tubing to increase heat transfer efficiency.

If the evaporator is quite small, there may be only one continuous circuit in the coil, but as the size increases, the increasing pressure drop through the longer circuit makes it necessary to divide the evaporator into several individual circuits emptying into a common header. The various circuits are usually fed through a distributor which equalizes the feed in each circuit in order to maintain high evaporator efficiency.

The spacing of fins on the refrigerant tubing will vary depending on the application. Low temperature coils may have as few as two fins per inch, while air conditioning coils may have up to twelve per inch or more. In general

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if the evaporator temperature is to be below 32° F. so that frost will accumulate, fin spacings of 4 per inch or less are commonly used, although closer fin spacings are sometimes used if efficient defrost systems are available. In air conditioning applications, icing of the coil is seldom a problem, and the limit on fin spacing may be dictated by the coil’s resistance to air flow.

Since the heat transfer efficiency of the coil increases with an increase in the mass flow of air passing through it, high velocities are desirable. However at face velocities greater than 500 to 600 FPM, water collecting on the coil from condensation will be blown off into the air stream, and except for specialized applications, these velocities are seldom exceeded.

PRESSURE DROP AND OTHER FACTORS IN EVAPORATOR DESIGN

As mentioned previously, pressure drop occurring in the evaporator results in a loss of system capacity due to the lower pressure at the outlet of the evaporator coil. With a reduction in suction pressure, the specific volume of the gas returning to the compressor increases, and the weight of the refrigerant pumped by the compressor decreases.

However there are other factors which must also be considered in evaporator design. If the evaporator tubing is too large, refrigerant gas velocities may become so low that oil will accumulate in the tubing and will not be returned to the compressor. The only means of assuring satisfactory oil circulation is by maintaining adequate gas velocities. The heat transfer ability of the tubing may also be greatly decreased if velocities are not sufficient to scrub the interior tubing wall, and keep it clear of an oil film. The goals of low pressure drop and high velocities are directly opposed, so the final evaporator design must be a compromise.

Pressure drops through the evaporator of approximately 1 to 2 psi are acceptable on most medium and high temperature applications, and 1/2 to 1 psi are common in low temperature evaporators.

EVAPORATOR CAPACITY

The factors affecting evaporator capacity are quite similar to those affecting condenser capacity.

1. Surface area or size of the evaporator .

2. Temperature difference between the evaporating refrigerant and the medium being cooled.

3. Velocity of gas in the evaporator tubes. In the normal

commercial range, the higher the velocity the greater the heat transfer rate.

4. The velocity and rate of flow over the evaporator surface of the medium being cooled.

5. Material used in evaporator construction.

6. The bond between the fins and tubing is quite important. Without a tight bond, heat transfer will be greatly decreased.

7. Accumulation of frost on evaporator fins. Operation at temperatures below freezing with blower coils will cause the formation of ice and frost on the tubes and fins. This can both reduce the air flow over the evaporator and reduce the heat transfer rate.

8. Type of medium to be cooled. Heat flows almost five times more effectively from a liquid to the evaporator than from air .

9. Dewpoint of the entering air. If the evaporator temperature is below the dewpoint of the entering air, latent as well as sensible cooling will occur.

TEMPERATURE DIFFERENCE AND DEHUMIDIFICATION

Since for a given installation, the physical characteristics are fixed, the primary variable as in the case of the condenser, is the temperature difference between the evaporating refrigerant and the medium being cooled, commonly called the TD. For a blower coil, the colder the refrigerant with respect to the temperature of the air entering the evaporator, the greater will be the capacity of the coil.

Temperature differences of 5° F. to 20° F. are commonly used. Usually for best economy, the TD should be kept as low as possible, since operation of the compressor will be more efficient at higher suction pressures.

The amount of moisture condensed out of the air is in direct relation to the temperature of the coil, and a coil operating with too great a differential between the evaporating temperature and the entering air temperature will tend to produce a low humidity condition in the refrigerated space. In the storage of leafy vegetables, meats, fruits, and other similar perishable items, low humidity will result in excessive dehydration and damage to the product. For perishable commodities requiring a very high relative humidity (approximately 90%) a TD from 8° F. to 12° F. is recommended, and for relative humidities slightly lower (approximately 80%) a TD from 12° F. to 16° F. is normally adequate.

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DEFROSTING OF BLOWER COILS

Ice and frost will accumulate continuously on coils operating below freezing temperatures, and air flow through the coil will be eventually blocked unless the frost is removed. To allow continuous operation on refrigeration applications where frost accumulation can occur, periodic defrost cycles are necessary.

If the air returning to the evaporator is well above 32° F., defrosting can be accomplished by allowing the fan to continue operation while the compressor is shut down, either for a preset time period or until the coil temperature rises a few degrees above 32° F., the melting temperature of the frost.

For low temperature applications, some source of heat must be supplied to melt the ice. Electric defrost systems

utilize electric heater coils or rods in the evaporator. Proprietary systems using water for defrosting are available. Hot gas defrosting is widely used, with the discharge gas from the compressor bypassing the condenser and discharging directly into the evaporator inlet. In hot gas defrost systems, the heat of compression or some source of stored heat provides defrost heat, and adequate protective devices such as re-evaporators or suction accumulators must be provided if necessary to prevent liquid refrigerant from returning to the compressor. Other systems may utilize reverse cycle defrosting, in which the flow of refrigerant is reversed to convert the evaporator temporarily into a condenser until the defrost period is complete.

To prevent refreezing of the melted condensate in the evaporator drain pan, a drain pan heater is required on low temperature systems.

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SECTION 7 CONTROL DEVICES, REFRIGERANT

In modern refrigeration practice, a wide variety of refrigerant control devices are used to obtain efficient economic operation. Small systems with manual control or simple “on-off” automatic control may require only one or two controls, but large systems with more elaborate automatic control may have a multitude of controls, the proper operation of each being essential to the satisfactory performance of the system.

In order to adjust a control for efficient performance, or recognize the effect of a malfunction, it is essential that the function, operation, and application of each refrigeration control be completely understood.

THERMOSTATIC EXPANSION VALVES

The most commonly used device for controlling the flow of liquid refrigerant into the evaporator is the thermostatic expansion valve. An orifice in the valve meters the flow into the evaporator, the rate of flow being modulated as required by a needle type plunger and seat, which varies the orifice opening.

The needle is controlled by a diaphragm subject to three forces. The evaporator pressure is exerted beneath the diaphragm tending to close the valve. The force of a superheat spring is also exerted beneath the diaphragm in the closing direction. Opposing these two forces is the pressure exerted by the charge in the thermal bulb, which is attached to the suction line at the outlet of the evaporator.

It is most convenient to visualize the action of the thermostatic expansion valve by considering the thermal bulb charge to be the same refrigerant as that being used in the system. With the unit in operation, the refrigerant in the evaporator is evaporating at its saturation temperature and pressure. So long as the thermal bulb is exposed to a higher temperature it will exert a higher pressure than the refrigerant in the evaporator, and therefore the net effect of these two pressures is to open the valve. The superheat spring pressure is a fixed pressure causing the valve to close whenever the net difference between the bulb pressure and the evaporator pressure is less than the superheat spring setting.

As the temperature of the refrigerant gas leaving the evaporator rises (an increase in superheat) the pressure exerted by the thermal bulb at the outlet of the coil increases, and the expansion valve flow increases; as the temperature of the leaving gas decreases (a decrease in superheat) the pressure exerted by the thermal bulb decreases, and the expansion valve closes slightly and the flow decreases.

With an evaporator and an expansion valve correctly sized for the load, the expansion valve feed will be quite stable at the desired superheat setting. An oversized expansion valve or an oversized evaporator can cause erratic feeding of the evaporator, which may result in large fluctuations in compressor suction pressure, and possible liquid return to the compressor.

Because of the pressure drop due to refrigerant flow through the evaporator, the evaporating pressure at the outlet of the evaporator coil will be lower than that at the expansion valve. If this pressure drop is of any magnitude, a higher superheat will be required to bring the forces acting on the valve diaphragm into equilibrium, and the evaporator will be partially starved. To compensate for pressure drop through the evaporator, an external equalizer connection is often used on the expansion valve. This introduces the evaporator outlet pressure under the valve diaphragm, rather than the evaporator inlet pressure, and the valve operation is then free from any influence due to evaporator pressure drop. Valves with external equalizer connections are recommended whenever the pressure drop through the evaporator

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is above 2 1/2 psi for high temperature applications, 1 1/2 psi in the medium temperature range, and 1/2 psi in the low temperature range. Valves with external equalizers must be employed when a pressure drop type of distributor is used.

Pressure limiting expansion valves are often used to limit the power requirement of the compressor. The valve is constructed in such a manner that it limits the suction pressure to a given maximum value, and restricts the refrigerant feed if the suction pressure rises above that point.

Gas charged pressure limiting valves have a limited charge, and at temperatures of the thermal bulb equivalent to its maximum operating pressure, all of the liquid charge has vaporized, and any further increase in temperature can only superheat the gas, but cannot exert additional pressure. Any increase in evaporator pressure will then act as a closing force on the expansion valve. The disadvantage of the gas charged valve is the possibility of the limited charge condensing in the head of the expansion valve, if the head is colder than the thermal bulb, causing the valve to lose control of the liquid feed. With gas charged valves, the thermal bulb must always be colder than the head of the valve, and the gas charged valve normally is used only on high temperature applications such as air conditioning.

Mechanical limiting valves are available, usually with a spring loaded double diaphragm type construction. If the evaporator reaches a preset pressure, the diaphragm collapses, and the valve feed is restricted until the pressure decreases sufficiently for the spring tension to restore the diaphragm to its normal operating position.

In order to achieve closer control for varying applications, expansion valves are available with different types of charge in the thermal bulb, each having different operating characteristics. The superheat spring is also normally equipped with an external adjusting screw so that it can be set for the desired superheat on a given application. Before adjusting any expansion valve, the exact characteristics of the valve should be thoroughly understood. The manufacturer’s catalog data must be consulted for detailed information on a given valve.

OTHER TYPES OF EXPANSION VALVES

The automatic expansion valve is really better described as a constant pressure expansion valve, since it modulates its feed to maintain a constant preset pressure in the evaporator. The automatic expansion valve was widely used at one time, but because of its tendency

to starve the evaporator on heavy loads, and flood the evaporator on light loads, it has been largely replaced by the thermostatic expansion valve and capillary tubes.

Hand expansion valves are sometimes used when an operator is available and manual liquid refrigerant feed is acceptable. A needle valve is adjusted as required to maintain the desired flow.

DISTRIBUTORS

When the refrigeration load is such that large evaporators are required, multiple refrigerant circuits are necessary to avoid excessive pressure drop through the evaporator. To insure uniform feed from the expansion valve to each of the various circuits, a refrigerant distributor is normally used. A typical distributor mounted on a direct expansion coil is shown in Figure 21, page 6-1.

As liquid refrigerant is fed through the expansion valve, a portion of the liquid flashes into vapor in order to reduce the liquid temperature to evaporator temperature. This combination of liquid and flash gas is fed into the distributor from the expansion valve, and is then distributed evenly through small feeder tubes, the number depending on the construction of the distributor and the number of circuits required to provide proper refrigerant velocity in the evaporator .

Without the distributor, the flow would separate into separate gas and liquid layers, resulting in the starving of some evaporator circuits. To avoid variations in circuit feed, extreme care must be taken to insure that tubing lengths are equal, so equal resistance is offered by each circuit.

There are two different approaches in the design of a distributor. A high-pressure drop distributor depends on the turbulence created by an orifice to achieve good distribution. A low-pressure drop distributor depends on a contour flow pattern with high velocity in the distributor throat to give proper distribution of the refrigerant flow. Both types of distributor give satisfactory performance when properly applied in accordance with the manufacturer’s instructions.

CAPILLARY TUBES

On small unitary equipment such as package air conditioners, domestic refrigeration equipment, and self-contained commercial refrigeration cases, capillary tubes are widely used for liquid refrigerant control. A capillary tube is a length of tubing of small diameter with the internal diameter held to extremely close tolerances. It is used as a fixed orifice to perform the same function

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CAPILLARY TUBE SELECTION R-22 HIGH TEMPERATURE

45° F. evaporating temperature (Preliminary Selection Only) Final Selection Should Be Determined by Unit Test

**Length to balance unit at 45° F. evaporating, 130° F. condensing, 10° F. Sub-cooling.

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CAPILLARY TUBE SELECTION R-22 MEDIUM TEMPERATURE

25°F, to 10°F. Evaporating Temperature (Preliminary Selection Only) Selection Should Be Determined by Unit Test

**Length to balance unit with 115°F. condensing, °F. sub-cooling in condenser, Heat Exchanger to give 15°F. sub-cooling.

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CAPILLARY TUBE SELECTION R-12 MEDIUM TEMPERATURE

25°F, to 10°F. Evaporating Temperature (Preliminary Selection Only) Final Selection Should Be Determined by Unit Test

**Length to balance unit with 115°F. condensing, 5°F. sub-cooling in condenser, Heat Exchanger to give 15°F. sub-cooling.

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CAPILLARY TUBE SELECTION R-22 LOW TEMPERATURE

15°F. to 25°F. Evaporating Temperature (Preliminary Selection Only) Final Selection Should Be Determined by Unit Test

*Length to balance unit at 110°F. condensing and 20°F. Liquid sub-cooling (15°F. in condenser, 15°F in heat exchanger)

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CAPILLARY TUBE SELECTION R-502 LOW TEMPERATURE

15°F. to 25°F. Evaporating Temperature (Preliminary Selection Only) Final Selection Should Be Determined by Unit Test

*Length to balance unit at 110°F. condensing and 20°F. Liquid sub-cooling (15°F. in condenser, 15°F in heat exchanger)

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as the expansion valve, to separate the high and low pressure sides of the system, and meter the proper feed of liquid refrigerant.

Since there are no moving parts, it is simple and trouble free if kept free of foreign material. A capillary tube is of very small diameter, and absolute freedom from foreign matter and moisture is essential, making a factory sealed unit a practical necessity.

Since the orifice is fixed, the rate of feed is relatively inflexible. Under conditions of constant load, and constant discharge and suction pressures, the capillary tube performs very satisfactorily. However, changes in the evaporator load or fluctuations in head pressure can result in under or over feeding of the evaporator.

A major advantage of the capillary tube in some systems is the fact that refrigerant continues to flow into the evaporator after the compressor stops operation, thus equalizing pressures on the high and low sides of the system. This allows the use of low starting torque motors.

The refrigerant charge is critical in capillary tube systems since normally there is no receiver to store excess refrigerant. Too much refrigerant will cause high discharge pressures and motor overloading, and possible liquid floodback to the compressor during the off cycle; too little will allow vapor to enter the capillary tube causing a loss in system capacity.

Due to its basic simplicity, the elimination of the need for a receiver, and the low starting torque requirement, a capillary tube system is the least expensive of all liquid control systems.

Sizing of a capillary tube is difficult to calculate accurately, and can best be determined by actual test on the system. Once determined, the proper size capillary tube can be applied to identical systems, so it is well adapted to production units. Figures 24, 25, 26, 27, and 28 give tentative selection data for capillary tubes.

FLOAT VALVES

On some specialized applications, it may be desirable to operate with completely flooded systems, that is, with the evaporator completely filled with liquid refrigerant. A typical application might be an industrial process cooling installation where a brine or liquid is piped through a chiller shell in which the refrigerant level is to be maintained. Special liquid level controls are available from expansion valve manufacturers. These normally are mounted in a secondary float chamber and modulate flow as necessary

to maintain a given liquid level. Such applications are quite specialized and the manufacturer’s instructions should be followed closely. Unless some means is provided for positive oil return, oil may accumulate in a float chamber causing lubrication difficulties.

Commercial or domestic applications using either high side or low side float chambers for liquid feed have been largely replaced by capillary tube and expansion valve control.

SOLENOID VALVES

A solenoid valve is an electrically controlled refrigerant flow control valve. It is not a modulating valve, and is either open or closed.

The valve consists of a body, a plunger with an iron core which seats in the valve orifice, and an electrical solenoid coil. A normally closed solenoid valve is closed when the coil is deenergized and the plunger is seated. When the solenoid coil is energized, the magnetic effect of the coil lifts the plunger and opens the valve. Normally open valves with a reverse type action are made, but are rarely used.

Solenoid valves are commonly used in refrigerant liquid and hot gas lines to stop refrigerant flow when not desired, or to isolate individual evaporators when

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multiple evaporators are used. On large installations, large numbers of solenoid valves may be necessary for satisfactory automatic control.

CRANKCASE PRESSURE REGULATING VALVES

This type of valve, commonly called a CPR valve or a holdback valve, limits the suction pressure at the compressor below a preset limit to prevent overloading of the compressor motor. The valve setting is determined by a pressure spring, and the valve modulates from fully open to fully closed in response to outlet pressure, closing on a rise in outlet pressure.

The crankcase pressure regulating valve should be located in the suction line between the evaporator and the compressor. Since the power requirement of the compressor declines with a fall in suction pressure, the CPR valve is normally used to prevent motor overloading on low temperature units during pulldown or defrost cycles. Use of the valve permits the application of a

larger displacement compressor without overloading a given size motor, but pressure drop through the valve may result in an unacceptable loss of system capacity unless the valve is adequately sized.

EVAPORATOR PRESSURE REGULATING VALVE

On systems with multiple evaporators operating at different temperatures, or on systems where the evaporating temperature cannot be allowed to fall below a given temperature, an evaporator pressure regulator valve is frequently used to control the evaporating temperature. This valve, often called an EPR valve, acts

similarly to the crankcase pressure regulator, except that it is responsive to inlet pressure. It should be located in the suction line at the evaporator outlet.

An EPR valve modulates from fully open to fully closed, closing on a fall in inlet pressure, and its sole function is to prevent the evaporator pressure from falling below a predetermined value for which the regulator has been set.

HOT GAS BYPASS VALVES

Hot gas bypass valves are used where it is desirable to modulate the compressor capacity and at the same time prevent the suction pressure from falling to objectionable low levels. These valves operate in the

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same fashion as crankcase pressure regulators since they are responsive to outlet pressure, modulate from fully open to fully closed, and open in response to a decrease in downstream pressure. The construction must be suitable to withstand the high temperature discharge gas from the compressor.

Hot gas valves are set to maintain a desired minimum pressure by spring tension, and may be either direct or pilot operated. They are normally equipped with an external equalizer connection, which operates in the same fashion as an external equalizer on an expansion valve to compensate for pressure drops in the lines. The external equalizer should be attached to the suction line at the point where it is desired to control the suction pressure. REVERSING VALVES

In recent years, usage of the “heat pump” principle to enable an air conditioning unit to supply both cooling and heating has become increasingly popular. Basically this involves switching the functions of the evaporator and condenser by a change in refrigerant flow as desired, so that the indoor coil becomes the evaporator for cooling purposes, and the condenser for heating usage. The outdoor coil in turn is a condenser during the cooling cycle, and an evaporator during the heating cycle.

To conveniently reverse the system operation, four-way reversing valves have been developed. By means of a slide action actuated by a solenoid, the connections from the compressor suction and discharge ports to the evaporator and condenser can be reversed at will.

Three-way valves are being increasingly used for hot gas defrosting. This valve enables the flow of hot gas from the compressor discharge valve to be shunted from the condenser to the evaporator for defrosting purposes, and then conveniently returned to the condenser when normal cooling is resumed.

CHECK VALVES

It is often desirable to prevent refrigerant from reversing its direction of flow during an off cycle, or during a change in the operating cycle. A simple spring loaded valve such as shown in Figure 34 allows flow in one direction only, and closes if pressures are such that reverse flow could occur. Check valves may be used in either liquid or gas lines, and are frequently used to prevent backflow of liquid refrigerant or hot gas in low ambient condenser controls, and in reverse cycle heat pumps. Check valves used in refrigeration systems should be spring loaded to prevent noise and chattering

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which may be caused by pulsations in refrigerant flow originating in the compressor. MANUAL SHUT-OFF VALVES

Manual shut-off valves are often used so that portions of the refrigeration system can be isolated for service or repairs. Special valves designed for refrigeration usage are required to avoid leakage.

COMPRESSOR SERVICE VALVES Compressor suction and discharge service valves are shut-off valves with a manual operated stem. Most service valves are equipped with a gauge port so that the refrigerant operating pressure may be observed.

When the valve is back-seated (the stem turned all the way out) the gauge port is closed and the valve is open. If the valve is front-seated (the stem turned all the way in) the gauge port is open to the compressor and the line connection is closed. In order to read the pressure while the compressor is in operation, the valve should be back-seated, and then turned in one or two turns in order to slightly open the connection to the gauge port. The compressor is always open to either the line or the gauge port, or both if the valve is neither front nor back-seated.

SCHRADER TYPE VALVE

The Schrader type valve is a recent development for convenient checking of system pressures where it is not economical, convenient, or possible to use the compressor service valves with gauge ports. The Schrader type valve is similar in appearance and principle to the air valve used on automobile or bicycle tires, and must have a cap for the fitting to insure leak-proof operation.

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This type of valve enables checking of the system pressure, or charging refrigerant without disturbing the unit operation. An adaptor is necessary for the standard serviceman’s gauge or hose connection to fit the Schrader type valve.

PRESSURE RELIEF VALVES

Safety relief valves are required by many local construction codes. Various types of relief valves are available, and the system requirement may be dictated by the local code requirement. Normally code requirements specify that the ultimate strength of the high side parts shall be a minimum of 5 times the discharge or rupture pressure of the relief valve, and that all condensing units with pressure vessels exceeding 3 cubic feet interval volume shall be protected by a pressure relief device. Discharge may be to the atmosphere, or it may be a discharge from the high pressure side of the system to the low pressure side.

A typical reseating type valve is shown in Figure 38. The valve opens at a preset pressure, and refrigerant is discharged until the pressure falls to the reseating point.

Some Copeland® brand compressors have reseating type pressure relief valves installed internally in the discharge chamber which allow excessive pressures to discharge to the suction chamber. A typical internal type valve is shown in Figure 39.

Rupture disc type relief devices have a thin disc which is designed to rupture at a given relief pressure, discharging the refrigerant to the atmosphere.

FUSIBLE PLUGS

A fusible plug is a safety device with a metal insert having a specified melting point. The allowable melting point is defined by code, but normally it is the saturation temperature of the refrigerant at a pressure no greater than 40% of the ultimate bursting pressure of the refrigerant containing vessel, or the critical temperature of the refrigerant, whichever is lower .

Fusible plugs are limited to units with pressure vessels not exceeding 3 cubic feet internal gross volume. They are used as a safety device in the event of fire, are responsive to temperature only and will not protect against excessively high pressures.

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WATER REGULATING VALVES

On water cooled condensers, a modulating water regulating valve is normally used to economize on water usage and to control condensing pressures within reasonable limits. Water valves may be either pressure or temperature actuated and act to throttle flow as necessary.

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SECTION 8 CONTROL DEVICES, ELECTRICAL

Both electrical and pneumatic controls are widely used for air conditioning and refrigeration system control. Pneumatic controls are primarily used on large central systems, while electric controls are used on applications of all sizes. Since electric controls are used almost exclusively in the commercial refrigeration field, this manual will cover only electric controls.

CONTROL DIFFERENTIAL

The basic function of most electrical control devices is to make or break an electric circuit which in turn controls a contactor, a solenoid coil, or some other functioning part of the system. Controls are available which may make or break a circuit on either a rise or fall in pressure or temperature. The type of action required depends on the function of the control and the medium being controlled.

The point at which a control closes a contact and makes a circuit is called the cut-in point. The point at which the control opens the switch and breaks the circuit is called the cut-out point. The difference between the cut-in and cut-out points is known as the differential.

A very small differential maintains close control but can cause short cycling of the compressor. A large differential will give a longer running cycle, but may result in fluctuations in the pressure or temperature being controlled, so the final operating differential must be a compromise.

The differential may be either fixed or adjustable, depending on the construction of the control. Adjustment of controls varies depending on the type and the manufacturer. On some controls, both the cut-in and cut-out points may be set at the desired points. On many pressure controls, the differential can be adjusted, and this in turn may affect either the cut-in or the cut-out point.

LINE VOLTAGE AND LOW VOLTAGE CONTROLS

Line voltage controls are designed to operate on the same voltage as that supplied to the compressor. Both 110 and 220 volt controls are quite commonly used, and 440 volt controls are available but are seldom used due to the danger from high voltage at the wiring connections.

Local codes often require low voltage controls, and a control circuit transformer may be used to reduce line

voltage to the control circuit voltage, usually 24 volts.

LOW PRESSURE AND HIGH PRESSURE CONTROLS

A low pressure control is actuated by the refrigerant suction pressure, and normally is used to cycle the compressor for capacity control purposes, or as a low limit control. The low pressure control often is used as the only control on small systems which can tolerate some fluctuations in the temperature to be maintained. The standard low pressure control makes on a rise in pressure, and breaks on a fall in pressure.

A high pressure control senses the compressor discharge pressure, and is normally used to stop the compressor in case of excessively high pressures. Since the allowable pressure limit varies with different refrigerants, the proper high pressure control for the refrigerant in the system must be used. A high pressure control makes on a fall in pressure and breaks on a rise in pressure. Either manual reset or automatic reset controls are available, the choice depending on the desired system operation.

Dual pressure controls are comprised of a low pressure and a high pressure control mounted in a single housing with a single switch operated by either control.

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CONDENSER FAN CYCLING CONTROL

In order to maintain air cooled condensing pressures at a satisfactory level during low ambient conditions, a condenser fan pressure control is frequently used. The control acts to break the circuit to the condenser fan on a drop in condensing pressure and makes the circuit on a rise in condensing pressure. Since this is the reverse of the action on a normal high pressure control, this is often described as a reverse acting high pressure control.

THERMOSTATS

A thermostat acts to make or break a circuit in response to a change in temperature. There are numerous types of thermostats ranging from a simple bimetallic switch to multiple switch controls operating from remote sensing bulbs. Thermostats may have a fixed control point or may have variable adjustments.

Normally a cooling thermostat will make on a rise in temperature and break on a fall in temperature, while a heating thermostat will make on a fall in temperature and break on a rise.

OIL PRESSURE SAFETY CONTROL

Special pressure controls have been developed to protect the compressor against loss of oil pressure. The

control is actuated by the difference in pressure between the outlet oil pressure of the oil pump and crankcase pressure. Since the inlet pressure of the oil pump is always crankcase pressure, the net difference in the two pressures is the net lubrication oil pressure.

Oil pressure safety controls are available with both adjustable and non-adjustable control settings, but the non-adjustable type is preferred to avoid difficulties arising from improper field adjustment.

If the oil pressure falls below safe limits, the control breaks to stop the compressor. As an added refinement, a time delay circuit is incorporated to delay the action of the control for a period up to 2 minutes to allow the compressor to establish oil pressure on start-up without nuisance tripping.

TIME CLOCKS

Frequently it is desirable to stop the compressor operation for a period of time to allow defrosting. In order to insure that this is done regularly at convenient times, a time clock can be used to either make or break wiring circuits at preset time intervals. Clocks are available for both 24 hour and 7 day cycles, and the defrost interval and time of initiation and termination can be adjusted as desired.

Various types of defrost control circuits are commonly used, such as time initiated, time terminated; time initiated, temperature terminated; or time initiated, pressure terminated. Normally on circuits with pressure or temperature termination, an overriding time termination is

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provided in the event the defrost cycle for some reason is abnormally prolonged. RELAYS

A relay consists of a set of contacts together with a magnetic coil mechanism which controls the contact position. The contacts may be normally open or normally closed when not energized, and a given relay may have from 1 to 5 or more sets of contacts. When the coil is energized, the contacts reverse their action and make or break various circuits as desired.

A relay may be used to control a large amperage load by means of a pilot circuit, to allow interlocking of controls on separate circuits, or for any application where remote control is required.

Most relays are of the potential type, and are actuated when the coil is energized with the proper voltage.

Current relays are actuated by a sufficient current flowing through the relay coil, and are normally used when it is desirable to make or break a circuit when a large change in current flow occurs. These are used in single phase motor starting circuits, and occasionally in safety circuits.

An impedance relay is similar to a normal potential relay except that the coil is wound so as to create a high resistance to current passage. When wired in parallel with a normal relay, the high impedance (resistance) of the relay will shunt the current to the normal circuit and the impedance coil will be inoperative. If the normal circuit is opened and the current must pass through the impedance relay, the relay coil will be energized and the impedance relay will operate. The voltage drop across the relay coil is so large that other magnetic coils in series with the impedance coil will not operate because of the resulting low voltage. Impedance relays are frequently used for safety lock-out circuits in the event of a motor protector trip.

TIME DELAY RELAY

Some relays are constructed with a time delay action so that the relay must be energized for a predetermined length of time before the magnetic coil can actuate the contacts. The time delay is normally non-adjustable, but relays are available with varying periods of delay.

This type of relay may be required for part winding start motors; in circuits to prevent short cycling, or for other specialized applications.

TRANSFORMERS

A transformer is an electrical device for transferring electrical energy from one circuit to another at a different voltage by means of electromagnetic induction. Transformers are frequently used in control circuits to step voltage down from line voltage to a lower control circuit voltage. There are no moving parts and the action of the transformer is determined by its coil windings.

The transformer output is limited by its size, but transformers are available for almost any output desired from a tiny alarm bell circuit to the giant transformers used on high voltage power transmission lines.

The selection of control circuit transformers can vitally affect the performance and life of many electrical components in a refrigeration or air conditioning system.

An inadequate transformer supplying abnormally low voltage to the control circuit will result in improper operation of contactors and/or motor starters due to chattering or sticking contacts, burned holding coils, or failure of contacts to properly close. Since any of these conditions can cause eventual system failure and possible damage to the compressor, control transformers must be properly sized.

Even though a proper size transformer has been selected, care must be taken to avoid excessive voltage drop in a low voltage control circuit. When using a 24 volt system with a remote thermostat, wire of sufficient current carrying capacity must be installed between the transformer and the thermostat .

According to NEMA standards, a solenoid or contactor must operate satisfactorily at a minimum of 85% of rated voltage. Allowing for a line voltage fluctuation of plus or minus 10% which can occur on electric utility systems, the voltage drop of the transformer and connecting wiring must be limited to 5% to insure a minimum of 85% of rated voltage at the magnetic device.

A transformer works on the magnetic induction principle, has no moving parts, and normally will have a long and trouble free life. However, overloading of a transformer results in excessive temperatures which will cause rapid deterioration of the insulation and eventual failure of the transformer coils.

A control circuit transformer will not overheat, nor will its secondary output voltage drop below 95% of its rated voltage if:

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1. The continuous VA (volt-amperes) capacity of the transformer is equal to or greater than any continuous VA load that can occur in the system.

2. The inrush VA capacity of the transformer is equal to or greater than the maximum VA load that can occur for any combination of sealed and/or inrush load. (A sealed load is the terminology used to describe a component load drawn through closed or sealed contacts, after the inrush current has returned to normal operating conditions.)

Selection Procedure

1. To select the proper transformer, the following information is necessary:

a. A list of all components in the control circuit.

b. The sealed VA and watts for each.

c. The inrush VA and inrush watts for each.

Note: The VA and watts of magnetic devices will vary with each manufacturer, so it is necessary to obtain exact information on the components to be used. Since not all of the inrush data is included in catalogs, it will be necessary to contact representatives of component manufacturers for information when designing control circuits.

2. The continuous VA requirement is determined by combining the sealed VA of all components in the circuit which can be energized at one time.

3. The inrush VA capacity of a transformer is determined by two factors — the VA inrush and the inrush load power factor. Each transformer manufacturer publishes rating charts showing the inrush capacity of each size transformer in terms of the per cent of rated load at varying secondary output voltages and varying power factors. Since output voltages lower than 95% are not acceptable, the only variable to be determined is the power factor.

The maximum VA inrush is found by combining the inrush VA that can occur with the maximum sealed VA that can occur simultaneously. To determine the inrush load power factor, divide the maximum inrush watts by the maximum inrush VA.

Examples Of Transformer Selection

Figure 44 is typical of one manufacturer’s curves showing the inrush capacity of three different transformers at

95% of rated output voltage for varying power factors. To illustrate the selection procedure, assume a control transformer is to be selected from Figure 44 for the following control circuit:

Example No.1 Sealed Sealed Inrush Inrush VA Watts VA Watts 1 - 60 amp contactor 25 7 165 124 1 - oil pressure safety switch 25 25 25 25 1 -10 amp fan starter 8 5 58 12

Assume the compressor is cycling on the contactor, and the oil pressure safety switch heater element is energized whenever the compressor oil pressure is below 15 psig. The fan starter is used to energize a fan circuit at the same time the compressor is energized. Therefore, the contactor and the fan starter would be a continuous load, and the inrush load would be the inrush VA of the contactor and the fan starter, plus the heater load of the oil pressure safety switch. A. Continuous VA requirement Contactor 25 VA Fan Starter 8 VA Total 33 VA

Any transformer with a rating of 33 VA or more would handle the continuous load, so the 60 VA transformer is satisfactory.

B. Inrush VA Requirement VA Watts Contactor 165 124 Fan Starter 58 12 Oil Pressure Safety Switch 25 25 Total 248VA 161 Watts 161 Watts Power factor equals 248 VA or .65

Although the continuous VA requirement could be met with the 60 VA transformer shown on Figure 44 it would require the 140 VA transformer to satisfy the inrush VA requirement, and the larger size must be used.

Example No.2

Assume the same conditions as in Example No.1 except that the compressor has two 60 amp contactors for across the line operation.

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A. Continuous VA requirement 2 Contactors 50 VA Fan Starter 8 VA Total 58 VA

The 60 VA transformer in Figure 44 is adequate for the continuous load.

B. Inrush VA requirement VA Watts 2 Contactors 330 VA 248 Watts Fan Starter 58 VA 12 Watts Oil Pressure Safety Switch 25 VA 25 Watts Total 413 VA 285 Watts

285 Watts Power factor equals 413 VA or .69

Referring to Figure 44 the 140 VA transformer does not have enough capacity for the inrush VA requirement, and the 200 VA transformer must be used.

These examples clearly indicate that the continuous VA requirement of the control circuit may be greatly exceeded by the inrush VA requirement, and both factors must be considered in selecting transformers.

Transformer manufacturers use a standard 20% power factor in their catalog literature for determining the maximum inrush VA rating of a transformer. Since the inrush power factor of contactor coils may be much higher than 20%, and since the resistive load of the oil pressure safety switch will increase the overall power factor, a typical compressor control circuit may have an inrush power factor greatly in excess of 20%. The allowable inrush VA load on a transformer decreases with an increase in the power factor (see Figure 44) and the use of an incorrect power factor may result in an undersized transformer. To properly size a transformer, the power factor must be calculated for the components to be used.

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SECTION 9 MOTORS

Electric motors are used as the power source on the great majority of refrigeration compressors, and practically all are now alternating current (A. C.) motors. The following discussion is limited to those motors of interest for driving refrigeration compressors.

Nearly all motors used for refrigeration applications are induction motors, the name coming from the fact that the current in the moving part of the motor is induced, the moving component having no connection to the source of current. The stationary part of an induction motor is called the stator, and the moving part the rotor. The stator windings are connected to the power source, while the rotor is mounted on the motor shaft, the rotation of the rotor providing the motor driving power source.

MOTOR TEMPERATURE

The first law of thermodynamics stated that energy cannot be either created or destroyed, but may be converted from one form into another. The motor receives electrical energy from the power source, but because of friction and efficiency losses, only a part of this input energy can be turned into mechanical output energy. The balance of the input energy is converted to heat energy, and unless this heat is dissipated, the temperature within the motor windings will rise until the insulation is destroyed. If a motor is kept free from contamination and physical damage, heat is practically the only enemy that can damage the windings.

The amount of heat produced in the motor depends both on the load and on motor efficiency. As the load is increased, the electrical energy input to the motor increases. The percentage of the power input converted to heat in the motor depends on motor efficiency, decreasing with an increase in efficiency, and increasing as efficiency decreases.

The temperature level, which a motor can tolerate, depends largely on the type of motor insulation and the basic motor design, but the actual motor life is determined by the operating conditions to which it is subjected during use. If operated in a proper environment, at loads within its design capabilities, a well designed motor should have an indefinite life. Continuous overloading of a motor resulting in consistently high operating temperatures will materially shorten its life.

OPEN TYPE MOTORS AND BELT DRIVES

Open type motors may be used for compressor drive with either a belt drive or a direct drive arrangement, but with the continued development of hermetic and accessible-hermetic motor-compressor design, the use of open type motors has declined rapidly. Open type motors for compressor drives should be selected conservatively since open type motors do not have the generous overloading safety factor found in most motors used in hermetic and accessible hermetic compressors.

V-belts are made in several different industry classifications, which have been standardized for interchangeability. For refrigeration use, fractional horsepower industrial belts and conventional industrial belts are the two types commonly used. The fractional horsepower belts are more flexible and are well adapted to small radius drives in the smaller horsepower range.

The conventional industrial belts have a higher horsepower rating for heavier loads. Fractional horsepower “4L” and conventional “A” belts have a comparable cross section and may be used on the same size pulley, while fractional horsepower “5L” and conventional “B” belts are similarly interchangeable.

For belt driven compressors, the compressor speed is determined by the size of the motor pulley, since the compressor pulley (flywheel) is normally fixed. The relative speeds of the motor and compressor are in direct relation to the diameter of the motor pulley and the compressor flywheel, but for accurate calculation, the pitch diameter of the pulley must be used rather than the outside pulley diameter (O. D.). The pitch diameter makes allowance for the fact that the V-belt rides partially inside the pulley O. D. For drives using “A” and “4L” section belts, the pitch diameter may be taken as pulley O. D. less 1/4”, and for “B” and “5L” section belt drives, pulley O. D. less 3/8”.

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The desired pulley size or the resulting compressor speed may then be calculated from the following relation:

(Motor Speed, RPM) x (Pitch Diameter, Motor Pulley, In.) Compressor Speed, RPM = (Pitch Diameter, Compressor Pulley, In.)

(Compressor Speed, RPM) x (Pitch Dia., Compressor Pulley, In.) Pitch Diameter, Motor Pulley, In. = (Motor Speed, RPM)

For example, to find the motor pulley diameter required when a 1750 RPM motor is to be used to drive a compressor having an 8” pitch diameter compressor pulley, when the desired compressor speed is 500 RPM, proceed as follows:

500 RPM Com. Speed x 8” P. D., Comp. Pulley P.D., Motor Pulley = 1750 RPM, Motor Speed

= 500 x 8 1750 = Approximately 2 1/4 inches P. D. For an “A” belt drive Motor Pulley O. D. = 2 1/4 inches P.D. + 1/4 inch = 2 1/2 inches O.D.

For a “B” belt drive Motor Pulley O. D. = 2 1/4 inches P.D. + 3/8 inch = 2 5/8 inches O.D.

HERMETIC MOTORS

In hermetic and accessible-hermetic motor-compressors, the motor is mounted directly on the compressor crankshaft, and is hermetically sealed within the compressor body. Aside from the economies inherent in this type of construction, the greatest advantage is that the motor can be cooled by a variety of means, such as air, water, or refrigerant vapor. Sealing the motor in the compressor body eliminates the troublesome problem of sealing the crankshaft so that the power may be transmitted without refrigerant leaks. By designing a motor for the specific application, and controlling motor temperature closely, a motor may be matched to a given load with the result that the motor output can be utilized at its maximum capability, while maintaining a generous safety factor considerably above that available with standard open type motors.

NAMEPLATE AMPERAGE

On open-type motors standard NEMA horsepower ratings are used to identify a motor’s power output capability. Because of industry practice, this nominal horsepower classification has carried over into hermetic motor identification, but it may be misleading when applied to this type of motor. With controlled cooling and motor protection sized for the exact load, a hermetic motor may be operated much closer to its maximum ability, so a given motor may be capable of much greater power output as a hermetic motor than an equivalent open motor. The amperage draw and watts of power required are much better indicators of hermetic motor operation.

Each Copeland® brand motor-compressor carries on its nameplate ratings for both locked rotor and full load amperes. The designation full load amperage persists because of long industry precedent, but in reality a much better term is nameplate amperage. On all welded compressors, on all new motors now being developed for Copelametic® compressors, and on most of the motors developed with inherent protection or internal thermostats, nameplate amperage has been arbitrarily established as 80% of the current drawn when the motor protector trips. The 80% figure is derived from standard industry practice of many years’ standing in sizing motor protective devices at 125 % of the current drawn at rated load conditions.

In order for the motor to meet Emerson Climate Technologies, Inc. standards, the maximum allowable current must be beyond the prescribed operating limits of the compressor, and is determined during qualification tests by operating the compressor at established maximum load conditions and lowering the supply voltage until the protector trip point is reached. Use of the standard 80% factor enables the service and installation engineer to safely size wiring, contactors, or other external line protective devices at 125% of the nameplate rating, since the motor-compressor protector will not allow the amperage to exceed this figure.

In most instances, the motor-compressor is capable of performing at nominal rating conditions at less than rated nameplate amperage. A given motor frequently is used in various compressor models for air-cooled,

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suction-cooled, water-cooled, high temperature, medium temperature, low temperature, R-12, R-22, or R-502 applications as required. Obviously on many applications there will be a greater safety factor than on others.

This does not mean that every compressor may be operated continuously at a load greatly in excess of its nameplate rating without fear of failure. Motor amperage is only one factor in determining a compressor’s operating limitations. Discharge pressure and temperatures, motor cooling, and torque requirements are equally critical. Safe operating limits have been established for each compressor, and are published on compressor specification sheets.

VOLTAGE AND FREQUENCY

Although electric energy distributed in the United States is 60 cycle, distribution voltages are not standardized. Single phase voltages may be 115, 208, 220, 230, or 240, and most utilities reserve the right to vary the supply voltage plus or minus 10% from the nominal rating. Three phase voltage may be 208, 220, 240, 440, 460, or 480, again plus or minus 10%. Unless motors are specifically designed for the voltage range in which they are operated, overheating may result.

Most motors may be operated at the voltage on the motor nameplate, plus or minus 10% without danger of overheating, but in order to allow more flexibility of operation, extended voltage range motors are now being developed where the usage warrants this action. For example, a three phase motor nameplated 208/240 volts is satisfactory for operation from 187 volts to 264 volts, and a three phase motor nameplated 440/480 volts may be operated from 396 volts to 528 volts.

In many parts of the world, the electrical power supply is 50 cycle rather that 60 cycle. If both voltage and frequency supplied to a motor vary at the same rate, operation of a given motor at the lower frequency condition within narrow limits is satisfactory in some cases. For example, a 440 volt, 3 phase, 60 cycle motor will operate satisfactorily on 380 volt, 3 phase, 50 cycle power supply.

However, in some 50 cycle applications, the relationship between voltage and frequency is such that standard 60 cycle motors cannot be used unless the motor characteristics are suitable for 50 cycle operation, and this can be determined only by test. On most single phase 50 cycle applications, specially wound 50 cycle motors are required.

THREE PHASE MOTORS

Three phase motors are wound with 3 separate windings or phases. Each of the windings is 120° out of phase with the other windings, which results in a very high starting torque motor requiring no supplemental mechanisms or devices for starting. The direction of rotation of the motor may be changed by reversing any two line connections.

Because three phase motors can utilize smaller wire sizes and therefore are smaller, they are used on almost all applications larger than 5 HP, and if three phase power is available, are frequently preferred for any load larger than fractional horsepower applications.

SINGLE PHASE MOTORS

A single phase motor has but one running winding or phase, and basically is not a self starting motor. Once started, it will run as a pure induction motor. In order to provide starting torque, a second winding called the starting winding is provided, which normally has greater resistance than the running winding. The different variations of single phase motors are primarily due to the different starting arrangements used.

If the start winding remains in the circuit during operation it would be damaged by excessive heat. Therefore, the starting winding is removed from the circuit as the motor approaches rated speed by either a potential relay, a current relay, or a centrifugal switch.

A current relay is normally open when de-energized, and the coil is wound so that the contacts will close when starting current is being drawn by the motor, but will drop out when the current approaches normal full load conditions. Therefore, the current relay is closed only during the starting cycle.

A potential relay is normally closed when de-energized, and the coil is designed to open the contacts only when sufficient voltage is generated by the start winding. Since the voltage or back-EMF generated by the start winding is proportional to the motor speed, the relay will open only when the motor has started and is approaching normal running speed. The illustrations show the schematic wiring with the motor in operation so the potential relay is in the energized position.

SPLIT PHASE MOTORS

On a split phase motor, the running winding and start winding are in parallel, and are spaced 90° apart. The

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run winding is wound with relatively heavy wire, while the start winding is wound with fine wire and has a much greater resistance than the run winding. The combination of greater resistance and physical displacement results in the start winding being slightly out of phase with the run winding, and thus produces sufficient magnetic force to cause the rotor to rotate.

Figure 45 illustrates a split phase motor schematic electrical diagram with a current relay to break the start winding circuit when the compressor has reached operating speed.

The starting torque of a split phase motor is low, the starting current is high, and the efficiency is relatively

low. As a result, this type of motor is generally limited to capillary tube systems in small fractional HP sizes.

CAPACITOR START - INDUCTION RUN MOTORS (CSIR)

This motor is similar to the split phase motor in construction except for the method used to obtain the phase displacement necessary for starting. In the split phase motor, the phase displacement is due to a higher resistance in the start winding. In the capacitor start motor, the necessary phase displacement is achieved through the use of a capacitor connected in series with the starting winding. A capacitor start-induction run schematic electric diagram is shown in Figure 46.

The starting winding is removed from the circuit after the motor is started by the potential relay or current relay as before. This type motor has a high starting torque, and therefore is suitable for applications where unequal pressures may be encountered on start up. Because of

the low power factor of this type motor, its use is usually limited to fractional horsepower applications.

CAPACITOR START - CAPACITOR RUN MOTORS (CSR)

By connecting a running capacitor in parallel with the starting capacitor (R to S terminals) as shown in Figure 47 the motor is strengthened because the start winding is loaded in phase with the main winding after the start capacitor is disconnected, which permits the starting winding to carry part of the running load. The running capacitor strengthens the motor, improves the power factor, reduces motor current, increases the efficiency, and decreases the temperature of the motor under design conditions. However, the motor must be designed for operation with a run capacitor, and a capacitor start-induction run motor usually is not suitable for conversion to capacitor start-capacitor run operation.

Normally current relays are not recommended for use with capacitor start-capacitor run motors because of the danger of the running capacitor discharging to the

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start capacitor through the current relay when it closes. The high voltage built up on the running capacitor can cause the current relay contacts to arc, possibly welding the contacts and causing compressor failure. If a current relay is used with a capacitor start-capacitor run motor, a resistor should be installed between the running capacitor and start capacitor to prevent a high current flow to the start capacitor on start-up. This condition does not occur on systems equipped with potential relays since the contacts are normally closed at start-up, and the voltage build-up on both start and run capacitors is similar.

The run capacitor is in the circuit continuously and is designed for continuous operation whereas the start capacitor is used only momentarily each time the motor starts and is designed for intermittent duty.

Capacitor start-capacitor run motors have high efficiency, high power factor, and high starting torque and are used in single phase motors from fractional HP through 5 HP in size.

PERMANENT SPLIT CAPACITOR MOTORS (PSC)

For some applications not requiring high starting torque, a motor with only a running capacitor is desirable. Because of the elimination of the start capacitor and

relay, this type of motor is economical and efficient, but has low starting torque. Its use is limited to systems on which pressures are equalized prior to start up, and is primarily used in air conditioning and small commercial applications.

A permanent split capacitor motor schematic electrical diagram is illustrated in figure 48. It is identical to the capacitor start-capacitor run diagram without the starting capacitor and relay. If increased starting torque is required, a starting capacitor and relay assembly can be added to this motor, making it identical to a capacitor start-capacitor run motor in operation.

DUAL VOLTAGE MOTORS

Certain Copeland® brand three phase motors are wound with two identical stator windings which are connected in parallel on 208 or 220 volt operation, and in series for 440/480 volt operation. Internal connections of this type of motor are shown in Figure 49.

These models have two parallel windings with nine leads which must be connected correctly for the voltage of the power supply. If the windings are connected out of phase, or if the jumper bars are not positioned correctly, motor overheating and possible failure can occur.

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TWO PHASE MOTORS

Two phase power is still used in a few isolated areas, and specially wound two-phase motors are required for use on this type of power supply. These motors have two parallel windings, and are similar to three phase motors in their operation. Capacitors and starting relays are not required. The motor is started directly across the line by means of a special 4 pole contactor . The phase windings are connected in parallel from the two phase three or four wire power supply.

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SECTION 10 STARTING EQUIPMENT AND MOTOR PROTECTORS

Since hermetic motors must operate under a wide range of operating conditions, and vary in size from fractional horsepower to 35 HP and larger, a wide variety of starting equipment is used.

CONTACTORS AND STARTERS

A contactor is a load current carrying device, which makes and breaks to start and stop the compressor motor. A starter is merely a contactor with motor protective devices mounted in a common enclosure.

On single phase motors up to 3 HP in size, the motor current is often low enough to be handled by the contacts in the low pressure control or thermostat, and no separate contactor is required. As the motor size increases, the amperage draw increases beyond the range of small control apparatus, and the motor current must be handled through the contacts of a starter or contactor, while the control makes and breaks a pilot circuit which energizes the coil of the contactor .

For motor-compressors whose power requirements are such that contactors are required, it is essential that the contactors used are adequately sized for the attached load. The rating of the contactor for both full load amperes and locked rotor amperes must be greater than the nameplate rating of the motor-compressor plus the nameplate rating of any fans or other accessories also operated through the contactor .

NEMA general purpose type contactors are built for the most severe industrial usage, and are designed for a minimum life of 2,000,000 cycles. Because they must be adaptable to any usage, general purpose contactors have a large safety factor, and as a result are both large and costly. For refrigeration and air conditioning applications, a life of 250,000 cycles is entirely adequate, so the physical construction can be lighter, and the cost of the contactor correspondingly less.

To meet the specific needs of the refrigeration and air conditioning industry, electrical equipment manufacturers have developed definite purpose contactors. These contactors are rated in amperes, and when selected properly for the load, are smaller and more economical than the general purpose contactor. Since compressor contactors are frequently subjected to quick recycling, the contacts must be large enough for satisfactory heat dissipation in order to prevent contactor overheating.

Overheating of the contacts may cause sticking and single phasing, and can cause a motor failure even though the motor overload protectors trip and open the control circuit.

In order to insure that definite purpose contactors are properly applied to Copeland® brand motor-compressors with pilot circuit protection, the contactor must meet Emerson Climate Technologies, Inc. minimum performance requirements.

CAPACITORS

An electrical capacitor is a device which stores electrical energy. They are used in electric motors primarily to displace the phase of the current passing through the start winding. While a detailed study of electrical theory is beyond the scope of this manual, capacitors in a motor circuit provide starting torque, improve running characteristics and efficiency, and improve the power factor.

The amount of electrical energy a capacitor will hold depends on the voltage applied. If the voltage is increased, the amount of electrical energy stored in the capacitor is increased. The capacity of a capacitor is expressed in microfarads (MFD) and is dependent on the size and construction of the capacitor.

The voltage rating of a capacitor indicates the nominal voltage at which it is designed to operate. Use of a capacitor at voltages below its rating will do no harm. Run capacitors must not be subjected to voltages exceeding 110% of the nominal rating, and start capacitors must not be subjected to voltages exceeding 130% of the nominal rating. The voltage to which a capacitor is subjected is not line voltage, but is a much higher potential (often called electromotive force or back EMF) which is generated in the start winding. On a typical 230 volt motor, the generated voltage may be as high as 450 volts, and is determined by the start winding characteristics, the compressor speed, and the applied voltage.

Capacitors, either start or run, can be connected either in series or parallel to provide the desired characteristics if the voltage and MFD are properly selected. When two capacitors having the same MFD rating are connected in series, the resulting total capacitance will be one half the rated capacitance of a single capacitor. The formula for determining capacitance (MFD) when capacitors are connected in series is as follows:

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1 = 1 + 1 MFD total MFD1 MFD2 For example, if a 20 MFD and a 30 MFD capacitor are connected in series, the resultant capacitance will be

1 1 1 MFDt = MFDl + MFD2

1 1 1 MFDt = 20 + 30

1 5 1 MFDt = 60 = 12

MFDt = 12 MFD

The voltage rating of similar capacitors connected in series is equal to the sum of the voltage of the two capacitors. However, since the voltage across individual capacitors in series will vary with the rating of the capacitor, for emergency field replacements it is recommended that only capacitors of like voltage and capacitance be connected in series to avoid the possibility of damage due to voltage beyond the capacitor limits.

When capacitors are connected in parallel, their MFD rating is equal to the sum of the individual ratings. The voltage rating is equal to the smallest rating of the individual capacitors.

It is possible to use any combination of single, series, or parallel starting capacitors, with single or parallel running capacitors (running capacitors are seldom used in series).

START CAPACITORS

Start capacitors are designed for intermittent service only, and have a high MFD rating. Their construction is of the electrolytic type in order to obtain the high capacity.

All standard Copeland® brand starting-capacitors are supplied with bleed-resistors securely attached and soldered to their terminals as shown in Fig. 51.

The use of capacitors without these resistors probably will result in sticking relay contacts and/or erratic relay operation – especially where short cycling is likely to occur.

This is due to the starting capacitor discharging through the relay contacts as they close, following a very short running cycle. The resistor will permit the capacitor charge to bleed down at a much faster rate, preventing arcing and overheating of the relay contacts.

The use of capacitors supplied by Emerson Climate Technologies, Inc. is recommended, but in case of an emergency exchange, a 15,000 -18,000 ohm, two watt resistor should be soldered across the terminals of each starting capacitor. Care should be taken to prevent their shorting to the case or other nearby metallic objects.

If sticking contacts are encountered on any starting relay the first item to check is the starting capacitor resistors. If damaged, or not provided, install new resistors, and clean the relay contacts or replace the relay.

Suitable resistors can be obtained from any radio parts wholesaler.

RUN CAPACITORS

Run capacitors are continuously in the operating circuit, and are normally of the oil filled type. The run capacitor capacitance rating is much lower than a start capacitor. Because of the voltage generated in the motor start winding, the run capacitor has a voltage across its terminals greater than line voltage.

The starting winding of a motor can be damaged by a shorted and grounded running capacitor. This damage usually can be avoided by proper connection of the running capacitor terminals.

The terminal connected to the outer foil (nearest the can) is the one most likely to short to the can and be grounded in the event of a capacitor breakdown. It is

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identified and marked by most manufacturers of running capacitors. See Fig. 52.

From the supply line on a typical 115 or 230 volt circuit, a 115 volt potential exists from the “R” terminal to ground through a possible short in the capacitor. (See wiring diagram Fig. 53.) However, from the “S” or start terminal, a much higher potential, possibly as high as 400 volts, exists because of the EMF generated in the start winding. Therefore, the possibility of capacitor failure is much greater when the identified terminal is connected to the “S” or start terminal.

THE IDENTIFIED TERMINAL SHOULD ALWAYS BE CONNECTED TO THE SUPPLY LINE, OR “R” TERMINAL, NEVER TO THE “S” TERMINAL. This applies to PSC as well as capacitor-start, capacitor-run motors.

If connected in this manner, a shorted and grounded running-capacitor will result in a direct short to ground from the “R” terminal and will blow line fuse No.1. The motor protector will protect the main winding from excessive temperature.

If, however, the shorted and grounded terminal is connected to the start winding terminal “S”, current will flow from the supply line through the main winding and through the start winding to ground. Even though the protector may trip, current will continue to flow through the start winding to ground, resulting in a continuing temperature rise and failure of the starting winding.

REDUCED VOLTAGE STARTING

Full voltage “across-the-line” starting is the least expensive way to start a three-phase motor, and all motors in Copeland® brand compressors are designed

for full voltage starting. However, due to power company limitations on starting current, some means of reducing the inrush starting current on larger horsepower motors is occasionally necessary. This is particularly true in other countries, and is necessary in some sections of the United States and Canada. The principal reason for these restrictions is to prevent light flicker, television interference, and undesirable side-effects on other equipment because of the momentary voltage dip. The reduced voltage start allows the power company voltage regulator to pick up the line voltage after part of the load is imposed, and thus avoids the sharper voltage dip that would occur if the whole load were thrown across the line.

Some electrical utilities may limit the inrush current drawn from their lines to a given amount for a specified period of time. Others may limit the current drawn on start-up to a given percent of locked rotor current.

Unloading the compressor can be helpful in reducing the starting and pull-up torque requirement, and will enable the motor to accelerate quickly. But regardless of whether the compressor is loaded or unloaded, the motor will still draw full starting amperage for a small fraction of a second. Since the principal objection usually is to the momentary inrush current drawn under locked rotor conditions when starting, unloading the compressor will not always solve the problem. In such cases some type of starting arrangement is necessary that will reduce the starting current requirement of the motor.

Starters to accomplish this are commonly known as reduced voltage starters, although in two of the most common methods the line voltage to the motor is not actually reduced. Since manual starting is not feasible for refrigeration compressors, the only type of starters to be considered are magnetic.

There are five types of magnetic reduced voltage starters, each of which has certain characteristics, which are desirable for specific applications.

1. Part winding 2. Star-Delta 3. Autotransformer 4. Primary Resistor 5. Reduced voltage step starting accessory

As the starting current is decreased, the starting torque also drops, and the selection of the starter to be used may be limited by the compressor torque requirement. The maximum torque available with reduced voltage starting is 64% of full-voltage torque, which can be obtained with an autotransformer starter, while part winding starters deliver approximately 45% of full-

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voltage torque, and star-delta starters only 33%. For Copeland® brand motor-compressors without unloaders, a starting torque of 45% of full-voltage torque or higher is recommended. The use of an unloaded start is helpful in critical applications, and for low torque starting such as encountered in star-delta starters, an unloaded start is essential if the compressor is to start under reduced voltage conditions.

However, it is not necessary for the motor-compressor to start and accelerate under the reduced voltage phase of starting to accomplish the objective of reducing the peak starting current. By energizing the motor a step at a time, the power company requirements may be satisfied, while at the same time by keeping the time delay between steps in starting to a minimum, damage to the compressor can be avoided. It is of course desirable for the compressor motor to start and accelerate under

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reduced voltage starting conditions to realize the major benefits of this type of starting arrangement.

1. Part Winding Start

This is not a true reduced voltage start, but it accomplishes the same job - limiting inrush current - by utilizing only part of the motor windings. Since it uses both the start and run contactors to carry the motor current during operation, it costs less than the other types.

To utilize part winding start, the motor must have a dual winding. Copeland® brand 208/220/440 volt, 3 phase motors are wound with two identical stator windings which are connected in parallel on 208 or 220 volt operation. For part winding start, the first step utilizes only one winding or 1/2 of the stator, and these motors may be used whenever part winding start is required on 208 or 220 volt power. Part winding start cannot be used on these motors when used on 440 volts, since the entire winding must be connected in series for 440 volt power .

Copelametic® model 4R and 6R compressors are currently available with dual wound motors, and some 4R and 6R models are available with specially wound motors for part winding start on 550 volts.

Basically all that is required for part winding start are two contactors, each capable of carrying the winding’s full load and locked rotor current requirement, with a time delay between the contactors. When the starter is energized, the first magnetic contactor closes and puts half of the motor winding across the line. A preset time delay relay is energized at the same time, and at the completion of the timing cycle, the second magnetic contactor closes and puts the second half of the motor winding in parallel with the first.

The normal Copeland® brand compressor motor protectors must be used. Where current sensing protectors are required, they must be installed in at least two phases of each contactor. Motors equipped with a Thermotector need no other external protection. To prevent tripping of the protectors during starting, the time delay between the first and second contactor must be well within the protector’s tolerance for locked rotor conditions, and a time delay device having a time cycle setting of one second ± 10% is required.

The exact current and torque characteristics of a motor will vary with design. For Copeland® brand motor-compressors starting on one winding, the motor will draw approximately 65% of the normal across-the-line starting current, and produce approximately 45% of the normal starting torque. Under heavily loaded conditions,

it is possible that the motor may not start until the second winding is energized, or if it does start, it may not accelerate. An unloaded start may be desirable under extreme conditions.

On part winding start applications, occasionally an electrical starting noise or “growl” of short duration may be noticed. This occurs when the first half winding starts the motor, but is unable to accelerate it beyond a few hundred rpm. As soon as the second winding is energized, the motor instantly accelerates, and the noise disappears. Since the time delay between windings is no greater than one second, the noise duration is very short.

The noise will vary with voltage, speed, pressure differential, motor horsepower, and will vary slightly from compressor to compressor. In addition, motors supplied by different sources may have slight differences in motor characteristics, and the resulting sound may be slightly different.

Occasionally service personnel mistake the starting noise for bearing drag. The starting noise is quite normal, will be more pronounced on larger motors, and does not harm the compressor In any way.

2. Star-Delta Starting

For star-delta starting (also referred to as wye-delta) a specially wound motor is required with both ends of each phase winding brought out to terminals. By means of contactors a motor designed for normal operation in delta is first connected in star, and after a predetermined time delay, the star connection is changed to delta. This starting arrangement is relatively simple and inexpensive, and is widely used in Europe.

Recently three phase, 50 cycle motors have been developed for most Copeland® brand compressors 7½ HP and larger specially wound for star-delta starting. Leads were brought out from both ends of each phase so that the motors could be connected in either star or delta. Motors are available for star-delta start connections on either 380 volt, 50 cycle, 3 phase or 220 volt, 50 cycle, 3 phase circuits.

When a motor designed for delta operation is connected in star, the voltage across each phase is reduced to 58% of normal, and the motor develops 1/3 of the normal starting torque. The inrush current in star is 1/3 of normal inrush current in delta.

Star-delta starting is suitable for low torque starting duty only. To insure starting on the star connection, some means of pressure equalization across the compressor

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prior to starting is essential. Unloading the compressor during the starting phase is also recommended.

In order to eliminate the objectionable flicker or jump in current occurring during the change from star to delta, closed transition starters employ an additional contactor and three resistors which are utilized to keep the motor connected to power through the resistors during the transition period. Since the transition period is less than 1/10 of a second in duration, the resistors can be relatively small. Closed transition starting is recommended to prevent objectionable surges of current.

Since the relation between line current and phase current will vary with the switch from star to delta connections, motor overload protectors must be mounted in the motor winding circuit. Compressors equipped with Thermotectors (fast response thermostats) require no additional external line protection since the Thermotector protects by sensing an increase in motor winding temperature. For compressors requiring external current sensing protectors, specially sized protectors are required, and the Emerson Climate Technologies, Inc. Application Engineering Department should be contacted for specifications.

3. Autotransformer Starters

Autotransformer type starters reduce the voltage across the motor terminals during the starting and accelerating period by first connecting the motor to taps on the transformer, and then after a time delay, switching the motor connection across the line.

Due to the lower starting voltage, the motor will draw less current and will develop less torque than if the motor were connected across the line.

Because of the transformer action, the current in the motor windings is actually greater than the line current by a proportion equal to the ratio of transformation, after allowing for the autotransformer excitation current. This results in a very flexible control system, since the starting current inrush can be effectively limited as desired, while the starting torque per ampere of the line current is the maximum available from any reduced voltage starter. The autotransformer starter is the most complex and the most expensive of the reduced voltage starters, but if high starting torque is required, it often is the only type that will perform acceptably.

Taps are provided on the transformer for various stages of voltage reduction, with reductions of 80% and 65% of full line voltage normally available on most models. Closed circuit transition is recommended to prevent high transient current when transferring from “start” to “run” conditions.

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Regular Copeland® brand motor protectors can be mounted in the leads to the compressor, or adequate protection is provided by inherent protectors. Since no special winding is required in the motor, the autotransformer starter can be used with any Copelametic® compressor. As with other types, the time delay must be very brief to avoid tripping the protector during the starting process.

4. Primary Resistor Starters

In many respects, the primary resistor starter is similar to the autotransformer type. The motor is connected to line current through heavy resistors during the initial starting step in order to reduce the voltage applied to the motor windings. After a time delay, the resistors are shorted out of the circuit, and the compressor is connected across the line.

However, since starting torque is proportional to the square of the voltage applied to the motor, the starting torque falls off rapidly with the reduction of applied motor voltage. The resistors act to prevent current surges, and provide smooth acceleration of the motor once starting is accomplished, since the voltage drop across the resistors decreases as the motor comes up to speed and the inrush starting current decreases.

No special motor windings are required, and regular Copeland® brand motor protectors can be used. As with the autotransformer starter, this system can be applied to any Copelametic® motor. The time delay must be limited to avoid tripping during starting.

5. Reduced Voltage Step Starting Accessory

The reduced voltage accessory panel was developed primarily as a low cost, special purpose auxiliary to solve the problem of light flicker caused by air conditioners of 3 HP and larger on single phase power lines. The cost of special transformers and additional equipment for the power companies made some type of voltage limitation device essential if service was to be continued to the large single phase loads. Basically the reduced voltage step starting accessory operates on the same principle as the primary resistor starter, but it is moderate in cost, is used in conjunction with the regular contactor, and is designed for consumer applications rather than industrial use.

The accessory inserts a resistance in series with the motor for approximately two seconds, after which a timing relay energizes a contactor and shorts the resistors out of the circuit. The resulting torque is low, and the motor quite possibly will not start with the resistance in the circuit, but the result is to break the inrush current into

two steps which reduces light flicker to a level which is not objectionable.

MOTOR PROTECTION

Since hermetic motors may have to handle great variations in load for extended periods, close tolerance protection must be provided to protect the motor in the event of an overload. Standard heater coils in general purpose starters do not trip fast enough to protect the motor under locked rotor conditions. Although fast trip heater coils were developed to give faster response, their variation due to ambient temperature changes makes them undependable under field conditions. Therefore specialized types of motor protection have been developed for refrigeration motor-compressors.

In the event the compressor fails to start, and an internal protector or thermostat trips, disconnecting the motor, it will normally reset very quickly after the initial trip. If several protector trips occur in succession, especially when the motor is very hot from operation at heavily loaded conditions, the motor temperature will rise to a point exceeding the protector setting, and an off period varying from 20 minutes to one hour may be required for the compressor motor to cool sufficiently so that the protector may reset. When this occurs, particularly on across-the-line internally sealed protectors, service personnel frequently assume that the motor has been damaged and is inoperative, when in reality the motor protection system is performing its intended duty. In the event a compressor is checked and found to be very hot and inoperative, allow at least one hour for the motor to cool, and recheck after the cooling period before changing the compressor .

Motor protection may be either of the line break or pilot circuit type. A line break protector incorporates contacts which actually open the line directly when the protector trips. A pilot circuit protector takes the motor off the line indirectly by opening the holding coil circuit of the contactor, but the compressor protection is still dependent on the contactor, since the compressor may be subject to damage in case the contacts of a contactor or starter have stuck or welded, despite the fact that the pilot circuit protector may open.

INTERNAL INHERENT LINE BREAK PROTECTOR

An internal inherent line break protector is a device carrying full load current, responsive both to current and/or temperature, which breaks line current if safe limits are exceeded.

For three phase motors, the internal inherent protector is connected in the center of a wye wound motor. It is

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located within the motor compartment on the motor winding, and there are no external connections. Because of its location, the protector is sensitive to temperature as well as current. When the protector opens, it breaks all three phases of the motor winding. Since this across-the-line device provides both over-current and locked rotor protection, a contactor may be used instead of a motor starter. Internal inherent protectors are among the best protection systems now available for hermetic and accessible-hermetic motor-compressors, but because of the size of the device required, in larger motors, and also because their usage is limited to single voltage motors, their use is currently restricted to 7½ HP motors and smaller.

Single phase internal inherent protectors usually consist of contacts mounted on a bimetal disc which is sensitive to both current passing through the protector and heat generated by the motor windings. They carry and break full line current in the same manner as three phase protectors, and have proved to be most satisfactory.

EXTERNAL INHERENT PROTECTOR

The external inherent protector is similar in construction and operation to the internal inherent protector, but the external protector is mounted on the compressor body and senses motor current and compressor body heat rather than motor winding heat. Because the external protector is not subjected to refrigerant pressure, its case is not hermetically sealed as is that of the internal inherent protector.

INTERNAL THERMOSTATS

On some motor-compressors, particularly larger horsepower sizes where inherent protectors cannot be used, internal thermostats are located in the motor winding. These are pilot circuit devices only, and react only to motor winding heat. When overheated they open the control circuit, thereby stopping the compressor. These thermostats cannot be replaced in the field and are protected against excessive current in the control circuit by fuses.

Because the temperature rise in motor windings during locked rotor conditions is both rapid and uneven, the thermostat often lags behind the winding temperature, and therefore some additional approved protective device is necessary to protect the compressor motor against locked rotor conditions.

EXTERNAL THERMOSTATS

On some older models of Copelametic® motor-compressors, an external thermostat is clamped to the

motor housing to indirectly sense motor temperature. This is a pilot circuit device, and is similar to the internal thermostat in operation. Its sensitivity to temperature is reduced, and consequently the protection provided is not as good as that of the internal type. External thermostats are no longer used by Emerson Climate Technologies, Inc. on current production.

CURRENT SENSITIVE PROTECTORS

External current sensing motor protectors are used in conjunction with internal thermostats to provide close tolerance locked rotor protection. They may be either thermal or magnetic in operation, carry full motor current, and are responsive to the current drawn by the motor. Normally these devices act to break the pilot circuit in the event of a motor overload, but calibrated circuit breaker types are available, which will break the line current to the compressor.

THERMOTECTOR

The Thermotector is a quick reacting thermostat imbedded in the motor windings which senses motor temperature. Its current carrying capacity limits its use to pilot circuit protection, but because of its fast response, it provides protection against overheating under locked rotor as well as running conditions. Therefore it can be used with a contactor without external current sensing protective devices, resulting in a simplified control circuit.

SOLID STATE PROTECTORS

Various electronic solid state devices are now under development for use in motor protection, and it is probable their usage will increase. The control system design will vary, but normally the sensing device is a temperature sensing element mounted on the motor windings in which the resistance changes with a change in motor temperature. The change in resistance when amplified by other solid state components acts to make or break the pilot circuit. As in the case with the Thermotector the quick response provides both running and locked rotor current protection.

FUSES AND CIRCUIT BREAKERS

On air conditioners having motor compressors with PSC motors, it is possible that nuisance tripping of household type circuit breakers may occur. PSC motors have very low starting torque, and if pressures are not equalized at start up, the motor may require several seconds to start and accelerate.

This is most apt to occur where a short cycle of the compressor can be caused by the thermostat making

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contact prematurely due to shock or vibration. Typically this can occur where the thermostat is wall mounted and can be jarred by the slamming of a door.

U. L. and most electrical inspection agencies now require that hermetic type refrigeration motor-compressors must comply with the National Electric Code maximum fuse sizing requirement. This establishes the maximum fuse size at 225% of the motor full load current, and by definition the motor-compressor nameplate amperage is considered full load current, unless this rating is superseded by another on the unit nameplate.

Since the motor protector may take up to 17 seconds to trip if the compressor fails to start, it is probable that a standard type fuse or circuit breaker sized on the basis of 225% of full load current may break the circuit prior to the compressor protector trip, since locked rotor current of the motor may be from 400% to 500% of nameplate amperage.

To avoid nuisance tripping, Emerson Climate Technologies, Inc. recommends that air conditioners with PSC motors be installed with branch circuit fuses or circuit breakers sized as closely as possible to the 225% maximum limitation, the fuse or circuit breaker to be of the time delay type with a capability of withstanding motor locked rotor current for a minimum of 17 seconds.

EFFECT OF UNBALANCED VOLTAGE AND CURRENT ON THREE PHASE MOTOR PROTECTION

When external current sensing motor protectors are used to protect a three phase motor-compressor against excessive current draw and resulting motor overheating, unbalanced motor currents can seriously affect the motor protection system. While it is generally recognized that a break in one phase of a three phase distribution system can result in excessive amperage draw because of the resulting single phasing condition, another and equally serious hazard is the effect on amperage of an unbalanced voltage in the power supply.

If single phasing occurs, the motor may stall unless lightly loaded, and once stopped, it will not start, resulting in locked rotor amperage draw. Under unbalanced voltage conditions, however, the motor will continue to operate, and motor protection may be dependent on the ability of the protectors to sense the abnormally high running current or the increase in the motor temperature.

A properly wound three phase motor connected to a supply source in which the voltages in each phase are balanced at all times will have identical currents in all three phases. The differences in motor windings in

modern motors are normally so small that the effect on amperage draw is negligible. Under an ideal condition, if the phase voltages were always equal, a single motor protector in just one line would adequately protect the motor against damage due to excessive amperage draw. As a practical matter, balanced supply voltages are not always maintained, so the three line currents will not always be equal.

Inherent line break motor protectors mounted at the center of the wye on wye wound motors provide protection against all forms of voltage variation. However, on larger motors the size of the protector makes inherent protectors impractical, and many larger models of Copeland® brand compressors have a combination pilot circuit protection system, consisting of internal thermostats and external current sensing protectors. Because internal thermostats are somewhat slow in reaction, and lag behind the actual motor temperature in the event of a fast temperature rise, locked rotor protection is provided by the external protector. Since in most cases adequate protection can be provided by two leg current sensing protection, and because of the size and cost of external protectors, the majority of compressors are installed with two leg protection, although the third protector can be supplied if desired. In order to determine if the motor will be adequately protected under various abnormal conditions, an understanding of the inter-relation of current and unbalanced voltage is essential.

When line voltages applied to a three phase induction motor are not the same, unbalanced currents will flow in the stator windings. The effect of unbalanced voltages is equivalent to the introduction of a “negative sequence voltage” which is exerting a force opposite to that created

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with balanced voltages. These opposing forces will produce currents in the windings greatly in excess of those present under balanced voltage conditions.

Voltage unbalance is calculated as follows:

% Voltage Unbalance = 100 x Max. Voltage Deviation from Average Volt. Average Voltage

For example, in Figure 57, assume voltage AB is 220 volts, BC is 230 volts, and AC is 216 volts.

Average Voltage = 216 + 220 + 230 3 = 222 Volts

Maximum Deviation = 230 - 222 = 8 Volts

100 X 8 % Voltage Unbalance = 222 = 3.6%

As a result of the voltage unbalance, the locked rotor current will be unbalanced to the same degree. However, the unbalance in load currents at normal operating speed may be from 4 to 10 times the voltage unbalance, depending on the load. With the 3.6% voltage unbalance in the previous example, load current in one phase might be as much as 30% greater than the average line current being drawn by the other two phases.

The NEMA Motors and Generators Standards Publication states that the percentage increase in temperature rise in a phase winding resulting from voltage unbalance will be approximately two times the square of the voltage unbalance.

% Increase in Temperature = 2 (Voltage Unbalance %)²

Using the voltage unbalance from the previous example, the % increase in temperature can be estimated as follows:

% Increase in Temperature = 2(3.6 x 3.6) = 25.9%

As a result of this condition, it is possible that one phase winding in a motor may be overheated while the other two have temperatures within normal limits. If only two motor protectors are being used, and the high current winding is not protected, ultimate motor failure may occur even though the protectors do not trip. Therefore, when installing external motor protectors for a motor in which only two of the three phases are to be protected, be sure the protectors are mounted in the phases with

the highest amperage draw.

A common source of unbalanced voltage on a three phase circuit is the presence of a single phase load between two of the three phases. (See Figure 58).

A large unbalanced single phase load, for example a lighting circuit, can easily cause sufficient variations in motor currents to endanger the motor. If at all possible, this condition should be corrected by shifting the single phase load as necessary. Supply voltages should be evenly balanced as closely as can be read on a commercial voltmeter.

A recent national survey by U.L. indicated that 36 out of 83 utilities surveyed, or 43%, allowed voltage unbalance in excess of 3%, and 30% allowed voltage unbalance of 5% or higher.

In the event of a supply voltage unbalance, the power company should be notified of such unbalance to determine if the situation can be corrected.

Unless the unbalance can be corrected, the only way to insure motor safety is to be sure the protectors are mounted in the high current phases when using two leg protection, or to use protectors in all three legs.

A simple single phase condition in the load circuit will cause the current in two of the three phases to increase, while there will be no current in the open phase.

A motor can be protected against this type of failure with only two protectors, since there will always be at least one protector in a line carrying the high single phase current.

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The effect of an open phase in the primary circuit of a power transformer depends on the type of transformer connection. Where both primary and secondary windings are connected in the same fashion, wye-wye or delta-delta, a fault in one phase of the primary will result in a low current in one phase of the secondary, and high currents in two phases, with results similar to the simple load circuit single phase condition.

But in wye-delta or delta-wye connected power transformers, an open circuit or single phase on the primary side of the transformer will result in a high current in only one phase of the motor with low currents in the other two phases.

Under locked rotor conditions, the high phase will draw an amperage slightly less than nameplate locked rotor current, while the other two legs will each draw approximately 50% of that amount. Under operating conditions, the current in the high phase could be in

excess of 200% of full load amperes, depending on load, while the current in the other two legs will be slightly greater than normal full load amperes.

Since the majority of power systems now use wye-delta or delta-wye transformer connections, occasional faults of this type can be expected.

Most commercial power systems are quite reliable, and if a primary single phase condition does occur in a system where the compressor has only two leg protection, the

motor would still be protected should the high current fall in either of the two protected legs. Therefore, the statistical chances of failure are small, so normally no special provisions for this type of protection are provided unless required by code or regulatory bodies. However, if a primary single phase fault occurs, it may last for several hours, and motors not adequately protected are quite likely to fail.

Most of the three phase protections systems used by Emerson Climate Technologies, Inc. will protect

against primary single phasing, but if there is any question concerning an application, the matter should be referred to the Emerson Climate Technologies, Inc. Application Engineering Department. However, where unbalanced supply voltages and single phase loads represent a continuing threat to three phase motor life, it is recommended that inherently protected motors be used, or if inherent protection is not available because of motor size, the motor-compressor should be equipped with 3 leg protection.

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SECTION 11ACCESSORIES

A number of accessory items are used in refrigeration circuits for specific purposes, and their requirement in a particular system depends on the application.

RECEIVERS

A receiver is primarily a liquid storage tank for refrigerant which is not in circulation. Small package systems utilizing capillary tubes for liquid refrigerant feed may have very small refrigerant charges, and if the operating load is fairly constant, careful design of the evaporator and condenser may allow the elimination of the receiver from the system. If the condenser has volume enough to provide storage space, a separate receiver is not required, and this is common design practice in water-cooled units with shell and tube condensers. However, on practically all air cooled units equipped with expansion valves, a separate receiver is required.

In order to provide space to store the refrigerant charge when maintenance is required on the system, the receiver should be large enough to hold the entire refrigerant charge. A valve at the receiver outlet is required in order to pump the refrigerant charge into the receiver, an operation commonly called pumping the system down.

The outlet of the receiver must be so located that a liquid seal is maintained at the outlet even though the level in the receiver tank may vary, to prevent any vapor from entering the liquid line. Therefore if the outlet is at the top, or if a side outlet is provided, a dip tube extending to approximately ½” from the bottom of the receiver is used.

HEAT EXCHANGERS

A heat exchanger is a device for transferring heat from one medium to another. In commercial refrigeration systems the general term of heat exchanger is used to describe a component for transferring heat from the liquid refrigerant to the refrigerant suction gas.

As mentioned previously, a heat exchanger is used to raise the temperature of the return gas to prevent frosting or condensation, to subcool the liquid refrigerant sufficiently to prevent the formation of flash gas in the liquid line, to evaporate any liquid refrigerant flooding through the evaporator, and to increase system capacity.

A typical heat exchanger is shown in Figure 61. Suction

gas flows through the large center tube, while liquid is piped through the smaller tube wrapped around the suction tubing. The cold suction vapor absorbs heat from the warm, high pressure liquid through the tube to tube metal contact. Internal fins are often provided

in the suction gas section to increase the heat transfer between the suction gas and the liquid refrigerant.

SUCTION ACCUMULATORS

If liquid refrigerant is allowed to flood through the system and return to the compressor before being evaporated, it may cause damage to the compressor due to liquid slugging, loss of oil from the crankcase, or bearing washout. To protect against this condition on systems

11-1


vulnerable to liquid damage such as heat pumps, truck refrigeration, or on any installation where liquid floodback can occur, a suction accumulator is often used.

The accumulator’s function is to intercept liquid refrigerant before it can reach the compressor crankcase. It should be located in the compressor suction line between the evaporator and the compressor, should have a capacity large enough to hold the maximum amount of liquid that might flood through, and must have provisions for a positive return of oil to the crankcase. Either a source of heat must be provided to evaporate the liquid refrigerant or a means must be provided to meter the liquid to the compressor at a safe rate. A positive oil return must also be provided so that oil does not trap in the accumulator.

Figure 62 illustrates a vertical accumulator with a U-tube suction connection.

OIL SEPARATORS

Although well designed systems are effective in preventing oil return problems, there are some cases where the use of oil separators may be necessary. They are most often required on ultra-low temperature systems, with flooded evaporators, or on other systems where inherent oil return problems are present.

An oil separator is basically a separation chamber for oil and discharge gas. There is always some oil in circulation in a refrigeration system and oil leaving the compressor is entrained in the hot discharge gas which is traveling at high velocity. The oil separator when used is installed in the discharge line between the compressor and the condenser. By means of baffles and a reduction of gas velocity in the oil separator chamber, most of the oil is separated from the hot gas, and is returned to the compressor crankcase by means of a float valve and connecting tubing. The efficiency of an oil separator varies with load conditions, and is never 100% effective even under ideal conditions. If system design causes oil logging, an oil separator may only delay lubrication difficulty rather than cure it.

DEHYDRATORS

Moisture is one of the basic enemies of a refrigeration system, and the moisture level in an operating system must be held to an acceptable low level to avoid system malfunctions or compressor damage. Even with the best precautions, moisture will enter a system any time it is opened for field service. Unless the system is thoroughly evacuated and recharged after exposure to moisture, the only other effective means of removing small amounts of moisture is with a dehydrator.

Dehydrators or driers, as they are commonly called, consist of a shell filled with a desiccant or drying agent, with an adequate filter at each end. Some driers are made in porous block form so that the refrigerant is filtered by the entire block. Driers are mounted in the refrigerant liquid line, so that all of the refrigerant in circulation must pass through the drier each time it circulates through the system. Most driers are so constructed that they can serve a dual function as both filter and drier.

Many different drying agents are used, but practically all modern driers are either of the throwaway type, or of the replaceable element type, and it is considered good practice to discard the used drier element each time the system is opened, and replace with a new drier or drier element.

SUCTION LINE FILTERS

In order to protect the compressor from contamination left in the system at the time of installation, suction line filters are widely used. The suction line filter is designed for permanent installation in the suction line, and may be of the sealed type, or may be equipped with a replaceable element so that the filter can be easily changed if necessary.

The replaceable element type is convenient for installing a system cleaning filter-drier element in the event of system contamination.

VIBRATION ELIMINATORS

In order to prevent the transmission of noise and vibration from the compressor through the refrigeration piping, vibration eliminators are frequently installed in both the suction and discharge lines. On small units, where small diameter soft copper tubing is used for the refrigerant

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lines, a coil of tubing may provide adequate protection against vibration. On larger compressors, flexible metallic hose is most frequently used. STRAINERS

Strainers, as the name implies, are mounted in refrigerant lines to strain any dirt, metal chips, etc. out of the refrigerant which might cause a malfunction in either the refrigerant control devices or the compressor. While the configuration of the strainer will vary, basically it is comprised of a shell with a fine mesh screen. Because of the small orifice in expansion and solenoid valves, strainers are normally mounted just upstream from them in the refrigerant liquid line.

SIGHT GLASS AND MOISTURE INDICATORS

A sight glass in the liquid line allows the operator or serviceman to observe the flow of liquid refrigerant. Bubbles or foaming in the sight glass indicate a shortage of refrigerant, or a restriction in the liquid line that is adversely affecting system operation. Sight glasses are widely used as a means of determining if the system is adequately charged when adding refrigerant.

Moisture indicators have been incorporated in sight glasses as shown in Figure 65.

The moisture indicator provides a warning signal for the serviceman in the event moisture has entered the system, indicating that the dehydrator should be changed or that other action should be taken to effectively dry the system.

DISCHARGE MUFFLERS

On systems where noise transmission must be reduced to a minimum, or where compressor pulsation might create vibration problems, a discharge muffler is frequently used to dampen and reduce compressor discharge noise. The muffler is basically a shell with baffle plates, with the internal volume required primarily dependent on the compressor displacement, although the frequency and intensity of the sound waves are also factors in muffler design.

CRANKCASE HEATERS

When the compressor is installed in a location where it will be exposed to ambient temperatures colder than the evaporator, refrigerant migration to the crankcase can be aggravated by the resulting pressure difference between the evaporator and compressor during off cycles. To protect against the possibility of migration, crankcase heaters are often employed to keep the oil in the crankcase at a temperature high enough so that any liquid refrigerant entering the crankcase will evaporate and create a pressure sufficient to prevent large scale migration.

Crankcase heaters may be of the insert type or can be mounted externally on the crankcase. The heater is a low wattage resistance element, normally energized continuously, and must be carefully selected to avoid overheating of the oil in the compressor.

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REFRIGERATION GAUGES

Pressure gauges, especially calibrated for refrigeration usage, are the primary tool of the serviceman in checking system performance. Gauges for the high pressure side of the system have scales reading from 0 psig to 300 psig (or for usage on higher pressures, from 0 psig to 400 psig). Gauges for the low pressure part of the system are termed compound gauges, since the scale is graduated

for pressures above atmospheric pressure in psig, and for pressures below atmospheric pressure in vacuum in inches of mercury. The compound gauge is calibrated from 30 inches of vacuum to pressures ranging from 60 psig to 150 psig depending on gauge design.

In addition to the pressure scales, equivalent saturation temperatures for commonly used refrigerants are usually shown on the gauge dial.

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EmersonClimate.com

Form No. AE 102 R2 (10/06))Emerson®, Emerson. Consider It Solved™, Emerson Climate Technologies™ and the Emerson Climate Technologies™ logo are the trademarks and service marks of Emerson Electric Co. and are used with the permission of Emerson Electric Co.Copelametic®, Copeland®, and the Copeland® brand products logo are the trademarks and service marks of Emerson Climate Technologies, Inc.All other trademarks are the property of their respective owners.Printed in the USA. © 1967 Emerson Climate Technologies, Inc. All rights reserved.

Part 3 - The Refrigeration Load



FOREWORD

The practice of refrigeration undoubtedly goes back as far as the history of mankind, but for thousands of years the only cooling mediums were water and ice. Today refrigeration in the home, in the supermarket, and in commercial and industrial usage is so closely woven into our everyday existence it is difficult to imagine life without it. But because of this rapid growth, countless people who must use and work with refrigeration equipment do not fully understand the basic fundamentals of refrigeration system operation.

This manual is designed to fill a need which exists for a concise, elementary text to aid servicemen, salesman, students, and others interested in refrigeration. It is intended to cover only the fundamentals of refrigeration theory and practice. Detailed information as to specific products is available from manufacturers of complete units and accessories. Used to supplement such literature—and to improve general knowledge of refrigeration—this manual should prove to be very helpful.


Part 3THE REFRIGERATION LOAD

Section 12. HEAT TRANSMISSION

Transmission Heat Load — Q .......................... 12-1Thermal Conductivity — k ................................ 12-1Thermal Resistivity — r .................................... 12-1Conductance — C ............................................ 12-1Thermal Resistance — R ................................. 12-1Surface File Resistance.................................... 12-1Overall Coefficient of Heat Transfer — U ......... 12-1Transmission Heat Load ................................... 12-2Values of Thermal Conductivity for Building Materials ...................................... 12-3Outdoor Design Data ........................................ 12-3Allowance for Radiation from the Sun .............. 12-8Recommended Insulation Thickness ................ 12-8Quick Calculation Table for Walk-in Coolers..... 12-9

Section 13. AIR INFILTRATION

Air Change Estimating Method ......................... 13-1Air Velocity Estimating Method ......................... 13-1Ventilating Air .................................................... 13-2Infiltration Heat Load ........................................ 13-2

Section 14. PRODUCT LOAD

Tables of Specific Product Data........................ 14-1Heat of Respiration ........................................... 14-1Sensible Heat Above Freezing ......................... 14-7Latent Heat of Freezing .................................... 14-8Sensible Heat Below Freezing ......................... 14-8Total Product Load............................................ 14-8Storage Data .................................................... 14-8

Section 15. SUPPLEMENTARY LOAD

Electric Lights and Heaters............................... 15-1Electric Motors .................................................. 15-1Human Heat Load ............................................ 15-1Total Supplementary Load ................................ 15-1

Section 16. EQUIPMENT SELECTION

Hourly Load ...................................................... 16-1Sample Load Calculation.................................. 16-1Relative Humidity and Evaporator TD .............. 16-2Compressor Selection ...................................... 16-2Component Balancing ...................................... 16-3The Effect of Change in Compressor Only on System Balance ........................... 16-9Quick Select Tables for Walk-in Coolers........... 16-11


INDEX OF TABLES

Table 4 Typical Heat Transmission Coefficients ................................................................................. 12-3Table 5 Summer Outdoor Design Data .............................................................................................. 12-5Table 6 Allowance for Sun Effect ........................................................................................................ 12-8Table 7 Recommended Minimum Insulation Thickness ..................................................................... 12-9Table 7A Quick Estimate Factors for Heat Transmission Through Insulated Walls .............................. 12-9Table 8 Average Air Changes per 24 Hours for Storage Rooms Due to Opening and Infiltration ...... 13-1Table 9 Heat Removed in Cooling Air to Storage Room Conditions .................................................. 13-2Table 10 Food Products Data ............................................................................................................... 14-1Table 11 Properties of Solids ................................................................................................................ 14-4Table 12 Properties of Liquids .............................................................................................................. 14-6Table 13 Storage Requirements and Properties of Perishable Products ............................................. 14-9Table 14 Storage Conditions for Cut Flowers and Nursery Stock ........................................................ 14-11Table 15 Space, Weight, and Density Data for Commodities Stored in Refrigerated Warehouses ..... 14-12Table 16 Heat Equivalent of Electric Motors ......................................................................................... 15-1Table 17 Heat Equivalent of Occupancy .............................................................................................. 15-1Table 18 Recommended Condensing Unit Capacity for Walk-in Coolers, 35°F. Temperature ............. 16-12Table 19 Recommended Condensing Unit Capacity for Walk-in Coolers, Low Temperature ............... 16-12


The heat gain through walls, floors and ceilings will vary with the type of construction, the area exposed to a dif-ferent temperature, the type of insulation, the thickness of insulation, and the temperature difference between the refrigerated space and the ambient air.

In catalog and technical literature pertaining to heat transfer, certain letter symbols are commonly used to denote the heat transfer factors, and a working knowl-edge of these symbols is frequently necessary to easily interpret catalog data.

TRANSMISSION HEAT LOAD — Q

The basic formula for heat transfer through some heat transfer barrier is:

Q = U x A x TD Q = Heat transfer, BTU/Hr U = Overall heat transfer coefficient BTU/(hour)(sq. ft.)(°F TD) A = Area in square feet TD = Temperature differential between sides of thermal barrier, for example, between outside design temperature and the refrigerated space temperature.

Q is the rate of heat flow, the quantity of heat flowing after all factors are considered.

THERMAL CONDUCTIVITY — k

Thermal conductivity, k, is defined as the rate of heat transfer that occurs through a material in units of BTU/(hr)(square foot of area)(°F TD) per inch of thick-ness. Different materials offer varying resistances to the flow of heat.

For example, the heat transfer in 24 hours through two square feet of material three inches in thickness hav-ing a thermal conductivity factor of .25 with an average temperature difference across the material of 70°F would be calculated as follows:

.25(k) x 2 sq. ft. x 24 hours x 70° TDQ = 3 inches thickness = 280 BTU Since the total heat transferred by conduction varies directly with time, area, and temperature difference, and varies inversely with the thickness of the material, it is readily apparent that in order to reduce heat transfer,

the thermal conductivity factor should be as small as possible, and the material as thick as possible.

THERMAL RESISTIVITY — r

Thermal resistivity is defined as the reciprocal of thermal conductivity of 1/k. “r” is of importance because resis-tance values can be added numerically.

R total = r1 + r2 + r3

Where r1, r2, and r3 are individual resistances. This makes the use of r convenient in calculating overall heat transfer coefficients.

CONDUCTANCE — C

Thermal conductance is similar to thermal conductiv-ity, except that it is an overall heat transfer factor for a given thickness of material, as opposed to thermal conductivity, k, which is a factor per inch of thickness. The definition is similar, BTU/(hour)(square foot of area)(°F TD).

THERMAL RESISTANCE — R

Thermal resistance is the reciprocal of conductance, 1/C in the same way that thermal resistivity is the reciprocal of conductivity.

SURFACE FILM RESISTANCE

Heat transfer through any material is affected by the surface resistance to heat flow, and this is determined by the type of surface, rough or smooth; its position, vertical or horizontal; its reflective properties; and the rate of airflow over the surface. Surface film conduc-tance, normally denoted by fi for inside surfaces and fo for outside surfaces is similar to conductance.

However, in refrigeration work with insulated walls, the conductivity is so low that the surface film conductance has little effect, and therefore, can be omitted from the calculation.

OVERALL COEFFICIENT OF HEAT TRANSFER — U

The overall coefficient of heat transfer, U, is defined as the rate of heat transfer through a material or compound structural member with parallel walls. The U factor, as it is commonly called, is the resulting heat transfer

Section 12HEAT TRANSMISSION

12-1


coefficient after giving effect to thermal conductivity, conductance, and surface film conductance, and is ex-pressed in terms of BTU/(hour) (square foot of area)(°F TD). It is usually applied to compound structures such as walls, ceilings, and roofs.

The formula for calculating the U factor is complicated by the fact that the total resistance to heat flow through a substance of several layers is the sum of the resistance of the various layers. The resistance of heat flow is the reciprocal of the conductivity. Therefore, in order to calculate the overall heat transfer factor, it is necessary to first find the overall resistance to heat flow, and then find the reciprocal of the overall resistance to calculate the U factor.

The basic relation between the U factor and the various conductivity factors is as follows:

1 X1 X2R Total = C + k1 + k2

1U= R Total

In the above equation, k1, k2, etc. are the thermal conductivities of the various materials used, C is the conductance if it applies rather than k1, and X1, X2, etc. are the thicknesses of the material.

For example, to calculate the U factor of a wall com-posed of two inches of material having a k1 factor of

.80, and two inches of insulation having a conductance of .16, the U value is found as follows:

1 X1

R Total = C + k1

1 2 = .16 + .80

= 6.25 + 2.5 = 8.75

1 1 U = R Total = 8.75

=.114 BTU/(hour)(sq. ft.)(°F TD)

TRANSMISSION HEAT LOAD

Once the U factor is known, the heat gain by transmis-sion through a given wall can be calculated by the basic heat transfer equation.

Assume a wall with a U factor of .114 as calculated in the previous example. Given an area of 90 square feet with an inside temperature of 0°F, an outside temperature of 80°F, the heat transmission would be:

Q = U x A x TD = .114 x 90 sq. ft x 80°TD = 812 BTU/hr

The entire heat gain into a given refrigerated space can be found in a similar manner by determining the U factor for each part of the structure surrounding the refrigerated space, and calculating as above.

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VALUES OF THERMAL CONDUCTIVITY FOR BUILDING MATERIALS

Extensive testing has been done by many laboratories to determine accurate values for heat transfer through all common building and structural materials. Certain materials have a high resistance to the flow of heat (a low thermal conductivity) and are therefore used as insulation to decrease the heat transfer into the refriger-ated space. There are many different types of insulation such as asbestos, glass fiber, cork, reflective metals, and the new foam materials. Most good insulating materials have a thermal conductivity (k) factor of ap-proximately .25 or less, and rigid foam insulations have been developed with thermal conductivity (k) factors as low as .12 to .15.

Heat transmission coefficients for many commonly used building materials are shown in Table 4.

OUTDOOR DESIGN DATA

Extensive studies have been made of weather bureau records for many years to arrive at suitable outdoor design temperatures. For air conditioning or refrigera-tion applications, the maximum load occurs during the hottest weather.

However, it is neither economical or practical to design equipment for the hottest temperature which might ever occur, since the peak temperature might occur for only a few hours over the span of several years. Therefore, the design temperature normally is selected as a tem-perature that will not be exceeded more than a given percentage of the hours during the four month summer season. Table 5 lists summer design temperatures, which will be equaled or exceeded only during 1% of the hours during the four summer months.

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ALLOWANCE FOR RADIATION FROM THE SUN

The primary radiation factor involved in the refrigeration load is heat gain from the sun’s rays. If the walls of the refrigerated space are exposed to the sun, additional heat will be added to the heat load. For ease in calcu-lation, an allowance can be made for the sun load in refrigeration calculations by increasing the temperature differential by the factors listed in Table 6.

This table is usable for refrigeration loads only, and is not accurate for air conditioning estimates.

RECOMMENDED INSULATION THICKNESS

As the desired storage temperature decreases, the refrigeration load increases, and as the evaporating temperature decreases, the compressor efficiency decreases. Therefore, from a practical and economic standpoint, the insulation thickness must be increased as the storage temperature decreases.

Table 7 lists recommended insulation thickness from the 1981 ASHRAE Handbook of Fundamentals. The recom-mendations are based on expanded polyurethane which has a conductivity factor of .16. If other insulations are used, the recommended thickness should be adjusted base on relative k factors.

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QUICK CALCULATION TABLE FOR WALK-IN COOLERS

As an aid in the quick calculation of heat transmission through insulated walls, Table 7A lists the approximate heat gain in BTU per 1°F. temperature difference per square foot of surface per 24 hours for various thick-nesses of commonly used insulations. The thickness of insulation referred to is the actual thickness of insulation, and not the overall wall thickness.

For example, to find the heat transfer for 24 hours through a 6’ x 8’ wall insulated with 4 inches of glass fiber when the outside is exposed to 95°F ambient temperature, and the box temperature is 0°F., calculate as follows:

1.9 factor x 48 sq. ft. x 95°TD = 8664 BTU

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SECTION 13AIR INFILTRATION

8 100Velocity = 100 FPM x 7 x 60 2.83 10=100 x 2.65 x 7.74

= 138 FPM

Estimated rate of Infiltration

138 FPM x 8 ft. x 4 ft. = 2210 cu. ft per min. 2

Infiltration velocities for various door heights and TDs are plotted in Figure 67.

If the average time the door is opened each hour can be determined, the average hourly infiltration can be calculated, and the heat gain can be determined as before.

Any outside air entering the refrigerated space must be reduced to the storage temperature, thus increasing the refrigeration load. In addition, if the moisture content of the entering air is above that of the refrigerated space, the excess moisture will condense out of the air, and the latent heat of condensation will add to the refrigera-tion load.

Because of the many variables involved, it is difficult to calculate the additional heat gain due to air infiltration. Various means of estimating this portion of the refrig-eration load have been developed based primarily on experience, but all of these estimating methods are subject to the possibility of sizable error, and specific applications may vary widely in the actual heat gain encountered.

AIR CHANGE ESTIMATING METHOD

The traffic in and out of a refrigerator usually varies with its size or volume. Therefore the number of times doors are opened will be related to the volume rather than the number of doors.

Table 8 lists estimated average air changes per 24 hours for various sized refrigerators due to door openings and infiltration for a refrigerated storage room. Note that these values are subject to major modification if it is definitely determined that the usage of the storage room is either heavy or light.

AIR VELOCITY ESTIMATING METHOD Another means of computing infiltration into a refriger-ated space is by means of the velocity of airflow through an open door. When the door of a refrigerated storage space is opened, the difference in density between cold and warm air will create a pressure differential causing cold air to flow out the bottom of the doorway and warm air to flow in the top. Velocities will vary from maximum at the top and bottom to zero in the center.

The estimated average velocity in either half of the door is 100 feet per minute for a doorway seven feet high at 60°F. TD. The velocity will vary as the square root of the height of the doorway and as the square root of the temperature difference.

For example the rate of infiltration through a door 8 feet high and 4 feet wide, with a 100°F. TD between the storage room and the ambient can be estimated as follows:

13-1


VENTILATING AIR

If positive ventilation is provided for a space by means of supply or exhaust fans, the ventilation load will replace the infiltration load (if greater) and the heat gain may be calculated on the basis of the ventilating air volume.

INFILTRATION HEAT LOAD

Once the rate of infiltration has been determined, the heat load can then be calculated from the heat gain

per cubic foot of infiltration as given in Table 9. For ac-curate calculations at conditions not covered by Table 9, the heat load can be determined by the difference in enthalpy between entering air and the storage room air conditions. This is most easily accomplished by use of the psychrometric chart, which will be discussed in detail in a subsequent section.

13-2


SECTION 14PRODUCT LOAD

HEAT OF RESPIRATION

Fruits and vegetables, even though they have been removed from the vine or tree on which they grew, are still living organisms. Their life processes continue for some time after being harvested, and as a result they give off heat. Certain other food products also undergo continuing chemical reactions which produce heat. Meats and fish have no further life processes and do not generate any heat.

The amount of heat given off is dependent on the specific product and its storage temperature. Table 10 lists various food products with pertinent storage data. Note that the heat of respiration varies with the storage temperature.

The product load is composed of any heat gain occur-ring due to the product in the refrigerated space. The load may arise from a product placed in the refrigerator at a temperature higher than the storage temperature, from a chilling or freezing process, or from the heat of respiration of perishable products. The total product load is the sum of the various types of product load which may apply to the particular application.

TABLES OF SPECIFIC PRODUCT DATA

The following tables list data on specific products that is essential in calculating the refrigeration product load. Table 10 covers food products, Table 11 solids, and Table 12 liquids.

14-1



SENSIBLE HEAT ABOVE FREEZING

Most products are at a higher temperature than the stor-age temperature when placed in a refrigerator. Since many foods have a high percentage of water content, their reaction to a loss of heat is quite different above and below the freezing point. Above the freezing point, the water exists in liquid form, while below the freezing point, the water has changed its state to ice.

As mentioned previously, the specific heat of a product is defined as the BTUs required to raise the temperature of one pound of the substance 1°F. The specific heats of various commodities are listed in Tables 10, 11, and 12. Note that in Table 10 the specific heat of the product above freezing is different than the specific heat below freezing, and the freezing point (listed in the first column) varies, but in practically all cases is below 32°F.

The heat to be removed from a product to reduce its

temperature above freezing may be calculated as fol-lows:

Q = W x c x (T1 - T2)

Q = BTU to be removed W = Weight of the product in pounds c = Specific heat above freezing T1 = Initial temperature, °F. T2 = Initial temperature, °F. (freezing or above)

For example, the heat to be removed in order to cool 1,000 pounds of veal (whose freezing point is 29°F.) from 42°F. to 29°F. can be calculated as follows:

Q = W x c x (T1 - T2) = 1000 pounds x .71 specific heat x (42-29) = 1000 x .71 x 13 = 9,230 BTU

14-7


LATENT HEAT OF FREEZING

The latent heat of fusion or freezing for liquids other than water is given in Table 12. Substances such as metals which contain no water do not have a freezing point, and no latent heat of fusion is involved in lowering their temperature. Most food products, however, have a high percentage of water content. In order to calculate the heat removal required to freeze the product, only the water need be considered. The water content percentage for various food products is given in Table 10, Column 2.

Since the latent heat of fusion or freezing of water is 144 BTU/lb., the latent heat of fusion for the product can be calculated by multiplying 144 BTU/lb. by the percentage of water content, and for ease in calculations this figure is given in Column 5 of Table 10. To illustrate, veal has a water percentage of 63%, and the latent heat of fusion listed in Column 5 for veal is 91 BTU/lb.

63% x 144 BTU/lb. = 91 BTU/lb.

The heat to be removed from a product for the latent heat of freezing may be calculated as follows:

Q = W x hif

Q = BTU to be removed W = Weight of product in pounds hif = latent heat of fusion, BTU/lb.

The latent heat of freezing of 1000 pounds of veal at 29°F. is:

Q = W x hif = 1000 lbs. x 91 BTU/lb. = 91,000 BTU

SENSIBLE HEAT BELOW FREEZING

Once the water content of a product has been frozen, sensible cooling again can occur in the same manner as that above freezing, with the exception that the ice in the product causes the specific heat to change. Note in Table 10 the specific heat of veal above freezing is .71, while the specific heat below freezing is .39,

The heat to be removed from a product to reduce its tem-perature below freezing may be calculated as follow:

Q = W x ci x (Tf - T3)

Q = BTU to be removed W = Weight of product in pounds ci = Specific heat below freezing Tf = Freezing temperature T3 = Final temperature

For example, the heat to be removed in order to cool 1,000 pounds of veal from 29°F. to 0°F. can be calcu-lated as follows:

Q = W x ci x (Tf - T3) = 1,000 lbs. x .39 specific heat x (29-0) = 1,000 x .39 x 29 = 11,310 BTU

TOTAL PRODUCT LOAD

The total product load is the sum of the individual calculations for the sensible heat above freezing, the latent heat of freezing, and the sensible heat below freezing.

From the foregoing example, if 1,000 pounds of veal is to be cooled from 42°F. to 0°F., the total would be:

Sensible Heat above Freezing 9,230 BTU Latent Heat of Freezing 91,000 BTU Sensible Heat Below Freezing 11,310 BTU Total Product Load 111,540 BTU

If several different commodities or crates, baskets, etc. are to be considered, then a separate calculation must be made for each item for an accurate estimate of the product load.

STORAGE DATA

Most commodities have conditions of temperature and relative humidity at which their quality is best preserved and their storage life is a maximum. Recommended stor-age conditions for various perishable products are listed in Table 13 and recommended storage conditions for cut flowers and nursery stock are listed in Table 14.

Data on various types of storage containers is listed in Table 15.

14-8


SECTION 15SUPPLEMENTARY LOAD

TOTAL SUPPLEMENTARY LOAD

The total supplementary load is the sum of the individual factors contributing to it. For example, the total supple-mentary load in a refrigerated storeroom maintained at 0°F. in which there are 300 watts of electric lights, a 3 HP motor driving a fan, and 2 people working continu-ously would be as follows:

300 Watts x 3.41 BTU/hr. 1,023 BTU/hr. 3 HP motor x 2,950 BTU/hr. 8,850 BTU/hr. 2 people x 1300 BTU/hr. 2,600 BTU/hr. Total Supplementary Load 12,473 BTU/hr.

In addition to the heat transmitted into the refrigerated space through the walls, air infiltration, and product load, any heat gain from other sources must be included in the total cooling load estimate.

ELECTRIC LIGHTS AND HEATERS

Any electric energy directly dissipated in the refrigerated space such as lights, heaters, etc. is converted to heat and must be included in the heat load. One watt hour equals 3.41 BTU, and this conversion ratio is accurate for any amount of electric power.

ELECTRIC MOTORS

Since energy cannot be destroyed, and can only be changed to a different form, any electrical energy transmitted to motors inside a refrigerated space must undergo a transformation. Any motor losses due to fric-tion and inefficiency are immediately changed to heat energy. That portion of the electrical energy converted into useful work, for example in driving a fan or pump, exists only briefly as mechanical energy, is transferred to the fluid medium in the form of increased velocity, and as the fluid loses its velocity due to friction, eventually becomes entirely converted into heat energy.

A common misunderstanding is the belief that no heat is transmitted into the refrigerated space if an electric motor is located outside the space, and a fan inside the space is driven by means of a shaft. All of the electrical energy converted to mechanical energy actually be-comes a part of the load in the refrigerated space.

Because the motor efficiency varies with size, the heat load per horsepower as shown in Table 16 has different values for varying size motors. While the values in the table represent useful approximations, the actual elec-tric power input in watts is the only accurate measure of the energy input.

HUMAN HEAT LOAD

People give off heat and moisture, and the resulting refrigeration load will vary depending on the duration of occupancy of the refrigerated space, temperature, type of work, and other factors. Table 17 lists the average head load due to occupancy, but stays of short duration, the heat gain will be somewhat higher.

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SECTION 16EQUIPMENT SELECTION

Once the refrigeration load is determined, together with the required evaporating temperature and the expected condensing temperature, a compressor can be intel-ligently selected for a given application.

For refrigerated fixtures or prefabricated coolers and cold storage boxes to be produced in quantity, the load is normally determined by test. If the load must be estimated, the expected load should be calculated by determining the heat gain due to each of the factors contributing to the total load. Many short methods of estimating are commonly used for small refrigerated walk-in storage boxes with varying degrees of accuracy. A great deal of judgment must be used in the application of any method.

HOURLY LOAD

Refrigeration equipment is designed to function continu-ously, and normally the compressor operating time is de-termined by the requirements of the defrost system. The load is calculated on a 24 hour basis, and the required hourly compressor capacity is determined by dividing the 24 hour load by the desired hours of compressor operation during the 24 hour period. A reasonable safety factor must be provided to enable the unit to recover rapidly after a temperature rise, and to allow for loading heavier than the original estimate.

When the refrigerant evaporating temperature will not be below 30°F., frost will not accumulate on the evaporator, and no defrost period is necessary. It is general practice to choose the compressor for such applications on the basis of 18 to 20 hour operation.

For applications with storage temperatures of 35°F. or higher, and refrigerant temperatures low enough to cause frosting, it is common practice to defrost by stop-ping the compressor and allowing the return air to melt the ice from the coil. Compressors for such applications should be selected for 16 to 18 hour operation.

On low temperature applications, some positive means of defrost must be provided. With normal defrost peri-ods, 18 hour compressor operation is usually accept-able, although some systems are designed for continu-ous operation except during the defrost period.

An additional 5% to 10% safety factor is often added to load calculations as a conservative measure to be sure the equipment will not be undersized. If data concerning the refrigeration load is very uncertain, this may be desir-

able, but in general the fact that the compressor is sized on the basis of 16 to 18 hour operation in itself provides a sizable safety factor. The load should be calculated on the basis of the peak demand at design conditions, and normally the design conditions are selected on the basis that they will occur no more that 1% of the hours during the summer months. If the load calculations are made reasonably accurately, and the equipment sized properly, an additional safety factor may actually result in the equipment being oversized during light load condi-tions, and can result in operating difficulties.

SAMPLE LOAD CALCULATION

The most accurate means of estimating a refrigeration load is by considering each factor separately. The follow-ing example will illustrate a typical selection procedure, although the load has been chosen to demonstrate the calculations required and does not represent a normal loading.

Walk-in cooler with 4 inches of glass fiber insulation, located in the shade.

Outside Dimensions, Height 8 ft., Width 10 ft., Length 40 ft., inside volume 3,000 cu. ft.

Floor area (outside dimensions) 400 sq. ft. on insulated slab in contact with ground.

Ambient temperature 100°F., 50% relative humidity Ground temperature 55°F. Refrigerator temperature 40°F.

1/2 HP fan motor running continuously

Two 100 watt lights, in use 12 hours per day.

Occupancy, 2 men for 2 hours per day.

In storage: 500 pounds of bacon at 50°F. 1000 pounds of string beans

Entering product: 500 pounds of bacon at 50°F. 15,000 pounds of beer at 80°F. To be reduced to storage temperature in 24 hours.

Heavy door usage.

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(A) HEAT TRANSMISSION LOAD

Sidewalls: 40’ x 8’ x 2 = 640 Ft2 x 60°TD x 1.9 (Table 7A) = 72,960 BTU

10’ x 8’ x 2 = 160 Ft2 x 60°TD x 1.9 = 18,240

Ceiling: 40’ x 10’ = 400 Ft2 x 60°TD x 1.9 = 45,600

Floor: 40’ x 10’ = 400 Ft2 x 15°TD x 1.9 = 11,400

Total 24 hour transmission load = 148,200

(B) AIR INFILTRATION

3000 Ft3 x 9.5 air changes (Table 8) x 2 usage factor x 2.11 factor (Table 9) 120,270 BTU

(C) PRODUCT LOAD 500 lbs. bacon x .50 sp. ht. (Table 10) x 10°TD = 2,500 BTU 15,000 lbs. beer x 1.0 sp. ht. (Table 10) x 40°TD = 600,000 BTU

500 lbs. lettuce x 2700 BTU/24 Hr/Ton (Table 10) = 675 BTU

1,000 lbs. beans x 9700 BTU/24 Hr/Ton (Table 10) = 4,850 BTU Total 24 hour Product Load 608,025 BTU

(D) SUPPLEMENTARY LOAD

200 Watts x 12 hours x 3.41 BTU/Hr 8,184 BTU

1/2 H.P. x 4250 BTU/Hr-Hr (Table 16) x 24 51,000 BTU

2 People x 2 Hrs/Day x 840 BTU/Hr (Table 17) 3,360 BTU

Total 24 hour Supplementary Load 62,544 BTU

(E) REQUIRED COMPRESSOR CAPACITY

24 Hour Load:

Heat Transmission 148,200 BTU Air Infiltration 120,270 Product 608,025 Supplementary 62,544 Total 24 Hour Load 939,039 BTU

Required compressor capacity:

Based on 16 hour operation 58,690 BTU/Hr.

RELATIVE HUMIDITY AND EVAPORATOR TD

Relative humidity in a storage space is affected by many variables, such as system running time, moisture infiltration, condition and amount of product surface ex-posed, air motion, outside air conditions, type of system control, etc. Perishable products differ in their require-ments for an optimum relative humidity for storage, and recommended storage conditions for various products are shown in Tables 13 and 14. Normally satisfactory control of relative humidity in a given application can be achieved by selecting the compressor and evaporator for the proper operating temperature difference or TD between the desired room temperature and the refriger-ant evaporating temperature.

The following general recommendations have proven to be satisfactory in most normal applications:

Desired TDTemperature Relative (RefrigerantRange Humidity to Air) 25°F. to 45°F. 90% 8°F. to 12°F.25°F. to 45°F. 85% 10°F. to 14°F.25°F. to 45°F. 80% 12°F. to 16°F.25°F. to 45°F. 75% 16°F. to 22°F.10°F. and below — 15°F. or less

COMPRESSOR SELECTION In order to select a suitable compressor for a given application, not only the required compressor capacity must be known, but also the desired evaporating and condensing temperatures.

Assuming a desired relative humidity of 80%, a 14° TD might be used, which in a 40°F. storage room result in evaporating temperature of 26°F. To provide some safety factor for line losses, the compressor should be selected for the desired capacity at 2°F. to 3°F. below the desired evaporating temperature.

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The condensing temperature depends on the type of condensing medium to be used, air or water, the design ambient temperature or water temperature, and the capacity of the condenser selected. Air cooled condens-ers are commonly selected to operate on temperature differences (TD) from 10°F. to 30°F. the lower TD nor-mally being used for low temperature applications, and higher TDs for high temperature applications where the compression ratio is less critical. For the purposes of this example, a design TD of 20°F. has been selected, and in 100°F. ambient temperatures, this would result in a condensing temperature of 120°F.

COMPONENT BALANCING

Commercially available components seldom will exactly match the design requirements of a given system, and since system design is normally based on estimated peak loads, the system may often have to operate at conditions other than design conditions. More than one combination of components may meet the per-formance requirements, the efficiency of the system normally being dependent on the point at which the system reaches stabilized conditions or balances under operating conditions.

The capacities of each of the three major system components, the compressor, the condenser, and the evaporator, are each variable but interrelated. The compressor capacity varies with the evaporating and condensing temperatures. For illustration purposes an air cooled condenser will be considered, and for a given condenser with constant air flow, its capacity will vary with the temperature difference between the condensing temperature and the ambient temperature.

The factors involved in the variation in evaporator capac-ity are quite complex when both sensible heat transfer and condensation are involved. For component bal-ancing purposes, the capacity of an evaporator where both latent and sensible heat transfer are involved (a wet coil) may be calculated as being proportional to the total heat content of the entering air, and this in turn is proportional to the wet bulb temperature. For wet coil conditions, evaporator capacities are normally available from coil manufacturers with ratings based on the wet bulb temperature of the air entering the coil. For condi-tions in which no condensation occurs (a dry coil) the evaporator capacity can be accurately estimated on the basis of the dry bulb temperature of the air enter-ing the coil.

Some manufacturers of commercial and low tempera-ture coils publish only ratings based on the temperature difference between entering dry bulb temperature and the evaporating refrigerant temperature. Although frost

accumulation involving latent heat will occur, unless the latent load is unusually large, the dry bulb ratings may be used without appreciable error.

Because of the many variables involved, the calculation of system balance points is extremely complicated. A simple, accurate, and convenient method of forecast-ing system performance from readily available manu-facturer’s catalog data is the graphical construction of a component balancing chart. The following example illustrates the use of such a chart in checking the pos-sible balance points of a system when selecting equip-ment. To illustrate the procedure, tentative selections of a compressor, condenser, and evaporator have been made for the sample load previously calculated.

Figure 69 shows the compressor capacity curves as published by Emerson Climate Technologies, Inc. on the Copeland® brand compressor specification sheet. It should be noted that Copeland® brand compressor capacity curves for Copelametic® compressors are based on 65°F. return suction gas. In order to realize the full compressor capacity, the suction gas must be raised to this temperature in a heat exchanger. If the suction gas returns to the compressor at a lower temperature, or if the increase in suction gas temperature occurs due to heat transfer into the suction line outside the refriger-ated space, the effective compressor capacity will be somewhat lower. In the example, the desired capacity was 58,690 BTU/hr. at 24°F. evaporating temperature and 120°F.condensing temperature, and this compres-sor was the closest choice available, having a capacity of 57,000 BTU/hr. at the design conditions.

Figure 70 shows the same compressor curves, with the condenser capacity curves for the tentative con-denser selection superimposed. From the condenser manufacturer’s data, condenser capacity in terms of compressor capacity at varying evaporating tempera-tures are plotted, and the condenser capacity curves can then be drawn. Note that the net condensing capacity decreases at lower evaporating temperatures due to the increased heat of compression.

It is now possible to construct balance lines for the com-pressor and condenser at various ambient temperatures as shown in Figure 71. For an ambient temperature of 100°F., point A would represent the balance point if the compressor were operating at a suction pres-sure equivalent to a 28°F. evaporating temperature and 120°F. condensing temperature. At this point the capacity of the condenser would exactly match that of the compressor at a 20° TD (condensing temperature minus ambient temperature). The balance point is de-termined by the intersection of the 20°F. TD condenser capacity curve with the compressor capacity curve for

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CONDENSER CAPACITY CURVES SUPERIMPOSEDON COMPRESSOR CAPACITY CURVES


a condensing temperature 20°F above the specified ambient temperature of 100°F., or 120°F. In a similar manner balance point B can be located by the inter-section of the 25°F. TD condenser capacity curve and the compressor capacity curve (estimated) for 125°F. condensing, and balance point C can be located by the intersection of the 15°F. TD condenser capacity curve with the compressor capacity curve (estimated) for 115°F. condensing. The line connecting points A, B, and C represents all the possible balance points when the system is operating with air entering the condenser at a temperature of 100°F. In a similar fashion, condenser-compressor balance lines can be determined for other ambient temperatures, and plotted as shown in Figure 72. (To simplify the illustration, condenser capacity curves have not been shown)

The tentative evaporator coil selected was rated by the manufacturer only in terms of BTU/hr per degree temperature difference between the entering dry bulb temperature and the refrigerant evaporating tempera-ture, and have a capacity of 4,590 BTU/hr/°TD. In Figure 73 evaporator capacity curves have been plotted and superimposed on the compressor capacity curves and the condenser-compressor balance lines. An evaporator capacity curve for each entering air temperature can be constructed by plotting any two points.

Point A represents the evaporator capacity at 14°TD which for an entering air temperature of 40°F. would require a refrigerant evaporating temperature of 26°F. However, an allowance must be made for line friction losses since the pressure in the evaporator will always be higher than the suction pressure at the compressor because of pressure drop in the suction line. Allowing 2°F. as an estimated allowance for line pressure drop, an evaporating temperature of 26°F. would result in a pressure at the compressor equivalent to a saturated evaporating temperature of 24°F. Therefore the capac-ity of the evaporator for a 14° TD and 40°F. entering air would be plotted at the corresponding compressor capacity at 24°F.

Point B represents the evaporator capacity at 10° TD, which for 40°F. entering air temperature requires a refrigerant evaporating temperature of 30°F., and after allowing for suction line losses, a corresponding compressor capacity at 28°F. A line can then be drawn through these two points, representing all possible capacities of the evaporator with 40°F. entering air and varying refrigerant evaporating temperatures. In a similar fashion, capacity curves can be constructed for other entering air temperatures.

The system performance can now be forecast for any condition of evaporator entering air temperature and

ambient temperature. With 100°F. ambient temperature and an evaporator entering air temperature of 40°F., the original design conditions, the system would have a capacity of 59,000 BTU/hr, a compressor suction pressure equivalent to an evaporating temperature of 26°F., and a condensing temperature of 120°F. Even under extreme load conditions of 50°F. entering air and 110°F. ambient, the condensing temperature would not exceed 133°F. These conditions are close enough to the original design requirement to insure satisfactory performance.

This type of graphical analysis can be quickly and eas-ily made by using the compressor specification sheet as the basic chart, and superimposing condenser and evaporator capacity curves.

THE EFFECT OF CHANGE IN COMPRESSOR ONLY ON SYSTEM BALANCE

Occasionally the exact replacement compressor may not be available, and the question arises as to whether an alternate compressor with either more or less ca-pacity might provide satisfactory performance. The graphical balance chart provides a convenient means of forecasting system performance.

Figure 74 is a revised balance chart for a system utilizing the same evaporator and condenser as in the previous example, but with a compressor having only 5/6 of the previous capacity. New compressor capacity curves for the smaller compressor have been plotted on the same capacity chart used previously. Since there is no change in the basic capacity of the condenser or evaporator, the condenser capacity and evaporator capacity curves are unchanged.

However, a new compressor-condenser balance line must be plotted, and to avoid excessive detail in the il-lustration, a balance line for 100° ambient temperature only has been shown.

A comparison can now be made between the system with the original compressor, Figure 73, and the system with the smaller compressor, Figure 74.

Original Revised System SystemAmbient Temperature 100°F. 100°F.Air Entering Evaporator 40°F. 40°F.Refrigerant Evaporating Temp. 26°F. 27°F.Condensing Temperature 120°F. 115°F.Capacity at 100°F. Ambient and 40°F. Entering Air, BTU/hr. 59,000 53,000

16-9



Note that although the compressor capacity was de-creased by 1/6 or 16 2/3%, the net system capacity decreased only about 10%. Since the condenser and evaporator were unchanged, the compressor could operate at more efficient conditions, with decreased condensing pressure and increased suction pressure.

The same type of analysis can be applied to determine the effect on system capacity if the compressor on a unit designed for 60 cycle operation is operated on 50 cycle power. However for the evaporator and condenser capacity to remain constant, the air flow across both evaporator and condenser must be unchanged. If the original balance chart was made on the basis of fans operating on 60 cycle power, and the fan air delivery is decreased by operation of the fan motors on 50 cycle power, then both the evaporator and condenser capac-ity curves must be changed to reflect the decrease in capacity.

Another type of application where this type of analysis may be valuable is on systems with fluctuating loads and compressors with capacity control features. Since the evaporator and condenser remain unchanged, the reduced compressor capacity can be plotted as dem-onstrated, and new balance points determined, taking into effect any changes in the temperature of the air entering the evaporator.

QUICK SELECTION TABLES FOR WALK-IN COOLERS

The most accurate means of determining the refrigera-tion load is by calculating each of the factors contributing to the load as was done in the previous example. How-ever, for small walk-in coolers, various types of short cut estimating methods are frequently used.

The transmission load will always be dependent on the external surface, and an actual calculation should be made where possible.

As an aid in rapid selection of a condensing unit for the normal walk-in cooler application, Tables 19 and 20 give recommended refrigeration capacities for various sized coolers. The condensing unit capacity must be equal to or greater than the capacity shown at the required refrigerant evaporating temperature after allowance for the desired evaporating and condensing TD.

The capacities given are for average applications. If the load is unusual, these tables should not be used. The low temperature tables do not include any allowance for a freezing load, and if a product is to be frozen, ad-ditional capacity will be required.

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Table 18

Table 19


EmersonClimate.com

Form No. AE 103 R3 (10/06)Emerson®, Emerson. Consider It Solved™, Emerson Climate Technologies™ and the Emerson Climate Technologies™ logo are the trademarks and service marks of Emerson Electric Co. and are used with the permission of Emerson Electric Co.Copelametic®, Copeland®, and the Copeland® brand products logo are the trademarks and service marks of Emerson Climate Technologies, Inc.All other trademarks are the property of their respective owners.Printed in the USA. © 1968 Emerson Climate Technologies, Inc. All rights reserved.

Part 4 - System Design



This is the fourth of a series of publications comprising the Emerson Climate Technologies, Inc. Refrigeration Manual. Although each separate part covers a specific area of refrigera-tion theory and practice, each successive publication presumes a basic understanding of the material presented in the previous sections.

Part 1 Fundamentals of Refrigeration Part 2 Refrigeration System Components Part 3 The Refrigeration Load Part 4 System Design

The application and design recommendations are intended only as a general guide. The exact requirements of a given installation can only be determined after the specific design criteria and desired operating conditions are known.


1

Part4SYSTEMDESIGN

Section17. BASICAPPLICATIONRECOMMENDATIONS

Fundamental Design Principles ...................... 17-1Compressor Selection .................................... 17-1System Balance.............................................. 17-1Refrigerant ...................................................... 17-2 Compressor Cooling ....................................... 17-2Compressor Lubrication ................................. 17-3Oil Pressure Safety Control ............................ 17-3Oil Separators................................................. 17 4Suction Line Accumulators ............................. 17-4Pumpdown System Control ............................ 17-5Crankcase Heaters ......................................... 17-5Crankcase Pressure Regulating Valves ......... 17-6Low Ambient Head Pressure Control ............. 17-6Liquid Line Filter-Drier .................................... 17 6Sight Glass and Moisture Indicator................. 17-7Liquid Line Solenoid Valve ............................. 17-7Heat Exchanger .............................................. 17-7Thermostatic Expansion Valves ..................... 17-7Evaporators .................................................... 17-8Suction Line Filters ......................................... 17-9High and Low Pressure Controls .................... 17-9Interconnected Systems ................................. 17-10Electrical Group Fusing .................................. 17-10

Section18. REFRIGERATIONPIPING

Basic Principles of Refrigeration Piping Design .......................................... 18-1Copper Tubing for Refrigerant Piping ............. 18-2Fittings for Copper Tubing .............................. 18-2Equivalent Length of Pipe............................... 18-2Pressure Drop Tables ..................................... 18-5Sizing Hot Gas Discharge Lines ..................... 18-5Sizing Liquid Lines.......................................... 18-14Sizing Suction Lines ....................................... 18-15Double Risers ................................................. 18-21Suction Piping for Multiplex Systems ............. 18-22Piping Design for Horizontal and Vertical Lines ........................................... 18-23Suction Line Piping Design at the Evaporator ............................................... 18-24Receiver Location ........................................... 18-25Vibration and Noise ........................................ 18-25Recommended Line Sizing Tables ................. 18-26

Section19. LOWTEMPERATURE SYSTEMS

Single Stage Low Temperature Systems ........ 19-1Two Stage Low Temperature Systems ........... 19-2Volumetric Efficiency ...................................... 19-2Two Stage Compression and Compressor Efficiency ............................ 19-2Compressor Overheating at Excessive Compression Ratios ................................ 19-5Basic Two Stage System ................................ 19-6Two Stage System Components .................... 19-6Piping on Two Stage Systems ........................ 19-9Cascade Refrigeration Systems ..................... 19-13

Section20. TRANSPORTREFRIGERATION

Compressor Cooling ....................................... 20-1Compressor Speed......................................... 20-1Compressor Operating Position ..................... 20-2Compressor Drive........................................... 20-2Refrigerant Charge ......................................... 20-2Refrigerant Migration ...................................... 20-2Oil Charge ...................................................... 20-3Oil Pressure Safety Control ............................ 20-3Oil Separators................................................. 20-3Crankcase Pressure Regulating Valve ........... 20-3Condenser ...................................................... 20-4Receiver ......................................................... 20-4Purging Air in a System .................................. 20-4Liquid Line Filter-Drier .................................... 20-4Heat Exchanger .............................................. 20-5Liquid Line Solenoid Valve ............................. 20-5Suction Line Accumulator ............................... 20-5Crankcase Heaters ......................................... 20-5Pumpdown Cycle............................................ 20-5Forced Air Evaporator Coils............................ 20-5Thermostatic Expansion Valves ..................... 20-6Defrost Systems ............................................. 20-6Thermostat ..................................................... 20-7High-Low Pressure Control ............................ 20-7Eutectic Plate Applications ............................. 20-7Refrigerant Piping ........................................... 20-10Vibration ......................................................... 20-10Electrical Precautions ..................................... 20-10Installation ...................................................... 20-11Field Troubleshooting on Transport Units ....... 20-12


Section21. CAPACITYCONTROL

Internal Capacity Control Valves ...................21-1External Capacity Control Valves ..................21-1Hot Gas Bypass.............................................21-1Bypass into Evaporator Inlet..........................21-3Bypass into Suction Line ...............................21-3Solenoid Valves for Positive Shut-offand Pumpdown Cycle....................................21-5Desuperheating Expansion Valve ..................21-5Typical Multiple-Evaporator Control System .......................................21-5Power Consumption with Hot Gas Bypass .....................................21-6

Section22. LIQUIDREFRIGERANTCONTROLINREFRIGERATIONANDAIRCONDITIONINGSYSTEMS

Refrigerant-Oil Relationship ..........................22-1Refrigerant Migration .....................................22-1

Liquid Refrigerant Flooding ...........................22-2Liquid Refrigerant Slugging ...........................22-2Tripping of Oil Pressure Safety Control .........22-2Recommended Corrective Action ..................22-2

Section23. ELECTRICALCONTROLCIRCUITS

Typical Lockout Control Circuit ......................23-1Control Circuit for Compressor Protection Against Liquid Refrigerant Flooding ...............................23-3Control Circuits to Prevent Short Cycling ......22-3Control Circuits for Compressors with Capacity Control Valves .........................23- 5


INDEXOFTABLES

Table 20A Ventilation Air Requirements for Machine Rooms CFM/1000 BTU/HR at 10° F. Air Temperature Rise ............................................................................................. 17-3

Table 21 Recommended Minimum Low Pressure Control Setting ........................................................ 17-9Table 22 Dimensions and Properties of Copper Tube ........................................................................... 18-3Table 23 Weight of Refrigerant in Copper Lines.................................................................................... 18-4Table 24 Equivalent Length in Feet of Straight Pipe For Valves and Fittings ........................................ 18-5Table 25 Pressure Drop Equivalent for 2° F. Change in Saturation Temperature at Various Evaporating Temperatures ................................................................................. 18-15Table 26 Maximum Recommended Spacing Between Pipe Supports for Copper Tubing .................... 18-26Table 27 Recommended Liquid Line Sizes ........................................................................................... 18-27Table 28 Recommended Discharge Lines Sizes ................................................................................... 18-28Table 29 Recommended Suction Line Sizes, R-12, 40° F. .................................................................... 18-29Table 30 Recommended Suction Line Sizes, R-12, 25° F. .................................................................... 18-30Table 31 Recommended Suction Line Sizes, R-12, 15° F. .................................................................... 18-31Table 32 Recommended Suction Line Sizes, R-12, –20° F. .................................................................. 18-31Table 33 Recommended Suction Line Sizes, R-12, –40° F. .................................................................. 18-32Table 34 Recommended Suction Line Sizes, R-22, 40° F. .................................................................... 18-32Table 35 Recommended Suction Line Sizes, R-22, 25°F. ..................................................................... 18-33Table 36 Recommended Suction Line Sizes, R-22, 15° F. .................................................................... 18-34Table 37 Recommended Suction Line Sizes, R-22, –20° F. .................................................................. 18-35Table 38 Recommended Suction Line Sizes, R-502, 25° F. .................................................................. 18-35Table 39 Recommended Suction Line Sizes, R-502, 15° F. .................................................................. 18-36Table 40 Recommended Suction Line Sizes, R-502, –20° F. ................................................................ 18-37Table 41 Recommended Suction Line Sizes, R-502, –40° F. ................................................................ 18-38Table 42 Efficiency Comparison of Single Stage vs. Two Stage Compression Typical Air Cooled Application with Refrigerant R-502 ................................................................ 19-6Table 43 Recommended Discharge Line Sizes for Two Stage Compressors ....................................... 19-10Table 44 Recommended Liquid Line Sizes for Two Stage Compressors.............................................. 19-10Table 45 Recommended Suction Line Sizes for Two Stage Compressors, –60° F. .............................. 19-11Table 46 Recommended Suction Line Sizes for Two Stage Compressors, –60° F. .............................. 19-11Table 47 Recommended Suction Line Sizes for Two Stage Compressors, –80° F. .............................. 19-12Table 48 Recommended Suction Line Sizes for Two Stage Compressors, –80° F. .............................. 19-12


SECTION17BASICAPPLICATIONRECOMMENDATIONS

FUNDAMENTALDESIGNPRINCIPLES

There are certain fundamental refrigeration design principles which are vital to the proper functioning of any system.

1. The system must be clean, dry, and free from all contaminants.

2. The compressor must be operated within safe temperature, pressure, and electrical limits.

3. The system must be designed and operated so that proper lubrication is maintained in the compressor at all times.

4. The system must be designed and operated so that excessive liquid refrigerant does not enter the com-pressor. Refrigeration compressors are designed to pump refrigerant vapor, and will tolerate only a limited quantity of liquid refrigerant.

5. Proper refrigerant feed to the evaporator must be maintained, and excessive pressure drop in the refrigerant piping must be avoided.

If these give steps are accomplished, then operation of the system is reasonably certain to be trouble free. If any one is neglected, then eventual operating problems are almost certain to occur. These basic fundamentals are closely inter-related, and must always be kept in mind with regard to the application of any component, or whenever any change in system operation is con-templated.

COMPRESSORSELECTION

The compressor must be selected for the capacity required at the desired operating conditions in accor-dance with the manufacturer’s recommendations for the refrigerant to be used. Standard Copeland® brand single stage compressors are approved for operation with a given refrigerant in one of the following operating ranges. Evaporat-ing TemperatureHigh Temperature 45° F. to 0° F. or 55° F. to 0° F.

Medium Temperature 25° F. to -5° F. Low Temperature 0° F. to -40° F.

Extra Low Temperature -20° F. to -40° F.

Operation at evaporating temperatures above the ap-proved operating range may overload the compressor motor. Operation at evaporating temperatures below the approved operating range is normally not a problem if the compressor motor can be adequately cooled, and discharge temperatures can be kept within allowable lim-its. Evaporating temperatures below -40° F. are normally beyond the practical lower limit of single stage operation because of compressor inefficiencies and excessive discharge gas temperatures. Because of problems of motor cooling or overloading, some motor-compressors may have approval for operation at limited condensing or evaporating temperatures within a given range, and if so, these limitations will be shown by limited perfor-mance curves on the specification sheet.

A given compressor may be approved in two different operating ranges with different refrigerants, for example, high temperature R-12 and low temperature R-502. Since the power requirements for a given displacement with both R-22 and R-502 are somewhat similar, in some cases a compressor may be approved in the same operating range for either of these refrigerants.

Two stage compressors may be approved for evapo-rating temperatures as low as -80° F., but individual compressor specifications should be consulted for the approved operating range.

Operation at temperatures below -80° F. is normally beyond the practical efficiency range of Copeland® brand two stage compressors, and for lower evaporating temperatures, cascade systems should be employed.

Compressors with unloaders have individually estab-lished minimum operating evaporating temperatures since motor cooling is more critical with these compres-sors. As the compressor is unloaded, less refrigerant is circulated through the system, and consequently less return gas is available for motor cooling purposes. Copeland® brand motor-compressors should never be operated beyond published operating limits without prior approval of the Emerson Climate Technologies, Inc. Application Engineering Department.

SYSTEMBALANCE

If the compressor or condensing unit selected for a given application is to satisfactorily handle the refrigeration load, it must have sufficient capacity. However, over ca-

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pacity can be equally as unsatisfactory as under capacity, and care must be taken to see that the compressor and evaporator balance at the desired operating conditions. Checking the proposed system operation by means of a compressor-evaporator-condenser balance chart as described in Section 16 is recommended.

If fluctuations in the refrigeration load are to be ex-pected, which could result in compressor operation at excessively low suction pressures, then some means of capacity control must be provided to maintain acceptable evaporating temperatures. If compressors with unloaders are not available or suitable, and if the load cannot be adequately handled by cycling the compressor, a hot gas bypass circuit may be required.

REFRIGERANT

Copeland® brand compressors are primarily designed for operation with Refrigerants 12, 22, and 502. Opera-tion with other refrigerants in cascade systems may be satisfactory if the proper motor and displacement combination is selected, adequate lubrication can be maintained, and if adequate compressor protection is provided.

R-502 is highly recommended for all single stage low temperature applications, and particularly where evaporating temperatures of -20° F. and below may be encountered. Because of the undesirable high discharge temperatures of R-22 when operated at high compres-sion ratios, R-22 should not be used in single stage low temperature compressors 5 HP and larger.

Different expansion valves are required for each refrig-erant, so the refrigerants are not interchangeable in a given system, and should never be mixed. If for some reason it is desirable to change from one refrigerant to another in an existing system, it is usually possible to convert the system by changing expansion valves and control settings providing the existing piping sizes and component working pressures are compatible. In some cases the existing motor-compressor may be satisfac-tory—for example, in converting from R-22 to R-502. If the conversion will result in higher power requirements as is the case in changing from R-12 to R502, then it may also be necessary to change the motor-compressor.

The refrigerant charge should be held to the minimum required for satisfactory operation, since an abnormally high charge will create potential problems of liquid re-frigerant control.

COMPRESSORCOOLING

Refrigerant cooled motor-compressors are dependent

on return suction gas for motor cooling, and to a con-siderable extent, on both air and refrigerant cooled motor-compressors, the discharge gas temperature is directly related to the temperature of the return suction gas. Discharge temperatures above 325° F. to 350° F. contribute to oil breakdown and valve plate damage, and to avoid compressor damage, operating temperatures must be kept below this level. Peak temperatures occur at the discharge valves, and normally the temperature of the discharge line will be from 50° F. to 100°F. be-low the temperature at the valve plate. Therefore the maximum allowable discharge line temperatures from 225°F. to 250°F.

Suction gas entering the compressor should be no higher than 65°F. under low temperature load condi-tions, or 90°F. under high temperature load conditions, and must never exceed 100°F. On some abnormally critical low temperature applications it may be desirable to insulate the suction lines and return the suction gas to the compressor at lower than normal temperatures to prevent the discharge temperatures from exceeding safe limits, but this is not normally necessary on com-mercial application where the saturated evaporating temperature is -40°F. or above. The low discharge temperature characteristics of R-502 have made pos-sible much more trouble free operation in single stage low temperature applications.

Air cooled motor-compressors must have a sufficient quantity of air impinging directly on the compressor body for motor cooling. Refrigerant cooled motor-compres-sors are cooled adequately by the refrigerant vapor at evaporating temperatures above 0°F., but at evaporating temperatures below 0°F., additional motor cooling by means of air flow is necessary.

On air cooled condensing units, adequate cooling can normally be accomplished by locating the compressor in the discharge air blast from the condenser fan. For proper cooling, the fan must discharge air directly against the compressor, since the compressor usually cannot be adequately cooled by air pulled through a compartment in which the compressor is located. If the compressor is not located in the condenser discharge air stream, cooling must be provided by means of an auxiliary fan discharging air directly again the compressor body. On Copeland® brand compressors with multiple heads such as the 4R and 6R models, auxiliary horizontal airflow may not provide satisfactory cooling, and vertical cool-ing fans are required.

Water cooled compressors are provided with a water jacket or are wrapped with a copper water coil, and wa-ter must be circulated through the compressor cooling circuit before entering the condenser.

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Two-stage compressors are equipped with a desuper-heating expansion valve for interstage cooling, and no auxiliary cooling is required.

If compressors or condensing units are located in a ma-chine room, adequate ventilation air must be provided to avoid an excessive temperature rise in the room. To allow for peak summer temperatures a 10°F. temperature rise is recommended, although a 15°F. rise in cooler ambients might be acceptable.

The most accurate calculation is to determine the total heat to be rejected by adding the compressor refriger-ating capacity at the design operating condition to the heat equivalent of the motor input. The CFM can then be calculated by the formula…

BTU/HRCMF = ºTD For example, determine the machine room ventilation for an air cooled condensing unit operating at -25°F. evaporator, 120°F. condensing with a net refrigeration capacity of 23,000 BTU/HR, 6,400 watts input to the compressor motor, and a 1 H.P. condenser fan motor.

Compressor capacity 23,000 BTU/HRHeat equivalent 6400 watts x 3.413 21,843 BTU/HRHeat equivalent 1 H.P. fan motor 3,700 BTU/HRTotal Heat to be Rejected 48,543 BTU/HR

48,543 BTU/HRCMF = 10º TD = 4.854 CFM With remote condensers, approximately 10% of the heat rejected is given off by the compressor casting and the discharge tubing, and the ventilation can be calculated accordingly.

For convenience, table 20A gives a quick estimate of the ventilation air requirement if only the compressor capacity is known.

COMPRESSORLUBRICATION

An adequate supply of oil must be maintained in the crankcase at all times to insure continuous lubrication. The normal oil level should be maintained at or slightly above the center of the sight glass while operating. An excessive amount of oil must not be allowed in the system as it may result in slugging and possible damage to the compressor valves.

Compressors leaving the factory are charged with naphthenic refrigerant oils. A complete list of acceptable refrigerants and lubricants are listed on form #93-11. The use of any other oil must be specifically cleared

with the Emerson Climate Technologies, Inc. Application Engineering Department. The naphthenic base oil has definite advantages over paraffinic base oils because separation of refrigerant from paraffinic oils occurs at substantially higher temperatures with the same oil-refrigerant concentration. When this separation or two phase condition exists the oil floats on top of the refrig-erant and the oil pump inlet at the bottom of the sump is fed almost pure refrigerant at start up. The resulting improper lubrication can result in bearing failure. Because of the lower separating temperature of naphthenic oil, the possibility of two-phasing is greatly reduced.

Copelametic® compressors are shipped with a generous supply of oil in the crankcase. However the system may require more or less oil depending on the refrigerant charge and the system design. On field installed sys-tems, after the system stabilizes at its normal operating conditions, it may be necessary to add or remove oil to maintain the desired level.

OILPRESSURESAFETYCONTROL

A major percentage of all compressor failures are caused by lack of proper lubrication. Improper lubrication or the loss of lubrication can be due to a shortage of oil in the system, logging of oil in the evaporator or suction line due to insufficient refrigerant velocities, shortage of refriger-ant, refrigerant migration or floodback to the compressor crankcase, failure of the oil pump, or improper operation of the refrigerant control devices.

Regardless of the initial source of the difficulty, the great majority of compressor failures due to loss of lubrication could have been prevented. Although proper system design, good preventive maintenance, and operation within the system’s design limitations are the only cure for most of these problems, actual compressor damage usually can be averted by the use of an oil pressure safety control.

An oil pressure safety control with a time delay of 120 seconds is a mandatory requirement of the Emerson Climate Technologies, Inc. warranty on all Copelametic® compressors having an oil pump. The control oper-

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ates on the differential between oil pump pressure and crankcase pressure, and the two minute delay serves to avoid shut down during short fluctuations in oil pres-sure during start-up.

A trip of the oil pressure safety switch is a warning that the system has been without proper lubrication for a period of two minutes. Repeated trips of the oil pressure safety control are a clear indication that something in the system design or operation requires immediate remedial action. On a well designed system, there should be no trips of the oil pressure safety control, and repeated trips should never be accepted as a normal part of the system operation.

The oil pressure safety control will not protect against all lubrication problems. It cannot detect whether the compressor is pumping oil or a combination of refriger-ant and oil. If bearing trouble is encountered on systems where the oil pressure safety control has not tripped, even though inspection proves it to be properly wired, with the proper pressure setting, and in good operating condition, marginal lubrication is occurring which prob-ably is due to liquid refrigerant floodback.

OILSEPARATORS

Proper refrigerant piping design and operation of the system within its design limits so that adequate refriger-ant velocities can be maintained are the only cure for oil logging problems, but an oil separator may be a definite aid in maintaining lubrication where oil return problems are particularly acute.

For example, consider a compressor having an oil charge of 150 ounces, with the normal oil circulation rate being 2 ounces per minute. This means that on a normal system with proper oil return at stabilized condi-tions, two ounces of oil leave the compressor through the discharge line every minute, and two ounces return through the suction line. If a minimum of 30 ounces of oil in the crankcase is necessary to properly lubricate the compressor, and for some reason oil logged in the system and failed to return to the compressor, the compressor would run out of oil in 60 minutes. Under the same conditions with an oil separator having an efficiency of 80%, the compressor could operate 300 minute or 5 hours before running out of oil.

As a practical matter, there seldom are conditions in a system when no oil will be returned to the compressor, and even with low gas velocities, some fraction of the oil leaving the compressor will be returned. If there are regular intervals of full load conditions or defrost periods when oil can be returned normally, an oil separator can help to bridge long operating periods at light load con-

ditions. Oil separators are mandatory on systems with flooded evaporators controlled by a float valve, on all two stage and cascade ultra-low temperature systems, and on any system where oil return is critical.

Oil separators should be considered as a system aid but not a cure-all or a substitute for good system de-sign. They are never 100% efficient, and in fact may have efficiencies as low as 50% depending on system operating conditions. On systems where piping design encourages oil logging in the evaporator, an oil separa-tor can compensate for system oil return deficiencies only on a temporary basis, and may only serve to delay lubrication difficulties.

If a system is equipped with a suction accumulator, it is recommended that the oil return from the separa-tor be connected to the suction line just ahead of the accumulator. This will provide maximum protection against returning liquid refrigerant to the crankcase. If the system is not equipped with a suction accumulator, the oil return line on suction cooled compressors may be connected to the suction line if more convenient than the crankcase, but on air cooled compressors, oil return must be made directly to the crankcase to avoid damage to the compressor valves.

If the separator is exposed to outside ambient tem-peratures, it must be insulated to prevent refrigerant condensation during off periods, resulting in return of liquid to the compressor crankcase. Small low wattage strap-on heaters are available for oil separators, and if any problem from liquid condensation in the separator is anticipated, a continuously energized heater is highly recommended.

SUCTIONLINEACCUMULATORS

If liquid refrigerant is allowed to flood through a refrigera-tion or air conditioning system and return to the compres-sor before being evaporated, it may cause damage to the compressor due to liquid slugging, loss of oil from the crankcase, or bearing washout. To protect against this condition on systems vulnerable to liquid damage a suction accumulator may be necessary.

The accumulator’s function is to intercept liquid re-frigerant before it can reach the compressor valves or crankcase. It should be located in the suction line near the compressor, and if a reversing valve is used in the system, the accumulator must be located between the reversing valve and the compressor. Provisions for posi-tive oil return to the crankcase must be provided, but a direct gravity flow which will allow liquid refrigerant to drain to the crankcase during shut-down periods must be avoided. The liquid refrigerant must be metered back

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to the compressor during operation at a controlled rate to avoid damage to the compressor.

Some systems, because of their design, will periodically flood the compressor with liquid refrigerant. Typically, this can occur on heat pumps at the time the cycle is switched from cooling to heating, or from heating to cool-ing. The coil which has been serving as the condenser is partially filled with liquid refrigerant, and when suddenly exposed to suction pressure, the liquid is dumped into the suction line. On heat pumps equipped with expansion valves, there may be further flooding due to the inability of the expansion valve to effectively control refrigerant feed for a short period after the cycle change until the system operation is again stabilized.

A similar situation can occur during defrost cycles. With hot gas defrost, when the defrost cycle is initiated, the sudden introduction of high pressure gas into the evapo-rator may force the liquid refrigerant in the evaporator into the suction line. If the defrost cycle is such that the evaporator can fill with condensed liquid during defrost, or on systems utilizing electric defrost without a pump-down cycle, an equally dangerous situation may exist at the termination of the defrost cycle.

On systems with a large refrigerant charge, or on any system where liquid floodback is likely to occur, a suc-tion line accumulator is strongly recommended. On heat pumps, truck applications, and on any system where liquid slugging can occur during operation, a suction line accumulator is mandatory for compressor protection unless otherwise approved by the Emerson Climate Technologies, Inc. Application Engineering Department. The actual refrigerant holding capacity needed for a given accumulator is governed by the requirements of the particular application, and the accumulator should be selected to hold the maximum liquid floodback an-ticipated.

PUMPDOWNSYSTEMCONTROL

Refrigerant vapor will always migrate to the coldest part of the system, and if the compressor crankcase can become colder than other parts of the system, refrigerant in the condenser, receiver, and evaporator will vaporize, travel through the system, and condense in the compressor crankcase.

Because of the difference in vapor pressures of oil and refrigerant, refrigerant vapor is attracted to refrigera-tion oil, and even though no pressure or temperature difference exists to cause a flow, refrigerant vapor will migrate through the system and condense in the oil until the oil is saturated. During off cycles extending several hours or more, it is possible for liquid refrigerant to al-

most completely fill the compressor crankcase due to the oil attraction. For example, in a system using R-12 refrigerant which is allowed to equalize at an ambient temperature of 70° F., the oil-refrigerant mixture in the crankcase will end up about 70% refrigerant before equilibrium is reached.

The most positive and dependable means of keeping refrigerant out of the compressor crankcase is the use of a pumpdown cycle. By closing a liquid line solenoid valve, the refrigerant can be pumped into the condenser and receiver, and the compressor operation controlled by means of a low pressure control. The refrigerant can thus be isolated during periods when the compressor is not in operation, and migration of refrigerant to the compressor crankcase is prevented.

Pumpdown control can be used on all thermostatic ex-pansion valve systems with the addition of a liquid line solenoid valve, provided adequate receiver capacity is available. Slight refrigerant leakage may occur through the solenoid valve, causing the suction pressure to rise gradually, and a recycling type control is recommended to repeat the pumpdown cycle as required. The occasional short cycle usually is not objectionable.

A pumpdown cycle is highly recommended whenever it can be used. If a non-recycling pumpdown circuit is required, then consideration should be given to the use of a crankcase heater in addition to the pumpdown for more dependable compressor protection.

CRANKCASEHEATERS

On some systems operating requirements, noise consid-erations, or customer preference may make the use of a pumpdown system undesirable, and crankcase heaters are frequently used to control migration.

By warming the oil, the absorption of refrigerant by the oil is minimized, and under mild weather conditions, any liquid refrigerant in the crankcase can be vaporized and forced out of the compressor. For effective protec-tion, heaters must be energized several hours before starting the compressor. It is recommended that they be energized continuously, independent of compressor operation. Improperly sized heaters can overheat the oil, and heaters used on Copeland® brand compres-sors must be specifically approved by the Emerson Climate Technologies, Inc. Application Engineering Department.

It would be a mistake to assume that crankcase heaters are a dependable cure for all migration problems. As the ambient conditions contributing to migration worsen, the ability of the crankcase heater to keep refrigerant out

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of the crankcase decreases. If the suction line slopes toward the compressor, and the temperature to which the suction line is exposed is sufficiently lower than the temperature of the oil, refrigerant may condense in the suction line and flow back to the compressor by gravity at a rate sufficient to offset the heat introduced by the heater. Heaters will not protect against liquid slugs or excessive liquid flooding. However, where operating conditions are not too severe, crankcase heaters can provide satisfactory protection against migration.

Where a pumpdown cycle is not used, crankcase heaters are mandatory on heat pumps, and on other air con-ditioning applications if the refrigerant charge exceeds the established limits for Copeland® brand compres-sors, unless tests prove the compressor is adequately protected by other means.

To prevent possible damage in shipment, crankcase heat-ers are not installed on compressors at the factory.

CRANKCASEPRESSUREREGULATINGVALVES

In order to limit the power requirement of the compressor to the allowable operating limit, a crankcase pressure regulating valve may be necessary. This most frequently occurs on low temperature compressors where the power requirement during pulldown periods or after defrost may be greatly in excess of the compressor motor’s capabilities. Copeland® brand compressors should not be operated at suction pressures in excess of the published limits on compressor specification sheets without approval of the Emerson Climate Technologies, Inc. Application Engineering Department.

Since any pressure drop in the compressor suction line lowers the system capacity, the CPR valve should be sized for a minimum pressure drop. In order to restrict pull down capacity as little as possible, the valve set-ting should be as high as the motor power requirement will allow.

Thermal expansion valves of the pressure limiting type are not recommended when a CPR valve is used, par-ticularly if the pressure settings are fairly close, because of the possibility of the action of the two valves coming in conflict in their response to system pressures.

LOWAMBIENTHEADPRESSURECONTROL

Within the operating limitations of the system, it is desir-able to take advantage of lower condensing tempera-tures whenever possible for increased capacity, lower discharge temperatures, and lower power requirements. However, too low a discharge pressure can produce serious malfunctions. Since the capacity of capillary

tubes and expansion valves is proportional to the dif-ferential pressure across the capillary tube or valve, a reduction in discharge pressure will reduce its capacity and produce a drop in evaporating pressure.

Low discharge pressures can result in starving the evaporator coil with resulting oil logging, short cycling on low pressure controls, reduction of system capacity, or erratic expansion valve operation.

Systems with water cooled condensers and cooling towers require water regulating valves, or some other means of controlling the temperature or the quantity of water passing through the condenser.

If air cooled air conditioning systems are required to operate in ambient temperatures below 60° F., a suitable means of controlling head pressures must be provided. Refrigeration systems are also vulnerable to damage from low head pressure conditions, and adequate head pressure controls should be provided for operation in ambient temperatures below 50°F.

Several proprietary control systems are available for low ambient operation, most of which maintain head pres-sure above a preset minimum by partially flooding the condenser and thus reducing the effective surface area. Methods of this type can control pressures effectively, but do require a considerable increase in refrigerant charge and adequate receiver capacity must be provided.

Air volume dampers on the condenser operated from refrigerant discharge pressure provide a simple, eco-nomical, and effective means of control which is widely used.

Adequate protection at lowest cost can often be pro-vided by a reverse acting high pressure control which senses discharge pressure, and acts to disconnect the condenser fan circuit when the head pressure falls below the control’s minimum setting. The proper adjustment of the off-on differential is particularly important to avoid excessive fan motor cycling, and the resulting fluctua-tions in discharge pressure may contribute to uneven expansion valve feeding. In cold ambient temperatures the condenser must be shielded from the wind.

LIQUIDLINEFILTER-DRIER

A liquid line filter-drier must be used on all field installed systems, and on all systems opened in the field for service. Filter-driers are highly recommended for all sys-tems, but are not mandatory on factory assembled and charged units where careful dehydration and evacuation is possible during manufacture. Precharged systems with quick connect fittings having a rupture disc are considered to be the equivalent of factory charged systems.

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Moisture can be a factor in many forms of system dam-age, and the reduction of moisture to an acceptable level can greatly extend compressor life and slow down harmful reactions. The desiccant used must be capable of removing moisture to a low end point and further should be of a type which can remove a reasonable quantity of acid. It is most important that the filter-drier be equipped with an excellent filter to prevent circulation of carbon and foreign particles.

SIGHTGLASSANDMOISTUREINDICATOR

A combination sight glass and moisture indicator is es-sential for easy field maintenance on any system, and is required on any field installed system unless some other means of checking the refrigerant charge is provided.

A sight glass is a convenient means of determining the refrigerant charge, showing bubbles when there is insufficient charge, and a solid clear glass when there is sufficient charge. However, the operator should bear in mind that under some circumstances even when the receiver outlet has a liquid seal, bubbles or flash gas may show in the sight glass. This may be due to a restriction or excessive pressure drop in the receiver outlet valve, a partially plugged drier or strainer, or other restriction in the liquid line ahead of the sight glass. If the expansion valve feed is erratic or surging, the increased flow when the expansion valve is wide open can create sufficient pressure drop to cause flashing at the receiver outlet.

Another source of flashing in the sight glass may be rapid fluctuations in compressor discharge pressure. For example, in a temperature controlled room, the sudden opening of shutters or the cycling of a fan can easily cause a reduction in the condensing temperature of 10°F. to 15°F. Any liquid in the receiver may then be at a temperature equivalent to the lower condensing pressure, and flashing will continue until the system has stabilized at the new condensing temperature.

While the sight glass can be a valuable aid in servic-ing a refrigeration or air conditioning system, a more positive liquid indicator is desirable, and the system performance must be carefully analyzed before placing full reliance on the sight glass as a positive indicator of the system charge.

LIQUIDLINESOLENOIDVALVE

A liquid line solenoid valve is recommended on all field installed systems with large refrigerant charges, particularly when the system has a charge in excess of three pounds of refrigerant per motor HP. The solenoid valve will prevent continued feed to the evaporator through the expansion valve or capillary tube when the

compressor is not operating, and will control migration of liquid refrigerant from the receiver and condenser to the evaporator and compressor crankcase.

If a pumpdown cycle is not used, the liquid line solenoid valve should be wired to the compressor motor terminals so that the valve will be de-energized when the motor is not operating.

HEATEXCHANGER

A liquid to suction heat exchanger is highly recom-mended on all refrigeration systems, and is required on package water chillers and water to water heat pumps because of the low operating superheat. On medium and low temperature applications, a heat exchanger increases system capacity, helps to eliminate flashing of liquid refrigerant ahead of the expansion valve, and aids both in preventing condensation on suction lines and in evaporating any liquid flooding through the evaporator.

On small systems, soldering the liquid and suction lines together for several feet makes an effective heat exchanger.

THERMOSTATICEXPANSIONVALVES

Thermostatic expansion valves must be selected and applied in accordance with the manufacturer’s instruc-tions. Either internally equalized or externally equalized valves will feed properly if applied correctly. If the thermal expansion valve is of the externally equalized type, the external equalizer line must be connected, preferably at a point beyond the expansion valve thermal bulb. Do not cap or plug the external equalizer connection as the valve will not operate without this connection.

Valve superheat should be preset by the valve manu-facturer, and field adjustment should be discouraged. Valves in need of adjustment should be set to provide 5°F. to 10°F. superheat at the thermal bulb location. Too high a superheat setting will result in starving the evaporator, and can cause poor oil return. Too low a superheat setting will permit liquid floodback to the compressor.

A minimum of 15°F. superheat at the compressor must be maintained at all times to insure the return of dry gas to the compressor suction chamber, and a minimum of 20°F. superheat is recommended. Note that this is not superheat at the expansion valve, but should be calcu-lated from pressure measured at the suction service valve and the temperature measured 18” from the compres-sor on the bottom of the horizontal run of suction line tubing. Lower superheat can result in liquid refrigerant

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flooding back to the compressor during variations in the evaporator feed with possible compressor damage as a result. Excessively wet refrigerant vapor continually returning to the compressor can reduce the lubricating qualities of the oil and greatly increase compressor wear, as well as resulting in a loss of capacity.

It is important that users realize that flash gas in the liquid line can seriously affect expansion valve control. So long as a head of pure liquid refrigerant is maintained at the expansion valve, its performance is relatively stable. But if flash gas is mixed with liquid refrigerant fed to the valve, a larger orifice opening is required to feed the same weight of liquid refrigerant. The only way the orifice opening can be increased is by an increase in superheat, and as the percentage of flash gas increases, the superheat increases, the valve opens wide, and the evaporator is progressively more starved.

If the valve has been operating with a large percentage of flash gas entering the expansion valve, and a head of pure liquid refrigerant is suddenly restored, the orifice opening will be larger than required for the load, and liquid will flood through the system to the compressor until the valve again regains control. Conventional expansion valves with the thermal bulb strapped to the suction line may be somewhat sluggish in response, and it may be several minutes before control can be restored to normal.

Typically, changes in the quality of liquid refrigerant feed-ing the expansion valve can occur quickly and frequently because of the action of head pressure control devices, sudden changes in the refrigeration load, hunting of the expansion valve, action of an unloading valve, or rapid changes in condensing pressure.

On systems with short suction lines and low superheat requirements, quick response thermal bulbs or wells in the suction line may be essential to avoid periodic floodback to the compressor.

Temperatures and pressure alone may not give a true picture of the actual liquid refrigerant control in a system. Excessive oil circulation has the effect of increasing the evaporating temperature of the refrigerant. The response of the expansion valve is based on the saturation pressure and temperature of pure refrigerant. In an operating sys-tem, the changed pressure-temperature characteristics of the oil rich refrigerant will give the expansion valve a false reading of the actual superheat, and can result in a somewhat lower actual superheat than apparently exists, causing excessive liquid refrigerant floodback to the compressor. The only real cure for this condition is to reduce oil circulation to a minimum. Normally excessive oil in the evaporator can only result from an excessive

system oil charge or other factors which could cause excessive oil circulation, or from low velocities in the evaporator which result in oil logging. In low tempera-ture applications where proper oil circulation cannot be maintained, an oil separator may be required.

Vapor charged valves are satisfactory for air condition-ing usage, and are desirable in many cases because of their inherent pressure limiting characteristics. For all refrigeration applications, liquid charged valves should be used to prevent condensation of the charge in the head of the valve and the resulting loss of control in the event the head becomes colder than the thermal bulb.

A pressure limiting type valve may be helpful in limiting the compressor load, and also prevents excessive liquid refrigerant floodback on start-up. On systems using hot gas defrost, the defrost load is normally greater than the refrigeration load, and some other means of limiting the compressor power input must be used if required.

The thermostatic expansion valve must be sized prop-erly for the load. Although a given valve normally has a wide operating capacity range, excessively undersized or oversized valves can cause system malfunctions. Undersized valves may starve the evaporator, and the resulting excessive superheat may adversely affect the system performance. Oversized valves can cause hunting, alternately starving and flooding the evaporator, resulting in extreme fluctuations in suction pressure.

The thermal bulb should normally be located on a hori-zontal section of the suction line, close to the evaporator outlet, on the evaporator side of any suction line trap or heat exchanger. Do not under any circumstances locate the thermal bulb in a location where the suction line is trapped since this can result in erratic feeding. Satisfactory performance can usually be obtained with the bulb strapped to the suction line at the 3 o’clock position. Mounting on the top of the suction line will de-crease sensitivity, and may allow possible liquid flooding. Mounting on the bottom of the suction line can cause erratic feeding due to the rapid temperature changes that can result from even small amounts of liquid refrigerant reaching the thermal bulb location. Particular attention should be given to the location of the thermal bulb on multiple evaporator systems to insure that the refriger-ant returning from one evaporator does not affect the control of another evaporator.

EVAPORATORS

Evaporators must be properly selected for the refrig-eration load. Too large an evaporator might result in low velocities and possible oil logging. Too small an evaporator will have excessive temperature differentials

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between the evaporating refrigerant and the medium to be cooled. The allowable TD between the entering air and the evaporating refrigerant may also be dictated by the humidity control required.

Internal volume of the evaporator tubing should be at a minimum to keep the system refrigerant charge as low as possible, so the smallest diameter tubing that will give acceptable performance should be used. Since pressure drop at low evaporating temperatures is criti-cal so far as capacity is concerned, multiple refrigerant circuits with fairly short runs are preferred. At the same time, it is essential that velocities of refrigerant in the evaporator be high enough to avoid oil trapping.

Vertical headers should have a bottom outlet to allow gravity oil drainage.

SUCTIONLINEFILTERS

A heavy duty suction line filter is recommended for every field installation. The filter will effectively remove contaminants from the system at the time of installation, and serves to keep the compressor free of impurities during operation. In the event of a motor burn, the filter will prevent contamination from spreading into the system through the suction line.

The suction line filter should be selected for a reason-able pressure drop, and should be equipped with a pressure fitting just ahead of the filter, preferably in the shell, to facilitate checking pressure drop across the filter during operation.

HIGHANDLOWPRESSURECONTROLS

Both high and low pressure controls are recommended for good system design on all air cooled systems 1 HP and larger, and are essential on all field installed air cooled systems and on all water cooled systems.

When used for low temperature unit operation control, the low pressure control must not be set below the minimum operating limits of the compressor or the system. One of the most frequent causes of motor overheating and inadequate lubrication is operation of the compressor at excessively low suction pressures. Product specification sheets list the approved compres-sor operating range, and recommended minimum low pressure control settings for various operating ranges are shown in Table 21.

High pressure controls may be either manual or automatic reset as desired by the customer. If of the manual reset type, provision must be made to prevent liquid refrigerant flooding through the system to the compressor in the event of a trip of the high pressure control.

Internal automatic reset pressure relief valves (Co-pelimit) are provided in most welded compressors 1 ¾ HP and larger. On factory assembled package systems, the internal Copelimit valve may satisfy U.L. and code requirements without the use of an external high pressure control. A similar high side to low side automatic reset pressure relief valve is installed in all Copelametic® compressors with displacements of 3,000 CFH or greater.

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On factory assembled and charged package systems, such as room air conditioners, where loss of charge protection is not considered critical, or where the motor protection device can provide loss of charge protection, low pressure controls may not be essential although recommended.

INTERCONNECTEDSYSTEMS

When the crankcases of two or more compressors are interconnected for parallel operation on a single refrigeration system, serious problems of oil return and vibration may be encountered unless the system is properly designed. The tandem compressor consisting of two individual compressors with an interconnecting housing replacing the individual stator covers provides a simple, trouble free solution to this problem.

Because of the potential operating problems, intercon-nection of individual compressors is not approved with the exception of factory designed, tested, and assembled units specifically approved by the Emerson Climate Tech-nologies, Inc. Application Engineering Department.

ELECTRICALGROUPFUSING

Individual circuit breakers or fuses should be provided for each compressor motor. Group fusing, where two or more compressors are installed on one fused discon-nect, is not recommended since an electrical failure in one compressor would not trip the fuse, and extensive electrical damage could result.

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SECTION18REFRIGERATIONPIPING

Probably the first skill that any refrigeration apprentice mechanic learns is to make a soldered joint, and run-ning piping is so common a task that often its critical importance in the proper performance of a system is overlooked. It would seem elementary in any piping system that what goes in one end of a pipe must come out the other, but on a system with improper piping, it is not uncommon for a serviceman to add gallons of oil to a system, and it may seemingly disappear without a trace. It is of course lying on the bottom of the tubing in the system, usually in the evaporator or suction line. When the piping or operating condition is corrected, the oil will return and those same gallons of oil must be removed.

Refrigeration piping involves extremely complex relation-ships in the flow of refrigerant and oil. Fluid flow is the name given in mechanical engineering to the study of the flow of any fluid, whether it might be a gas or a liquid, and the inter-relationship of velocity, pressure, friction, density, viscosity, and the work required to cause the flow. These relationships evolve into long mathematical equations which form the basis for the fan laws which govern fan performance, and the pressure drop tables for flow through piping. But 99% of the theories in fluid flow textbooks deal with the flow of one homogenous fluid, and there is seldom even a mention of a combi-nation flow of liquid, gas, and oil such as occurs in any refrigeration system. Because of its changing nature, such flow is just too complex to be governed by a simple mathematical equation, and practically the entire working knowledge of refrigeration piping is based on practical experience and test data. As a result, the general type of gas and liquid flow that must be maintained to avoid problems is known, but seldom is there one exact an-swer to any problem.

BASICPRINCIPLESOFREFRIGERATIONPIPINGDESIGN

The design of refrigeration piping systems is a continuous series of compromises. It is desirable to have maximum capacity, minimum cost, proper oil return, minimum power consumption, minimum refrigerant charge, low noise level, proper liquid refrigerant control, and perfect flexibility of system operation from 0 to 100% of system capacity without lubrication problems. Obviously all of these goals cannot be satisfied, since some are in direct conflict. In order to make an intelligent decision as to just what type of compromise is desirable, it is essential that the piping designer clearly understand the basic effects on system performance of the piping design in the different parts of the system.

In general, pressure drop in refrigerant lines tends to decrease capacity and increase power requirements, and excessive pressure drops should be avoided. The mag-nitude of the pressure drop allowable varies depending on the particular segment of piping involved, and each part of the system must be considered separately. There are probably more tables and charts available covering line pressure drop and refrigerant line capacities at a given pressure drop than on any other single subject in the field of refrigeration.

It is most important, however, that the piping designer realize that pressure drop is not the only criteria that must be considered in sizing refrigerant lines, and that often refrigerant velocities rather than pressure drop must be the determining factor in system design. In addition to the critical nature of oil return, there is no better invitation to system difficulties than an exces-sive refrigerant charge. A reasonable pressure drop is far more preferable than over-sized lines which can contain refrigerant far in excess of the system’s needs. An excessive refrigerant charge can result in serious problems of liquid refrigerant control, and the flywheel effect of large quantities of liquid refrigerant in the low pressure side of the system can result in erratic opera-tion of the refrigerant control devices.

The size of the service valve supplied on a compressor, or the size of the connection on a condenser, evapora-tor, accumulator, or other accessory does not determine the size of line to be used. Manufacturers select a valve size or connection fitting on the basis of its application to an average system, and such factors as the type of application, length of connecting lines, type of system control, variation in load, and other factors can be major factors in determining the proper line size. It is quite possible the required line size may be either smaller or larger than the fittings on various system components. In such cases, reducing fittings must be used.

Since oil must pass through the compressor cylinders to provide lubrication, a small amount of oil is always circulating with the refrigerant. Refrigeration oils are soluble in liquid refrigerant, and at normal room tem-peratures they will mix completely. Oil and refrigerant vapor, however, donot mix readily, and the oil can be properly circulated through the system only if the mass velocity of the refrigerant vapor is great enough to sweep the oil along. To assure proper oil circulation, adequate refrigerant velocities must be maintained not only in the suction and discharge lines, but in the evaporator circuits as well.

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Several factors combine to make oil return most critical at low evaporating temperatures. As the suction pres-sure decreases and the refrigerant vapor becomes less dense, the more difficult it becomes to sweep the oil along. At the same time as the suction pressure falls, the compression ratio increases, and as a result compres-sor capacity is reduced, and the weight of refrigerant circulated decreases. Refrigeration oil alone becomes the consistency of molasses at temperatures below 0°F., but so long as it is mixed with sufficient liquid refrigerant, it flows freely. As the percentage of oil in the mixture increases, the viscosity increases.

At low temperature conditions all of these factors start to converge, and can create a critical condition. The density of the gas decreases, the mass velocity flow decreases, and as a result more oil starts accumulating in the evaporator. As the oil and refrigerant mixture be-comes more viscous, at some point oil may start logging in the evaporator rather than returning to the compressor, resulting in wide variations in the compressor crankcase oil level in poorly designed systems.

Oil logging can be minimized with adequate velocities and properly designed evaporators even at extremely low evaporating temperatures, but normally oil separators are necessary for operation at evaporating temperatures below -50°F. in order to minimize the amount of the oil in circulation.

COPPERTUBINGFORREFRIGERANTPIPING For installations using R-12, R-22, and R-502, copper tubing is almost universally used for refrigerant piping. Commercial copper tubing dimensions have been stan-dardized and classified as follows:

Type K Heavy Wall Type L Medium Wall Type M Light Wall

Only types K or L should be used for refrigerant piping, since type M does not have sufficient strength for high pressure applications. Type L tubing is most commonly used, and all tables and data in this manual are based on type L dimensions.

It is highly recommended that only refrigeration grade copper tubing be used for refrigeration applications, since it is available cleaned, dehydrated, and capped to avoid contamination prior to installation. Copper tubing commonly used for plumbing usually has oils and grease or other contaminants on the interior wall, and these can cause serious operating problems if not removed prior to installation.

Table 22 lists the dimensions and properties of standard commercial copper tubing in the sizes commonly used in refrigeration systems, and Table 23 lists the weight of various refrigerants per 100 feet of piping in liquid, suction and discharge lines.

FITTINGSFORCOPPERTUBING

For brazed or soldered joints, the required elbows, tees, couplings, reducers, or other miscellaneous fittings may be either forged brass or wrought copper. Cast fittings are not satisfactory since they may be porous and often lack sufficient strength.

EQUIVALENTLENGTHOFPIPE

Each valve, fitting, and bend in a refrigerant line con-tributes to the friction pressure drop because of its interruption or restriction of smooth flow. Because of the detail and complexity of computing the pressure drop of each individual fitting, normal practice is to establish an equivalent length of straight tubing for each fitting. This allows the consideration of the entire length of line, including fittings, as an equivalent length of straight pipe. Pressure drop and line sizing tables and charts are normally set up on the basis of a pressure drop per 100 feet of straight pipe, so the use of equivalent lengths allows the data to be used directly.

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The equivalent length of copper tubing for commonly used valves and fittings is shown in Table 24.

For accurate calculations of pressure drop, the equiva-lent length for each fitting should be calculated. As a practical matter, an experienced piping designer may be capable of making an accurate overall percentage allowance unless the piping is extremely complicated. For long runs of piping of 100 feet or greater, an allow-ance of 20% to 30% of the actual lineal length may be adequate, while for short runs of piping, an allowance as high as 50% to 75% or more of the lineal length may be necessary. Judgment and experience are necessary in making a good estimate, and estimates should be checked frequently with actual calculations to insure reasonable accuracy.

For items such as solenoid valves and pressure regulating valves, where the pressure drop through the valve is relatively large, data is normally available from the manufacturer’s catalog so that items of this nature can be considered independently of lineal length calculations.

PRESSUREDROPTABLES

Figure 76, 77, and 78 are combined pressure drop charts for refrigerants R-12, R-22, and R-502. Pressure drops in the discharge line, suction line, and liquid line can be determined from these charts for condensing temperatures ranging from 80°F. to 120°F.

To use the chart, start in the upper right hand corner with the design capacity. Drop vertically downward on the line

representing the desired capacity to the intersection with the diagonal line representing the operating condition desired. Then move horizontally to the left. A vertical line dropped from the intersection point with each size of copper tubing to the design condensing temperature line allows the pressure drop in psi per 100 feet of tubing to be read directly from the chart. The diagonal pres-sure drop lines at the bottom of the chart represent the change in pressure drop due to a change in condensing temperature.

For example, in Figure 78 for R-502, the dotted line represents a pressure drop determination for a suction line in a system having a design capacity of 5.5 tons or 66,000 BTU/hr operating with an evaporating tem-perature of -40°F. The 2 5/8” O.D. suction line illustrated has a pressure drop of 0.22 psi per 100 feet at 85°F. condensing temperature, but the same line with the same capacity would have a pressure drop of 0.26 psi per 100 feet at 100°F. condensing, and 0.32 psi per 100 feet at 120°F. condensing.

In the same manner, the corresponding pressure drop for any line size and any set of operating conditions within the range of the chart can be determined.

SIZINGHOTGASDISCHARGELINES

Pressure drop in discharge lines is probably less criti-cal than in any other part of the system. Frequently the effect on capacity of discharge line pressure drop is over-estimated since it is assumed the compressor discharge pressure and the condensing pressure are the same. In fact, there are two different pressures, the compressor discharge pressure being greater than the condensing pressure by the amount of the discharge line pressure drop. An increase in pressure drop in the discharge line might increase the compressor discharge pressure materially, but have little effect on the condens-ing pressure. Although there is a slight increase in the heat of compression for an increase in head pressure, the volume of gas pumped is decreased slightly due to a decrease in volumetric efficiency of the compres-sor. Therefore the total heat to be dissipated through the condenser may be relatively unchanged, and the condensing temperature and pressure may be quite stable, even though the discharge line pressure drop and therefore the compressor discharge pressure might vary considerably.

The performance of a typical Copelametic® compres-sor, operating at air conditioning conditions with R-22 and an air cooled condenser indicates that for each 5 psi pressure drop in the discharge line, the compres-sor capacity is reduced less than ½ of 1%, while the power required is increased about 1%. On a typical low

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temperature Copelametic® compressor operating with R-502 and an air cooled condenser, approximately 1% of compressor capacity will be lost for each 5 psi pres-sure drop, but there will be little or no change in power consumption.

As a general guide, for discharge line pressure drops up to 5 psi, the effect on system performance would be so small as to be difficult to measure. Pressure drops up to 10 psi would not be greatly detrimental to system performance provided the condenser is sized to maintain reasonable condensing pressures.

Actually a reasonable pressure drop in the discharge line is often desirable to dampen compressor pulsation, and thereby reduce noise and vibration. Some discharge line mufflers actually derive much of their efficiency from pressure drop through the muffler.

Discharge lines on factory built condensing units usually are not a field problem, but on systems installed in the field with remote condensers, line sizes must be selected to provide proper system performance.

Because of the high temperatures existing in the dis-charge line, oil flows freely, and oil circulation through both horizontal and vertical lines can be maintained satisfactorily with reasonably low velocities. Since oil traveling up a riser usually creeps up the inner surface of the pipe, oil travel in vertical risers is dependent on the velocity of the gas at the tubing wall. The larger the pipe diameter, the greater will be the required velocity at the center of the pipe to maintain a given velocity at the wall surface. Figures 79 and 80 list the maximum recommended discharge line riser sizes for proper oil return for varying capacities. The variation at different condensing temperatures is not great, so the line sizes shown are acceptable on both water cooled and air cooled applications.

If horizontal lines are run with a pitch in the direction of flow of at least ½” in 10 feet, there is normally little problem with oil circulation at lower velocities in horizontal lines. However, because of the relatively low velocities required in vertical discharge lines, it is recommended wherever possible that both horizontal and vertical dis-charge lines be sized on the same basis.

Toillustratetheuseofthechart,assumeasystemoperatingwithR-22at40°F.evaporatingtemperaturehasacapacityof100,000BTU/hr.TheintersectionofthecapacityandevaporatingtemperaturelinesatpointXonFigure80indicatethedesigncondi-tion.Since this isbelow the21/8”O.D. line, themaximumsizethatcanbeusedtoinsureoilreturnupaverticalriseris15/8”O.D.

Oil circulation in discharge lines is normally a problem only on systems where large variations in system capac-ity are encountered. For example, an air conditioning system may have steps of capacity control allowing it to operate during periods of light load at capacities possibly as low as 25% or 33% of the design capacity. The same situation may exist on commercial refrigera-tion systems where compressors connected in parallel are cycled for capacity control. In such cases, vertical discharge lines mustbe sized to maintain velocities above the minimum necessary to properly circulate oil at the minimum load condition.

For example, consider an air conditioning system using R-12 having a maximum design capacity of 300,000 BTU/hr with steps of capacity reduction up to 66%. Although the 300,000 BTU/hr condition could return oil up a 3 1/8” O.D. riser, at light load conditions the system would have only 100,000 BTU/hr capacity, so a 2 1/8” O.D. riser must be used. In checking the pressure drop chart, Figure 76, at maximum load conditions, a 2 1/8” O.D. pipe will have a pressure drop of approximately 3 psi per 100 feet at a condensing temperature of 120°F.

One other limiting factor in discharge line sizing is excessive velocity which can cause noise problems. Velocities of 3,000 FPM or more may result in high noise levels, and it is recommended that maximum velocities be kept well below this level. Figures 81 and 82 give equivalent discharge line gas velocities for varying capacities and line sizes over the normal refrigeration and air conditioning range.

Because of the flexibility in line sizing that the allowable pressure drop makes possible, discharge lines can al-most always be sized satisfactorily without the necessity of double risers. If modifications are made to an existing system which result in the existing discharge line being oversized at light load conditions, the addition of an oil separator to minimize oil circulation will normally solve the problem.

To summarize, in sizing discharge lines, it is recom-mended that a tentative selection of line size be made on the basis of a total pressure drop of approximately 5 psi plus or minus 50%, the actual design pressure drop to a considerable degree being a matter of the designer’s judgment. Check Figure 79 or 80 to be sure that velocities at minimum load conditions are adequate to carry oil up vertical risers, and adjust vertical riser size if necessary. Check Figure 81 or 82 to be sure velocities at maximum load are not excessive.

Recommended discharge line sizes for varying capaci-ties and equivalent lengths of line are given in Table 28.

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SIZINGLIQUIDLINES Since liquid refrigerant and oil mix completely, velocity is not essential for oil circulation in the liquid line. The primary concern in liquid line sizing is to insure a solid liquid head of refrigerant at the expansion valve. If the pressure of the liquid refrigerant falls below its saturation temperature, a portion of the liquid will flash into vapor to cool the liquid refrigerant to the new saturation tem-perature. This can occur in a liquid line if the pressure drops sufficiently due to friction or vertical lift.

Flash gas in the liquid line has a detrimental effect on system performance in several ways. It increases the pressure drop due to friction, reduces the capacity of the expansion device, may erode the expansion valve pin and seat, can cause excessive noise, and may cause erratic feeding of the liquid refrigerant to the evaporator.

For proper system performance, it is essential that liquid refrigerant reaching the expansion device be subcooled slightly below its saturation temperature. On most systems the liquid refrigerant is sufficiently subcooled as it leaves the condenser to provide for normal system pressure drops. The amount of subcool-ing necessary, however, is dependent on the individual system design.

On air cooled and most water cooled applications, the temperature of the liquid refrigerant is normally higher than the surrounding ambient temperature, so no heat is transferred into the liquid, and the only concern is the pressure drop in the liquid line. Besides the friction loss caused by flow through the piping, a pressure drop equivalent to the liquid head is involved in forcing liquid to flow up a vertical riser. A head of two feet of liquid refrigerant is approximately equivalent to 1 psi. For example, if a condenser or receiver in the basement of a building is to supply liquid refrigerant to an evapora-tor three floors above, or approximately 30 feet, then a pressure drop of approximately 15 psi must be provided for in system design for the liquid head alone.

On evaporative or water cooled condensers where the condensing temperature is below the ambient air temperature, or on any application where liquid lines must pass through hot areas such as boiler or furnace rooms, an additional complication may arise because of heat transfer into the liquid. Any subcooling in the condenser may be lost in the receiver or liquid line due to temperature rise alone unless the system is properly designed. On evaporative condensers where a receiver and subcooling coil are used, it is recommended that the refrigerant flow be piped from the condenser to the receiver and then to the subcooling coil. In critical applications it may be necessary to insulate both the receiver and the liquid line.

On the typical air cooled condensing unit with a conven-tional receiver, it is probable that very little subcooling of liquid is possible unless the receiver is almost completely filled with liquid. Vapor in the receiver in contact with the subcooled liquid will condense, and this effect will tend toward a saturated condition.

At normal condensing temperatures, the following relation between each 1°F. of subcooling and the cor-responding change in saturation pressure applies.

EquivalentChange inSaturationRefrigerant Subcooling Pressure

R-12 1° F. 1.75 psi

R-22 1° F. 2.75 psi

R-502 1° F. 2.85 psi

To illustrate, 5°F. subcooling will allow a pressure drop of 8.75 psi with R-12, 13.75 psi with R-22, and 14.25 psi with R-502 without flashing in the liquid line. For the previous example of a condensing unit in a basement requiring a vertical lift of 30 feet or approximately 15 psi, the necessary subcooling for the liquid head alone would be 8.5°F. with R-12, 5.5°F. with R-22, and 5.25°F. with R-502.

The necessary subcooling may be provided by the condenser used, but for systems with abnormally high vertical risers, a suction to liquid heat exchanger may be required. Where long refrigerant lines are involved, and the temperature of the suction gas at the condensing unit is approaching room temperatures, a heat exchanger located near the condenser may not have sufficient temperature differential to adequately cool the liquid, and individual heat exchangers at each evaporator may be necessary.

In extreme cases, where a great deal of subcooling is required, there are several alternatives. A special heat exchanger with a separate subcooling expansion valve can provide maximum cooling with no penalty on system performance. It is also possible to reduce the capacity of the condenser so that a higher operating condens-ing temperature will make greater subcooling possible. Liquid refrigerant pumps may also be used to overcome large pressure drops.

Liquid line pressure drop causes no direct penalty in power consumption, and the decrease in system capac-ity due to friction losses in the liquid line is negligible. Because of this the only real restriction on the amount of liquid line pressure drop is the amount of subcooling

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available. Most references on pipe sizing recommend a conservative approach with friction pressure drops in the 3 to 5 psi range, but where adequate subcooling is available, many applications have successfully used much higher design pressure drops. The total friction includes line losses through such accessories as sole-noid valves, filter-driers, and hand valves.

In order to minimize the refrigerant charge, liquid lines should be kept as small as practical, and exces-sively low pressure drops should be avoided. On most systems, a reasonable design criteria is to size liquid lines on the basis of a pressure drop equivalent to 2°F. subcooling.

A limitation on liquid line velocity is possible damage to the piping from pressure surges or liquid hammer caused by the rapid closing of liquid line solenoid valves, and velocities above 300 FPM should be avoided when they are used. If liquid line solenoids are not used, then higher velocities can be employed. Figure 83 gives liquid line velocities corresponding to various pressure drops and line sizes.

To summarize, in sizing liquid lines, it is recommended that the selection of line size be made on the basis of a total friction pressure drop equivalent to 2°F. subcool-ing. If vertical lifts or valves with large pressure drops are involved, then the designer must make certain that sufficient subcooling is available to allow the necessary pressure drop without approaching a saturation condi-tion at which gas flashing could occur. Check Figure 83 to be sure velocities do not exceed 300 FPM if a liquid line solenoid is used.

Recommended liquid line sizes for varying capacities and equivalent lengths of line are given in Table 27.

SIZINGSUCTIONLINES

Suction line sizing is the most critical from a design and system standpoint. Any pressure drop occurring due to frictional resistance to flow results in a decrease in the pressure at the compressor suction valve, compared with the pressure at the evaporator outlet. As the suc-tion pressure is decreased, each pound of refrigerant returning to the compressor occupies a greater volume, and the weight of the weight of the refrigerant pumped by the compressor decreases. For example, a typical low temperature R-502 compressor at -40°F. evaporating temperature will lose almost 6% of its rated capacity for each 1 psi suction line pressure drop.

Normally accepted design practice is to use as a design criteria a suction line pressure drop equivalent to a 2°F. change in saturation temperature. Equivalent pressure

drops for various operating conditions are shown in Table 25.

TABLE25PRESSUREDROPEQUIVALENTFOR2°F.

CHANGEINSATURATIONTEMPERATUREATVARIOUSEVAPORATINGTEMPERATURES

Evaporating PressureDrop,PSITemperature R-12 R-022 R-502 45ºF 2.0 3.0 3.3 20ºF 1.35 2.2 2.4 0ºF 1.0 1.65 1.85 -20ºF .75 1.15 1.35 -40ºF .5 .8 1.0

Of equal importance in sizing suction lines is the ne-cessity of maintaining adequate velocities to properly return oil to the compressor. Studies have shown that oil is most viscous in a system after the suction vapor has warmed up a few degrees from the evaporating temperature, so that the oil is no longer saturated with refrigerant, and this condition occurs in the suction line after the refrigerant vapor has left the evaporator. Move-ment of oil through suction lines is dependent on both the mass and velocity of the suction vapor. As the mass or density decreases, higher velocities are required to force the oil along.

Nominal minimum velocities of 700 FPM in horizontal suction lines and 1500 FPM in vertical suction lines have been recommended and used successfully for many years as suction line sizing design standards. Use of the one nominal velocity provided a simple and convenient means of checking velocities. However, tests have shown that in vertical risers the oil tends to crawl up the inner surface of the tubing, and the larger the tubing, the greater velocity required in the center of the tubing to maintain tube surface velocities which will carry the oil. The exact velocity required in vertical lines is dependent on both the evaporating temperature and the line size, and under varying conditions, the specific velocity required might be either greater or less than 1500 FPM.

For better accuracy in line sizing, revised maximum recommended vertical suction line sizes based on the minimum gas velocities shown in the 1980 ASHRAE Handbook have been calculated and are plotted in chart form for easy usage in Figures 84 and 86. These revised recommendations superseded previous vertical suction riser recommendations. No change has been made in the 700 FPM minimum velocity recommenda-tion for horizontal suction lines, and Figures 85 and 87 cover maximum recommended horizontal line sizes for proper oil return.

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To illustrate, again assume a system operating with R-12 at 40°F. evaporating temperature has a capacity of 100,000 BTU/hr. On Figure 84, the intersection of the evaporating temperature and capacity lines indicate that a 2 1/8” O.D. line will be required for oil return in the vertical suction risers.

Even though the system might have a much larger design capacity, the suction line sizing must be based on the minimum capacity anticipated in operation under light load conditions after allowing for the maximum reduction in capacity from capacity control if provided.

Since the dual goals of low pressure drop and high velocities are in direct conflict, obviously compromises must be made in both areas. As a general approach, in suction line design, velocities should be kept as high as possible by sizing lines on the basis of the maximum pressure drop that can be tolerated, but in no case should gas velocity be allowed to fall below the minimum levels necessary to return oil. It is recommended that a tenta-tive selection of suction line sizes be made on the basis of a total pressure drop equivalent to a 2°F. change in the saturated evaporating temperature. Check Figures 84 or 86 to be sure that velocities in vertical risers are satisfactory. Where refrigerant lines are lengthy, it may be desirable to use as large tubing as practical to minimize pressure drop, and Figure 85 or 87 should be checked to determine the maximum permissible horizontal line size. The final consideration must always be to maintain velocities adequate to return oil to the compressor, even if this results in a higher pressure drop than is normally desirable.

Recommended suction line sizes for varying capaci-ties and equivalent lengths of line are given in Tables 29 to 41.

DOUBLERISERS

On systems equipped with capacity control compressors, or where tandem or multiple compressors are used with one or more compressors cycled off for capacity control, single suction line risers may result in either unaccept-ably high or low gas velocities. A line properly sized for light load conditions may have too high a pressure drop at maximum load, and if the line is sized on the basis of full load conditions, then velocities may not be adequate at light load conditions to move oil through the tubing. On air conditioning applications where somewhat higher pressure drops at maximum load conditions can be tolerated without any major penalty in overall sys-tem performance, it is usually preferable to accept the

additional pressure drop imposed by a single vertical riser. But on medium or low temperature applications where pressure drop is more critical and where separate risers from individual evaporators are not desirable or possible, a double riser may be necessary to avoid an excessive loss of capacity.

A typical double riser configuration is shown in Figure 88. The two lines should be sized so that the total cross-sectional area is equivalent to the cross-section area of a single riser that would have both satisfactory gas velocity and acceptable pressure drop at maximum load conditions. The two lines normally are different in size, with the larger line trapped as shown, and the smaller line must be sized to provide adequate velocities and acceptable pressure drop when the entire minimum load is carried in the smaller riser.

In operation, at maximum load conditions gas and en-trained oil will be flowing through both risers. At minimum load conditions, the gas velocity will not be high enough to carry oil up both risers. The entrained oil will drop out of the refrigerant gas flow, and accumulate in the “P” trap, forming a liquid seal. This will force all of the flow up the smaller riser, thereby raising the velocity and assuring oil circulation through the system.

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For example, assume a low temperature system as follows:

Maximum capacity 150,000 BTU/hr. Minimum capacity 50,000 BTU/hr. Refrigerant R-502 Evaporating Temperature -40°F. Equivalent length of piping, horizontal 125 ft. Vertical Riser 25 ft. Desired design pressure drop (equivalent to 2°F.) 1 psi

A preliminary check of the R-502 pressure drop chart, Figure 78, indicates for a 150 foot run with 150,000 BTU/hr capacity and a total pressure drop of approximately 1 psi, a 3 1/8” O.D. line is indicated. At the minimum capacity of 50,000 BTU/hr, Figure 87 shows a 3 5/8” O.D. horizontal suction line is acceptable, but Figure 84 indicates that the maximum vertical riser size is 2 1/8” O.D. Referring again to the pressure drop chart, Figure 78, the pressure drop for 150,000 BTU/hr through 2 1/8” O.D. tubing is 4 psi per 100 feet, or 1.0 psi for the 25 foot suction riser. Obviously, either a compromise must be made in accepting a greater pressure drop at maximum load conditions, or a double riser must be used.

If the pressure drop must be held to a minimum, then the size of the double riser must be determined. At maximum load conditions, a 3 1/8” O.D. riser would maintain adequate velocities, so a combination of the sizes approximating the 3 1/8” O.D. line can be selected for the double riser. The cross sectional area of the line sizes to be considered are:

3 1/8” O.D. 6.64 sq. in.

2 5/8” O.D. 4.77 sq. in.

2 1/8” O.D. 3.10 sq. in.

1 5/8” O.D. 1.78 sq. in.

At the minimum load condition of 50,000 BTU/hr., the 1 5/8” O.D. line will have a pressure drop of approxi-mately .5 psi, and will have acceptable velocities, so a combination of 2 5/8” O.D. and 1 5/8” O.D. tubing should be used for the double riser.

In a similar fashion, double risers can be calculated for any set of maximum and minimum capacities where single risers may not be satisfactory.

SUCTIONPIPINGFORMULTIPLEXSYSTEMS

It is common practice in supermarket applications to operate several fixtures, each with liquid line solenoid valve and expansion valve control, from a single com-pressor. Temperature control of individual fixtures is normally achieved by means of a thermostat opening and closing the liquid line solenoid valve as necessary. This type of system, commonly called multiplexing, requires careful attention to design to avoid oil return problems and compressor overheating.

Since the fixtures fed by each liquid line solenoid valve may be controlled individually, and since the load on each fixture is relatively constant during operation, in-dividual suction lines and risers are normally run from each fixture or group of fixtures controlled by a liquid line solenoid valve for minimum pressure drop and maximum efficiency in oil return. This provides excellent control so long as the compressor is operating at its design suction pressure, but there may be periods of light load when most or all of the liquid line solenoids are closed. Un-less some means of controlling compressor capacity is provided, this can result in compressor short cycling or operation at excessively low suction pressures, which can result not only in overheating the compressor, but in reducing the suction pressure to a level where the gas becomes so rarefied it can no longer return oil properly in lines sized for much greater gas density.

Because of the fluctuations in refrigeration load caused by closing of the individual liquid line solenoid valves, some means of compressor capacity control must be provided. In addition, the means of capacity control must be such that it will not allow extreme variations in the compressor suction pressure.

Where multiple compressors are used, cycling of indi-vidual compressors provides satisfactory control. Where multiplexing is done with a single compressor, a hot gas bypass system has proven to be the most satisfac-tory means of capacity reduction, since this allows the compressor to operate continuously at a reasonably constant suction pressure while compressor cooling can be safely controlled by means of a desuperheating expansion valve.

In all cases, the operation of the system under all pos-sible combinations of heavy load, light load, defrost, and compressor capacity must be studied carefully to be certain that operating conditions will be satisfactory.

Close attention must be paid to piping design on mul-tiplex systems to avoid oil return problems. Lines must be properly sized so that the minimum velocities neces-sary to return oil are maintained in both horizontal and

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vertical suction lines under minimum load conditions. Bear in mind that although a hot gas bypass maintains the suction pressure at a proper level, the refrigerant vapor being bypassed is not available in the system to aid in returning oil.

PIPINGDESIGNFORHORIZONTALANDVERTICALLINES

Horizontal suction and discharge lines should be pitched downward in the direction of flow to aid in oil drainage, with a downward pitch of at least ½ inch in 10 feet. Refrigerant lines should always be as short and should run as directly as possible.

Piping should be located so that access to system com-ponents is not hindered, and so that any components which could possibly require future maintenance are easily accessible. If piping must be run through boiler rooms or other areas where they will be exposed to abnormally high temperatures, it may be necessary to insulate both the suction and liquid lines to prevent excessive heat transfer into the lines.

Every vertical suction riser greater than 3 to 4 feet in height should have a “P” trap at the base to facilitate oil return up the riser as shown in Figure 89. To avoid the accumulation of large quantities of oil, the trap should be of minimum depth and the horizontal section should be as short as possible. Prefabricated wrought copper traps are available, or a trap can be made by using two street ells and one regular ell. Traps at the foot of hot gas risers are normally not required because of the easier movement of oil at higher temperatures. However, it is recommended that the discharge line from the compressor be looped to the floor prior to be-ing run vertically upwards to prevent the drainage of oil back to the compressor head during shut down periods. See Figure 90.

For long vertical risers in both suction and discharge lines, additional traps are recommended for each full length of pipe (approximately 20 feet) to insure proper oil movement.

In general, trapped sections of the suction line should be avoided except where necessary for oil return. Oil or liquid refrigerant accumulating in the suction line during the off cycle can return to the compressor at high velocity as liquid slugs on start up, and can break compressor valves or cause other damage.

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SUCTIONLINEPIPINGDESIGNATTHEEVAPORATOR If a pumpdown control system is not used, each evapo-rator must be trapped to prevent liquid refrigerant from draining back to the compressor by gravity during the off cycle. Where multiple evaporators are connected to a common suction line, the connections to the common suction line must be made with inverted traps to prevent drainage from one evaporator from affecting the expan-sion valve bulb control of another evaporator.

Where a suction riser is taken directly upward from an evaporator, a short horizontal section of tubing and a trap should be provided ahead of the riser so that a suitable mounting for the thermal expansion valve bulb is available. The trap serves as a drain area, and helps to prevent the accumulation of liquid under the bulb which could cause erratic expansion valve operation. If the suction line leaving the evaporator is free draining or if a reasonable length of horizontal piping precedes the vertical riser, no trap is required unless necessary for oil return.

Typical evaporator connections are illustrated in Figure 91.

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RECEIVERLOCATION

Gas binding at the receiver can occur when the receiver is exposed to an ambient temperature higher than the condensing temperature. Heat transfer through the re-ceiver shell causes some of the liquid in the receiver to evaporate, creating a pressure in the receiver high than in the condenser. This forces liquid refrigerant to back up into the condenser until its efficiency is reduced to the point where the condensing pressure again exceeds the pressure in the receiver.

The best remedy for this problem is to make sure the receiver is always exposed to ambient temperatures lower than the condensing temperature. If this is not possible, the receiver should be insulated to minimize heat transfer. Various types of venting arrangements for receivers have been proposed, but these require extreme care in circuiting to avoid flow problems. When the receiver is vented back to the condenser, the only force causing flow from the condenser to the receiver is gravity. Vented piping arrangements are complicated at best, and should be avoided if possible.

Even though there may not be sufficient heat transfer into a receiver to cause gas binding, on systems where the condensing temperature is lower than the ambient (for example water cooled or evaporative condensers with remote receivers) the liquid refrigerant may be warmed sufficiently in the receiver to lose most and possibly all of its subcooling. As mentioned previously, special subcooling coils or insulation may be required for proper operation under these conditions.

If a difference in temperature exists between two parts of an idle refrigeration system with interconnecting piping, this actually creates a little built-in static refrigeration system. The liquid refrigerant at the high temperature point will slowly vaporize, travel through the system as vapor, and recondense at the lowest temperature point. This most often is a matter of concern with a roof mounted remote condenser when the compressor is located in an inside machine room. If the system is idle, the sun is shining on the condenser, and the machine room is cool, then liquid is going to move out of the condenser and back down the discharge line to the machine room. Occasionally inverted traps are made in the discharge line at the condenser in the belief they will prevent this type of reverse flow. Actually with even a few degrees temperature difference, an inverted loop 20 feet high would be of no value.

However, if the receiver is located either in the machine room, or at some other point where it will not be exposed to the roof heat, reverse flow from the condenser seldom is a source of operating difficulty. The amount of refrig-

erant actually returning down the discharge line will be minimized and rarely if ever will this cause compressor damage if good piping practice is followed. It is possible to mount a check valve in the discharge line near the condenser as a means of preventing refrigerant back-flow of this nature, but check valves in this location are noisy, expensive, and subject to damage, and should be employed only if absolutely essential.

VIBRATIONANDNOISE

No matter how well the compressor is isolated, some noise and vibration will be transmitted through the pip-ing, but both can be minimized by proper design and support of the piping.

On small units a coil of tubing at the compressor may provide adequate protection against vibration. On larger units, flexible metallic hose is frequently used. When the compressor is supported by vibration absorbing mounts allowing compressor movement, refrigerant lines should not be anchored solidly at the unit, but at a point beyond the vibration absorber, so the vibration can be isolated and not transmitted into the piping system.

The noise characteristics of a large refrigeration or air conditioning system, particularly when installed with long refrigerant lines and remote condensers, are not predict-able. Variations in piping configuration, the pattern of gas flow, line sizes, operating pressures, the compressor and unit mounting, all can affect the noise generated by the system. Occasionally a particular combination of gas flow and piping will result in a resonant frequency which may amplify the sound and vibration to an undesirable level. Gas pulsation from the compressor may also be amplified in a similar manner.

If gas pulsation or resonant frequencies are encountered on a particular application, a discharge line muffler may be helpful in correcting the problem. The purpose of a muffler is to dampen the pulses of gas in the discharge line and to change the frequency to a level which is not objectionable. A muffler normally depends on mul-tiple internal baffles and/or pressure drop to obtain an even flow of gas. In general, the application range of a muffler depends on the mass flow of gas through the muffler, so the volume and density of the refrigerant gas discharged from the compressor are both factors in muffler performance.

A given muffler may work satisfactorily on a fairly wide range of compressor sizes, but is also quite possible that a given system may require a muffler with a particular pressure drop to effectively dampen pulsations. On problem applications, trial and error may be the only final guide. While larger mufflers are often more efficient in reducing the overall level of compressor discharge noise,

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in order to satisfactorily dampen pulsations, smaller mufflers with a greater pressure drop are usually more effective. Adjustable mufflers are often helpful since they allow tuning of the muffler pressure characteristics to the exact system requirement.

Occasionally, a combination of operating conditions, mounting and piping arrangement may result in a reso-nant condition, which tends to magnify compressor pul-sation and cause a sharp vibration, although noise may not be a problem. For larger Copelametic® compressors, discharge muffler plates have been developed for use when necessary to dampen excessive pulsation. The muffler plate fits between the discharge valve and the compressor body and has a number of muffling holes to provide the proper characteristics for the particular compressor displacement. The muffling holes break up the pattern of gas flow and create sufficient restriction to reduce the gas pulsation to a minimum.

When piping passes through walls or floors, precautions should be taken to see that the piping does not touch any structural members and is properly supported by hangers in order to prevent the transmission of vibra-tion into the building. Failure to do so may result in the building structure becoming a sounding board.

Table 26 gives the maximum recommended spacing for pipe supports.

RECOMMENDEDLINESIZINGTABLES

Tables 27 to 41 give recommended line sizes for single stage applications at various capacities and for equivalent

lengths of pipe based on the design criteria discussed previously. (For piping recommendations on two stage systems, refer to Section 19).

Vertical suction line sizes have been selected on the basis of a total vertical rise up to 30 feet. For longer risers, individual calculations should be made since the increased pressure drop may require different line sizes and possibly the use of double risers in place of the single riser shown.

Discharge line sizes have been calculated on the basis of a nominal pressure drop of 5 psi. Vertical line sizes have been selected so that minimum velocities neces-sary to carry oil up the riser will be maintained under the reduced load conditions shown, and velocities have been checked to see that they do not exceed 2,700 FPM at maximum load conditions. Because of the relatively small variation in discharge line velocity over the normal refrigeration and air conditioning range, the line sizes shown may be safely used for evaporating temperatures from -40°F. to 45°F., and condensing temperatures from 80°F. to 130°F.

Liquid line sizes have been calculated on the basis of a nominal pressure drop equivalent to 2 ½° F. subcooling, and velocities have been checked to see that they do not exceed 250 FPM. Liquid lines from the condenser to receiver have been selected on the basis of 100 FPM velocity in accordance with standard industry practice in order to allow a free draining line with gas equaliza-tion where piping allows. As in the case with discharge lines, the relatively small variation in liquid line velocities over the normal refrigeration and air conditioning range allows use of the recommended line sizing for evaporat-ing temperatures from -40°F. to 45°F., and condensing temperatures from 80°F. to 130°F.

Suction line sizes have been calculated on the basis of a nominal pressure drop equivalent to a 2°F. change in the saturated evaporating temperature. Both horizontal and vertical line sizes have been checked to see that the necessary minimum velocities are maintained under the reduced load conditions shown. Line sizes have been calculated for various evaporating temperatures, and may be safely applied for condensing temperatures from 80°F. to 130°F.

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SECTION19LOWTEMPERATURESYSTEMS

SINGLESTAGELOWTEMPERATURESYSTEMS

Low temperature single stage systems become increas-ingly critical from a design and application standpoint as the desired evaporating temperature is decreased. The combination of high compression ratios, low oper-ating temperatures, and rarefied return gas can cause lubrication and overheating problems, and make the compressor more vulnerable to damage from moisture and contaminants in the system.

The compressor selection, suction temperature, and application must be such that the temperature of the discharge line measured within 1” to 6” of the discharge service valve does not exceed 230°F. for Refrigerants 12, 22, and 502. Under these conditions, the estimated average temperature at the discharge port (measured at the valve retainer on the valve plate) will be approximately 310°F. for R-12 and R-502, and 320°F. for R-22.

The compressor displacement, pressure limiting devices, and quantity of cooling air or water must be selected to prevent the motor temperature from exceeding the limits stated below:

A. 210°F. when protected by inherent protectors affected by line current and motor temperature.

B. 190°F. when protected by motor starters.

The temperature of the motor should be determined by the resistance method and should be determined when the compressor is tested in the highest ambient in which it is expected to operate, at 90 per cent of rated voltage, with 90°F. return suction gas temperature. For longer motor life, operating temperatures of 170°F. to 190°F. are highly recommended.

In order to prevent the discharge and motor tempera-tures from exceeding recommended limits, it is very desirable, and in some instances absolutely necessary, to insulate the suction lines and return the suction gas to the compressor at a lower than normal tempera-ture. This is particularly important with suction-cooled compressors when R-22 is used. (Approximately 30°F. superheat suggested.)

Suction cooled compressors require auxiliary cooling by means of an air blast on the compressor for operation below 0°F evaporator temperature.

Either the evaporator must be properly designed, or a pressure limiting device such as a pressure limiting expansion valve or crankcase pressure regulating valve must be provided to prevent motor overloading during pulldown periods, or after defrost.

Emerson Climate Technologies, Inc. now recommends R-502 for all single stage low temperature applications where evaporating temperatures of -20°F. and below may be encountered. Now that R-502 is readily available, R-22 should not be used in single stage low temperature compressors, 5 H.P. and larger. The lower discharge temperatures of R-502 have resulted in much more trouble-free operation.

An adequate supply of oil must be maintained in the crankcase at all times to insure continuous lubrication. If the refrigerant velocity in the system is so low that rapid return of the oil is not assured, an adequate oil separator must be used. The normal oil level should be maintained at or slightly above the center of the sight glass. An excessive amount of refrigerant or oil must not be allowed in the system as it may result in excessive liquid slugging and damage to the compressor valves, pistons, or cylinders.

The formation or make up of the lines must be so de-signed that oil trapping will not exist. The highest velocity possible without encountering excessive pressure drop is recommended.

Care must be taken to prevent the evaporating temperature from dropping so far below the normal system operating point that the refrigerant veloc-ity becomes too low to return oil to the compressor. The low pressure control cut-out setting should not be below the lowest published rating point for the compressor, without prior approval of the Emerson Climate Technologies, Inc. Application Engineering Department.

The smallest practical size tubing should be used in condensers and evaporators in order to hold the system charge to a minimum. When large refrigerant charges are unavoidable, recycling pumpdown control should be used.

If air cooled condensing units are required to operate in low ambient temperatures, the use of some means of head pressure control to prevent the condensing pressure from falling too low is highly recommended to maintain normal refrigerant velocities. Several commonly used types of control are described in Section 17.

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An adequate filter-drier of generous size must be in-stalled in the liquid line, preferably in the cold zone. The desiccant used must be capable of removing moisture to a low end point and be capable of removing a rea-sonable quantity of acid. It is most important that the filter-drier be equipped with an excellent filter to prevent circulation of carbon and foreign particles. A permanent suction line filter is highly recommended to protect the compressor from contaminants which may be left in the system during installation.

A combination liquid sight glass and moisture indicator should be installed for easy field maintenance.

After complete assembly, all systems should be thor-oughly evacuated with a high grade vacuum pump and dehydrated to assure that no air or moisture remains in the system. The compressor motor must not be operated while the high vacuum pump is in operation, otherwise motor damage is very likely to occur.

The system should be charged with clean dry refrigerant only through a dehydrator. Other substances such as liquid dehydrants or alcohol must not be used.

TWOSTAGELOWTEMPERATURESYSTEMS

Two stage systems because of their basic design and operation are inherently more efficient and encounter fewer operating hazards at low operating temperatures than single stage equipment. The two stage compressor has its limitations. At evaporating temperatures below -80°F. it loses efficiency and motor heating becomes an increasing problem. The lowest approved operating range is -80°F. and at lower evaporating temperatures a cascade system is recommended. But for applications with evaporating temperatures in the -20°F. to -80°F. range, the two stage compressor efficiency is high, the discharge temperatures are low, and field experience with properly applied two stage compressors has been excellent.

The two stage system is somewhat more complex and sophisticated than a simple single stage system, and many of the operating problems encountered on two stage systems stem from the fact that too often they have been applied without sufficient appreciation of the safeguards which must be taken in system design.

VOLUMETRICEFFICIENCY

Three definitions given previously are of importance in analyzing two stage systems.

The compression ratio is the ratio of the absolute dis-

charge pressure (psia) to the absolute suction pressure (psia).

The absolute pressure is gauge pressure plus atmo-spheric pressure, which at sea level is standardized at 14.7 pounds per square inch.

Volumetric efficiency is defined as the ratio of the actual volume of the refrigerant gas pumped by the compressor to the volume displaced by the compressor pistons.

Figure 92 illustrates a typical single stage volumetric efficiency curve. Note that as the compression ratio increase, the volumetric efficiency decreases.

Two factors cause a loss of efficiency with an increase in compression ratio. The density of the residual gas remaining in the cylinder clearance space after the compression stroke is determined by the discharge pres-sure—the greater the discharge pressure the greater the density. Since this gas does not leave the cylinder on the discharge stroke, it re-expands on the suction stroke, thus preventing the intake of a full cylinder of vapor from the suction line. As the compression ratio increases, the more space in the cylinder on the intake stroke is filled by the residual gas.

The second factor in the loss of efficiency is the high temperature of the cylinder walls resulting from the heat of compression. As the compression ratio increases, the heat of compression increases, and the cylinders and head of the compressor become very hot. Suction gas entering the cylinder on the intake stroke is heated by the cylinder walls and expands, resulting in a reduced weight of gas entering the compressor.

Obviously, a single stage compressor has its limita-tions as compression ratios increase. The effective low limit of even the most efficient single stage system is approximately -40°F. evaporating temperature. At lower evaporating temperatures, the compression ratio becomes so high that capacity falls rapidly, the com-pressor may no longer be handling a sufficient weight of return gas for proper motor cooling, and because of decreased gas density, oil may no longer be properly circulated through the system.

TWOSTAGECOMPRESSIONANDCOMPRESSOREFFICIENCY

In order to increase operating efficiency at low evapo-rating temperatures, the compression can be done in two steps or stages. For two stage operation, the total compression ratio is the product of the compression ratio of each stage. In other words, for a total compression

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ratio of 16 to 1, the compression ratio of each stage might be 4 to 1; or compression ratios of 4 to 1 and 5 to 1 in separate stages will result in a total compression ratio of 20 to 1.

Two stage compression may be accomplished with the use of two compressors with the discharge of one pumping into the suction of the second, but because of the difficulty of maintaining proper oil levels in the two crankcases, it is more satisfactory to use one compres-sor with multiple cylinders. On Copeland® brand two stage compressors, the ratio of low stage to high stage displacement is 2 to 1. The greater volume of the low stage cylinders is necessary because of the difference in specific volume of the refrigerant vapor at low and interstage pressures. While the compression ratios of the two stages are seldom exactly equal, they will be approximately the same. A typical 6 cylinder two stage compressor with its external manifold and desuperheat-ing expansion valve is shown in Figure 93, and a typical 3 cylinder two stage compressor with external manifold is shown in Figure 94.

Figure 95 shows a comparison of five different volu-metric efficiency curves. The three straight lines are typical single stage curves—one for an air conditioning compressor, one for a typical multi-purpose compres-sor, and one for a low temperature compressor. There are some variations in compressor design involved, but the primary difference in characteristics is due to clearance volume.

The two vertical curved lines represent the compara-tive efficiency of a two stage compressor. Actually each separate stage would have a straight line character-istic similar to the single stage curves, but to enable comparison with single stage compressors, the overall volumetric efficiency has been computed on the basis of the total displacement of the compressor, not just the low stage displacement.

The solid black curve represents the efficiency of a two stage compressor without a liquid subcooler. Note that the efficiency is relatively constant over a wide range of total compression ratios, and that the crossover in efficiency with the best low temperature single stage compressor is at a compression ratio of approximately 13 to 1. In other words, at compression ratios lower than 13 to 1, a single stage compressor will have more capacity than a two stage compressor of equal displace-ment without liquid subcooling.

The dotted curve represents the efficiency of the same two stage compressor with a liquid subcooler. In the subcooler, the liquid refrigerant being fed to the evapo-

rator is first subcooled by liquid refrigerant fed through the interstage desuperheating expansion valve, and a much greater share of the refrigeration load has been transferred to the high stage cylinders. Since the high stage cylinders operate at a much higher suction pressure, the refrigeration capacity there is far greater per cubic foot of displacement than in the low stage cylinders. In effect the capacity of the compressor has been greatly increased without having to handle any additional suction gas returning from the evaporator. Note that with the liquid subcooler, the crossover point in efficiency as compared with a single stage compres-

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sor is at a compression ratio of approximately 7 to 1. In other words, at compression ratios lower than 7 to 1, the single stage compressor will have more capac-ity for equal displacement, but at compression ratios higher than 7 to 1, the two stage compressor will have more capacity.

Table 42 lists comparative operating data at varying evaporating temperatures for a Copeland® brand com-pressor available either as a single stage or two stage compressor. Although the displacement, refrigerant, and motor are the same, the rapidly increasing advantage of two stage operation as the evaporating temperature decreases is plainly shown.

COMPRESSOR OVERHEATING AT EXCESSIVECOMPRESSIONRATIOS

In addition to efficiency, the extremely high temperatures created by operation at abnormally high compression ratios makes the use of single stage compressors im-

practical for ultra-low temperature applications. Figure 96 shows a valve plate with carbon formation due to oil breakdown from excessive heat. Excessive cylinder temperatures can also cause rapid piston and cylinder wear, cylinder scoring, and early failure of the compres-sor.With two stage compressors, the interstage expan-sion valve maintains safe operating temperatures and this type of damage is prevented.

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BASICTWOSTAGESYSTEM

The basic flow of refrigerant in a six cylinder two stage compressor is shown schematically in Figure 97. The suction gas returning from the evaporator enters the four low stage cylinders directly from the suction line. Since the discharge gas from the first stage cylinders is heated from compression, it must be cooled by the desuperheating expansion valve before entering the motor chamber. The desuperheated refrigerant vapor, now at interstage pressure, enters the high stage cyl-inders, is compressed, and is then discharged to the condenser.

Figure 98 is a schematic view of a typical two stage system showing the various components necessary for proper operation.

TWOSTAGESYSTEMCOMPONENTS

a.LiquidLineSolenoidValve

To prevent leakage during the off period, a solenoid valve must be placed in the liquid supply line immediately

ahead of the desuperheating expansion valve. It should be wired so as to be open when the motor is running and closed when not running. A toggle switch placed in the electric line to the solenoid valve will facilitate service during pumpdown.

A 100 mesh strainer must be installed in the liquid line feeding the desuperheating valve, up stream from the solenoid valve, to protect both valves from contami-nants.

b.OilSeparator

At ultra-low temperatures, the decrease in density of the refrigerant suction vapor and the increasing viscosity of the refrigeration oil make oil return during extended pe-riods of operation or during light load periods extremely difficult. In order to minimize oil circulation and safely bridge the operating periods between defrost periods when oil will be returned, oil separators are standard on all Copeland® brand two stage condensing units, and are strongly recommended on all two stage compressors.

The oil separator will provide some increase in refrigerat-

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ing capacity due to the increased heat transfer capability of the evaporator surface resulting from the reduced oil in circulation. It will also act as a muffler to reduce discharge pulsation and system noise transmission. c.SuctionLineAccumulator

Since the suction gas is returned directly to the low stage cylinders without going through the motor cham-ber, the two stage compressor is vulnerable to damage if excessive liquid floods back from the evaporator. To prevent damage from slugging, adequate suction line accumulators are mandatory on any system prone to return slugs of liquid or oil to the compressor. This may be especially critical on systems with hot gas defrost.

d.SuctionLineFilter

A good suction line filter is recommended on any field installed system to prevent damage from copper filings, solder, flux, bits of steel wool, or other contamination

left in the system. The nominal cost of the filter is good insurance for the most vulnerable part of the system.

e.LiquidSightGlass

A liquid line sight glass should be installed in the liq-uid line just ahead of the desuperheating expansion valve to provide a positive check for shortage of liquid refrigerant.

When a liquid subcooler is used, the regular liquid line sight glass (additional to the one ahead of the desuper-heating expansion valve) should be installed between the receiver and the subcooler. If installed beyond the subcooler it will not be dependable since it may not show bubbles even when the system is short of refrigerant.

f.CrankcasePressureRegulatingValve

Two stage compressors will overload if allowed to oper-ate for extended periods with high suction pressures. The suction pressure on some systems can be limited

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to a satisfactory point by the size and type of evapora-tor used, or by the use of pressure limiting expansion valves. However, the load after defrost is often the most critical, and pressure limiting expansion valves will not protect against an overload at this time. If the system can overload at any time during its operating cycle, a crankcase pressure regulating valve must be used.

g.DesuperheatingExpansionValve

Expansion valves currently supplied as original equip-ment with Copeland® brand two stage compressors are of the non-adjustable superheat type. In the event of valve failure, standard field replacement valves with adjustable superheat which have been approved by Emerson Climate Technologies, Inc. may be used. Improper valve selection can result in compressor overheating and possible damage due to improper liquid refrigerant control.

It is recommended that the branch liquid line to the desuperheating expansion valve be taken from the bottom of the main liquid line. Never install a tee with a branch off the top of the main liquid line, since this can result in improper refrigerant feed, and possible motor overheating.

h.DefrostCycle

With electric defrost, the compressor is not running dur-ing the defrost cycle, so no special precautions other than those normally required with single stage systems are necessary.

However, motor cooling on two stage compressors is dependent on an adequate feed of liquid refrigerant from the desuperheating expansion valve. If a hot gas defrost system is used, it is imperative that a solid head of liquid is maintained at the desuperheating expansion valve at all times, and that the compressor be adequately

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protected against liquid refrigerant returning from the evaporator after condensation during the defrost cycle. Since hot gas defrost systems vary widely in design, it is not possible to make a general statement as to what special controls may be required. Most manufacturers have thoroughly pretested their systems, but on field installations, restrictor valves to maintain head pressure, additional refrigerant charge, suction accumulators, or other special controls may be necessary. In general, an electric defrost system is much less complicated and therefore usually more dependable on field installed two stage systems.

i.CondenserCapacity

When two stage compressors were first introduced into commercial usage for supermarkets, some users assumed that a condenser designed for a 10 HP single stage compressor would also be suitable for a 10 HP two stage unit. They failed to take into account the increased efficiency of the compressor, and apparently did not check the relative compressor capacities when selecting condensers. At -40°F. evaporating temperature, a two stage compressor may have almost twice the capac-ity of a low temperature single state compressor. As a result some field problems were encountered on early two stage units because of lack of condensing capacity. This is really not a basic problem, but it is a pitfall to be aware of when working with those who are accustomed to thinking of condensing units and condensers in terms of horsepower.

j.LiquidRefrigerantSubcooler

Two stage systems may be operated either with or with-out liquid subcoolers. The function of the subcooler is to cool the liquid refrigerant being fed to the evaporator by the evaporation of refrigerant fed through the desu-perheating expansion valve. This transfers a greater portion of the refrigeration load to the high stage cylin-ders, and because of the greater compressor capacity at higher suction pressures, the system capacity is greatly increased.

The temperature of the liquid refrigerant being fed to the evaporator is reduced in the subcooler to within ap-proximately 10°F. of the interstage saturated evaporating temperature, and the increase in system capacity can only be realized if the subcooled liquid is maintained at this low temperature, and heat transfer into the liquid line is prevented. Normally this requires insulation of the liquid line.

When selecting expansion valves for two stage systems with liquid subcoolers, the designer must bear in mind

that the expansion valve will have greatly increased capacity due to the low temperature of the refrigerant entering the valve. Unless this is taken into consider-ation, the increased refrigerating effect per pound of refrigerant may result in an oversized expansion valve with resulting erratic operation.

PIPINGONTWOSTAGESYSTEMS

The exact oil return characteristics of any system are difficult to forecast, and there is very little published data available on the design of refrigeration piping for ultra-low temperature systems. It is quite probable that any ultra-low temperature system will trap some oil dur-ing operation, even with the most conservative piping design, and the evaporator design may have a major influence on oil circulation.

In order to minimize oil circulation and prolong the operat-ing periods between defrost periods or other heavy load conditions which will return oil normally, oil separators are almost invariably required on two stage systems. Even with oil separators (which are never 100% efficient), it may be necessary to increase the number and frequency of defrost periods if oil is lost in the system.

Users frequently fail to realize that two stage systems with and without liquid subcooling present differences in piping design requirements. Two stage systems without liquid subcoolers are similar to single stage systems in that the liquid refrigerant temperature approaches the condensing temperature. On two stage systems with liquid subcoolers the refrigeration effect per pound of refrigerant is greatly increased because of the cold liquid refrigerant entering the evaporator. Therefore the pounds of refrigerant circulated for a given capacity will be greatly reduced, and line velocities correspond-ingly less. Most standard line sizing and pressure drop tables are based on liquid refrigerant temperatures equivalent to normal air and water cooled condensing temperatures, and do not apply to two stage systems with liquid subcoolers. Tables 43 through 48 give recommended line sizes for two stage systems. Line sizing has been calculated where possible on the basis of normal single stage pres-sure drop criteria, but line sizes have been selected to maintain the same mass velocity flow as that which has been found to be acceptable in the normal commercial range. It is essential that the piping designer realize that no piping design on ultra-low stage systems can guar-antee proper oil return, and that the requirement for an oil separator and possibly frequent defrost periods still remains, depending on the system characteristics.

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CASCADEREFRIGERATIONSYSTEMS

Multiple stage refrigeration can also be accomplished by using separate systems with the evaporator of the high stage serving as the condenser of the low stage by means of a heat exchanger. This type of system, termed a cascade system, is extremely flexible, and is well adapted to extremely low temperature systems, or to any system where the total compression ratio is very large. Since different refrigerants can be used in the separate systems, refrigerants with characteristics suitable for the specific application can be used. Cas-cade systems in multiples of two, three, or even more separate stages make possible refrigeration at almost any desired evaporating or condensing temperature.

Cascade systems have many hazards and potential problems not normally encountered in single stage re-frigeration, and successful system design and application require specialized knowledge and experience.

Figure 99 is a schematic diagram of a typical cascade system consisting of two stages. The cascade condenser is basically a direct expansion heat exchanger, acting as the evaporator of the high stage and the condenser of the low stage.

Various refrigerants can be and are used in cascade systems, with R-12, R-22, or R-502 frequently used in the high stage. The absolute pressures necessary to obtain evaporating temperatures below -80°F. with R-12, R-22, and R-502 are so low that the specific volume of the refrigerant becomes very high, and the resulting compressor displacement requirement is so great that the use of these refrigerants in the low stage becomes uneconomical. R-13, ethane, and a new refrigerant, R-23/13 (R-503) are frequently used for low stage applications.

R-13 is commonly used for evaporating temperatures in the -100°F. to -120°F. range since its pressure at those evaporating temperatures is such that its use is practi-cal with commonly available refrigeration compressors. However the critical temperature of R-13 is 84°F. and the critical pressure is 561 psia. This means it cannot be liquefied at temperatures above 84°F regardless of pressure, and the equilibrium pressure of a mixture of gas and liquid at 84°F. is 561 psia. In order to prevent excessive pressures in the system during non-operat-

ing periods, an expansion tank as shown in Figure 99 must be provided so that the entire refrigerant charge can exist as a vapor during off periods without exposing the compressor crankcase or the piping to excessive pressures. (Normally non-operating system pressures should be held to 150 psig or below).

Normally the expansion tank is located in the low pres-sure side of the system, with a relief valve from the high pressure side of the system discharging into the tank. The sizing of the tank is determined from the total refrigerant charge, the internal volume of the system, the maximum pressure desired, and the design ambient temperature. The specific volume of the vapor at the design storage conditions can be determined from the pressure enthalpy diagram of the refrigerant, such as shown in Figure 100. For example, at a temperature of 120°F. and a pressure of 140 psia, the specific volume of R-13 is .40 cubic feet per pound. If the system charge is 10 pounds of refrigerant, then the internal volume of the system including the expansion tank must be at least 4 cubic feet.

Because of the pressure relation of the system charge to the internal volume, the refrigerant charge in cascade systems is usually critical. Charging by means of a sight glass is unsafe. Either the exact charge must be mea-sured into the system, or the low stage may be charged with vapor to a stabilized non-operating pressure of 150 to 175 psi in maximum ambient conditions.

The pull down load may be many times the load at de-sign operating conditions, and some means of limiting the compressor loading during the pull down period is normally required, since it is seldom economical to size either the compressor motor or the condenser for the maximum load. Pressure limiting expansion valves or crankcase pressure regulating valves are acceptable if they are sized properly. Frequently a control system is designed to lock out the low stage system until the high stage evaporating temperature is reduced to the operating level so that excessive low stage condensing pressures do not occur on start up.

Various means of capacity control are employed, usu-ally by means of hot gas bypass. Care must be taken to insure proper compressor motor cooling and to avoid liquid floodback to the compressor.

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The cascade condensing temperature varies with individual system design. With normal high stage con-densing temperatures, either air cooled or water cooled, and evaporating temperatures in the -90°F. to -140°F. range, high state evaporating temperatures from 30°F. to -30°F. are commonly used. A difference of 10°F. to 20°F. between the high stage evaporating temperature and the low stage condensing temperature results in a reasonably sized cascade condenser. The compression ratios of the high stage and low stage should be approxi-mately equal for maximum efficiency, but small variations will not materially affect system performance.

Copeland® brand compressors are not tested with the refrigerants normally used in the low stage of cascade systems. Although many models of Copeland® brand compressors have been successfully applied for many years on cascade systems, the responsibility for the selection and application of the compressor must be that of the system designer, based on his testing, experience, and design approach.

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SECTION20TRANSPORTREFRIGERATION

Truck and trailer refrigeration is an increasingly impor-tant segment of the refrigeration industry. Despite the fact that transport applications face many operating problems peculiar to their usage, there exists very little application data pertaining to this field.

Many compressor failures in transport refrigeration usage are the result of system malfunction rather than the result of mechanical wear. It is clear that substantial savings in operating cost, and tremendous improvements in unit performance and life would be possible if the causes of compressor failure could be removed. Primarily the problem boils down to one of making sure that the com-pressor has adequate lubrication at all times.

Part of the problem of identifying the cause of failure stems from the fact that far too few users realize that ultimate failure of a compressor resulting from lack of lubrication frequently takes place at a time when there is an adequate supply of oil in the crankcase. This is due to continued deterioration of the moving parts resulting from the original or repeated damage in the past. It is not uncommon for a damaged compressor to operate satisfactorily all winter and then fail in the spring when subjected to heavier loads.

Another source of field problems is the fact that many units are installed by personnel who may not have ad-equate training, equipment, or experience. Often units, particularly those in common carrier service, may be serviced in emergencies by servicemen not familiar with the unit, or indeed, with transport refrigeration generally.

Because of the installation and service hazards, it is extremely important that the unit be properly designed and applied to minimize, and if possible, prevent service problems.

COMPRESSORCOOLING

Air-cooled motor-compressors must have a sufficient quantity of air passing over the compressor body for motor cooling. Refrigerant-cooled motor-compres-sors are cooled adequately by the refrigerant vapor at evaporating temperatures above 0°F. saturation, but at evaporating temperatures below 0°F. additional motor cooling by means of air flow is necessary.

Normally the condenser fan if located so that it discharges on the compressor will provide satisfactory cooling. For proper cooling, the fan must discharge air directly against the compressor. The compressor cannot be adequately

cooled by air pulled through a compartment in which the compressor is located. If the compressor is not located in the condenser discharge air stream, adequate air circulation must be provided by an auxiliary fan.

COMPRESSORSPEED

Open type compressors operating from a truck engine by means of a power take-off or by a belt drive are subject to extreme speed ranges. A typical truck engine may idle at 500 RPM to 700 RPM, run at 1,800 RPM at 30 MPH, and run at 3,600 RPM to 4,000 RPM over the highway at high speeds. Whatever the power take-off or belt ratio, this means the compressor must operate through a speed ratio range of 6 to 1 or greater unless it is disconnected from the power source by some means.

The compressor speed must be kept within safe limits to avoid loss of lubrication and physical damage. Op-eration within the physical limitations of the compressor may be possible, for example from 400 RPM to 2,400 RPM. It may be possible to use a cut-out switch to disconnect the compressor from the power source at a given speed. The compressor manufacturer should be contacted for minimum and maximum speeds of specific compressors.

If the compressor is of the accessible-hermetic type, there is no problem concerning speed so long as the electrical source is operating at the voltage and frequency for which the motor was designed. If the speed of the generator is varied in order to obtain variable speed operation, the voltage and frequency on the normal al-ternating current generator will vary proportionally. Since the compressor speed and motor load will vary directly with the frequency, it is often possible to operate over a wide speed range with satisfactory results.

However, it should be born in mind that increasing the frequency and voltage of the generator above the level for which the compressor motor was designed will in-crease the load on the compressor, may overload the motor, and can result in bearing or other compressor damage. Operation at speeds too low to provide ad-equate compressor lubrication must also be avoided, although normally lubrication can be maintained on Copelametic® compressors down to 600 RPM and possibly lower speeds.

Each new application involving operation of the com-pressor at a voltage and frequency differing from its nameplate rating should be submitted to the Emerson

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Climate Technologies, Inc. Application Engineering Department for approval.

One other problem that may arise with operation from a variable speed generator is the operation of electrical contactors, relays, etc. on voltages below or above their nameplate rating. Field test have shown that the winding design and physical construction of electrical compo-nents can cause wide variation in voltage tolerance. The drop-out voltage of various types of commercially available 220 volt contactors may vary from 145 volts to 180 volts depending on construction. If it is planned to operate at variable voltage and frequencies, the electrical components which are to be used should be extensively tested at the electrical extremes in cooperation with the manufacturer to insure proper operation.

COMPRESSOROPERATINGPOSITION

Occasionally compressor failures will occur due to loss of lubrication caused by parking the truck on too steep a slope. The resulting tilt of the compressor may cause the oil level to fall below the pick-up point of the oil flinger or oil pump.

Operation of the unit while the truck is parked on steep inclines should be avoided. If this is unavoidable, then consideration should be given to mounting the compres-sor so that oil will tend to flow to the oil pick-up point. Since this will vary on different model compressors, and the individual parking arrangement will affect the direc-tion of the compressor pitch, each application must be considered individually.

In severe cases, consult with the compressor manufacturer.

COMPRESSORDRIVE

Direct drive form an engine, either gasoline or diesel, to a compressor requires very careful attention to the cou-pling design. Alignment between the engine drive shaft and the compressor crankshaft is critical both in parallel and angular planes. Even slight angular misalignment can cause repetitive compressor crankshaft breakage. Because of the sharp impulses from the engine firing, a flexible coupling giving some resiliency is required. The coupling should be capable of compensating for slight parallel or angular misalignment and should also allow some slight endplay movement of the crankshafts. Nylon splines, neoprene bushings, and flexible disc type couplings have all been used successfully.

For a compressor driven from a power take-off by means of a shaft and two universal joints, the crosses in the

U-joints must be kept parallel to each other. Where possible, the compressor rotation should be in the same direction whether on electric standby or driven from the engine.

In driving a compressor with V-belts, care must be taken to avoid excessive belt tension and belt slap. A means for easily adjusting belt tension should be provided. It may be necessary to provide an idler pulley to dampen belt movement on long belt drives. Care should be taken to mount the compressor so that the compressor shaft is parallel with the engine crankshaft.

REFRIGERANTCHARGE

Refrigerant R-12 is used in most transport systems at the present time, but R-502 is well suited for low tem-perature applications, and its use is increasing. Since R-502 creates a greater power requirement for a given compressor displacement than R-12, the motor-com-pressor must be properly selected for the refrigerant to be used. Different expansion valves are required for each refrigerant, so the refrigerants are not interchange-able in a given system and should never be mixed. Receivers for R-502 require higher maximum working pressures than those used with R-12, so normally it is not feasible to attempt to convert an existing R-12 unit for the use of R-502.

The refrigerant charge should be held to the minimum required for satisfactory operation. An abnormally high refrigerant charge will create potential problems of liquid refrigerant migration, oil slugging, and loss of compres-sor lubrication due to bearing washout or excessive refrigerant foaming in the crankcase.

Systems should be charged with the minimum amount of refrigerant necessary to insure a liquid seal ahead of the expansion valve at normal operating temperatures. For an accurate indication of refrigerant charge, a sight glass is recommended at the expansion valve inlet, and a combination sight glass and moisture indicator is es-sential for easy field maintenance checking. It should be born in mind that bubbles in the refrigerant sight glass can be caused by pressure drop or restrictions in the liquid line, as well as inadequate liquid subcooling. Manufacturer’s published nominal working charge data should be used only as a general guide, since each installation will vary in its charge requirements.

REFRIGERANTMIGRATION

Refrigerant migration is a constant problem on transport units because of the varying temperatures to which the different parts of the system are exposed. On eutectic

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plate applications, liquid refrigerant will be driven from the condensing unit to the plates during the day’s opera-tion, with the threat of floodback on start-up. On both plate and blower units not in operation, the body and evaporator immediately after operation will be colder than the condensing unit, causing migration to the evapora-tor. During daytime hours the body and evaporator will warm up, and because of body insulation will remain much warmer than the compressor during the night hours when the ambient temperature falls, resulting in a pressure differential sufficient to drive the refrigerant to the compressor crankcase.

Excessive refrigerant in the compressor crankcase on start-up can cause slugging, bearing washout, and loss of oil from the crankcase due to foaming. Dilution of oil with excessive refrigerant results in a drastic reduction of the lubricating ability of the oil. Adequate protective measures must be taken to keep migration difficulties at a minimum. Consideration should be given to keeping the refrigerant charge as low as possible, using a pump down cycle, use of a suction accumulator, and the use of a liquid line solenoid valve.

OILCHARGE

Compressors leaving the factory are chargedwith naphthenic 150 viscosity refrigeration oil.Acompletelistofacceptablerefrigerantsandoilsisavailableon form#93-11.Thenaphthenicoilhasdefinite advantages over paraffinic oils because of less tendency toseparate from the refrigerantatreducedtemperatures.

Compressors are shipped with a generous supply of oil. However, the system may require additional oil depend-ing on the refrigerant charge and system design. After the unit stabilizes at its normal operating conditions on the initial run-in, additional oil should be added if necessary to maintain the oil level at the ¾ full level of the sight glass in the compressor crankcase. The high oil level will provide a reserve for periods of erratic oil return.

OILPRESSURESAFETYCONTROL

A major percentage of all compressor failures are caused by lack of proper lubrication. Only rarely is the lack of lubrication actually due to a shortage of oil in the system or failure of the oiling system. More often the source of the lubrication failure may be refrigerant floodback, oil trapping in the coils, or excessive slug-ging on start up.

To prevent failures from all these causes, the Emerson Climate Technologies, Inc. warranty requires that an ap-

proved manual reset type oil pressure safety control with a time delay of 120 seconds be used on all Copelametic® compressors having an oil pump. The control operates on the differential between oil pump pressure and crankcase pressure, and the time delay serves to avoid shut down during short fluctuations in oil pressure during start-up. A non-adjustable control is strongly recommended, but if an adjustable type control is used, it must be set to cut out at a net differential pressure of 9 psig. Oil pressure safety controls are available with alarm circuits which are energized should the oil pressure safety control open the compressor control circuit.

OILSEPARATORS

Proper refrigerant velocities and good system design are the only cure for oil trapping problems. Oil separa-tors are vulnerable to damage from float valve vibration, and for that reason are not commonly used on transport units. Oil separators are not normally recommended for over-the-road use on trailers, but they have been used successfully in some city operations on ice cream truck applications.

The oil separator traps a major part of the oil leaving the compressor, and since the oil is returned directly to the crankcase by means of a float valve, oil circulation in the system is minimized. On low temperature systems, oil separators may be of value in holding the amount of oil in circulation to a level which can be adequately re-turned to the compressor by the refrigerant in the system. However, on systems where piping design encourages oil logging in the evaporator circuit, an oil separator may only serve to delay lubrication difficulties.

The oil separator should be insulated to prevent refriger-ant condensation and return of liquid to the compressor crankcase. A convenient means of returning oil to the compressor, and still providing maximum protection against liquid return is to connect the oil return line to the suction line just before the suction accumulator.

CRANKCASEPRESSUREREGULATINGVALVE

In order to limit the load on the compressor, a crankcase pressure regulating valve may be necessary. During periods when the valve is throttling, it acts as a restrictor, and on start-upor during a hot gas defrost cycle, it acts as an expansion valve in the line. The preferred location for the CPR valve is ahead of the suction line accumula-tor. The accumulator will trap liquid refrigerant feeding back and allow it to boil off or feed the compressor at a metered rate to avoid compressor damage. However, location of the accumulator ahead of the CPR valve is acceptable if the accumulator has adequate capacity to

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prevent liquid floodback to the compressor.

The CPR valve should be sized for a minimum pres-sure drop to avoid loss of capacity, and should never be set above the published operating range of the compressor.

CONDENSER

Condenser construction must be rigid and rugged, and the fin surface should be treated for corrosion resistance unless the metal is corrosion resistant. The area in which the condenser is mounted affects its design. Condensers mounted on the skirt of a truck or beneath a trailer receive a great deal of road splash, while those mounted high on the nose of a truck or trailer are in somewhat cleaner atmosphere. If the condenser is mounted beneath a trailer facing in the direction of travel, a mud guard should be provided. The type of tube and fin construction affects the allowable fin spacing, but in general, fin spacing of no more than 8 fins to the inch is recommended, although some manufacturers are now using fin spacing as high as 10 and 12 per inch.

Since the unit will operate for extended periods when the vehicle is parked, ram air from the movement of the vehicle cannot be considered in designing for adequate air flow, but the condenser fan should be located so that the ram air affect aids rather than opposes condenser air flow. It also should be born in mind that often many trucks or trailers will be operating side by side at a load-ing dock, and the air flow pattern should be such that one unit will not discharge hot air directly into the intake of the unit on the next vehicle.

Since the space available for condenser face area is limited in transport refrigeration applications, the con-denser tube circuiting should be designed for maximum efficiency.

Low head pressure during cold weather can result in lubrication failure of compressors. With trucks operating or parked outside or in unheated garages in the winter months, this condition can frequently occur. A decreased pressure differential across the expansion valve will reduce the refrigerant flow, resulting in decreased refrigerant velocity and lower evaporator pressures, permitting oil to trap in the evaporator. Frequently the feed will be decreased to the point that short-cycling of the compressor results. The use of a reverse acting pressure control for cycling the condenser fan, or some other type of pressure stabilizing device to maintain reasonable head pressure is highly recommended.

RECEIVER

Because of field installation and repair, all units should be equipped either with a receiver or an adequately sized condenser so that the refrigerant charge is not critical. Valves should be provided so that the system can be pumped down. A positive liquid level indicator on the receiver will aid in preventing over-charging, and high and low test cocks have been used satisfactorily for this purpose. The size of the receiver should be held to the minimum required for safe pump down.

It is recommended that a charging valve be provided in the liquid line. While not essential, it is a fact that most servicemen will charge liquid rather than vapor into a system, and a charging valve makes this possible without damage to the compressor.

On units in operation over-the-road, powered either from the truck engine or a separate engine power source, the receiver may be subjected to temperatures higher than the condensing temperature because of the heat given off by the engine. This can result in abnormally high condensing pressures because of liquid refrigerant being forced back into the condenser, excessive refrigerant charge requirements, and flashing of liquid refrigerant in the liquid line. If excessive heating of the receiver can occur, provisions should be made for ventilation of the receiver compartment with ambient air, or the receiver should be insulated.

PURGINGAIRINASYSTEM

Occasionally due to improper installation or maintenance procedures, a unit will not be completely evacuated, or air will be allowed to enter the system after evacu-ation. The noncondensable gases will exert their own pressures in addition to refrigerant pressure, and will result in head pressures considerably above the normal condensing pressure.

Aside from the loss of capacity resulting from the higher head pressure, the presence of air in the system will greatly increase the rate of corrosion and can lead to possible carbon formation, copper plating, and/or mo-tor failure.

If it is discovered that air has been allowed to contami-nate the system, the refrigerant should be removed, and the entire unit completely evacuated with an efficient vacuum pump.

LIQUIDLINEFILTER-DRIER

On all transport refrigeration systems, because of the

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uncertainties of installation and service, a liquid line filter-drier is essential. It is recommended that the filter-drier be oversized by at least 50% for the refrigerant charge because of the many opportunities during field maintenance for moisture to enter the system. It should have flare connections for easy replacement.

HEATEXCHANGER

A heat exchanger should be considered mandatory on all units. It improves the performance, insures liquid refrigerant at the expansion valve, and helps assure the return of dry gas. Normally it should be located inside the refrigerated space to avoid loss of capacity, but it can be located externally if insulated.

LIQUIDLINESOLENOIDVALVE

When, because of the design of the system, the refrig-erant charge cannot be held to a level which can be safely handled by the compressor should refrigerant migration occur, a normally closed liquid line solenoid may be required. On 3 HP systems with refrigerant charges exceeding 15 pounds, and on 5 HP systems with refrigerant charges exceeding 20 pounds, a liquid line solenoid is recommended, and some manufactur-ers make liquid line solenoids mandatory on all units 1 ½ HP and larger.

The valve should be wired in parallel with the compres-sor so that it will be closed when the system is not in operation. It should be installed between the receiver and the expansion valve, and should have a filter-drier or strainer mounted just upstream from it in the liquid line. A soft-seated valve, of non-stick coating or similar material, is preferred for better control during over-the-road operation.

SUCTIONLINEACCUMULATOR

A suction line accumulator is considered mandatory on all systems 2 HP and larger in size, and is recom-mended for all units. The purpose of the accumulator is to intercept any liquid refrigerant which might flood through the system before it reaches the compressor, particularly on start-up or on hot gas defrost cycles. Because crankcase heaters or a pumpdown cycle are not always operative on transport units, the accumula-tor is the best protection that can be provided for the compressor.

Provisions for positive oil return to the crankcase must be provided, but a direct gravity flow is not acceptable since this would allow liquid refrigerant to drain to the crankcase during shutdown periods. Capacity of the

accumulator usually should be a minimum of 50% of the system charge, but the required size will vary with system design. Tests are recommended during the de-sign phase of any new unit to determine the minimum capacity for proper compressor protection.

An external source of heat is desirable to accelerate the boiling of the liquid refrigerant in the accumulator so that it may return to the compressor as gas. Mounting in the condenser air stream or near the compressor will normally be satisfactory.

CRANKCASEHEATERS

Because of the interruptible power source inherent in transport refrigeration, it is difficult to insure continuous operation of the heaters. A continuous drain on the truck battery would not be acceptable.

Crankcase heaters will help when connected to a continu-ous power source, but cannot be relied on for complete protection against damage from liquid migration.

PUMPDOWNCYCLE A pumpdown cycle is the best means of protecting the compressor from refrigerant damage, particularly if an excessively large charge cannot be avoided. As in the case of crankcase heaters, the fact that power may not always be available makes a pumpdown system unreliable. It is quite possible that the power to the unit might be shut off at any moment with the unit in opera-tion and refrigerant in the coils. If pumpdown control is used, special operating precautions should be taken to insure complete pumpdown before the electric power is disconnected.

FORCEDAIREVAPORATORCOILS

Air velocities across the coil should not exceed 500-600 FPM in order to avoid blowing water from the coil onto the load. Care should be taken to insure even air distri-bution across the coil, since uneven airflow can cause uneven loading of the refrigerant circuits. Fin spacing exceeding 6 per inch is not recommended because of the rapid build-up of frost on the fins. However, some users and manufacturers recommend spacing as low as 3 or 4 fins per inch, while others report satisfactory experience with spacings as high as 8 per inch provided proper defrost controls are used.

Delivered air velocity should be adequate to insure good air circulation in the vehicle. Noise level is not a design limitation in a van, so velocities up to 1,500 FPM or higher can be used.

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Internal volume of the refrigerant tubes should be kept to a minimum to keep the refrigerant volume as low as possible. Since pressure drop at low temperatures is critical so far as capacity is concerned, multiple refriger-ant circuits with fairly short runs are preferred. Pressure drop in the evaporator should be no more than 1 to 2 psig. At the same time, it is essential that velocities of refrigerant in the evaporator be high enough to avoid oil trapping. 5/8” evaporator tubes are acceptable, but ½” are preferred, and 3/8” tubing has been used successfully. Vertical headers should have a bottom outlet to allow gravity oil draining.

An evaporator face guard should be provided to protect the fins and tubing from cargo damage. Ample air inlet area should be provided, with access from both sides and the bottom if possible, to prevent blocking of air to the evaporator by cargo stacked in the vehicle.

THERMOSTATICEXPANSIONVALVES

Because of the wide range of load conditions and the premium on pulldown time in the transport field, it has been common practice for some manufacturers to oversize expansion valves used on transport units, particularly on units equipped with blower evaporator coils. If the expansion valve is oversized too greatly, surging of the refrigerant feed will result with possible floodback and erratic operation. If this occurs, a smaller valve must be used.

A liquid charged type valve is essential to retain control, since the head may frequently be colder than the sens-ing bulb. Vapor charged expansion valves should not be used on transport refrigeration systems.

Valve superheat should be preset by the valve manu-facturer and field adjustment should be discouraged. However, valves in need of adjustment should be set to provide 5°F. to 10°F. superheat at the evaporator. Too high a superheat setting will result in starving the evaporator and poor oil return. Too low a superheat set-ting will permit liquid floodback to the compressor.

Pressure limiting type valves are sometimes used to limit the compressor load according to the allowable suction pressure. Since oil return to the compressor is extremely slow during the pulldown period due to the throttling action of this type of valve, MOP valves are generally not recommended for transport applications, and a crankcase pressure regulating valve is recom-mended if the compressor load must be limited.

It should be born in mind that the pressure across the valve affects its maximum capacity and its rate of feed.

Therefore, the valve operation and the amount of su-perheat may be materially affected by changes of head pressure caused by changes in the ambient temperature. Some means of stabilizing head pressure is desirable to provide a uniform expansion valve feed.

DEFROSTSYSTEMS

A defrost system, either electrical, reverse cycle, or hot gas, is essential for satisfactory operation of any low temperature transport unit equipped with forced air evaporators. If trucks are to be used as weekend stor-age containers at temperatures close to 32°F., return air as a defrosting medium may result in load temperature fluctuations.

An electrical defrost system is feasible when the unit is operating from an engine generator set or from a station-ary electrical supply. The reverse cycle defrost using a four-way valve is exceedingly fast and effective, but may be sensitive to any foreign material in the system. Hot gas defrost using the heat of compression is effective only if some means of maintaining head pressure on the compressor is available, or if refrigerant condensing in the evaporator can be re-evaporated. Partial flooding of the condenser has been used, but this results in car-rying a very large charge of refrigerant in the system. Some proprietary systems using heat from the engine cooling water or heat from the engine exhaust have been used with success.

Drain pan heaters are required on low temperature instal-lations to prevent the build up of ice in the drain pan. To prevent the defrost heat from entering the cargo space, the evaporator fan should be stopped during defrost, or a damper installed in the air outlet.

Automatic start of the defrost cycle is recommended to avoid excessive accumulation of frost on the evaporator, and automatic termination should be provided to avoid returning overheated gas to the compressor. Since vibra-tion will cause maintenance problems on time clocks, a control responsive to fan air pressure is frequently used for defrost initiation, and a temperature responsive con-trol for defrost termination. Another method of automatic defrost control that has been used satisfactorily is a two element control sensing return air and coil tempera-tures, and operating on the differential between the two temperatures.

A suction accumulator is considered mandatory with any system using a hot gas or reverse cycle defrost system. The use of steam or hot water for cleaning or defrost

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purposes should be avoided unless a suction accumula-tor of adequate size is used to intercept the liquid driven out of the plates or evaporator by the heat.

THERMOSTAT

If the unit is controlled by a thermostat, a snap action type is essential to prevent chattering of the contacts. It is recommended that enclosed type switches be sealed against moisture. A calibrated adjustment with a set tem-perature indicator is highly desirable. The construction of the control should be such that it will withstand road shock and vibration. A liquid charged sensing bulb is desirable for fast response and accuracy of control.

HIGH-LOWPRESSURECONTROL

A combination high and low pressure control is recom-mended for all systems. If a thermostat is used for unit control, and a pumpdown system is not used, a low pres-sure control of the manual reset type should be wired in series with the thermostat to serve as a safety cut-off in the event of loss of refrigerant charge or other abnormal conditions resulting in low suction pressures.

When used for low temperature unit operational control, the low pressure control should be provided with a low differential for accurate control. For accuracy, refrigera-tion gauges must be used in setting cut-in and cut-out points, since the indicator on the face of the control is not sufficiently accurate for control purposes.

Motor-compressors with single phase motors having inherent protection, 2 HP and smaller, can be operated directly on a pressure control, but larger HP compres-sors usually require a contactor since oil pressure safety controls require a pilot circuit, as they cannot carry the running current.

EUTECTICPLATEAPPLICATIONS

Eutectic plate applications are subject to both oil logging in the evaporator and liquid floodback to the compressor on start-up unless care is taken in system layout and installation. Since either of these conditions can result in compressor failure, adequate steps must be taken to protect the compressor.

In order to avoid trapping oil, high refrigerant velocity must be maintained through the evaporator tubing. Since the velocity is dependent on the volume of refrigerant in circulation, plates should be connected in series as required to provide an adequate refrigeration load for each expansion valve circuit.

The following table may be used as a guide in deter-mining the minimum eutectic plate surface that must be connected to one expansion valve to insure velocities sufficient to return oil to the compressor. The recommen-dations are based on refrigerant evaporating tempera-tures 15°F. below the plate eutectic temperature, place manufacturers’ catalog data and recommendations, and a leaving gas velocity of 1,500 FPM. For easy field calculation, the eutectic plate surface shown is for one side of the plate only, e.g. a 24” x 60” plate would have 10 square feet of surface.

RecommendedPlateSurfaceForEachExpansionValveCircuit

Low Temperature Medium Temperature Tubing Plates Below Plates Above Diameter 0°F. Eutectic 0°F. Eutectic Minimum Maximum Minimum Maximum

5/8” O.D. 12 sq. ft. 32 sq. ft. 15 sq. ft. 32 sq. ft.



Basically the circuiting and valving of a truck plate system should be designed so that velocities in each refrigeration circuit will be above a given minimum (for adequate oil return) and below a given maximum (for a pressure drop that does not cause excessive capacity penalty). It is recommended that circuits approaching the maximum should be used whenever possible.

For example, if in a given truck for low temperature use, plates with below 0°F. eutectic solution were used, circuits might be selected as follows:

Given: 2 - 24” x 120” plates @ 20 sq. ft. each 2 - 24” x 60” plates @ 10 sq. ft. each 1 - 30” x 60” plate @ 12.5 sq. ft.

5/8”O.D.Tubing

CircuitA 1 - 24” x 120” plate 20 sq. ft.

B 1 - 24” x 120” plate 20 sq. ft.

C-Series (1 - 30” x 60” plate 12.5 sq. ft. ( 2 - 24” x 60” plates 20 sq. ft.

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3/4”O.D.Tubing

CircuitA-Series ( 1 - 30” x 60” plate 12.5 sq. ft. ( 2 - 24” x 60” plates 20 sq. ft.

B-Series 2 - 24” x 120” plates 40 sq. ft.

7/8”O.D.Tubing

Circuit

A-Series ( 1 - 30” x 60” plate 12.5 sq. ft. ( 2 - 24” x 60” plates 20 sq. ft.

B-Series 2 - 24” x 120” plates 40 sq. ft.

Normally the eutectic plates are selected by the system designer for the particular truck and application require-ment. In order to keep the refrigerant charge within ac-ceptable limits, it is important that both the total number of plates and the plate internal refrigerant volume be kept to a absolute minimum required to accomplish the desired refrigeration.

Because of the large refrigerant charge required for plates, and the variable nature of the load imposed on the compressor, plate circuits are subject to extreme varia-tions in refrigerant velocity. It has been our experience that proper velocities are of much greater importance than low pressure drop in determining the heat transfer rate between the refrigerant and the eutectic solution. Many users, following normal commercial refrigeration practice where it is assumed that refrigerant charges are low and velocities are consistently high, have placed an undue importance on low pressure drop in selecting and circuiting plates, and as a result have unknowingly created lubrication problems in their systems while gain-ing little or nothing in capacity performance. In many instances capacity has actually been reduced due to loss of proper refrigerant control.

A common misconception is that the use of separate expansion valves on each plate will give increased ca-pacity and more rapid pulldown. This is not necessarily so. The use of more expansion valves will result in a lower pressure drop through the refrigerant circuit which might aid capacity slightly, but in most cases the resulting improper control actually decreases capacity.

On two plates, for example, the use of two expansion valves would result in two sections of tubing being used as drier area in order to obtain the necessary superheat for proper operation of the expansion valves. If only one expansion valve were used, only one length of tubing

for this superheating function would be required, and the effective refrigeration area would be increased. The use of one expansion valve on multiple plates results in a much higher velocity, and as a result the scrubbing action of the refrigerant on the walls of the tube causes a much higher rate of heat transfer. Our experience would indicate, particularly at low evaporating temperatures, that very possibly multiple plates operating on one expansion valve will have more capacity and a better pulldown than the same plates operating with individual expansion valves.

A similar misconception is that the use of larger O.D. tubing in plates will result in a lower pressure drop and therefore increase capacity. As in the case of expan-sion valves, the use of smaller tubing, although possibly resulting in a slightly higher pressure drop, will greatly increase refrigerant velocity, increase the heat transfer rate as a result, and again our experience indicates on low temperature plates that capacity may actually be increased because of the smaller tubing. The smaller tub-ing requires a smaller refrigerant charge, and therefore also decreases the problem of refrigerant migration.

Expansion valves on plate circuits should be no larger than 1 ton size, and ½ ton valves will give better control on smaller circuits in the medium temperature range. The piping and thermal sensing bulbs should be located so that each valve operates independently and is not influ-enced by the return line controlled by another valve.

Field experience indicates that due to the throttling action of an MOP valve after shut-down or defrost periods, oil may not be returned to the crankcase at a fast enough rate to maintain compressor lubrication in the event oil is lost from the compressor on start-up due to liquid re-frigerant foaming in the crankcase. Therefore, pressure limiting type expansion valves are not recommended for plate circuits.

Because of the amount of oil trapped in the plates dur-ing operation, additional oil normally must be added to the compressor during the initial pulldown cycle, or after the unit reaches its normal operating conditions. Sufficient oil should be added to maintain the oil level at approximately the ¾ full level of the compressor oil sight glass.

As the eutectic solution becomes frozen, the boiling action of the refrigerant slows, and a higher percentage of liquid refrigerant lies in the bottom of the evapora-tor tubing. When the unit cycles off, or the power is disconnected, the plates may be partially filled with liquid refrigerant and oil. At some later time when the compressor is again started, the liquid will flood back to the compressor.

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To protect against liquid floodback, a suction accumu-lator is mandatory on units of 2 HP and larger, and is recommended on all transport units. If a crankcase pressure regulating valve is used, the accumulator should be located if possible between the CPR valve and the compressor in order to provide the maximum protection.

A liquid line solenoid valve can be helpful in minimizing migration from the condenser and receiver to the evapo-rator and compressor during periods when the unit is not in operation. If the system refrigerant charge is not excessive, a liquid line solenoid may not be required, but some manufacturers feel they should be mandatory on all plate systems 1 ½ HP in size and larger.

All plate applications should be equipped with the follow-ing:

a. Properly sized expansion valves.

b. A liquid to suction heat exchanger for maximum efficiency.

c. A liquid line filter-drier.

d. A combination sight glass and moisture indicator for easy maintenance.

e. An oil pressure safety control on all compressors having oil pumps.

f. A reverse acting pressure control to stop the con-denser fan in order to maintain satisfactory compres-sor head pressure during cold weather operation.

g. Suction line accumulator (2 HP and larger).

One of the major problems is low temperature eutectic plate applications is the practice of the operator or serviceman of reducing the low pressure cut-out below the operating limits of the refrigeration system, possibly to such a low setting that the resulting refrigerant ve-locities are too low to return oil to the compressor. This practice has been stimulated by the demand for lower and lower ice cream temperatures, and the serviceman often fails to realize the hazard he is creating. The in-creased compression ratio is not a problem in a properly designed compressor so long as adequate lubrication is maintained. But once the eutectic solution is frozen, the decrease in evaporator load causes the compres-sor suction pressure to drop rapidly, and at extremely low suction pressures, compressor capacity falls off rapidly. From -25°F. to -40°F. the capacity may decrease by 50% in the best R-12 low temperature compressor,

and from -25°F. to -50°F. the reduction in capacity may be as high as 75%. As a result, there may no longer be adequate refrigerant velocity in the evaporator circuit to return oil to the crankcase. At such low capacities, the expansion valve may no longer be able to properly control the liquid refrigerant feed.

Repeated extended periods of operation below the operating range of the system are almost certain to result in eventual compressor failure. There is no cure for this situation except adequate education of the user. To provide proper protection for the compressor the low pressure control should be set to cut out at approximately 10°F. below the normal evaporating temperature.

For example, a system equipped with plates containing -8°F. or -9°F. eutectic solution will normally operate with a refrigerant evaporating temperature of approximately -25°F., and the low pressure cut-out should be set at the equivalent of -35°F. or 8” of vacuum on R-12 refrig-erant, and 5 psig on R-502. If the system is controlled by a thermostat sensing the truck air temperature, the thermostat should be set no lower than the plate eutectic temperature.

The user must realize that a compressor’s application is limited by the rest of the system. Because of the inherent problems of oil return presented by the shape and mounting characteristics necessitated in truck ap-plications, and the large amounts of tubing which must be used in plate construction, the minimum satisfactory evaporating temperature for both R-12 and R-502 is approximately -40°F. Lowpressure controlsonallplatesystemsmustbesettocutoutatorabovethe equivalent pressure setting; for R-12, 11” ofvacuum,andforR-502,5psig.

In order to maintain the evaporating temperature within acceptable limits, it is essential that the combination of condensing unit and plates be properly balanced. The selection of too small a condensing unit may result in a freezing rate that is too slow. But of equal and possibly greater importance, the selection of too large a condens-ing unit may result in an excessively large temperature difference between the plate eutectic temperature and the refrigerant evaporating temperature. This condi-tion most frequently occurs when a large condensing unit is selected in order to achieve a quick pulldown, or to shorten the time necessary to freeze the eutectic solution. Since the minimum satisfactory evaporating temperature is approximately -40°F., the condensing unit should be selected so that the normal operating evaporating temperature on low temperature plates is not below -30°F. to -35°F.

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REFRIGERANTPIPING

Normal good piping practice should be followed in installing refrigerant lines for split systems. A silver solder alloy should be used for making connections to the compressor and for long runs of tubbing where vibration may be a problem, and a high temperature silver solder alloy only must be used on compressor discharge lines. For other connections, 95/5 solder is acceptable, and makes possible easier field repair. 50/50 solder should not be used since it does not have sufficient strength for transport usage. Acid core type solder should not be used.

VIBRATION

The greatest single hazard of transport refrigeration usage is damage from vibration and shock. Although shock tests on the nose of trailers have recorded very high shock levels, the great majority of all failures from this source are due to the cumulative effect of small vibrations. Any line, capillary tube, or structural member that is subjected to continuous sharp vibration, or that rattles against a neighboring member in operation is almost certain to fail within a fairly short period of time. It cannot be stressed too strongly that normal commercial construction of condensing units and evaporators for the usual commercial application is not adequate for over the road usage.

Emerson Climate Technologies, Inc. manufactures a line of condensing units specially designed for transport usage. The frames are ruggedly constructed, and all components are mounted to minimize vibration.

When compressors are installed in a system manufac-turer’s condensing unit, care must be taken to see than the compressor is bolted down firmly. Neoprene or other resilient shock mounts may be used, but spring mounting is not acceptable. Internally spring mounted compres-sors are not suitable for transport applications due to the danger of internal damage from severe shocks, and continuous spring movement.

Vibration eliminators should be mounted in the com-pressor discharge and suction lines. A very common fault is the installation of a vibration absorber between two sections of rigid piping, in which case the vibration absorber may be as rigid as the piping. Metal vibration eliminators should never be mounted in such a fashion that they are subjected to stress in either compression or extension. An improperly installed vibration elimina-tor can actually cause line failure. Flexible refrigerant lines such as Aeroquip, Stratoflex, or Anchor which are specifically designed for use with the appropriate

refrigerant may be used in place of metallic vibration absorbers. Metallic vibration absorbers should have joints adequately sealed to prevent condensation from freezing and damaging the joints.

Welding is preferable to bolting in fastening structural members. Sheet metal screws, and other metal fas-teners not securely held by lock washers or lock nuts are not dependable. All wiring and piping should be protected with grommets where passing through sheet metal holes.

Evaporator and condenser tube sheets, when used for mounting, should be of solid, one piece construction, and may require heavier gauge construction than used in normal commercial practice for strength purposes. Coil tube sheets should be manufactured with collars, as raw edge holes can cut the tubing due to vibration.

ELECTRICALPRECAUTIONS

Electrical failures are a common field maintenance prob-lem due to the wet environment, shock and vibration, and the possibility of improper power from an engine generator set.

For the safety of operating and maintenance personnel, the electrical system should be grounded to theframe,andtheframeinturngroundedbymeansofachainormetallinktothegroundifageneratorsetismountedonthevehicle. All components should be grounded from one to the other, such as the generator set to condensing section to evaporator section. Cablestoremotesourcesofpowershouldcarryanextrawireforgroundingpurposesatthesupplyplug.

At the time of manufacturer, each system should be given a high potential test to insure against electrical flaws in the wiring. All relays and terminals should be protected against the weather, and all wiring should be covered with protective loom to guard against abrasion. All switches should be of the sealed type, recommended by the manufacturer for use in wet environments. Plug type line connectors should be of the waterproof type. Electrical cables connecting split units should have a watertight cable cover, or should be run in conduit. All wiring should be fastened securely to prevent chafing, and should be clearly identified by wire marking and/or following the color code specified by the National Elec-trical Code.

Adequately sized extension cords, plugs, and recep-tacles must be used to avoid excessive voltage drop. Voltage at the compressor terminals must be within 10% of the nameplate rating, even under starting conditions.

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Many single phase starting problems on small delivery trucks can be traced to the fact that power is supplied to the compressor from household type wiring circuits through long extension cords, neither of which are sized properly for the electrical load. Single phase open type motors which are used for belt driving a compressor during over-the-road operation must be equipped with a relay to break the capacitor circuit, rather than a centrifugal switch. The variable speed operation expe-rienced during truck operation may cause a centrifugal switch to fail because of excessive wear at low operat-ing speeds. All start capacitors must be equipped with bleed resistors to permit the capacitor charge to bleed off rapidly, preventing arcing and overheating of the relay contacts.

Whenunitsareoperatedfromseveralpowersources,besureallplugsandreceptaclesarewiredinthesamesequence,so that thecompressor rotationwillnotbereversed.

INSTALLATION

A large number of field failures that now occur could be prevented by proper installation practice. To assure trouble free operation, every effort should be made to carry out the following minimum procedures.

1. Read the manufacturer’s instructions.

2. Be sure that structural or reinforced members are provided to mount the units.

3. Thoroughly clean all copper lines before assembling. Do not use steel wool for cleaning since the metal slivers may cause electrical problems in the com-pressor. If the tubing is not precleaned and capped, pull a rag saturated with refrigerant oil through the tube and blow out with nitrogen prior to connecting lines to the evaporator and condenser.

4. Use only a suitable silver solder alloy or 95/5 solder in making soldered joints.

5. When brazing lines, circulate inert gas such as dry nitrogen through the line to prevent oxidation.

6. Install piping in the wall or floor of the vehicle, or provide an adequate guard.

7. After the lines are installed, pressurize to 150 psig, and leak test. The use of an electronic leak detector is recommended for greater sensitivity. As a final check, the system should be sealed for 12 hours after pulling a deep vacuum. If the vacuum will not hold, the system should be rechecked for leaks,

repaired, and retested to insure that it is ready for evacuation and charging.

8. Use a good high vacuum pump to evacuate the system and leave the pump on the system for a minimum of 4 hours. Evacuate to less than 1,500 microns, and break the vacuum with refrigerant to 5 psig. Repeat the evacuation process, and break with refrigerant as before. Evacuate a final time to 500 microns or less and the system is ready for charging.

WARNING: Topreventmotordamagedonotuse the motor-compressor to evacuate thesystem.Amotor-compressorshouldneverbestartedoroperatedwhilethesystemisunderadeepvacuum,orseriousdamagemayresultbecauseof the reduceddielectricstrengthoftheatmospherewithinthemotorchamber.

9. Charge the unit with refrigerant, either vapor through the suction valve, or preferably liquid through a liquid line charging valve if provided. Thecompressormustneverbechargedwithliquidrefrigerantthroughthesuctionside.

10. If using an engine-generator as a power source,

start the engine and check the generator output voltage to be sure it is correct.

11. Check the voltage at the compressor terminals, start the unit, check the amperage draw of the compres-sor, and the rotation of the fans to be sure the unit is phased properly.

12. Observe the discharge and suction pressures. If an abnormal pressure develops, stop the unit immedi-ately and check to see what is causing the difficulty. Take corrective action if required.

13. Observe the refrigerant oil level and check the oil pressure, if the compressor is equipped with a posi-tive displacement oil pump. If the oil level becomes dangerously low during the pulldown period, add oil to the compressor. After the unit reaches normal operating conditions, add oil if necessary to bring the level to a point ¾ full in the crankcase sight glass.

14. Check all manual and automatic controls.

15. After a minimum of two hours of operation, make another leak test.

16. After the unit has reached the proper operating conditions, and all controls have been checked, run the unit overnight on automatic control to be

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sure operation is satisfactory. Check oil level in the compressor, and add oil if necessary.

17. When unit is delivered to the customer, be sure that operating personnel have proper written instruc-tions on operating and maintenance procedures. The responsible sales personnel should verbally explain the operation of the unit to the user, and wiring diagrams and operating instructions should be permanently carried on the vehicle, either by means of a decal or in an envelope properly protected form loss or damage.

FIELDTROUBLESHOOTINGONTRANSPORTUNITS

The great majority of all low temperature compressor failures in transport refrigeration can be traced to lubrica-tion problems. No compressor can operate satisfactorily unless oil logging and liquid floodback can either be prevented or safely controlled by safeguard devices in the refrigeration system.

The following check-off list covers possible corrective action on units experiencing field difficulties. For a more detailed discussion of each item, refer to the appropri-ate section in this manual. The need for any particular modification would of course depend on the individual application.

1. Eutectic plate circuiting for high refrigerantvelocity

It is essential for proper oil return to the compressor that high refrigerant velocities be maintained through the evaporator circuits. For a rough rule of thumb on plates with 7/8” or ¾” O.D. tubing, there should be no less than 3 small or 2 large plates in series on one expansion valve. On plates with 5/8” O.D. tubing, there should be no less than 2 small or one large plate on one expansion valve.

2. Expansionvalvesoneutecticplates

Expansion valves should be no larger than 1 ton capacity in size, liquid or cross charged, and inter-nally equalized.

3. RefrigerantCharge

The refrigerant charge must be held to a minimum to avoid refrigerant migration problems. Use a sight glass to check for a liquid seal at the expansion valve at low temperature operating conditions.

4. LiquidLineSolenoidValve

If excessive refrigerant migration to eutectic plates is occurring during over-the-road operation, a liquid line solenoid valve may be required to properly control large refrigerant charges.

5. SuctionLineAccumulator

A suction line accumulator is the best protection that can be provided to guard against liquid floodback. Itshouldbemandatoryonalltruckapplications2HPandlarger. For maximum efficiency, it should be installed close to the compressor, and if a CPR valve is used, between the compressor and the CPR valve. The accumulator must have provisions for positive oil return.

6. HeadPressureControl

In winter operation, head pressures may drop so low that inadequate feeding of the expansion valve may result, and the evaporator may be starved. A reverse acting high pressure control should be used to cycle the condenser fan if head pressure drop below 80 psig on R-12 operation, or 125 psig on R-502 operation, unless other acceptable means of controlling head pressure are provided.

7. OilLevelinCrankcase

Whenthecompressorswithoutoilpumpsareusedontruckapplications,theoillevelshouldbe maintained high in the compressor sightglasstoassureareserveoflubricatingoilforperiodsoferraticoilreturn.Theusershouldbewarnedthatthecompressormaynotbegettingadequatelubricationiftheoilleveldropsbelowthebottomofthesightglass.Onlyanaphthenicoilshouldbeusedwhichhasaviscosityof150,a pour point of -35°F. and a floc point of -70°F. Thisoilhasprovensatisfactoryforalllowtem-peratureapplications.

8. OilPressureSafetyControl

On all compressors having positive displacement type oil pumps, an oil pressure safety control is required.

9. HighPressureCut-Out

Several manufacturers have produced units with no high pressure control. Failure of the condenser fan motor may result in excessive head pressures,

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the compressor can operate under such conditions, the compressor pays the penalty. The user must realize that a compressor’s application is limited by the rest of the system. Because of the inherent problems of oil return presented by the shape and mounting characteristics necessitated in truck ap-plications, and the large amounts of tubing which must be used in plate construction, the minimum satisfactory evaporating temperature for both R-12 and R-502 is approximately -40°F. Lowpressurecontrolsontransportsystemsmustbesettocutoutatorabove11”ofvacuumonR-12,and5psigonR-502.

12.LocationofTruckwhileSystemisOperating

Trucks must be parked on a reasonably level surface while the refrigeration unit is in operation. Short pe-riods of operation on an incline such as experienced in over-the-road operation are not a problem, but long periods of operation while the truck is parked on a steep incline or on the side of a hill may rob the compressor of lubrication if the oil level flows away from the pick up point of the oil flinger or oil pump.

and subsequent compressor failure. A high pressure control is essential.

10.LiquidLineFilter-DrierandHeatExchanger

These should be standard on all units.

11. LowPressureControlSetting

A major educational effort is required to point out to the user the dangers of by-passing the low pressure cut-out, or setting it at dangerously low levels.

When eutectic plates are completely frozen, the compressor suction pressure falls very rapidly, with a consequent sharp drop in compressor capacity, and resulting lubrication difficulties, since velocity in the plates may no longer be sufficient to return oil to the compressor. Users, particularly ice cream distributors, frequently try to reduce the van body temperature to the lowest temperature possible as an added safety factor for the day’s operation.

Since the system is normally not designed for the extremely low evaporating temperatures at which

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SECTION21CAPACITYCONTROL

On many refrigeration and air conditioning systems, the refrigeration load will vary over a wide range. This many be due to differences in product load, ambient temperature, usage, occupancy, or other factors. In such cases compressor capacity control is a necessity for satisfactory system performance.

The simplest form of capacity control is “on-off” opera-tion of the compressor. This works acceptably with small compressors, but for larger compressors, it is seldom satisfactory, because of fluctuations in the controlled temperature. Under light load conditions it can result in compressor short cycling. On refrigeration applications where ice formation is not a problem, users frequently reduce the low pressure cut-out setting to a point beyond the design limits of the system in order to prevent short cycling. As a result, the compressor may operate for long periods at extremely low evaporating temperatures. Both of these conditions can cause compressor damage and ultimate failure.

Two different types of unloading are used on Copeland® brand compressors, internal and external.

INTERNALCAPACITYCONTROLVALVES

A schematic illustration of the Copeland® brand internal unloading valve is shown in Figure 101.

In the normal operating position with the solenoid valve de-energized, the needle valve is seated on the lower port, and the unloading plunger chamber is exposed to suction pressure through the suction pressure port. Since the face of the plunger is open to the suction chamber, the gas pressures across the plunger are equalized, and the plunger is held in the open position by the spring.

When the solenoid valve is energized, the needle valve is seated on the upper port, and the unloading plunger chamber is exposed to discharge pressure through the discharge pressure port. The differential between discharge and suction pressure forces the plunger down, sealing the suction port in the valve plate, thus preventing the entrance of suction vapor into the un-loaded cylinders.

With the suction port sealed, the cylinder pumps down into a vacuum until it reaches a point where no pump-ing action occurs.

EXTERNALCAPACITYCONTROLVALVES

For Copeland® brand three cylinder compressors, a solenoid operated external bypass valve is used for unloading, as shown in Figure 101.

Copelametic® compressors with external capacity con-trol have a bypass valve so arranged that the unloaded cylinder is isolated from the discharge pressure created by the unloaded cylinders. The bypass valve connects the discharge ports of the unloaded cylinder to the com-pressor suction chamber. Since the piston and cylinder do not work other than pumping vapor through the by-pass circuit, and handle only suction vapor, the problem of cylinder overheating while unloaded is practically eliminated. At the same time, the power consumption of the compressor motor is greatly reduced because of the reduction in work performed.

Because of the decreased volume of suction vapor returning to the compressor from the system and avail-able for motor cooling, the operating range of unloaded compressors must be restricted. In general, Copeland® brand compressors with capacity control are recom-mended only for high temperature applications, but in some instances they can be satisfactorily applied in the medium temperature range. Because of the danger of overheating the compressor motor on low temperature systems, either cycling the compressor or hot gas bypass is recommended.

HOTGASBYPASS

Compressor capacity modulation by means of hot gas bypass is recommended where normal compressor cycling or the use of unloaders may not be satisfactory. Basically this is a system of bypassing the condenser with compressor discharge gas to prevent the compressor suction pressure from falling below a desired setting.

All hot gas bypass valves operate on a similar principle. They open in response to a decrease in downstream pressure, and modulate from fully open to fully closed over a given range. Introduction of the hot, high pres-sure gas into the low pressure side of the system at a metered rate prevents the compressor from lowering the suction pressure further.

The control setting of the valve can be varied over a wide range by means of an adjusting screw. Because of the reduced power consumption at lower suction pres-sures, the hot gas valve should be adjusted to bypass at

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the minimum suction pressure within the compressor’s operating limits which will result in acceptable system performance.

If a refrigeration system is properly designed and in-stalled, field experience indicates that maintenance may be greatly reduced if the compressor operates continuously within the system’s design limitations as opposed to frequent cycling. Electrical problems are minimized, compressor lubrication is improved, and liquid refrigerant migration is avoided.

Therefore, on systems with multiple evaporators where the refrigeration load is continuous, but may vary over a wide range, hot gas bypass may not only provide a convenient means of capacity control, it may also result in more satisfactory and more economical operation.

BYPASSINTOEVAPORATORINLET

On single evaporator, close connected systems, it is frequently possible to introduce the hot gas into the evaporator inlet immediately after the expansion valve. Distributors are available with side openings for hot gas inlet. Bypassing at the evaporator inlet has the effect of creating an artificial cooling load. Since the regular system thermostatic expansion valve will meter its feed as required to maintain its superheat setting, the refriger-ant gas returns to the compressor at normal operating temperatures, and no motor heating problem is involved. High velocities are maintained in the evaporator, so oil

return is aided. Because of these advantages, this type of control is the simplest, least costly and most satisfac-tory bypass system. This type of bypass is illustrated in Figure 102.

BYPASSINTOSUCTIONLINE

Where multiple evaporators are connected to one compressor, or where the condensing unit is remote from the evaporator it may be necessary to bypass hot gas into the refrigerant suction line. Suction pressures can be controlled satisfactorily with this method, but a desuperheating expansion valve is required to meter liquid refrigerant into the suction line in order to keep the temperature of the refrigerant gas returning to the compressor within allowable limits. It is necessary to thoroughly mix the bypassed hot gas, the liquid refrig-erant, and the return gas from the evaporator so that the mixture entering the compressor is at the correct temperature. A mixing chamber is recommended for this purpose, and a suction line accumulator can serve as an excellent mixing chamber while at the same time protecting the compressor from liquid floodback. See Figure 103 for typical installation.

Another commonly used method of mixing is to arrange the piping so that a mixture of discharge gas and liquid refrigerant is introduced into the suction line at some distance form the compressor, in a suction header if possible. Figure 104 illustrates this mixing method.

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SOLENOIDVALVESFORPOSITIVESHUT-OFFANDPUMPDOWNCYCLE

In order to allow the system to pumpdown, a solenoid valve must be installed ahead of the hot gas bypass valve. Since the hot gas valve opens on a decrease of downstream pressure, it will be open any time the sys-tem pressure is reduced below its setting. If the system control is such that this solenoid valve is closed during the normal cooling cycle, it may also prevent possible loss of capacity due to leakage.

A solenoid valve is also recommended ahead of the desuperheating expansion valve to prevent leakage and allow pumpdown. Both of the solenoid valves should be of the normally closed type, and wired so they are de-energized when the compressor is not operating.

DESUPERHEATINGEXPANSIONVALVE

If a desuperheating expansion valve is required, it should be of adequate size to reduce the temperature of the discharge gas to the proper level under maximum bypass conditions. The temperature sensing bulb of the expansion valve must be located so that it can sense the temperature of the gas returning to the compressor after the introduction of the hot gas and the desuperheating liquid. Suction gas entering the compressor should be no higher than 65°F. under low temperature load conditions, or 90°F. under high temperature load conditions.

On low temperature applications where hot gas bypass is used to prevent the compressor suction pressure from falling below safe operating levels, valves with unusually high superheat settings may be required. For example, suppose a control was desired to prevent a system using R-502 from operating below -35°F. The temperature of the gas returning to the compressor must be prevented from exceeding 65°F. Therefore, when the desuperheating expansion valve is feeding, it will sense on one side of its diaphragm, the system pressure equivalent to -35°F. or 6.7 psig, and in order to maintain 65°F. return gas, it will require a superheat setting of 65°F. plus 35°F. or 100°F. Expansion valves with special charges are available from expansion valve manufacturers with superheat settings over ex-tremely wide ranges, although these will not normally be available in a local wholesaler’s stock. Contact the expansion valve manufacturer’s local representative for assistance in selecting valves with nonstandard superheat settings.

TYPICALMULTIPLE-EVAPORATORCONTROLSYSTEM

A typical hot gas bypass control system with three evaporators is illustrated in Figure 103 together with a schematic electric control system for cycling control of the compressor. The double pole thermostats close on a demand for refrigeration, and as long as any one evaporator is demanding cooling the compressor oper-ates, and the hot gas bypass valve modulates flow as

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necessary to prevent the suction pressure from falling below a fixed set point.

If all evaporators are satisfied, all of the thermostats are open, and all liquid line solenoid valves and the hot gas solenoid valve are de-energized, and therefore closed. The compressor will then cycle off on low pressure control until a thermostat again closes.

In order to protect the compressor against danger from liquid flooding in the event of a trip of a compressor safety device, provision must be made in the wiring circuit to de-energize the hot gas and the desuperheating liquid line solenoid valves if the compressor is inoperative. On a pumpdown system, this can be accomplished by means of a solenoid valve control relay as shown in Figure 103.

If continuous compressor operation is desired, single pole thermostats can be used, and the hot gas and desuperheating liquid line solenoid valves should be connected directly to the load side of the compressor contactor. In the event all three evaporators are satisfied, the compressor will operate on 100% hot gas bypass until cooling is again required.

Compressors equipped with inherent protection can cycle on the inherent protection can cycle on the inher-ent protector independently of the contractor. To avoid flooding the compressor with liquid refrigerant in the

event the inherent protector should trip, the hot gas solenoid valve and the liquid line solenoid valve should be connected through a current sensing relay such as the Penn R-10A, as shown in Figure 105.

POWERCONSUMPTIONWITHHOTGASBYPASS

Since the power consumption as well as the capacity of a compressor is reduced with a decrease in com-pressor suction pressure, the control system should be such that the system is allowed to reach its lowest satisfactory operating suction pressure before hot gas is bypassed. Where major reductions in capacity are required, operating economy may be best achieved by handling the load with two compressors. One can be cycled for a 50% reduction in both capacity and power, while the capacity of the compressor remaining on the line is modulated by hot gas control.

It is not necessarily true that continuous compressor operation with hot gas bypass will result in a higher power bill than cycling operation for a given load. Almost all utilities make a monthly demand charge based on peak loads. Since the peak motor demand occurs when locked rotor current is drawn on start-up, the utility demand charge may reflect motor starting requirements rather than the true running load. With continuous operation, once the motors are on the line, starting peaks may be eliminated and the reduction in the demand charge may offset the increased running power consumption.

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SECTION22LIQUIDREFRIGERANTCONTROLINREFRIGERATIONANDAIRCONDITIONINGSYSTEMS

attracted by oil and will vaporize and migrate through the system to the compressor crankcase even though no pressure difference exists to cause the movement. On reaching the crankcase the refrigerant will condense into a liquid, and this migration will continue until the oil is saturated with liquid refrigerant. The amount of refrigerant that the oil will attract is primarily dependent on pressure and temperature, increasing rapidly as the pressure increases and approaching a maximum at saturated pressures and temperatures in the normal room temperature range.

When the pressure on a saturated mixture of refrigerant and oil is suddenly reduced, as happens in the compres-sor crankcase on start-up, the amount of liquid refrigerant required to saturate the oil is drastically reduced, and the remainder of the liquid refrigerant flashes into vapor, causing violent boiling of the refrigerant and oil mixture. This causes the typical foaming often observed in the compressor crankcase at start-up, which can drive all of the oil out of the crankcase in less than a minute. (Not all foaming is the result of refrigerant in the crankcase - agitation of the oil will also cause some foaming.)

One condition that is somewhat surprising when first encountered by field personnel is the fact that the intro-duction of excessive liquid refrigerant into the compressor crankcase can cause a loss of oil pressure and a trip of the oil pressure safety control even though the level of the refrigerant and oil mixture may be observed high in the compressor crankcase sight glass. The high percent-age of liquid refrigerant entering the crankcase not only reduces the lubricating quality of the oil, but on entering the oil pump intake may flash into vapor, blocking the entrance of adequate oil to maintain oil pump pressure, and this condition can continue until the percentage of refrigerant in the crankcase is reduced to a level which can be tolerated by the oil pump.

Liquid refrigerant problems can take several different forms, each having its own distinct characteristics.

REFRIGERANTMIGRATION

Refrigerant migration is the term used to describe the accumulation of liquid refrigerant in the compressor crankcase during periods when the compressor is not operating. It can occur whenever the compressor becomes colder than the evaporator, since a pressure differential then exists to force refrigerant flow to the colder area. Although this type of migration is most pronounced in colder weather, it can also exist even at relatively high ambient temperatures with remote type condensing units for air conditioning and heat pump applications.

One of the major causes of compressor failure is dam-age caused by liquid refrigerant entering the compressor crankcase in excessive quantities. Since improper control of liquid refrigerant can often cause a loss of lubrication in the compressor, most such compressor failures have been classified as lubrication failures, and many people fail to realize that the problem actually originates with the refrigerant.

A well designed, efficient compressor for refrigeration, air conditioning and heat pump duty is primarily a va-por pump designed to handle a reasonable quantity of liquid refrigerant and oil. To design and build a pump to handle more liquid would require a serious compromise in one or more of the following: size, weight, capacity, efficiency, noise, and cost.

Regardless of design there are limits to the amount of liquid a compressor can handle, and these limits depend on factors such as internal volume of the crankcase, oil charge, type of system and controls, and normal operat-ing conditions. Proper control of liquid refrigerant is an application problem, and is largely beyond the control of the compressor manufacturer.

The potential hazard increases with the size of the re-frigerant charge and usually the cause of damage can be traced to one or more of the following:

1. Excessive refrigerant charge.

2. Frosted evaporator.

3. Dirty or plugged evaporator filters.

4. Failure of evaporator fan or fan motor.

5. Incorrect capillary tubes.

6. Incorrect selection or adjustment of expansion valves.

7. Refrigerant migration.

REFRIGERANT-OILRELATIONSHIP

In order to correctly analyze system malfunctions, and to determine if a system is properly protected, a clear understanding of the refrigerant-oil relationship is essential.

One of the basic characteristics of a refrigerant and oil mixture in a sealed system is the fact that refrigerant is

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Anytime the system is shut down and is not operative for several hours, migration to the crankcase can occur regardless of pressure due to the attraction of the oil for refrigerant.

If excessive liquid refrigerant has migrated to the com-pressor crankcase, severe liquid slugging may occur on start-up, and frequently compressor damage such as broken valves, damaged pistons, bearing failures due to loss of oil from the crankcase, and bearing washout (refrigerant washing oil from the bearings) can occur.

LIQUIDREFRIGERANTFLOODING

If an expansion valve should malfunction, or in the event of an evaporator fan failure or clogged air filters, liquid refrigerant may flood through the evaporator and return through the suction line to the compressor as liquid rather than vapor. During the running cycle, liquid flooding can cause excessive wear of the moving parts because of the dilution of the oil, loss of oil pressure resulting in trips of the oil pressure safety control, and loss of oil from the crankcase. During the “off” cycle after running in this condition, migration of refrigerant to the crankcase can occur rapidly, resulting in liquid slugging when restarting.

LIQUIDREFRIGERANTSLUGGING

Liquid slugging is the term used to describe the passage of liquid refrigerant through the compressor suction and discharge valves. It is evidenced by a loud metallic clatter inside the compressor, possibly accompanied by extreme vibration of the compressor.

Slugging can result in broken valves, blown head gas-kets, broken connecting rods, broken crankshafts, and other major compressor damage.

Slugging frequently occurs on start-up when liquid refrigerant has migrated to the crankcase. On some units, because of the piping configuration or the loca-tion of components, liquid refrigerant can collect in the suction line or evaporator during the off cycle, returning to the compressor as solid liquid with extreme velocity on start-up. The velocity and weight of the liquid slug may be of sufficient magnitude to override any internal anti-slug protective devices of the compressor.

TRIPPINGOFOILPRESSURESAFETYCONTROL

One of the most common field complaints arising from a liquid flooding condition is that of a trip of the oil pressure safety control after a defrost period on a low

temperature unit. The system design on many units allows refrigerant to condense in the evaporator and suction line during the defrost period, and on start-up this refrigerant floods back to the compressor crankcase, causing a loss of oil pressure and recurring trips of the oil pressure safety control.

One trip or a few trips of the oil pressure safety control may not result in serious damage to the compressor, but repeated short periods of operation without proper lubrication are almost certain to result in ultimate com-pressor failure. Trips of the oil pressure safety control under such circumstances are frequently viewed by the serviceman as nuisance trips, but it cannot be stressed too strongly that they are warning trips, indicating the compressor has been running without oil pressure for 2 minutes, and that prompt remedial action is required.

RECOMMENDEDCORRECTIVEACTION

The potential hazard to a refrigeration or air conditioning system is in almost direct proportion to the size of the refrigerant charge. It is difficult to determine the maximum safe refrigerant charge of any system without actually testing the system with its compressor and other major components. The compressor manufacturer can deter-mine the maximum amount of liquid the compressor will tolerate in the crankcase without endangering the working parts, but has no way of knowing how much of the total system charge will actually be in the compressor under the most extreme conditions. The maximum amount of liquid a compressor can tolerate depends on its design, internal volume, and oil charge. Where liquid migration, flooding, or slugging can occur, corrective action should be taken, the type normally being dictated by the system design and the type of liquid problem.

1. MinimizeRefrigerantCharge The best compressor protection against all forms

of liquid refrigerant problems is to keep the charge within the compressor limits. Even if this is not pos-sible, the charge should be kept as low as reasonably possible.

Use the smallest practical size tubing in condensers, evaporators, and connecting lines. Receivers should be as small as possible.

Charge with the minimum amount of refrigerant re-quired for proper operation. Beware of bubbles showing in the sight glass caused by small liquid lines and low head pressures. This can lead to serious overcharg-ing.

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2. PumpdownCycle

The most positive and dependable means of properly controlling liquid refrigerant, particularly if the charge is large, is by means of a pumpdown cycle. By clos-ing a liquid line solenoid valve, the refrigerant can be pumped into the condenser and receiver, and the compressor operation controlled by means of a low pressure control. The refrigerant can thus be isolated during periods when the compressor is not in opera-tion, and migration to the compressor crankcase is prevented. A recycling type of pumpdown control is recommended to provide protection against possible refrigerant leakage through control devices during the off cycle. With the so-called one time pumpdown, or non-recycling type of control, sufficient leakage may occur during long off periods to endanger the compressor.

Although the pumpdown cycle is the best possible protection against migration, it will not protect against flooding during operation.

3. CrankcaseHeaters

On some systems, operating requirements, cost, or customer preference may make the use of a pump-down cycle undesirable, and crankcase heaters are frequently used to retard migration.

The function of a crankcase heater is to maintain the oil in the compressor at a temperature higher than the coldest part of the system. Refrigerant enter-ing the crankcase will then be vaporized and driven back into the suction line. However, in order to avoid overheating and carbonizing of the oil, the wattage input of the crankcase heater must be limited, and in ambient temperatures approaching 0°F., or when exposed suction lines and cold winds impose an added load, the crankcase heater may be overpowered, and migration can still occur.

Crankcase heaters when used are normally energized continuously, since its takes several hours to drive the refrigerant from the crankcase once it has entered and condensed in the oil. They are effective in combating migration if conditions are not too severe, buttheywill not protect against liquid floodback.

4. SuctionAccumulators

On systems where liquid flooding is apt to occur, a suction accumulator should be installed in the suction line. Basically the accumulator is a vessel which serves as a temporary storage container for liquid refrigerant

which has flooded through the system, with a provision for metered return of the liquid to the compressor at a rate which the compressor can safely tolerate.

Flooding typically can occur on heat pumps at the time the cycle is switched from cooling to heating, or from heating to cooling, and a suctionaccumulatorismandatoryonallheatpumpsunlessotherwiseapprovedbytheEmersonClimateTechnologies,Inc.ApplicationEngineeringDepartment.

Systems utilizing hot gas defrost are also subject to liquid flooding either at the start or termination of the hot gas cycle. Compressors on low superheat ap-plications such as liquid chillers and low temperature display cases are susceptible to occasional flooding from improper refrigerant control. Truck applications experience extreme flooding conditions at start up after long non-operating periods.

On two stage compressors the suction vapor is returned directly to the low stage cylinders without passing through the motor chamber, andasuctionaccumulatorshouldbeusedtoprotectthecom-pressorvalvesfromliquidslugging.

Since each system will vary with respect to the total refrigerant charge and the method of refrigerant con-trol, the actual need for an accumulator and the size required is to a large extent dictated by the individual system requirement. If flooding can occur, an accumu-lator must be provided with sufficient capacity to hold the maximum amount of refrigerant flooding which can occur at any one time, and this can be well over 50% of the total system charge in some cases. If accurate test data as to the amount of liquid floodback is not available, then 50% of the system charge normally can be used as a conservative design guide.

5. OilSeparators

Oil separators cannot cure oil return problems caused by system design, nor can they remedy liquid refrig-erant control problems. However, in the event that system control problems cannot be remedied by other means, oil separators may be helpful in reducing the amount of oil circulated through the system, and can often make possible safe operation through critical periods until such time as system control can be re-turned to normal conditions. For example, on ultra low temperature applications or on flooded evaporators, oil return may be dependent on defrost periods, and the oil separator can help to maintain the oil level in the compressor during the period between defrosts.

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SECTION23ELECTRICALCONTROLCIRCUITS

Electrical control circuits may be quite simple or extremely complicated, depending on the control requirements of the particular system. Most wiring diagrams furnished with refrigeration equipment are of the pictorial type, and show the wiring as nearly as possible in the manner in which it is installed. Normally the different components are shown, together with terminal designations and wire colors. The pictorial diagram is essential as a guide to proper wiring.

Schematic wiring diagrams are useful in analyzing and explaining the performance of a control circuit, since the schematic diagram shows the various parts of the circuit in a functional manner only, thus reducing the diagram to its simplest form.

Both types of diagram may be used to describe the same control circuit.

TYPICALLOCKOUTCONTROLCIRCUIT

A typical wiring diagram of a compressor control circuit with part winding motor start, and a 10 minute lockout circuit in the event of a compressor protector trip is shown in Figure 106. The pictorial diagram is shown in the upper half of the illustration, while the schematic diagram is shown at the bottom.

In this circuit, which is designed for fully automatic op-eration, fast cycling of the compressor from the opera-tion of the motor protectors is eliminated by the use of a 10 minute time delay in conjunction with double pole impedance relay.

Basically an impedance relay is similar to a normal relay except that the coil has been wound so as to create a

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high resistance to current passage. If wired in parallel with a circuit having lower resistance, the high imped-ance (resistance) of the relay will shunt the current to the alternate circuit and the impedance relay will be inoperative. If the alternate circuit is opened and the current must pass through the impedance relay, the relay coil is energized and the relay operates. The voltage drop across the impedance relay is so large that other magnetic coils in series with the relay will not operate because of the resulting low voltage.

In the control circuits shown, the coil of the impedance relay is connected in parallel with the automatic pro-tectors, and in series with the holding coil of the motor starter or contactor. In the event of an overload on the compressor, the stator thermostats or the overload protectors open, and the control circuit current must pass through the impedance relay coil, energizing the impedance relay. Because of the high impedance of this coil, the voltage to the holding coil of the motor starter or contactor falls below the voltage required for opera-tion, and the contactor or starter opens, removing the compressor from the line. When the impedance relay is energized, a single pole double through 10 minute time delay relay is energized through a set of normally open contacts on the impedance relay. A set of normally closed contacts on the time delay relay break the control circuit and prevent a normal circuit being re-established through the automatic protectors while the time delay relay is operative. A set of normally open contacts on the time delay relay close when the relay is energized to maintain a circuit to the time delay coil. After a 10 minute internal, a cam on the time delay relay trips, automatically returning the circuit to normal operation. In the event the stator thermostats have not reset, or the overload condition again occurs, the circuit will continue repeating the 10 minute lockout cycle.

CONTROLCIRCUITFORCOMPRESSORPROTECTIONAGAINSTLIQUIDREFRIGERANTFLOODING

On systems with large refrigerant charges, compressor damage can occasionally be caused by liquid refrigerant flooding the compressor crankcase should the compres-sor be non-operative due to a trip of a safety device. This can occur even if the control circuit provides for a continuous pumpdown cycle.

Typically this can happen if the compressor trips either on the motor overload protectors or on the oil pressure safety control. The compressor would then be non-operative, but if the thermostat or other control device is calling for cooling, the liquid line solenoid valve will be open and liquid refrigerant will continue to feed into the evaporator, eventually flooding through to the com-

pressor. When the safety device is either manually or automatically reset and the compressor is restarted, the crankcase will be filled with liquid refrigerant.

On larger horsepower compressors, this can be a seri-ous problem, both because of the potential cost of the possible damage to the compressor and the amount of refrigerant involved. Flooding of the compressor under non-operative conditions can be prevented by the use of a reverse acting low pressure control as shown in Figure 107.

The liquid line solenoid valve is controlled by the ther-mostat, but to complete the circuit through the solenoid valve, the contacts in the reverse acting low pressure control must be closed. Since the dual pressure control completes the compressor contactor circuit when the suction pressure reaches 50 psig, the reverse acting control will remain closed during normal system opera-tion. However, should the compressor contactor circuit be broken by any of the safety devices so that the compressor could not start, the reverse acting control will open on a rise in pressure when the evaporator pressure rises above 90 psig. Opening of the reverse acting control de-energizes the liquid line solenoid valve, and stops the liquid refrigerant feed.

When the compressor is again restored to operation, the suction pressure is reduced, the reverse acting control again closes, and operation proceeds normally.

The pressure settings shown are tentative settings for an R-22 air conditioning system, and actual set-tings must be determined after reviewing the system’s normal operating range. The important factor is that the reverse acting low pressure control must be set to open well above the setting at which the dual pressure control closes.

CONTROLCIRCUITSTOPREVENTSHORTCYCLING

Short cycling often occurs on air conditioning and re-frigeration equipment due to a shortage of refrigerant, leaking solenoid valves, incorrect pressure control set-tings, thermostat chatter, or other causes. Short cycling causes overheating of the compressor and contactor, may cause nuisance tripping of the motor protectors, and in some cases has resulted in welded contactor points and motor failure.

Figure 108 shows a control circuit similar to the lockout circuit discussed previously, with the addition of a pump-down control circuit with a 45 second time delay to delay starting after closing of the dual pressure control. When the operational control is closed, the normally closed

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liquid line solenoid valve is energized. The resulting flow of liquid refrigerant into the low pressure side of the system increases the suction pressure, causing the contacts of the low pressure control to make, energizing the time delay relay.

After 45 seconds the time delay relay makes, complet-ing the main control circuit “XY” through the compres-sor contactor holding coil. In the event of an overload in the compressor circuit, the 10 minute lockout circuit functions as described previously. In the event the op-erational control or dual pressure control chatter or close immediately after opening, the time delay will prevent re-energizing the circuit for 45 seconds.

CONTROLCIRCUITSFORCOMPRESSORSWITHCAPACITYCONTROLVALVES

To avoid damage to the compressor from refrigerant migration, and to allow proper operation on pumpdown systems, it is essential that capacity control solenoid valves be de-energized when the compressor is not operating.

In control circuits operating at line voltage, the solenoid valve and control can be connected to the load side of the contactor as shown in Figure 109.

On large installations, the control circuit may have a power source independent of the compressor power supply. In such cases, the unloading solenoid valve and control may be connected in parallel with the compres-sor contactor coil as in Figure 110.

There are thousands of variations and types of control circuits, and the above examples are shown merely to illustrate typical circuits frequently encountered in refrig-eration work. The basic circuits shown can be adapted as necessary depending on the individual requirement.

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1675 W. Campbell Rd..Sidney, OH 45365

EmersonClimate.com


Part 5 - Installation and Service



This is the fifth and last of a series of publications comprising the Emerson Climate Tech-nologies, Inc. Refrigeration Manual.

Part 1 — Fundamentals of RefrigerationPart 2 — Refrigeration System ComponentsPart 3 — The Refrigeration LoadPart 4 — System DesignPart 5 — Installation and Service

The installation and service information is intended as a guide to good installation practice, and as an aid in analyzing system malfunctions. The section on service fundamentals is designed to serve as an introduction to various service procedures for beginning service men, students, salesmen, and others, needing a basic understanding of service tech-niques.


Part 5INStaLLatION aND SErVICE

Section 24. INStaLLatION

Recommended Installation Procedures......................................24-1Fundamentals of Evacuation and Dehydration ....................................24-6Brazing Connections on Welded Motor Compressors ........................24-11Installation of Suction and Discharge Line Vibration Absorbers .................24-12Typical Installation Specifications ..........24-13

Section 25. SErVICING COPELaND® BraND COMPrESSOrS

Nameplate Identification ........................25-1Identification of Port Locations in Heads of Copelametic® Motor-Compressors ........................25-5Identification of Motor Terminals on Single Phase Compressors ............25-5Proper Valve Plate and Head Gaskets for 3, 4, and 6 Cylinder Compressors...................................25-6Copeland® Brand Oil Pumps ................25-10Typical Copelametic® Compressor Construction ....................................25-20Maintenance Accessibility on Copelametic® Compressors ...........25-20Field Troubleshooting ............................25-23

Section 26. FUNDaMENtaLS OF SErVICE OPEratION

Contaminants ........................................26-1Handling of Refrigerant Containers .......26-1Safe Handling of Compressed Gases When Testing or Cleaning Refrigeration Systems.....................26-3Handling Copper Tubing ........................26-6Brazing Refrigerant Lines ......................26-6Service Valves .......................................26-8The Gauge Manifold ..............................26-9Purging Non-Condensables ..................26-10System Pumpdown................................26-11Refrigerant Leaks ..................................26-11Evacuation .............................................26-13Charging Refrigerant into a System ......26-14Removing Refrigerant from a System ...26-17Handling Refrigeration Oil .....................26-18Determining the Oil Level ......................26-18Adding Oil to a Compressor ..................26-19Removing Oil from a Compressor .........26-20Handling Filter-Driers.............................26-21Compressor Burnouts—What to Do ......26-21Compressor Failures That Could Have Been Prevented .....................26-24Preventive Maintenance ........................26-29

Section 27. USEFUL ENGINEErING Data


INDEX OF taBLES

Table 49 Boiling Point of Water at Varying Pressures ................................................................24-8Table 50 Comparison of Gauge and Absolute Pressures at Varying Altitudes ...........................24-8Table 51 Melting Points of Typical Commercial Brazing Compounds ........................................24-12Table 52 Service Diagnosis Chart ..............................................................................................25-29Table 53 Temperature Scales .....................................................................................................27-1Table 54 International Rating Conditions ...................................................................................27-1Table 55 Thermal Units ..............................................................................................................27-2Table 56 Fahrenheit—Centigrade Temperature Conversion Chart ............................................27-3Table 57 Properties of Saturated Steam ....................................................................................27-4Table 58 Decimal Equivalents, Areas, and Circumferences of Circles .......................................27-5Table 59 Conversion Table — Inches into Millimeters ................................................................27-6Table 60 Conversion Table — Decimals of an Inch into Millimeters ...........................................27-7Table 61 Conversion Table — Millimeters into Inches ................................................................27-7Table 62 Conversion Table — Hundredths of Millimeter into Inches ..........................................27-9Table 63 Metric Prefixes .............................................................................................................27-9Table 64 Length ..........................................................................................................................27-10Table 65 Area .............................................................................................................................27-10Table 66 Weight, Avoirdupois .....................................................................................................27-10Table 67 Volume, Dry .................................................................................................................27-11Table 68 Volume, Liquid .............................................................................................................27-11Table 69 Density .........................................................................................................................27-11 Table 70 Pressure ......................................................................................................................27-11Table 71 Velocity ........................................................................................................................27-12Table 72 Heat, Energy, Work ......................................................................................................27-12Table 73 Solid and Liquid Expendable Refrigerants ..................................................................27-12


SECtION 24INStaLLatION

rECOMMENDED INStaLLatION PrOCEDUrES

It is quite probable that a majority of operating failures on field installed systems can be traced to careless or inadequate installation procedures. The following instructions have been prepared to help the installation and/or service engineer sys-tematically cover the many points which must be considered to provide each installation with trouble free performance.

These instructions are general in nature, and have been primarily for field installed and connected sys-tems normally utilizing compressors 2 horsepower in size or larger. However, the procedures can be applied to almost any type of field installed system, utilizing only those procedures which apply to the specific installation.

Design and application

A location for the compressor should be selected which provides good ventilation, even when re-mote condensers are to be used, since the motor-compressor and discharge lines give off heat. Air cooled compressors must be provided with forced convection air cooling.

Air cooled condensers must be located to insure adequate air for condensing purposes. Care must be taken to avoid recirculation of air from one con-denser to another.

Water cooled units must be provided with an ad-equate supply of water to maintain desired condens-ing temperatures. In order to avoid concentration of impurities, fungus, and scaling in cooling towers and evaporative condensers, a continuous waste bleed to a drain of approximately 2 gallons per hour per horsepower must be provided so that a continuous addition of fresh make-up water will be required.

Units and compressors must be level to insure proper lubrication.

Refrigerant suction lines must be sized to maintain adequate velocities for proper oil return.

Handling and receiving of Equipment

Responsibility should be assigned to a dependable individual at the job site to receive material. Each shipment should be carefully checked against the bill of lading. The shipping receipt should not be signed until all items listed on the bill of lading have been accounted for.

Check carefully for concealed damage. Any short-age or damages should be reported to the delivering carrier. Damaged material becomes the delivering carrier’s responsibility, and should not be returned to the manufacturer unless prior approval is given to do so.

When uncrating, care should be taken to prevent damage. Heavy equipment should be left on its shipping base until it has been moved to the final location.

The packing list included with each shipment should be carefully checked to determine if all parts and equipment have been received. Any accessories such as starters, contactors or controls should be fastened to the basic unit to avoid loss and prevent possible interchanging with other units.

Installation, Electrical

The supply power, voltage, frequency, and phase must coincide with the compressor nameplate. All wiring should be carefully checked against the manufacturer’s diagrams. Field wiring must be connected in accordance with the National Electric Code, or other local codes that may apply.

Check to insure proper:

(a) Wire Sizes to handle the connected load.

(b) Fuses recommended for compressors. (See Emerson Climate Technologies, Inc. Electrical Handbook)

(c) Magnetic starters, contactors, and motor pro-tection devices approved by Emerson Climate Technologies, Inc.

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(d) Operation of oil pressure safety control.

(e) Direction of rotation and speed of fans and/or water pumps.

(f) Wiring with no grounded lines or controls.

Installation, refrigerant Piping

Take extreme care to keep refrigeration tubing clean and dry prior to installation. The following procedures should be followed:

(a) Do not leave dehydrated compressors filter-driers open to the atmosphere any longer than is absolutely necessary. (One or two minutes maximum suggested.)

(b) Use only refrigeration grade copper tubing, properly sealed against contamination. Water tubing often contains wax and other trouble-some contaminants.

(c) Permanent suction line filters and liquid line filter-driers are recommended in all field installed systems.

(d) Suction lines should slope ˚ inch per 10 feet towards the compressor.

(e) Suitable P-type oil traps should be located at the base of each suction riser to enhance oil return to the compressor.

(f) When brazing refrigerant lines, an inert gas

should be passed through the line at low pres-sure to prevent scaling and oxidation inside the tubing. Dry nitrogen is preferred.

(g) Use only a suitable silver solder alloy or 95/5 solder on suction and liquid lines, and a high temperature silver solder alloy only on discharge lines.

(h) In order to avoid damage to the internal joints in vibration eliminators, line connections to vibration eliminators should be made with a silver solder alloy such as Easy-Flo having a melting temperature of 900°F. to 1200°F.

(i) Limit the soldering paste or flux to the minimum required to prevent contamination of the s o l -der joint internally. Flux only the male portion of the connection, never the female. After brazing, remove surplus flux with a damp cloth.

(j) If vibration absorbers are to be installed in suction or discharge lines they must be ap-plied according to the manufacturer’s recommendations. With Copelametic® motor-compressors, the preferred position is parallel to the crankshaft, as close to the compressor as possible. Vibration eliminators may be in-stalled in a vertical position if joints are sealed against trapping of condensation which might damage the vibration absorber bellows due to freezing. Filling of the joints with soft solder as a means of sealing is recommended. Installa-tion of the vibration absorber in a horizontal plane at right angles to the crankshaft is not acceptable since the resulting stress from compressor movement may cause failure of the bellows or of the refrigerant line.

(k) Two evacuation valves are necessary. One should be in the suction line and one in the liquid line at or near the receiver.

(l) After all lines are connected, the entire system must be leak tested. The complete system should be pressurized to not more than 175 psig with refrigerant and dry nitrogen (or dry CO2). The use of an electronic type leak detector is highly recommended because of its greater sensitivity to small leaks. As a further check it is recommended that prior to charging, the system be evacuated to a pressure of 1 PSIA or less, and sealed for 12 hours. Any leakage of air into the system will cause the vacuum reading to decrease. If an air leak is indicated, the system should again be leak tested, and leaks repaired. For a satisfactory installation, the system must be leak tight.

(m) After the final leak test, refrigerant lines ex-posed to high ambient conditions should be insulated to reduce heat pick-up and prevent the formation of flash gas in the liquid lines. Suction lines should be insulated, if exposed, to prevent condensation.

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Installation, Plumbing

Good practice requires the following:

(a) Lines should be sloped adequately to drain by gravity any water accumulated from condens-ing, defrosting, or cleaning operations.

(b) All plumbing connections should be made in accordance with local plumbing codes.

(c) Condensate lines from refrigerated fixtures must be trapped and run to an open drain. They must not be connected directly to the sewer system.

IF tHE SYStEM IS WatEr-COOLED:

(d) Water pipe sizes should be adequate to provide the required flow at the lowest inlet pressure anticipated.

(e) Control devices such as solenoid valves, modulating valves, or hand valves that could cause hydraulic hammer should be protected by a stand-pipe and air pocket to absorb this shock. Electrical or pressure oper-ated water control valves should be installed between the water supply and the condenser inlet—never between the condenser and the drain. If water supply pressure is excessive, a pressure reducing valve must be used since the allowable working pressure of water valves is normally 150 psig. Pressures above this level can also cause damage to the condenser.

(f) The water pump must be checked for rotation and proper performance.

(g) Check for water leaks.

Evacuation

A good high vacuum pump should be connected to both the low and high side evacuation valves with copper tube or high vacuum hoses (¼” ID minimum). If the compressor has service valves, they should remain closed. A high vacuum gauge capable of registering pressure in microns should be attached to the system for pressure readings.

A shut off valve between the gauge connection and the vacuum pump should be provided to allow the system pressure to be checked after evacuation. Do not turn off vacuum pump when connected to an evacuated system before closing shut off valve.

The vacuum pump should be operated until a pressure of 1,500 microns absolute pressure is reached—at which time the vacuum should be broken with the refrigerant to be used in the sys-tem through a drier until the system pressure rises above “0” psig.

Repeat this operation a second time.

Open the compressor service valves (if supplied) and evacuate the entire system to 500 microns absolute pressure.

Raise the pressure to 2 psig with the refrigerant and remove the vacuum pump.

Under no conditions is the motor-compressor to be started or operated while the system is under a high vacuum. To do so may cause serious dam-age to the motor windings because of the reduced dielectric strength of the atmosphere within the motor chamber.

Check-Out and Start Up

After the installation has been completed, the fol-lowing points should be covered before the system is placed in operation.

(a) Check electrical connections. Be sure they are all tight.

(b) Observe compressor oil level before start-up. The oil level should be at or slightly above the center of the sight glass. Use only oil approved by Emerson Climate Technologies, Inc.

(c) Remove or loosen shipping retainers under motor-compressors. Make sure hold down nuts on spring mounted compressors are not touching the compressor feet.

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(d) Check high and low pressure controls, water valves, pressure regulating valves, oil pressure safety controls, and all other safety controls, and adjust if necessary.

(e) Check thermostat for normal operation.

(f) Suitable tags or other means should be provided to indicate refrigerant used in the system. Some Copeland® brand condensing unit nameplates have two detachable corner tabs. One should be removed so that the nameplate indicates the refrigerant used.

(g) Wiring diagrams, instruction bulletins, etc., at-tached to motor-compressors or condensing units should be read and filed for future refer-ence.

(h) Make the proper refrigerant connections and

charge the unit with the refrigerant to be used. Weigh the refrigerant drum before charging so an accurate record can be kept of the weight of refrigerant put in the system. If the refriger-ant must be added to the system through the suction side of the compressor, charge in vapor form only. Liquid charging must be done in the high side only.

(i) Observe system pressures during charging and initial operation. Do not add oil while the system is short of refrigerant, unless oil level is dangerously low.

(j) Continue charging until system has sufficient

refrigerant for proper operation. Do not over-charge. Remember that bubbles in a sight glass may be caused by a restriction as well as a shortage of refrigerant.

(k) Do not leave unit unattended until the system has reached normal operating conditions and the oil charge has been properly adjusted to maintain the oil level at the center of the sight glass.

Operational Check-Out

After the system has been charged and has op-erated for at least two hours at normal operating

conditions without any indication of malfunction, it should be allowed to operate over-night on auto-matic controls. Then a thorough recheck of the entire system operation should be made as follows:

(a) Check compressor head and suction pressures. If not within system design limits, determine why and take corrective action.

(b) Check liquid line sight glass and expansion valve operation. If there are indications that more refrigerant is required, leak test all con-nections and system components and repair any leaks before adding refrigerant.

(c) When applicable, observe oil level in com-pressor crankcase sight glass, and add oil as necessary to bring level to center of the sight glass.

(d) Thermostatic expansion valves must be checked for proper superheat settings. Feeler bulbs must be in positive contact with the suc-tion line. Valves with high superheat settings produce little refrigeration and poor oil return. Too little superheat causes low refrigeration capacity and promotes liquid slugging and compressor bearing washout. Liquid refrigerant must be prevented from reaching the crankcase. If proper control cannot be achieved with the system in normal operation, a suction accumu-lator must be installed in the suction line just ahead of the compressor to prevent liquid refrigerant from reaching the compressor.

(e) Using suitable instruments, carefully check line voltage and amperage at the compressor terminals. Voltage must be within 10% of that indicated on the compressor nameplate. If high or low voltage is indicated, notify the power company. The current normally should not exceed 120% of the nameplate rating. If amper-age draw is excessive, immediately determine the cause and take corrective action. On three phase motor-compressors, check to see that a balanced load is drawn by each phase.

(f) All fan motors on air cooled condensers, evaporators, etc. should be checked for proper

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rotation. Fan motor mounts should be carefully checked for tightness and proper alignment. If belt drives are used, check the belt tension. All motors requiring lubrication should be oiled or greased as necessary.

(g) Check defrost controls for initiation and termina-tion setting, and length of defrost period. Check crankcase heaters if used.

(h) Check winter head pressure controls for pres-sure setting.

(i) Check crankcase pressure regulating valves, if any, for proper setting.

(j) Adjust water valves on water cooled systems to maintain desired condensing temperatures. Check water pumps for proper rotation.

(k) Install instruction card and control system dia-gram for use of store manager or owner.

Identification

Each refrigerated fixture and cooler coil should be numbered starting at No. 1. These numbers should be not less than ̊ ” in height and should be stenciled or marked neatly on the fixture in an inconspicu-ous location easily available to the serviceman. The compressors or condensing units serving the fixtures should be marked with the numbers of the cases and coils served with figures not less than 1” in height.

Service record

A permanent data sheet should be prepared on each installation, with a copy for the owner and the original for the installing contractor’s files. If another firm is to handle service and maintenance, additional copies should be prepared as necessary.

The form of the data sheet may vary, but a complete record of sizes and identification of all components used in the installation, together with any pertinent information should be included. Following is a sug-gested check-off list:

(a) Compressor manufacturer, model, and serial number.

(b) Equipment manufacturer, model, and serial number.

(c) Design operating temperatures.

(d) Condensing unit model, and serial number. (If package condensing unit.)

(e) If remote condenser, type, manufacturer, model, fan data.

(f) Refrigerant and weight of charge.

(g) Electrical service, volts, cycles, phase, wire size.

(h) Control circuit, voltage, fuse size.

(i) Contactor or starter, manufacturer, model, size, part number.

(j) Compressor motor protection, type, size, part number.

(k) Data on capacitors, relays, or other electrical components.

(l) Pressure control, type, size, model number, setting.

(m) Oil pressure safety control, type, model num-ber.

(n) Defrost control, type, manufacturer, model number, setting.

(o) Data on miscellaneous refrigeration compo-nents such as pressure controls, winterizing controls, oil separators, crankcase heaters, solenoids, valves, etc.

(p) Liquid line drier, manufacturer, size, model number, connections.

(q) Schematic diagram of refrigerant piping.

(r) Final settings on all pressure, regulating, and safety controls.

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FUNDaMENtaLS OF EVaCUatION aND DEHYDratION

Although millions of dollars have been spent on refrigeration research, many of the reactions in-side air conditioning and refrigeration systems are still a mystery. We do know that the presence of moisture, heat, and oxygen under certain condi-tions can result in many forms of system damage. Corrosion, sludging, copper plating, oil breakdown, carbon formation, and eventual compressor failure can be caused by these contaminants.

The absence of any one of the three, or its reduction to an acceptable level can greatly extend compres-sor life and slow down harmful reactions. If all three can be controlled, then a sound foundation has been made for a trouble free installation.

Copeland® brand compressors are carefully tested to determine limits within which operation is possible without creating excessive heat in the compressor. But under the best operating conditions, heat is going to be produced as a natural consequence of compression of the refrigerant gas. Discharge temperatures in excess of 200°F. are unavoidable. Therefore, major efforts must be directed at prevent-ing moisture and air from entering the system.

Moisture In a refrigeration System

Moisture exists in three forms; as a solid when it is frozen into ice, as liquid water, and as a vapor or gas. It is extremely rare that moisture will enter a refrigeration system in the form of ice or water. It is the invisible water vapor that exists in the air around us that creates the real hazard.

The ability of air to hold water vapor increases with the temperature of the air. On a hot, humid summer day, the air may be actually loaded with moisture. Relative humidity is the term commonly used to express the percentage of saturation, that is, the existing moisture content of the air expressed as a percentage of the maximum moisture that the air could contain at a given temperature.

The relative humidity determines the dew point, or the temperature at which moisture will condense out of the air. Condensation occurs on the outside of a cold glass of water in a warm room, and it can

occur in exactly the same fashion inside a cold evaporator which has been opened and exposed to the atmosphere.

Despite the fact that water vapor exists as part of the air around us, it acts quite independently of the air. Vapor pressure is independent of air pres-sure, and its speed of movement is astonishing. This means that water vapor cannot be stopped by air movement.

Obviously it is impossible to prevent water vapor from entering the system anytime it is opened to the atmosphere. However, if the temperature of the exposed part of the system is above the dew point, or if the time of exposure is short, the amount of moisture actually entering the system will be small. If a new drier is installed in the liquid line each time the system is opened for maintenance, the drier will normally have sufficient capacity to lower the moisture in the system to a safe level.

However, at the time of original installation, or after exposure for long periods during maintenance, the amount of moisture in the system may be greater than a drier’s effective capacity. In such cases, evacuation is the only effective means of removing large quantities of moisture from the system, and to successfully dehydrate a system by evacuation, pressures within the system must be reduced to levels which will cause the trapped moisture to vaporize.

air In a refrigeration System

The air we breathe is primarily composed of nitrogen and oxygen. Both elements remain in a gaseous form at all temperatures and pressures encountered in commercial refrigeration and air conditioning systems. Therefore, although these gases can be liquefied under extremely low temperatures, they may be considered as non-condensable in a refrigeration system.

Scientists have discovered that one of the basic laws of nature is the fact that in a combination of gases, each gas exerts its own pressure indepen-dently of others, and the total pressure existing in a system is the total of all the gaseous pressures present. A second basic characteristic of a gas is that if the space in which it is enclosed remains

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constant, so that it cannot expand, its pressure will vary directly with the temperature. Therefore, if air is sealed in a system with refrigerant, the nitrogen and oxygen will each add their pressure to the system pressure, and this will increase as the temperature rises.

Since the air is non-condensable, it will usually trap in the top of the condenser and the receiver. Dur-ing operation the compressor discharge pressure will be a combination of the refrigerant condensing pressure plus the pressure exerted by the nitrogen and oxygen. The amount of pressure above normal condensing pressures that may result will depend on the amount of trapped air, but it can easily reach 40 to 50 psig or more. Any time a system is running with abnormally high head pressure, air in the system is a prime suspect.

Nitrogen is basically an inert gas and does not easily enter into chemical reactions. Oxygen, however, is just the reverse, and at the slightest opportunity will combine with other elements. Rust, corrosion, and burning are all common oxidation processes.

In the refrigeration system, oxygen and moisture quickly join in a common attack on the refrigerant and oil, and can cause corrosion, copper plating, acid formation, sludging, and other harmful reac-tions. Tests have shown that in the presence of heat, the combination of air and moisture is far more apt to cause breakdown of the refrigerant and oil mixture than greatly increased amounts of moisture alone.

Pressure - temperature - Evaporating relationships

Anyone familiar with refrigeration knows that re-frigerants follow a definite fixed pressure-tempera-ture relationship, and that at a given pressure the refrigerant will boil or vaporize at a corresponding saturation temperature. Water follows exactly the same pattern, and this is the basis for dehydration by evacuation.

The pressure which determines the boiling points of refrigerants and water is absolute pressure, nor-mally expressed in terms of psia, which is defined as the pressure existing above a perfect vacuum.

The atmosphere surrounding the Earth is composed of gases, primarily oxygen and nitrogen, extending many miles above the surface of the Earth. The weight of that atmosphere pressing down on the Earth creates the atmospheric pressure we live in. At a given point, the atmospheric pressure is relatively constant except for minor changes due to changing weather conditions. For purposes of standardization and as a basic reference for comparison, the atmospheric pressure at sea level has been universally accepted, and this has been established at 14.7 pounds per square inch, which is equivalent to the pressure exerted by a column of mercury 29.92 inches high.

At very low pressures, it is necessary to use a smaller unit of measurement since even inches of mercury are too large for accurate reading. The micron, a metric unit of length, is commonly used for this purpose, and when we speak of microns in evacuation, we are referring to absolute pressure in units of microns of mercury. Relationships of the various units of measurement are as follows:

1 pound per sq. in. = 2.03 inches mercury1 inch mercury = .491 pounds per sq. in.1 inch mercury = 25,400 microns mercury1 inch = 25,400 microns1 millimeter = 1,000 microns1 micron = .001 millimeter

The refrigeration serviceman’s bourdon tube gauge reads 0 pounds per square inch when not con-nected to a pressure producing source. Therefore the standard relationship has been established that absolute pressure is equal to gauge pressure plus 14.7 psi. Pressures below 0 psig are actually negative readings on the gauge, and are referred to as inches of vacuum. The gauge is calibrated in the equivalent of inches of mercury.

Factors affecting Vacuum Pump Performance

A vacuum pump suitable for refrigeration work must not only be capable of pulling a high vacuum, but must be capable of maintaining that vacuum on the system for prolonged periods. As moist air is pumped through the vacuum pump, the moisture will seek to condense in the vacuum pump oil sump, and once the oil is saturated, water vapor

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The above table clearly illustrates the reduction of the boiling point of water with a reduction of pressure. It is clear that at normal room temperatures, dehydration by evacuation requires pressures below 0.40 psia, which means a corresponding vacuum reading at sea level of 29.2 inches of mercury. At pressures above that, boiling simply would not take place. From a practical standpoint, much lower pressures are necessary to create a temperature difference to the boiling water so that heat transfer can take place, and also to offset pressure drop in the connecting lines, which is extremely critical at very low pressures. Pres-sures from 1,500 to 2,000 microns are required for effective dehydration, and equipment to accomplish this is normally described as being designed for high vacuum work. Heat should be applied to systems which are known to contain free water to aid in evacuation.

It is important to remember that gauge pressures are only relative to absolute pressure. The table shows relationships existing at various elevations assuming that standard atmospheric conditions prevail. Obvi-ously, a given gauge pressure at varying elevations may actually reflect a wide variation in actual absolute pressures.


escaping from the oil may prevent the pump from achieving a high vacuum. Unless the pump is specifically designed to prevent this condition, the oil may become saturated before one evacuation job is completed.

In order to prevent condensation, some vacuum pumps have a vented exhaust or gas ballast feature. Basically this involves allowing a small bleed of atmospheric air to enter the second stage of a two stage pump, or the discharge chamber of a single stage pump prior to the discharge stroke to prevent condensation of water during compression.

Since reciprocating pumps lose efficiency at vacuums greater than 27 inches of mercury, rotary pumps are primarily used for high vacuum work. Single stage vacuum pumps are available which are capable of pulling a very high vacuum, but in general they are vulnerable to oil contamination, and if the exhaust is vented to protect the oil, then the pump’s efficiency is reduced. Although single stage pumps may be quite satisfactory for small systems, for best high vacuum performance in refrigeration usage a two stage vacuum pump with gas ballast on the second stage is recommended.

Even at extremely low pressures, it is essential that the system to be evacuated is at a tempera-ture high enough to insure boiling of any water to be removed. With pressures of 2000 microns and below, normal room temperatures of 70°F. to 80°F. are adequate. Evacuation of temperatures below 50°F. is not recommended.

If a great deal of moisture must be removed from a system by the vacuum pump, the oil may become saturated with moisture despite the gas ballast feature or the best pump design. Once this has oc-curred, the only solution is to change the oil in the vacuum pump. Even with the best vacuum pump, frequent oil changes are necessary to maintain ef-ficiency. It is recommended that the oil be changed before each major evacuation.

If there is any possibility that large amounts of water may be trapped in a system, the lines should be blown out with dry nitrogen prior to attaching the vacuum pump. This will not only aid in prolonging the life of the pump, it will materially decrease the time required to evacuate the system. If it is known

that a system is saturated with water, for example after the rupture of tubes in a water cooled con-denser, a special low temperature moisture trap should be installed in the suction line ahead of the vacuum pump intake. Suitable traps are available from vacuum pump manufacturers.

One factor that is not fully appreciated by most servicemen is the critical nature of the pressure drop that occurs due to restrictions in the line dur-ing evacuation. For field evacuation with portable vacuum pumps, lines connecting the vacuum pump to the system should be a minimum of ¼ in. I.D. on small systems, and on larger systems at least ½ in. I.D. copper tubing should be used. Evacuat-ing valves are recommended for every system. These should be installed in both the suction and liquid lines, and should be at least as large as the connecting lines. The typical serviceman’s manifold and charging hose will cause sufficient restriction to prevent a high vacuum being reached, and compressor service valves are also unsatisfac-tory for high vacuum work. If restrictions exist in the connecting lines, gauges at the vacuum pump will reflect pump pressure, but will not give a true picture of pressures in the system.

The speed with which a system may be evacuated depends on both the displacement of the vacuum pump and the size of the connecting lines and fit-tings. A good high vacuum pump has a very high pumping efficiency down to absolute pressures of 1,000 microns and below, possibly as high as 85% to 90% or more. This means that a vacuum pump with 1 CFM displacement may still be capable of pumping up to .9 CFM with a suction pressure of 1,000 microns and discharging to atmosphere.

However, a vacuum pump’s performance can be greatly reduced by the size of connecting lines and fittings. In the low or medium vacuum range, this may not greatly affect a pump’s efficiency, but at pressures below 5,000 microns the pump’s net capacity can decrease rapidly. The following comparison is based on one pump manufacturer’s catalog information on pumping speed of rotary vacuum pumps.

It is interesting to note that more efficiency can be gained by increasing the connecting line size on a 1 CFM pump from ¼ in. I.D. to a larger size than

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can be gained by putting a 5 CFM pump on the same ¼ in. connection.

Calculations to determine pull down time are quite complicated, since the pump’s efficiency changes with the reduced pressure, and the size and length of the connecting lines may greatly affect the per-formance of a given pump. The following estimate of pull down time is based on one manufacturer’s catalog data, but because of the assumptions that must be made in the calculation, the figures are at best an approximation.

The above table provides a good comparison of relative pump performance. It is quite clear that if a connecting line no larger than ¼ in. I.D. is to be used, there is little to be gained by going to a larger vacuum pump. For large systems it is obvious that both a good sized vacuum pump and a large con-necting line are necessary if the required time is to be held to a minimum. The pull down time will vary directly with the internal volume of a given system, so for smaller systems the 1 CFM pump may be

perfectly satisfactory.

Measurement of Vacuum

As indicated earlier, the refrigeration serviceman’s gauge reads pressure only in relation to absolute pressure, and a given gauge reading may cover a wide range of actual pressures. For this reason, and also because the ordinary bourdon tube compound gauge is not designed for the extreme accuracy required in evacuation work, a special vacuum gauge is required for high vacuum readings.

For accurate pressure readings in the micron range for refrigeration use, a thermocouple vacuum gauge is recommended. This type of gauge is relatively inexpensive, easy to operate, rugged enough for field use, and requires little or no maintenance. The advantage of this gauge where moisture may be encountered in a system is that it measures not only the pressure due to residual gases, but also the pressure contributed by any water vapor remaining in the system. The McLeod type gauge is widely used in laboratory work, and is highly ac-curate for readings where moisture is not a factor, but it is not recommended for use in refrigeration work since it will not measure the pressure due to water vapor.

triple Evacuation

In order to insure a complete evacuation, Emerson Climate Technologies, Inc. recommends a triple evacuation, twice to 1,500 microns and the final time to 500 microns. The vacuum should be broken to 2 psig each time with the same type of refrigerant to be used in the system.

It is quite possible that the original evacuation, if not continued for a sufficient period of time may not completely remove all of the air and moisture from the system. Breaking the initial vacuum with dry refrigerant allows the fresh refrigerant to absorb and mix with any residual moisture and air, and the succeeding evacuation will remove a major portion of any remaining contaminants. If for example, each evacuation removed only 98% of the contents of the system, and any remaining contaminants mixed thoroughly with the refrigerant used to break the vacuum, after the triple evacuation the remaining contaminant percentage would be 2%

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EStIMatED tIME rEqUIrED FOr SYStEM PULL DOWN

based on 5 cubic feet internal volume


x 2% x 2% or .0008%. The residual contaminants have been reduced to such a low level they no longer are a danger to the system. This illustrates why triple evacuation is increasingly important if the vacuum pump is not of the highest efficiency, or if the evacuation time is not adequate to insure complete evacuation.

Many manufacturers use process pressures of 50 to 100 microns. However, in field evacuation, pressures in this range are very difficult to reach, particularly if refrigerant has been allowed to mix with oil in the system. The refrigerant will escape from the oil very slowly, and the time required to reach such low pressures might be quite unreason-able. The triple evacuation method to a pressure of 500 microns is practical under field conditions, and represents a specification that can be met.

For manufacturers having process equipment, the use of dry air with a dew point below -60°F. in place of refrigerant for dehydration in connection with a triple evacuation to the pressures described above is also highly recommended.

To evacuate a system properly requires time and care. Any slight carelessness in protecting the sealed system can undo all the precautions taken previously. But the slight extra effort required to make an evacuation properly and completely will pay big dividends in reduced maintenance and trouble free operation.

BraZING CONNECtIONS ON WELDED MOtOr-COMPrESSOrS

Suction and discharge line connections to welded motor-compressors are normally made by brazing the refrigerant lines directly into stub tubes on the compressor with a silver brazing alloy. Occasionally the joint between the stub tube and the steel shell is damaged by overheating during factory or field installation when the refrigerant line connections are made. This type of damage can be avoided by proper care during the brazing operation.

The connection between the stub tube and the shell is made with a 35% silver brazing alloy which has a melting range of 1125°F. to 1295°F. The temperature of this joint must be kept below this range during the line brazing operation to avoid damage.

Figure 111 illustrates a typical suction line connec-tion. The torch flame should be used primarily on the refrigerant line, with only enough heat applied to the stub tube to make the connection properly. Heat will be conducted into the joint area from the refrigerant line. The torch fame should have a greenish “feather” extending from the tip of the inner blue cone as illustrated in Figure 112. Heat should be applied to both sides of the tube, and the flame should be moved continuously in a circular motion to distribute the heat, and prevent overheating of the tubing. Compressors with damaged joints usu-ally show evidence of the torch flame having been allowed to burn directly on the compressor shell and the stub tube-shell joint.

Emerson Climate Technologies, Inc. recommends that a low melting point alloy such as Easy-Flo or Easy-Flo 45 be used in making the line joint rather than a higher melting point alloy such as Sil-Fos. The heat necessary to make a Sil-Fos joint is somewhat greater than required for Easy-Flo, making it more difficult to avoid overheating. Another advantage of a lower temperature brazing alloy is the reduced annealing effect which takes the place, thus result-ing in a stronger joint.

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To assure a sound, leak tight tubing connection without overheating, the surfaces must be properly cleaned and a suitable flux must be used. A low temperature brazing flux that is fully liquid and ac-tive below the flow point of the silver brazing alloy is required. Only the male connection should be fluxed, and only enough flux should be used to adequately cover the surface. Excess flux allowed to enter the system can cause starting failures on PSC motors, plug filters and valves, and may cause other complications due to chemical reactions.

INStaLLatION OF SUCtION aND DISCHarGE LINE VIBratION aBSOrBErS

In order to prevent the transmission of noise and vibration from the compressor through the refrigera-tion piping, vibration eliminators are often required in the suction and discharge lines. On small units where small diameter soft copper tubing is used for the refrigerant lines, a coil of tubing may provide adequate protection against vibration. On larger

units, flexible metallic hose is frequently used.

Metallic vibration absorbers should be selected to have the same or greater internal diameter than the connecting piping. Because of the convolu-tions of the inner wall of the absorber, excessive refrigerant gas velocity can cause whistling and noise problems.

Unless properly installed, stress resulting from line movement may cause failure of the vibration absorber, and possibly can lead to line breakage. Because of its construction, a metallic vibration absorber can easily adjust to movement in a radial direction, but it must not be subjected to stress in either compression or extension. Some manufac-turers recommend using two vibration absorbers at right angles, but normally this is not necessary on Copeland® brand compressors.

Emerson Climate Technologies, Inc. recommends installation parallel to the crankshaft, as close to the compressor as possible. The starting torque of the motor will tend to rock the compressor from side to side when starting, and mounting parallel to the crankshaft will allow the absorber to easily adjust to the movement.

Vibration absorbers may be installed in a vertical position if the joints are sealed against trapping of condensation which might damage the bellows due to freezing. Filling of the joints with soft solder as a means of sealing is recommended. Flexible metal hoses are available with a neoprene jacket which protects the absorber against any possible damage from condensation or moisture.

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on supermarket or other large commercial refrig-eration installations, and is readily adaptable to different types of applications.

Installation of an angle 45° from the vertical and parallel to the crankshaft is acceptable, although horizontal or vertical installation is preferred. A 45° angle installation at right angles to the compressor crankshaft can actually act as a brace, causing compression stress, and is not acceptable.

Installation in the horizontal plane at right angles to the crankshaft is not acceptable, since compressor movement would tend to either compress or extend the absorber, and early failure of the absorber or connecting fittings could result.

The line connected to the end of the absorber op-posite the source of vibration should be firmly an-chored to a solid member. No movement will then be transmitted into the refrigerant lines beyond. Where a vertical or 45° mounting is used, the piping must be arranged so that sufficient allowance for movement is made. As a convenient means of checking the installation, a spring mounted compressor should be free to bottom solidly on the mounting pad or mounting snubber without stressing the absorber. The refrigerant lines should be in proper alignment prior to installation of the absorber, and sufficient space should be allowed so that it can be installed without being either stressed or compressed.

Internal joints of metallic vibration absorbers are often made up with a brazing compound which has a melting point of approximately 1,300°F. In order to avoid damage to the internal joints, line connections should be made with a silver solder alloy having a melting temperature below 1,200°F.

tYPICaL INStaLLatION SPECIFICatIONS

On large field installed refrigeration and air con-ditioning systems, it is advisable to have a written specification covering the work to be done and the responsibilities of each party. The specification is an aid in assuring a clear understanding of the contractor’s responsibility prior to the start of the job, so that disputes and disagreements may be eliminated.

Specifications may vary from a short paragraph covering the scope of the work to a detailed de-scription of the work to be done. The following specification is typical of the type frequently used

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Typical Specification

Large Commercial refrigeration and air Conditioning Installation

1. Definition of Terms

1.1 “Contractor” shall mean the refrigeration installation contractor.

1.2 “Owner” is .

1.3 “Manufacturer” shall mean the company or companies which will supply various equipment such as fixtures, compressors, coils, etc.

1.4 “Refrigeration Installation” shall mean the necessary labor and all parts and accessories neces-sary to complete the work outline in this specification.

2. Scope of Work

2.1 These specifications are intended to cover the installation of compressors, condensers, coils, condensing units, fixtures, and all other fittings, devices, and accessories required to complete the refrigeration systems as shown or called for on the refrigeration plans and schedules. The omission from these specifications or from the refrigeration plans and schedules of express ref-erence to any parts necessary for the complete installation is not to be construed as releasing the contractor from responsibility for furnishing such parts.

2.2 For details of installation refer to the fixture plan, refrigeration schedule, floor plan, plumbing plan, electric plan, air conditioning, heating, and ventilating plan, manufacturer’s installation instructions, and to applicable codes and ordinances.

2.3 The Contractor shall furnish and install any necessary refrigerant piping, fittings, vibration elimina-tors, line valves, solenoid valves, crankcase pressure regulating valves, thermostatic expansion valves, dehydrators, strainers, sight glasses, moisture indicators, refrigerant, oil, filters, insulation and all fittings and accessories necessary to make a complete installation unless otherwise speci-fied, together with all labor required to complete the installation and perform the service covered by this specification. The Contractor is responsible for unloading, assembling, and installing all fixtures, coolers, coils, compressors, condensing units, air conditioners, condensers, and other refrigeration equipment unless otherwise specified. The Contractor shall also arrange for the removal of crating and packing materials, and shall leave the uncrating area and the compressor room clean and neat.

2.4 The Contractor shall familiarize himself with the project, and shall cooperate with other contractors doing work on the building. If any conflict, interference, or discrepancies come to the attention of the contractor, he shall notify the owner immediately before proceeding any further with the installation.

2.5 No additional payment over and above the contract price will be made unless the Contractor receives a written order by the Owner or his representative for the addition.


2.6 Equipment and services shall be furnished as follows:

to Be Furnished By Owner Contractor Others Refrigerated fixtures Coils for coolers Air conditioning units Air conditioning temperature controls Air cooled condensers Compressors Condensing units Refrigeration system controls Coolers & freezers (walk in) Coolers & freezers (reach in) Ventilation and exhaust fans and controls Cooling tower and controls Plumbing Sheet metal, duct work, dampers, etc. Motor starters and protectors Electrical wiring, disconnect switches and connections

3. Fees, Permits, Licenses, and Insurance

3.1 All necessary permits and licenses incident to the work and required by local ordinance shall be secured and paid for by the contractor. All equipment shall be installed in strict compliance with all local building codes and ordinances.

3.2 The Contractor shall not commence work under this contract until he has obtained all the insur-ance required hereunder, and has filed certificates to that effect with the Owner. The Contractor shall indemnify and hold harmless the Owner for any and all claims, suits, losses, damages, or expenses on account of bodily injury, sickness, disease, death, and property damage as a result of the Contractor’s operations, acts, omissions, neglect or misconduct in connection with this project. Insurance coverage shall include but is not limited to

(a) Contractor’s Public Liability Insurance(b) Contractor’s Contingent Liability Insurance(c) Property Damage Insurance(d) Automotive Public Liability Insurance(e) Automotive Property Damage Insurance

4. Refrigerant Piping Materials

4.1 Unless otherwise specified, all refrigeration piping shall be refrigeration grade Type L or Type K hard drawn degreased sealed copper tubing. Alternate proposals may be submitted for the use of Type L refrigeration grade soft copper tubing for long underfloor runs only providing runs are straight and free from kinks and bends.

4.2 Extreme care shall be taken to keep all refrigerant piping clean and dry. It shall be kept sealed except when cutting or fabricating. Each length shall be inspected and swabbed with a cloth soaked in refrigeration oil if any dirt, filings, or visible moisture are present.

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4.3 All sweat-type fittings shall be wrought copper or forged brass. All elbows and return bends shall be of the long radius type. If flare fittings are required, they shall be of the frost proof type, (except on connections not subject to condensation), and constructed of forged brass. Soldered joints are preferred and shall be used wherever practical.

5. Refrigerant Piping Installation

5.1 Tubing shall be installed in a neat, workmanlike manner with horizontal runs sloped toward the compressor at a rate of 1” per 20’. All lines shall be supported at intervals of not more than 8’ and suitably anchored. Rubber grommets shall be used between tubing and clamps to prevent line chafing.

5.2 Where vertical risers of more than 5 feet occur in a suction line, the riser shall be trapped at the bottom.

5.3 Where a branch suction line enters a main suction line it shall enter at the top. Piping shall be arranged so refrigerant or oil cannot drain from the suction line into the coil.

5.4 Individual fixture or unit suction and liquid lines shall be of the size recommended by the Manu-facturer as shown in the applicable installation and service instructions. Liquid and hot gas refrigerant lines shall be sized in accordance with good industry practice to avoid excessive pressure drops. Branch and main suction lines shall be sized to maintain adequate velocities to properly return oil to the compressor under minimum load conditions at the lowest saturated suction pressure to be expected.

5.5 All joints in the compressor discharge line shall be brazed with a suitable high temperature silver solder alloy containing not less than 15% silver. Use only a suitable silver solder alloy on all copper to copper connections in the suction line and liquid line. At any copper to brass joint where damage could occur from excess heat use 95/5 solder. Use a solder with at least 35% silver content on all copper to steel, brass to steel, or steel to steel joints. During the brazing operation, dry nitrogen must be bled through the piping at very low pressure to prevent oxidation and scaling.

5.6 In order to avoid damage to the internal Silfos joints in vibration eliminators, line connections to vibration eliminators are to be made with a silver solder alloy such as Easy-Flo having a melting temperature of 900°F. to 1,200°F. (well below the 1,300°F. melting point of Silfos).

5.7 To prevent contamination of the line internally, limit the soldering paste or flux to the minimum required. Flux only the male portion of the connection, never the female.

5.8 Suction lines from low temperature cases shall be insulated where run below the floor level. All exposed suction lines, both low and medium temperature, shall be insulated as necessary to prevent condensation.

5.9 Insulation shall be of the cellular type such as Armstrong “Armaflex” or equal, shall fit the tubing snugly, and shall be applied and sealed in accordance with the Manufacturer’s instructions.

5.10 The refrigerant piping shall be adequately protected. Permanent guards shall be installed as required to protect the piping and fittings from damage. Metal pipe sleeves shall be provided where tubing passes through a concrete wall or floor, and the space around the tubing shall be filled with a mastic insulating compound.

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5.11 Arrange the piping so that normal inspection and servicing of the compressor and other equip-ment is not hindered. Do not obstruct the view of the crankcase oil sight glass, or run piping so that it interferes with removal of the compressor or other components.

5.12 Tubing installed in trenches or conduit under the floor must be level to prevent oil trapping. Guard against deformation or damage from trucks carrying heavy loads, or cement being poured.

6. Installation of Accessories

6.1 Vibration eliminators shall be installed in the suction and discharge lines of all compressors with spring or flexible mounting. The vibration eliminator must be applied according to the Man-ufacturer’s recommendations. For Copelametic® compressors, the vibration eliminator should be mounted parallel to the crankshaft, as close to the compressor as possible. Installation in a horizontal plane at right angles to the crankshaft is not acceptable, since the resulting stress from compressor movement may cause failure of the vibration absorber. If installed in a vertical position, the eliminator joints must be sealed against dripping from condensation to protect from freezing.

6.2 A solder type combination liquid sight glass and moisture indicator shall be installed in each system and located for easy visibility.

6.3 If liquid line driers are not otherwise specified, they shall be of the filter-drier type, and of the size recommended by the Manufacturer. Drier cartridges shall not be installed until the second evacuation has been completed.

6.4 Two evacuation fittings are necessary. One should be in the suction line at the inlet side of the suction line filter, and one should be in the liquid line at the outlet side of the filter-drier. If prop-erly valved, the connection in the liquid line may serve as a charging valve. After evacuation and charging, the fittings are to be capped or removed. Connections should be at least 3/8” and preferably ˚” in size.

6.5 A permanent suction line filter shall be installed in each compressor suction line. A pressure fitting must be provided ahead of the filter, preferably in the shell, to facilitate checking the pressure drop. If the pressure drop across the filter is in excess of 1 psig after the initial 24 hours of opera-tion, the suction line filter cartridge shall be replaced, or if the filter is of the sealed permanent type, the filter shall be replaced.

7. Drain Connections

7.1 Unless otherwise specified, condensate drains from coils and cases to the floor drain will be the responsibility of the Contractor. No drain line shall be smaller than the coil drain pan connection. All drain lines shall be hard copper tubing except for those in reach-in coolers. Lines should be sloped adequately to drain by gravity any water accumulated from condensing, defrosting, or cleaning operations. All condensate lines from refrigerated fixtures must be trapped and run to an open drain. They must not be connected directly to the sewer system. If necessary for clean-ing, threaded unions shall be provided in the most accessible location near the fixture.

8. Testing, Evacuation, and Charging

8.1 The Contractor shall notify the Owner 24 hours in advance of any test so that the Owner and/or Manufacturer’s representative may be present for the test if desired.

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8.2 When the refrigeration connections have been completed, the system shall be tested at a mini-mum of 150 psig with the compressor suction and discharge valves closed, and all other valves in the system open. (If local codes require higher test pressures, such codes must be complied with). Leak testing shall be performed with an electronic leak detector, unless the use of a halide torch is specifically authorized by the Owner. Refrigeration piping will not be acceptable unless it is gas tight. If any leaks are found, isolate the defective area, discharge the gas and repair the leaks, and then repeat the test. When testing has been completed, release all pressure freely to the atmosphere.

8.3 The system shall be evacuated with a vacuum pump specifically manufactured for vacuum duty, having a capability of pulling a vacuum of 50 microns or less. Evacuation of the system must never be done by the use of the refrigeration compressor. The pump should be connected to both the low and high side evacuation valves with copper tube or high vacuum hoses. (1/4” I.D. minimum). The compressor service valves should remain closed. A high vacuum gauge capable of registering pressure in microns should be attached to the system for pressure readings. Hermetic or accessible-hermetic motor compressors must not be operated during evacuation because of the reduced dielectric strength of the atmosphere within the motor chamber. To check system pressure, a hand valve must be provided between the pressure gauge and the vacuum pump which can be closed to isolate the system and check the pressure.

8.4 Evacuate each system to an absolute pressure not exceeding 1,500 microns. Install a drier of the required size in the liquid line, open the compressor suction and discharge valves, and evacuate to an absolute pressure not exceeding 500 microns. Leave the vacuum pump running for not less than two hours without interruption. Raise the system pressure to 2 psig with refrigerant, and remove the vacuum pump.

8.5 Refrigerant shall be charged directly from the original drums through a combination filter-drier. Each drier may be used for a maximum of three cylinders of refrigerant, and then must be replaced with a fresh drier. Charge the system by means of a charging fitting in the liquid line. Weigh the refrigerant drum before charging so that an accurate record can be kept of the weight of refrigerant put in the system. If refrigerant is added to the system through the suction side of the compressor, charge in vapor form only.

9. Start-Up

9.1 Compressors and condensing units will normally be delivered to the job with sufficient oil for the average installation. Check all compressors for proper oil level, and if necessary add suf-ficient oil to bring the level to the center of the crankcase sight glass. Use only the refrigeration oil recommended by the compressor manufacturer. All oil must be delivered to the job in factory sealed, unopened containers.

9.2 Before operating any motor or other moving parts, they are to be lubricated with the proper oil or grease as necessary.

9.3 Remove or loosen shipping retainers under motor compressors. Make sure hold down nuts on

spring mounted compressors are not touching the compressor feet, and are not more than 1/16” above the mounting foot.

9.4 Check high and low pressure control cut-in and cut-out points. Check water valve settings. Adjust if necessary.

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9.5 After the compressor is started, continue charging until system has sufficient refrigerant for proper operation. Do not overcharge. During start-up, no compressor is to be left operating unattended and unwatched until the system is properly charged with refrigerant and oil.

9.6 Do not add refrigeration oil while the system is short of refrigerant unless oil level is dangerously low. If oil has been added during charging, carefully check the compressor crankcase sight glass after reaching a normal operating condition to be sure the system does not contain an excessive amount of oil which can cause slugging or loss of refrigerating capacity.

9.7 The temperature controls shall be set to maintain the following temperatures in the center of the fixture before stocking:

FIXTURE TEMPERATURE °F.

(Minimum) (Maximum) Meat walk-in cooler 31 33 Meat holding cooler 29 31 Self-Service meat counter 31 33

Dairy walk-in cooler 36 38 Self-Service dairy case 36 38 Produce walk-in cooler 38 40 Self-Service produce counter 38 40 Self-Service beverage case 38 40 Frozen food storage cooler -15 -10 Self-Service frozen food case -5 0 Self-Service ice cream case -15 -10 Meat preparation room 54 56

10. Operation and Check-Out

10.1 The Contractor shall be responsible for the proper adjustment of all controls in the system, in-cluding the controls on each refrigeration circuit, air temperature controls in the machine room, remote condenser or water tower controls, water regulating valves, or such other controls as may be required.

10.2 The Contractor shall check the compressor overload protectors with the manufacturer’s speci-fications, and inform the Owner if they are incorrect.

10.3 The Contractor shall furnish a competent refrigeration service mechanic to check and make any necessary adjustments to the controls during the time the fixtures are being stocked. The mechanic shall remain at the store for at least 8 hours during the first day the store is open for business beginning 1 hour before opening time.

11. Identification and User Instruction

11.1 Each refrigerated fixture and cooler coil should be numbered starting at No. 1. These numbers shall be not less than 1” in height and shall be stenciled or marked neatly on the fixture in an inconspicuous location easily available to the serviceman. The compressors or condensing units serving the fixtures should be marked with the numbers of the cases and coils served with figures not less than 1 ½” in height.

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11.2 All switches, starters, and controls shall be identified as to the fixture or condensing unit they serve.

11.3 The Contractor shall turn over to the Owner one copy of all manufacturer’s literature furnished with each piece of equipment. Within 30 days after the store is opened, the Contractor shall instruct the store management on the proper operation, care and upkeep of all equipment.

11.4 A permanent data sheet shall be prepared on each installation with two copies for the Owner and the original for the installing Contractor’s files. The data sheet shall contain a complete record of sizes and identification of all components used in the installation together with any pertinent information. The data sheet should include but is not limited to the following:

A. Compressor manufacturer, model, and serial number.B. Fixture manufacturer, model, and serial number.C. Design operating temperatures.D. Condensing unit model, and serial number. (If package condensing unit)E. If remote condenser, type, manufacturer, model, fan data.F. Refrigerant and weight of charge.G. Electrical service, volts, phase, cycles, wire size.H. Control circuit, voltage, fuse size.I. Contactor or starter, manufacturer, model, size, part number.J. Compressor motor protection, type, size, part number.K. Data on capacitors, relays, or other electrical components.L. Pressure control, type, size, model number, setting.M. Oil pressure safety control, type, model number.N. Defrost control, type, manufacturer, model number, setting.O. Data on miscellaneous refrigeration system components such as pressure controls, winterizing

controls, oil separators, crankcase heaters, solenoid valves, valves, etc.P. Liquid line drier, manufacturer, size, model number, connections.Q. Schematic diagram of refrigerant piping.

12. Warranty and Guarantees

12.1 All equipment and material supplied and installed by the Contractor shall be guaranteed for one year from the date of the store opening. The Contractor shall provide the necessary labor, materi-als, and incidental expenses to maintain the equipment in proper operation for a period of one year from the date the store opens for business, without additional cost to the owner. (Tempera-ture rises caused by improper stocking or abnormal air currents shall not be the responsibility of the Contractor). The service shall not include repairs or replacements due to damage by fire, earthquake, tornado, the elements or act of God, or damage caused by misuse of the system by the Owner, power failures, broken glass, or lightning.

12.2 Official acceptance of the completed job shall be when the job is complete in every detail and has been run under load conditions with satisfactory performance for a period of at least one week.

12.3 In the event any equipment furnished by the Owner is found to be defective, the Owner will com-pensate the Contractor for the labor and material used in replacing the equipment or repairing the defects.

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12.4 The first year service shall include at least three complete lubrications at approximately 4 month intervals. At the time the equipment is lubricated, each system shall be checked for proper ad-justment, and any necessary repairs or corrections shall be made.

12.5 Approximately 30 days prior to the expiration of the one year warranty period, the Contractor shall make a final inspection, checking each system for proper adjustment, and correcting any deficiencies, and shall write the Owner a letter certifying that each system is free of leaks and is operating at the specified temperature.

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SECtION 25SErVICING COPELaND® BraND COMPrESSOrS

Emerson Climate Technologies, Inc. manufac-turers both welded and accessible hermetic (Copelametic®) motor-compressors. Welded compressors cannot be repaired internally in the field, and service operations on these compressors are limited to external electrical components and normal system repairs.

Copelametic® motor-compressors are specifically designed for field accessibility if required. Remov-able heads, stator covers, bottom plates and hous-ing covers allow access for easy field repairs in the event of compressor damage.

The description of service operations that follows is general in nature, but those sections dealing with internal maintenance apply only to Copelametic® compressors.

NaMEPLatE IDENtIFICatION

The model number designation on Copeland® brand compressors and condensing units provides a basic identification of the electrical and physical characteristics. The model numbering system for Copelametic® compressors is shown in Figure 114, for welded compressors in Figure 115, and for condensing units in Figure 116.

For example, model number 4RH1-2500-TMK-105 identifies a Copelametic® motor-compressor as follows:

4 Identifies compressor familyR Identifies refrigerant cooledH Identifies 3020 CFH displacement1 Identifies basic physical characteristics2500 Identifies nominal 25 HPT Identifies three phaseM Identifies Thermotector motor protectionK Identifies 208/220/440/3/60 motor wind-

ing105 Identifies specific bill of material iden-

tifying valves or other optional features

The serial number provides both an identification number and a record of the date of manufacture. It is comprised of 8 digits. The first two identify the

year of manufacture. The third digit is a code letter identifying the month of manufacture, the twelve months of the year being denoted by the first twelve letters of the alphabet (A for January, B for February, etc.). The last five digits are assigned in numerical order during each month’s production.

The manufacturer of the motor used in the motor-compressor is also shown by a code letter preceding the serial number. Code letters are as follows:

C Century D Delco E Emerson G General Electric S A. O. Smith W Wagner

To illustrate, a typical serial number might be C 69G19417. This would indicate:

C Century Motor 69G Manufactured in July, 1969 19417 Identification number

The motor electrical characteristics are also stamped on the nameplate. The motor may be operated at voltages plus or minus 10% of the nameplate rating.

Most Copeland® brand motor-compressors have a basic nameplate rating for both locked rotor and full load amperes based on motor test data. The designation full load amperage persists because of long industry precedent, but in reality a much better term is nameplate amperage. On all welded compressors, all new motors now being developed for Copelametic® compressors, and on most of the motors developed with inherent protection or internal thermostats, nameplate amperage has been arbitrarily established as 80% of the current drawn when the motor protector trips. The 80% figure is derived from standard industry practice of many years’ standing in sizing motor protective devices at 125% of the current drawn at normal load conditions.

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In order for the motor to meet Emerson Climate Technologies, Inc. standards, the trip point must be beyond the prescribed operating limits of the compressor, and is determined during qualification tests by operating the compressor at established maximum load conditions and lowering the supply voltage unit the trip point is reached. Use of the standard 80% factor enables the service and instal-lation engineer to safely size wiring, contactors, or other external line protective devices at 125% of the nameplate rating, since the motor-compressor protector will not allow the amperage to exceed this figure.

In most instances, the motor-compressor is ca-pable of performing at nominal rating conditions at less than rated nameplate amperage. Because of standardization, one motor frequently is used in various compressor models for air-cooled, suction-cooled, water-cooled, high temperature, medium temperature, low temperature, R-12, R-22, or R-502 applications as required. Obviously on many applications there will be a greater safety factory than on others.

When Copeland® brand motor-compressors are listed with U. L., the basic compressor nameplate rating is listed as a maximum. This allows O.E.M. users to list a lower unit nameplate rating should the unit electrical load be less than the original compressor rating. Frequently this permits the use of smaller fuses and wire sizes.

In order to avoid any conflict in the nameplate ratings of the compressor and the unit in small packaged equipment, some welded compressors now have no full load rating stamped on the nameplate, and are assigned an “80% of trip amps” rating on specifica-tion sheets. All welded compressors carry a locked rotor rating on the nameplate, and all Copelametic® compressors have both a locked rotor and a full load amperage rating on the nameplate.

IDENtIFICatION OF POrt LOCatIONS IN HEaDS OF COPELaMEtIC® MOtOr-COMPrESSOrS

In addition to the service ports normally available on suction and discharge compressor service valves, on Copelametic® compressors high and low pres-sure ports are provided in the compressor head.

These provide a convenient connection for high and low pressure controls, and unlike the ports in service valves, cannot be accidentally closed off.

The port locations in various compressor models are shown in Figure 117.

IDENtIFICatION OF MOtOr tErMINaLS ON SINGLE PHaSE COMPrESSOrS The terminal plates on Copelametic® compres-sors are stamped with the terminal identification, and identifying the common, run, and start termi-nals is seldom a problem. This is also true where tee blocks are used on welded compressors, but many welded compressors are manufactured with a Fusite terminal which may have no permanent identification.

Fig. 118 shows the various motor terminal configu-rations used by Copeland® brand products.

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Type A illustrates the individual terminal posts used on smaller horsepower Copelametic® compres-sors. Type A follows no standard industry pattern and applies only to Copelametic® compressors. The terminals are in the order shown when viewed from the stator cover end of the compressor (the end on which the terminal box is mounted).

Type B is a tee block used on larger horsepower compressors, both Copelametic® and welded. Type C is a Fusite connection, normally used with push-on type terminals.

Both Type B and Type C for production conve-nience and easy identification, follow the general industry rule of identifying common, start, and run terminals, always in that order, in the same fashion as reading a book. In other words, reading from left to right, and from top to bottom, the terminals are always C, S, and R.

PrOPEr VaLVE PLatE aND HEaD GaSKEtS FOr 3, 4, aND 6 CYLINDEr COMPrESSOrS

Occasionally when quick delivery of either new or replacement motor-compressors is required from a wholesaler’s stock, and the exact model is not on hand, compressor heads may be changed in the field in order to utilize available stock compres-sors.

WarNINGWhen compressor heads are changed to con-vert standard, capacity control, or two-stage compressors to some other model, the correct gaskets must be used to insure proper perfor-mance and prevent damage to the compressor.

the correct head gasket must exactly match the inner face of the head being used.

Standard Compressor Heads

Figure 119 is an inside view of a typical standard Copelametic® compressor head, showing the inner webbing. The discharge port is located in the valve plate in the area indicated, and the proper gasket matches the inner face of the head.

External Capacity Control Heads

Figure 120 is an inside view of a Copelametic® head equipped with an external unloading valve. The valve is mounted on a discharge port located in the top of the head. The normal discharge port area is fenced off by the “Y” in the inner webbing.

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The proper gasket exactly matches the inner face of the head, the gasket for the external unloading head being externally identified by the tab shown in Figure 120.

Since the area enclosed by the “Y” in the webbing is exposed to discharge pressure from the other cylinders, any leakage from the discharge port in the valve plate into the discharge chamber of an unloaded head can flow directly back to the suction chamber. Such leakage can cause the compres-sor suction pressure to rise immediately when the compressor is pumped down if the unloader valve is not tightly seated.

When a standard heat is replaced with a head equipped with an external unloading valve, the gasket must be changed and the correct gasket must be installed to prevent overheating of the compressor.

In the event unloading is not desired on a cyl-inder bank equipped with a head designed for unloading, both the cylinder head and gasket must be replaced. the correct gasket must be installed to prevent damage to the compres-sor.

A new internal type unloader is currently under development which will also require a special head, but the inner face of the head will be the same as a standard head, and the standard gasket may be used for the unloaded head as well.

two Stage Heads, 3 Cylinder

On two stage compressors, special heads are necessary to provide the necessary separation of the two stages of compression. Figure 121 is an inside view of a typical Copelametic® head for a 3 cylinder two stage compressor.

Refrigerant vapor is returned directly from the suc-tion line to the port in the cylinder head opening into the low stage suction chamber, and is then discharged by the low stage cylinders into the low stage discharge chamber. The gas (at interstage pressure) then enters the interstage manifold, is desuperheated by liquid refrigerant fed by the de-superheating expansion valve, and is discharged into the compressor motor chamber. The high stage suction gas follows the normal suction gas flow path from the motor chamber to the high stage suction chamber, and is then discharged to the condenser through the high stage discharge chamber.

The proper gasket exactly matches the inner web-bing of the head, and must be used to prevent leakage between stages and possible overheating of the compressor motor.

two Stage Heads, 6 Cylinder Compressor

On 6 cylinder two stage compressors, different heads must be used on the high and low stage cylinders. When viewed from the bearing housing end of the compressor (the end on which the oil pump is mounted) the center and right cylinder banks are low stage, and the left cylinder bank is high stage.

Figure 123 is an inside view of a typical low stage head for a 6 cylinder compressor, while Figure

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124 shows an inside view of a typical high stage head.

Refrigerant vapor is returned from the suction line to the normal discharge chamber on the compressor.

The vapor enters the low stage suction chamber through the port at the end of the valve plate, and is discharged from the low stage cylinders into the low stage discharge chamber. The gas (at interstage pressure) then enters the interstage manifold, is desuperheated by liquid refrigerant fed by the de-superheating expansion valve, and is discharged into the compressor motor chamber.

The low stage head is made with the area over the normal suction port blocked off. The proper gasket exactly matches the inner face of the head with the exception that the gasket outlines the solid area, but does not cover it completely.

The high stage head on two stage 6 cylinder com-pressors is similar to the head used on an unloaded head on 6 cylinder compressors, and takes the same head gasket. The high stage suction gas follows the normal suction gas flow path from the motor chamber to the high stage suction chamber,

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and is then discharged to the condenser through the high stage discharge chamber.

Identification of Head Gaskets

As a means of easily identifying head gaskets, and as a guide to proper installation, tabs have been pro-vided on gaskets used on capacity control and two stage heads on 3, 4, and 6 cylinder compressors. In the even there is a question as to whether the proper gasket has been installed, the external tab provides a convenient means of checking without having to remove the compressor head.

Standard head gaskets have no tab, and follow the configuration of the head. The position of the tab when the gasket is properly installed on external capacity control and two stage compressors is il-lustrated in Figures 125, 126, 127, and 128.

any time a compressor head is changed, the proper gaskets must be used to prevent dam-age to the compressor. Compressor failures or compressor damage due to use of improper

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gaskets will be considered as misuse not cov-ered by the Emerson Climate technologies, Inc. warranty, and regular replacement charges will apply.

COPELaND® BraND OIL PUMPS

On all Copelametic® compressors 5 HP and larger in size, and on 3 HP “NR” models, compressor lubrication is provided by means of a positive dis-placement oil pump. The pump is mounted on the bearing housing, and is driven from a slot in the crankshaft into which the flat end of the oil pump drive shaft is fitted.

Oil is forced through a hole in the crankshaft to the compressor bearings and connecting rods. A spring loaded ball check valve serves as a pres-sure relief device, allowing oil to bypass directly to the compressor crankcase if the oil pressure rises above its setting.

Since the oil pump intake is connected directly to the compressor crankcase, the oil pump inlet pres-sure will always be crankcase pressure, and the oil pump outlet pressure will be the sum of crankcase pressure plus oil pump pressure. Therefore, the net oil pump pressure is always the pump outlet pressure minus the crankcase pressure. When the compressor is operating with the suction pressure in a vacuum, the crankcase pressure is negative and must be added to the pump outlet pressure to determine the net oil pump pressure. A typical compound gauge is calibrated in inches of mercury for vacuum readings, and 2 inches of mercury are approximately equal to 1 psi.

For example: Pump Net Oil Outlet PumpCrankcase Pressure Pressure Pressure50 psig 90 psig 40 psi8” vacuum 36 psig 40 psi(equivalent to a reading of minus 4 psig)

In normal operation, the net oil pressure will vary depending on the size of the compressor, the tem-perature and viscosity of the oil, and the amount of clearance in the compressor bearings. Net oil pressures of 30 to 40 psi are normal, but adequate

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lubrication will be maintained at pressures down to 10 psi. The bypass valve is set at the factory to prevent the net pump pressure from exceeding 60 psi.

Every oil pump is given a 100% operating inspection at the factory prior to shipment. The pump is installed in a test stand and must lift oil through unprimed oil lines to a height not less than 12 inches, pick up and develop a full flow of oil within 30 seconds, must not exceed an established maximum power requirement, must develop a minimum of 40 psi pressure with the main outlet closed, and must pump a specified quantity of oil at standard test conditions. Operating pressures and reversal of the pump are checked on the test stand, and on larger compressors are checked again after the pump is installed in a compressor.

The oil pump may be operated in either direction, the reversing action being accomplished by a friction plate which shifts the inlet and outlet ports. After prolonged operation in one direction, wear, corro-sion, varnish formation, or burrs may develop on the reversing plate, and this can prevent the pump from reversing. Therefore, on installations where compressors have been in service for some time, care must be taken to maintain the original phasing of the motor if for any reason the electrical con-nections are disturbed. On transport refrigeration applications where power may be provided from both generators and dock power, both sources of power must be phased alike when connected to the unit in order to prevent reversing the compres-sor rotation.

The presence of liquid refrigerant in the crankcase can materially affect the operation of the oil pump. Violent foaming on start up can result in a loss of oil from the crankcase, and a resulting loss of oil pressure until oil returns to the crankcase. If liquid refrigerant or a refrigerant rich mixture of oil and refrigerant is drawn into the oil pump, the resulting flash gas may result in large variations and possibly a loss of oil pressure. Crankcase pressure may vary from suction pressure since liquid refrigerant in the crankcase can pressurize the crankcase for short intervals, and the oil pressure safety control low pressure connection should always be connected to the crankcase.

During a rapid pull-down of the refrigerant evaporat-ing pressure, the amount of refrigerant in solution in the crankcase oil will be reduced, and may cause flash gas at the oil pump. During this period the oil pump must pump both the flash gas and oil, and as a result the oil pressure may decrease temporarily. This will merely cause the oil pump to bypass less oil, and so long as the oil pressure remains above 9 psi, adequate lubrication will be maintained. As soon as a stabilized condition is reached, and liquid refrigerant is no longer reaching the pump, the oil pressure will return to normal.

The oil pressure safety control high pressure con-nection should be made to the oil pressure port on the oil pump as shown. On the initial start-up of a system, or if at anytime abnormal noise causes any question regarding lubrication, it is recommended that a gauge be attached to the Schrader type valve so that the oil pressure can be observed while the compressor is in operation. The Schrader type valve is for pressure checking only, and is normally closed, so the oil pressure safety control must never be connected to this port.

The oil pump face plate is held in place by the two bolts shown in Figure 129. (Note that these are smaller than the six mounting bolts). The face plate seats on an “O” ring seal and should not be removed. Do not put a gasket between the face

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tYPICaL COPELaND® BraND OIL PUMP


plate and the oil pump body, or the oil pump will be rendered inoperative.

The bolt holding the spring loaded bypass assembly in place should not be removed. The bypass pres-sure is not adjustable, and the bolt is provided for

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access during original assembly or factory mainte-nance, but it is not intended for field repairs. If the bolt is removed, the spring or other components are easily lost or damaged, rendering the oil pump inoperative.

Copeland® brand oil pumps are identified with a letter stamped into the casting as shown in Figure 129 and 130 and all are identical except for the pilot diameter.

Oil pumps identified with an “L” have a 1-15/16” O.D. pilot diameter, are designed for use on all 4R and 6R compressors, and will not fit any other compressor.

“S” oil pumps have a 1 ˚” O.D. pilot diameter, and were standard on all two and three cylinder com-pressors having oil pumps for many years. The “S” model oil pump is being replaced in current production with the “A” model oil pump, because of a change in bearing design on N, M, and 9 model compressors.

The “A” model oil pump has a 1 1/8” O.D. pilot diam-eter. It differs from the “S” and “L” oil pumps in that it is designed to register in the bearing rather than the bearing housing. This makes possible a new style line bored oil pump housing bearing providing accurate alignment of the oil pump.

A larger capacity oil pump with a double impeller has been developed for larger displacement com-pressors, but it is interchangeable with standard oil pumps with the exception that longer mounting bolts are required.

Field replacement of Oil Pumps

If it is determined that an oil pump is not function-ing properly, replace the oil pump and not the compressor.

The oil pump is mounted on the compressor bear-ing housing by means of the six bolts shown in Figure 129. Compressor bearing housings are not interchangeable on most compressor bodies and should not be removed.

Gaskets installed between the oil pump and the compressor body are shown in Figure 132. The tab on the gasket has been added solely for aid in

identification and alignment. The gasket must be in-stalled with the tab in the position shown (11 o’clock position) when viewed facing the compressor, and the slotted hole must always be to the installer’s left in the 9 o’clock position. If the gasket is installed in any other position, the oil ports will be blocked. Gaskets for “L” oil pumps are not interchangeable with gaskets for “A” and “S” pumps.

Some older models of Copelametic® compressors are equipped with Tuthill oil pumps, and these may be furnished on service replacement compressors. The Copeland® brand oil pump is perfectly inter-changeable with the Tuthill pump, and the same gaskets may be used.

WarNINGthe oil pump pilot shoulder must register snugly in either the bearing housing or bearing (depending on compressor design) to insure centering the oil pump. See Figure 131. If not properly registered, the resulting misalignment can result in excessive wear and possible failure of the oil pump. tolerances are very critical for proper operation and extreme care must be taken to insure that proper oil pump, and adaptor if required, is used. the following replacement procedures must be followed to insure trouble free operation.

1. Replacement of “L” oil pumps (4R and 6R com-pressors). Use only “L” oil pumps. Occasionally a serviceman will mount an “A” oil pump on a 4R or 6R compressor by mistake, and since the pilot shoulder will not register on the bear-ing housing, excessive play and misalignment of the shafts will develop resulting in failure of the oil pump. The pump should register in the bearing housing.

2. Replacement of “S” oil pumps with “A” replace-ment kit (M, N, 9 model compressors). Adap-tors have been developed for the “A” oil pump so it can be used as a replacement on all two and three cylinder Copelametic® compressors having oil pumps, regardless of the compres-sor pilot diameter. The “A” replacement kit can be used to replace “S” oil pumps on all older model M, N, and 9 compressors which have bearing housing pilot diameters, by using the 1 ½” O.D. adaptor.

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3. Replacement of “A” oil pump on NRL, NRM, NRN, 9R, 9T, 9W model compressors. The above compressor models with “A” oil pumps have an extended bearing, with a 1 ¼” I.D. nominal register. The “A” replacement kit can be used to replace the original “A” pump by using the 1 ¼” O.D. adaptor.

4. Replacement of “A” oil pump on MR, MW, NRA, NRB, NRD, NRE model compressors. The above compressor models with “A” oil pumps have an extended bearing with a 1⅛” nominal I.D. register. The “A” replacement kit can be used to replace the original “A” pump without the use of any adaptor.

tYPICaL COPELaMEtIC® COMPrESSOr CONStrUCtION

The exploded views illustrate typical Copeland® brand compressor construction details. Individual components will vary with different compressor models, but the basic method of assembly is similar.

MaINtENaNCE aCCESSIBILItY ON COPELaMEtIC® COMPrESSOrS

The heads may be removed on all Copelametic® compressors by removing the head bolts as shown in Figure 139.

The valve plate is then accessible and may be removed as shown in Figure 140. Note that the suction valve reeds are retained in position by dowel pins in the body.

If the motor-compressor is not seized internally it is normally possible to move the pistons by exerting force on the top of the piston, as illustrated in Figure 141. In the event of a broken connecting rod, the piston may “float” in the cylinder during operation. The connecting rod is broken if the piston can be depressed with little or no pressure without affecting the position of the other piston or pistons.

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Figure 138

tYPICaL DEtaILS OF MOtOr PrOtECtOrS USED ON COPELaMEtIC® MOtOr-COMPrESSOrS


Figure 142 shows the location of the suction strainer screen on small Copelametic® compressors. If a restriction somewhere in the low pressure side of the system is indicated, it is advisable to check the strainer for restriction.

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The compressor motor terminals and external pro-tectors are accessible by removing the terminal box lid as shown in Figure 143. The terminal box can be removed by removing the attaching screws.

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Figure 144 is a view of the motor end of a small compressor with the stator cover removed. Note that the external protector, on top of the body is held in position so that it has perfect contact with the compressor body when in the proper position.


This particular model uses an oil flinger for lubri-cation. The ends of the oil flinger run through the oil, some of which is picked up by the “V” near the ends of the arms. The flinger deposits the oil at the top of the stator cover, which then drains into the oil well shown on the stator cover in Figure 145. Note that the oil tube which centers in the hollow crankshaft then permits oil to run from the well through the crankshaft to provide lubrication to the moving parts.

Figure 146 shows arrows on the motor housing cover which must point upward in order that the oil well will be in a proper position to trap the oil.

FIELD trOUBLESHOOtING

One of the basic difficulties in preventing compres-sor failures arises in determining the actual reason for the failure. The compressor is the functioning heart of the refrigeration system, and regardless of the nature of a system malfunction, the com-pressor must ultimately suffer the consequences. Since the compressor is the component that fails, there often is a tendency to blame any failure on the compressor without determining the actual cause of the malfunction. In far too many cases, the actual cause of failure has not been discovered and corrected and the result has been recurring failures that could have been prevented.

If the service engineer is to help in eliminating the causes of compressor failure, then he must thoroughly understand both the operation of the system and the possible causes of failure that might occur, and he must be on the alert for any signs of system malfunction.

If a motor compressor fails to start and run properly, it is important that the compressor be tested to determine its condition. It is possible that external electrical components may be defective, the protec-tor may be open, a safety device may be tripped, or other conditions may be preventing compressor operation. If the motor compressor is not the source of the malfunction, replacing the compressor will only result in the unnecessary expenditure of time and money, while the basic problem remains.

If the service engineer closes his eyes to a basic system malfunction, or an improper control setting,

or an operating condition he knows is not right, he is not fooling the system or the compressor; he is only fooling himself.

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Every service man should have this motto embla-zoned in his mind “Do the job right the first time.” If you can’t find time to do it right, how can you find time to do it over again?

Schematic Wiring Diagram, Single Phase Motors

Actual field wiring diagrams may vary consider-ably in style or format, but Figure 147 is a simple schematic illustration of the basic wiring connec-tions and compressor motor winding relationships in a single phase motor. The diagram as shown illustrates a capacitor start, capacitor run motor, but the same diagram can apply to a permanent split capacitor motor if the starting capacitors and relay are removed, and can apply to a capacitor start, induction run motor if the running capacitor is removed.

A thorough understanding of the basic wiring connections is essential to successfully diagnose field electrical problems on single phase compres-sors.

If the Compressor Will Not run

1. If there is no voltage at the compressor terminals, follow the wiring diagram (Figure 148) and check back from the compressor to the power supply to find where the circuit is interrupted.

Check the controls to see if the contact points are closed (low pressure control, high pressure control, thermostat, oil pressure safety control, etc.). If a contactor is used check to see if the contacts are closed. Check for a blown fuse, open disconnect switch, or loose connection.

2. If power is available at the compressor termi-nals, and the compressor does not run, check the voltage at the compressor terminals while attempting to start the compressor (see Figure 149).

If the voltage at the compressor terminals is below 90% of the nameplate voltage, it is possible the motor may not develop sufficient torque to start. Check to determine if wire sizes are adequate, electrical connections are loose, the circuit is overloaded, or if the power supply is adequate.

3. On units with single phase PSC motors, the suction and discharge pressures must be equal-ized before starting because of the low starting torque of the motor. Any change in the refriger-ant metering device, the addition of a drier, or other changes in the system components may delay pressure equalization and create starting difficulties. If PSC motor starting problems are being encountered, the addition of a capacitor start kit is recommended.

4. On single phase compressors, a defective capacitor or relay may prevent the compressor starting. If the compressor attempts to start, but is unable to do so, or if there is a humming sound, check the relay to see if the relay contacts are damaged or fused. The relay points should be closed during the initial starting cycle, but should open as the compressor comes up to speed.

Remove the wires from the starting relay and capacitors. Use a high voltage ohmmeter to check for continuity through the relay coil. Re-

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place the relay if there is no continuity. Use an ohmmeter to check across the relay contacts. Potential relay contacts are normally closed when the relay is not energized, current relay contacts are normally open. If either gives an incorrect reading, replace the relay.

Any capacitor found to be bulging, leaking, or

damaged should be replaced.

Make sure capacitors are discharged before checking. Check for continuity between each capacitor terminal and the case. Continuity indicates a short, and the capacitor should be replaced.

Substitute “a known to be good” start capacitor if available. If compressor then starts and runs properly, replace the original start capacitor. On PSC motors, substitute “a known to be good” run capacitor if available. If compressor then starts and runs properly, replace the original run capacitor.

If a capacitor tester is not available, an ohmmeter may be used to check run and start capacitors for shorts or open circuits. Use an ohmmeter set to its highest resistance scale, and connect prods to capacitor terminals.

(a) With a good capacitor, the indicator should first move to zero, and then gradually in-crease to infinity.

(b) If there is no movement of the ohmmeter indicator, an open circuit is indicated.

(c) If the ohmmeter indicator moves to zero, and remains there or on a low resistance reading, a short circuit is indicated. Defec-tive capacitors should be replaced.

5. If the correct voltage is available at the com-pressor terminals, and no current is drawn, remove all wires from the terminals and check for continuity through the motor windings. On single phase motor compressors, check for continuity from terminals C to R, and C to S. On three phase compressors, check for con-tinuity between the terminals for connections to phases 1 and 2, 2 and 3, and 1 and 3. On

compressors with line break inherent protectors, an open overload protector can cause a lack of continuity. If the compressor is warm, wait one hour for compressor to cool and recheck. If continuity cannot be established through all motor windings, the compressor should be replaced.

Check the motor for ground by means of a continuity check between the common terminal and the compressor shell. If there is a ground, replace the compressor.

6. If the compressor has an external protector, check for continuity through the protector or protectors. (See Figure 150)

All external and internal inherent protectors on Copelametic® compressors can be replaced in the field. On larger compressors with thermo-stats, thermotectors, or solid state sensors, in the motor windings (D, H, M, S protection), the internal protective devices cannot be replaced

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and the stator or compressor must be changed if the internal protectors are defective or dam-aged.

If the Motor Compressor Starts But trips repeatedly On the Overload Protector

1. Check the compressor suction and discharge pressures while the compressor is operating. (See Figure 151.) Be sure the pressures are within the limitations of the compressor. If pres-sures are excessive, it may be necessary to clean the condenser, purge air from the system, add a crankcase pressure regulating valve, modify the system control, or take such other action as may be necessary to avoid excessive pressures.

An excessively low suction pressure may in-dicate a loss of charge, and a suction cooled motor compressor may not be getting enough refrigerant vapor across the motor for proper cooling.

On units with no service gauge ports where pressures can be checked, check condenser to be sure it is clean and fan is running. Excessive temperatures on suction and discharge line may also indicate abnormal operating conditions.

2. Check the line voltage at the motor terminals while the compressor is operating. (See Figure 149.) The voltage should be within 10% of the nameplate voltage rating. If outside those limits, the voltage supply must be brought within the proper range, or a motor compressor with dif-ferent electrical characteristics must be used.

3. Check the amperage drawn while the compres-sor is operating. (See Figure 152.) Under nor-mal operating conditions, the amperage drawn will seldom exceed 110% of the nameplate amperage and should never exceed 120% of the nameplate amperage. High amperage can be caused by low voltage, high head pressure, high suction pressure, low oil level, compressor mechanical damage, defective running capaci-tors, or a defective starting relay.

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On three phase compressors, check amperage in each line. One or two high amperage legs on a three phase motor indicates an unbal-anced voltage supply, or a winding imbalance. If all three legs are not drawing approximately equal amperage, temporarily switch the leads to the motor to determine if the high leg stays with the line or stays with the terminal. If the high amperage reading stays with the line, the problem is in the line voltage supply. If the high amperage reading stays with the terminal, the problem is in the motor.

If the amperage is sufficiently unbalanced to cause a protector trip, and the voltage supply is unbalanced, check with the power company to see if the condition can be corrected. If the voltage supply is balanced, indicating a defec-tive motor phase, the compressor should be replaced.

4. Check for a defective running capacitor or start-

ing relay in the same manner described in the previous section.

5. Check the wiring against the wiring diagram in the terminal box. On dual voltage motors, check the location of the terminal jumper bars to be sure phases are properly connected. (See Figure 153.)

6. Overheating of the cylinders and head can be caused by a leaking valve plate. To check, close the suction service valve and pump the compressor into a vacuum. Stop the compres-sor and crack the suction valve to allow the pressure on the suction gauge to build up to 0 psig. Again close the valve. If the pressure on the gauge continues to increase steadily, the valve plate is leaking. Remove the head and check the valve plate, replace if necessary. (See Figure 154.)

7. If all operating conditions are normal, the voltage supply at the compressor terminals balanced and within limits, the compressor crankcase temperature within normal limits, and the am-perage drawn within the specified range, the motor protector may be defective, and should be replaced.

If the operating conditions are normal and the compressor is running excessively hot for no observable reason, or if the amperage drawn is above the normal range and sufficient to repeatedly trip the protector, the compressor has internal damage and should be replaced.

If the Compressor runs But Will Not refrigerate

1. Check the refrigerant charge. If sight glass is available, it should show clear liquid. Check the evaporator surface to determine if it is evenly cold throughout, or if partially starved. A lack of charge may be indicated by light, fluffy frost at the expansion valve and evaporator inlet. Add refrigerant if necessary.

2. Check the compressor suction pressure. An abnormally low pressure may indicate a loss of refrigerant charge, a malfunctioning expan-sion valve or capillary tube, a lack of evaporator capacity possibly due to icing or low air flow, or a restriction in the system.


Often a restriction in a drier or strainer can be identified by frost or a decrease in temperature across the restriction due to the pressure drop in the line. This will be true only if liquid refrigerant is in the line at the restricted point, since any temperature change due to restriction would be caused by the flashing of liquid into vapor as the pressure changes.

Any abnormal restrictions in the system must be corrected.

3. Check the compressor discharge pressure. An abnormally high discharge pressure can cause loss of capacity, and can be caused by a dirty condenser, a malfunctioning condenser fan, or air in the system.

4. If the suction pressure is high, and the evapo-rator and condenser are functioning normally, check the compressor amperage draw. An

amperage draw near or above the nameplate rating indicates normal compressor operation, and it is possible the compressor or unit may have damaged valves or does not have suf-ficient capacity for the application.

An amperage draw considerably below the nameplate rating may indicate a broken suction reed or broken connecting rod in the compres-sor. Check the pistons and valve plate on an accessible compressor. If no other reason for lack of capacity can be found, replace a welded compressor.

Service Diagnosis Chart

Table 52 is a service diagnosis chart which can serve as a checklist of possible causes for various system malfunctions. While unusual conditions may occasionally occur, the chart covers the common types of malfunctions normally encountered.


SECtION 26FUNDaMENtaLS OF SErVICE OPEratIONS

The installation and maintenance of refrigeration equipment is one of the most exacting and demand-ing tasks in the service field. In addition to the care necessary in working with equipment built with fine precision to the close tolerances required, refriger-ants introduce an additional hazard. Servicemen often tend to underestimate how much care is required to properly protect a system.

So long as a refrigerant is tightly imprisoned and properly controlled, it can be made to perform useful work. But it doesn’t do it willingly. Given the slightest opportunity, it will escape. If joined by such common substances as moisture or air, it combines with them to form acids and attack the system. And, if left uncontrolled for even a few hours, it can migrate through the system, often with fatal results to the compressor on start-up. When handling refrigerants, the serviceman can never relax, he must always be alert and on guard.

CONtaMINaNtS

Absolute cleanliness is essential in a refrigeration system. In order to insure a reliable, trouble free unit, there are no compromises.

Unlike most other mechanical equipment, refrig-eration systems are vulnerable to attack from two common contaminants, air and water, which can-not be seen. Yet if either or both are present in a system, they quickly join in a common attack on the refrigerant and oil, and can cause corrosion, copper plating, acid formation, sludging, and other harmful reactions.

Antifreeze solutions or other additives may cre-ate undesirable chemical reactions in a system. Additives of any type are not recommended and should not be used.

It is amazing, and sometimes almost unbelievable, to see the many foreign materials that have entered a refrigeration system and end up in the compressor. Filings, shavings, dirt, solder, flux, metal chips, bits of steel wool, mortar, sand from sandcloth, wires from cleaning brushes, lengths of copper tubing—all have been encountered. Examination of returned

compressors indicates that many early failures could have been prevented if the contaminants had been removed from the system at the time of installation. This type of problem is encountered most often on field installed systems, and it seems inescapable that many of the contaminants found in systems could get there only from carelessness during installation.

When brazing copper tubing and fittings, copper oxide is invariably formed on the inside of the tube unless nitrogen or some other inert gas is circulated through the tubing during the brazing operation. That oxide can become a powdered abrasive, plugging oil passages, scoring bearings, plugging filters, and causing other injurious effects.

Reasonable care during installation and service can keep contamination in a system at a safe and acceptable level.

1. Take care to keep tubing clean and dry.

2. Pass an inert gas through the tubing when brazing refrigerant tubes.

3. Take extreme care to keep foreign materials out of the system when it is opened for service.

4. Suction line filters and liquid line filter-driers should be installed in all field installed sys-tems.

5. Thoroughly evacuate the system at the time of original installation, or after exposure for long periods, during maintenance.

6. Install a new filter-drier in the liquid line each time the system is opened for service.

HaNDLING OF rEFrIGEraNt CONtaINErS

The pressure created by liquid refrigerant in a sealed storage container is equal to its satura-tion pressure at the liquid temperature so long as there is vapor space available. If however, the container is over-filled, or if in the case of gradual and uniform overheating the liquid expands until

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the container becomes liquid full, hydrostatic pres-sure builds up rapidly to pressures far in excess of saturation pressures. Figure 155 illustrates the dangerous pressures that can be created under such circumstances, which can result in possible rupture of the refrigerant container such as illus-trated in Figure 156.

The chart in Figure 155 illustrates the pressure-tem-perature relationship of liquid refrigerant before and after a cylinder becomes liquid-full under gradual and uniform heating. The true pressure-temperature relationship exists up to the point where expansion volume is no longer available within the cylinder.

If a refrigerant cylinder becomes liquid-full, hy-drostatic pressure builds up rapidly with only a small increase in temperature. Excessive pressure build-up can cause cylinder rupture as pictured. Under uniform conditions of heating, the cylinder illustrated ruptured at approximately 1,300 pounds per square inch gauge pressure. If heat is applied with a torch to a local area, cylinder wall may be weakened at this point and the danger of rupture would be increased. In a controlled test a cylinder such as the one pictured flew over 40 ft. in the air upon rupture—a dramatic demonstration of the danger of over-heating cylinders.

Interstate Commerce Commission regulations prescribe that a liquefied gas container shall not be liquid full when heated to 55°C. (131°F.). If cylinders are loaded in compliance with this regulation, at temperatures above 131°F. liquid refrigerant may completely fill a container because of expansion of the liquid at increasing temperatures. Fusible metal plugs are designed to protect the cylinder in case of fire, but will not protect the cylinder from

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gradual and uniform overheating. Fusible metal plugs begin to soften at 157°F., but hydrostatic pressure developed at 157°F. is far in excess of cylinder test pressure.

The following safety rules should be followed at all times when handling cylinders of compressed gas.

1. Never heat a cylinder above 125°F.

2. Never store refrigerant cylinders in the direct sunlight.

3. Never place an electric resistance heater in direct contact with a refrigerant cylinder.

4. Never apply a direct flame to a cylinder.

5. When refilling small cylinders, never exceed the weight stamped on the refrigerant cylinder.

6. Do no drop, dent, or otherwise abuse cylin-ders.

7. Always keep the valve cap and head cap in place when the cylinder is not in use.

8. Always open all cylinder valves slowly.

9. Secure all cylinders in an upright position to a stationary object with a strap or chain when they are not mounted in a suitable stand.

The common fluorocarbon refrigerants (R-12, R-22, R-502) were originally developed by Dupont as “Freon” refrigerants, but different manufacturers use different trade names for the same refrigerant. For example R-12 is the common industry designation for the refrigerant Dichlorodifluoromethane, but it may be marketed as Freon 12, Genetron 12, Iso-tron 12, Ucon 12, etc. Refrigerant containers are usually color coded as follows:

R-11 Orange R-12 White R-22 Green R-502 Purple

SaFE HaNDLING OF COMPrESSED GaSES WHEN tEStING Or CLEaNING rEFrIGEratION SYStEMS

When the use of an inert gas is required for high pressure test purposes or to flush a contaminated system, Emerson Climate Technologies, Inc. rec-ommends the use of either dry nitrogen (N2) or dry carbon dioxide (CO2). At 70°F., dry nitrogen in “K” cylinders may be under a pressure of 2200 psig or more, and carbon dioxide at the same temperature may be under a pressure in excess of 830 psig. Extreme caution must be exercised in the use of highly compressed gases, since careless or im-proper handling can be very dangerous.

Oxygen or acetylene should never be used for pressure testing or cleanout of refrigeration sys-tems, as the use of either may result in a violent explosion. Free oxygen will explode on contact with oil, and acetylene will explode spontaneously when put under pressure unless dissolved in a special holding agent such as used in acetylene tanks.

WarNING - HIGH PrESSUrE COMPrESSED GaSES SHOULD NEVEr BE USED IN rEFrIGEratION SYStEMS WItHOUt a rELIaBLE PrESSUrE rEGULatOr aND PrESSUrE rELIEF VaLVE IN tHE LINES aS DESCrIBED HErEIN.

recommended test Pressures

All new Copelametic® and welded compressors are now designed with a crankcase ultimate bursting pressure in excess of 850 psig, and production samples are periodically checked hydrostatically to insure this standard being maintained. Many older models of Copeland® brand compressors and all belt driven compressor crankcases were designed for a minimum of 650 psig bursting pres-sure. However, the ultimate burst test is a strength test only, and both leaks and distortion can occur at high pressures even though the crankcase may not rupture.

Every Copelametic® compressor crankcase is subjected to a 300 psig pressure at the factory, and every Copeland® brand compressor is leak tested at a minimum of 175 psig.

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Because of the possibility of damage in transit, and the hazard of rupture with compressed gas or air, plus the fact that many manufacturers do not design for crankcase pressures as high as those for Copeland® brand products, it is recommended that all crankcase test pressures and all leak test pressures be limited to a maximum of 175 psig. U. L. safety standard for condensing units normally can be met by testing the complete unit at the required low side pressure of 150 psig.

In the event high side test pressures are required, the crankcase must be protected from the high pressure, not only as a safety measure, but also

to prevent possible distortion of the crankcase resulting in noise or mounting problems.

High side pressure conditions are dictated by the intended usage. Emerson Climate Technologies, Inc. minimum high side test pressures for unit ap-plications are as follows, but the maximum is not to exceed 500 psig.

Copeland® Brand Unit High Side MinimumApplication Leak Test PressureR-12, Air or Water Cooled 335 psigR-22 and R-502, Water Cooled 335 psigR-22 and R-502, Air Cooled 450 psig

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Figure 157 illustrates what can happen to a compressor if exposed to pressures in excess of the compressor’s ultimate strength. This type of damage most frequently occurs when servicemen attempt to purge or pressurize a refrigeration system with high pressure compressed gases without a pressure regulator.


recommended Procedure for Leak or Pressure tests

Figure 158 illustrates the gauges and valves that must be installed in the supply line for proper safety of personnel and equipment when testing with high pressure gases.

1. Separate relief valves for high and low side tests are required, one preset for 175 psig for low side tests, including the crankcase; the other preset at the required high side pressure.

2. When testing at pressures above 175 psig, the compressor and low pressure components must be disconnected from the system. Should it be impractical to disconnect the compressor during high side pressure test, an adequate means of pressure relief must be provided on the compressor crankcase to prevent dam-age in the event the high pressure gas should leak back into the crankcase. A bleed line, if provided, should be larger than the line from the gas cylinder.

3. With the compressed gas cylinder in the upright position, admit the dry nitrogen or dry carbon dioxide slowly until the desired system pressure is obtained.

4. Close the cylinder valve. Check the system pressure gauge, and adjust as necessary to obtain the proper pressure.

5. Proceed with test, and when complete, system pressure should be reduced to 0 psig, compres-

sor reconnected, the system evacuated, and then charged with the proper kind and amount of refrigerant.

recommended Procedure For Purging Con-taminated Systems

Evacuation is the only dependable and effective means of removing air and moisture from a system to the required low level. If air is trapped in the compressor, it is practically impossible to remove from the compressor crankcase by purging. In case of a motor burn, Emerson Climate Technologies, Inc. recommends only the filter-drier system clean-ing procedure.

However, in the event a system is badly contami-nated (for example, if a water line ruptures in a water cooled condenser) it may be desirable to purge the system with dry compressed gas prior to starting the final cleaning process. This not only can speed the cleaning procedure, but can reduce the con-taminants to a level that can be handled effectively by the necessary high vacuum equipment.

1. Disconnect the compressor and remove the low pressure components (expansion valves, capillary tubes, controls, etc.) from the system. Install suitable jumpers in place of expansion valves, capillary tubes, etc. and cap fittings from which controls were disconnected. A pressure relief device preset at 175 psig must be installed in the supply line. (See Figure 158).

2. Dry nitrogen, dry carbon dioxide may now be introduced into the system. The pressure regulator should be set to limit the pressure to 100 psig.

3. Purge gas through the system until all free contamination has been removed.

4. Close the cylinder gas valve, remove the pres-sure supply line, remove the jumpers, and reconnect the compressor and the low pressure components.

5. Install adequate filter-driers in the suction and liquid lines, pressure test, evacuate, and com-plete the system cleaning as necessary.

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HaNDLING COPPEr tUBING

Copper tubing is made for many types of usage, but tubing intended for plumbing or water pipe may contain waxes or oils on the interior surface that can be extremely detrimental in a refrigeration system.

Dehydrated and sealed coil of soft copper tubing as it comes from the manufacturer. Proper handling of tubing is necessary to obtain clean, dry systems.

Use only copper tubing especially cleaned and dehydrated for refrigeration usage. Soft copper tubing is available in rolls with the ends sealed, and hard drawn tubing is available capped and dehydrated. Keep the tubing capped or sealed until ready for installation, and reseal any tubing returned to storage.

Examples of hard drawn copper tubing for refrigeration service. Note caps on ends to keep interior surfaces clean during storage.

In the event hard drawn tubing is left open and does get dirty, draw a rag soaked in refrigeration oil through the tubing prior to usage.

BraZING rEFrIGEraNt LINES

Refrigeration systems must be leak free, and the ability to properly braze joints in tubing is an es-sential skill of the refrigeration serviceman.

Tubing should be cleaned and burnished bright before brazing. Care in cleaning is essential for good gas-tight connections. Particular attention should be given to preventing metal particles or abrasive material from entering the tubing.

A suitable low temperature brazing flux that is fully liquid and active below the flow point of the brazing alloy is required. Because of their nature, brazing fluxes are quite active chemically, and must be kept out of the system. Only the male connection should be fluxed, and only enough flux should be used to adequately cover the surface.

Applying flux to cleaned tubing before soldering. Flux should be applied sparingly and kept away from tube end.

When heat is applied to copper in the presence of air, copper oxide is formed. This oxide can be extremely harmful to a refrigeration system. To prevent its formation, an inert gas such as dry

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Courtesy E. I. DuPont de Nemours & Co.

Copper tube and fittings should be thorough-ly cleaned down to bare metal before making soldered or brazed joint. Care in cleaning will largely insure good, gas-tight connections. Note tubing is pitched downward to prevent entry of abrasive particles.

Making silver soldered joint with fitting look-ing down. Whenever possible, soldered joints should be made in this manner to keep flux and solder from getting inside. Note also that dry nitrogen is being swept through the tubing while soldering to prevent oxide formation.


nitrogen should be swept through the line at low pressure during the brazing operation. Always use a pressure regulating valve in the line connecting the nitrogen cylinder to the system.

The tubing should be properly supported so that no strain is placed on the joints during brazing and cooling, and so that expansion and contraction will not be restricted.

Apply heat evenly to the tube and fitting until the flux begins to melt. The way heat is applied can either draw flux into the joint or prevent its entry. Apply heat around the circumference of the fitting to draw the brazing alloy into the joint to make a mechanically strong and tight joint.

Never apply heat to a line under refrigerant pres-sure. The line may rupture, and the escaping re-frigerant pressure may throw blazing oil or molten solder through the air. Refrigerants when exposed to an open flame may break down into irritating or poisonous gases.

Immediately after the brazing alloy has set, apply a wet brush or cloth to the joint to wash off the flux. All flux must be removed for inspection and pressure testing.

SErVICE VaLVES

With the exception of small, unitary, sealed systems utilizing welded compressors, almost all refrigeration and air conditioning systems have service valves for operational checking and maintenance access. Normally on accessible-hermetic compressors, the compressor is equipped with suction and discharge valves having service ports. Some systems may have service valves on line connections, receiver valves, or charging valves.

Figure 164 illustrates a typical compressor service valve, but valves of similar construction may be used for base valves, receiver valves, or charging valves. Note that there is a common connection that is always open, a line connection, and a gauge port.

When the valve is back-seated (the stem turned all the way out) the gauge port is closed and the valve is open. If the valve is front-seated (the stem turned all the way in) the gauge port is open to

the common connection and the line connection is closed. In order to read the pressure while the valve is open, the valve should be back-seated, and then turned in one or two turns in order to slightly open the connection to the gauge port.

Some valves of the same general type intended for process access only may have only the line and gauge connections with the common port omitted. The action of the valve seat is unchanged. The line connection is closed when the valve is front-seated, the gauge connection is closed when the valve is back-seated.

Figure 165 illustrates a Schrader type valve similar in appearance and principle to the air valve used on automobile or bicycle tires.

The Schrader type valve is a recent development for convenient checking of system pressures where it is not economical, convenient, or possible to use the compressor valves with gauge ports.

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This type of valve enables checking of the system pressure, or charging refrigerant without disturbing the unit operation. An adaptor is necessary for the standard serviceman’s gauge or hose connection to fit the Schrader type valve.

tHE GaUGE MaNIFOLD

The most important tool of the refrigeration ser-viceman is the gauge manifold. It can be used for checking system pressures, charging refrigerant, evacuating the system, purging non-condensables, adding oil, and for many other purposes.

Basically the gauge manifold consists of compound and high pressure gauges mounted on a manifold with hand valves to isolate the common connec-tion, or open it to either side as desired. Figure 167 shows a schematic view of a gauge manifold with both valves closed. Figure 168 illustrates the same manifold with the common connection open to the high pressure connection. The ports above and below each valve are interconnected so the gauges will always register when connected to a pressure source.

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The left hand gauge is normally a compound or suction pressure gauge. The right hand gauge is the high or discharge pressure gauge. Flexible hoses are used to make connections from the manifold to the system.

Gauges are fine instruments and should be treated as such. Do no drop, keep in adjustment, and do not subject gauges to pressures higher than the maximum pressure shown on the scale.

Connecting the gauge manifold to a system is one of the most common service functions. To avoid introducing contaminants into the system, the hose connections must always be purged with refrigerant before connecting the manifold. A consistent proce-dure should always be followed by the serviceman in making the connections. For an operating system containing refrigerant, proceed as follows:

First, back-seat the service valves to which the gauges are to be connected so that the gauge ports are isolated. Be sure both manifold valves are closed (front-seated).

If operating conditions are such that the suction pressure is certain to be above 0 psig, tighten hose connections to both service valves. Be sure com-mon hose connection on manifold is open.

Crack (open slightly) the high pressure manifold valve. Then crack the high pressure service valve

thus allowing refrigerant to bleed through the discharge and common hoses. Allow refrigerant to bleed for a few seconds, and then close the high pressure valve on the manifold. Repeat the same procedure with the low pressure valves. The manifold is then connected to the system ready for use.

In the case of system where the low side pressure might be in a vacuum, all purging must be done from the high pressure service valve. Back-seat the service valves and tighten the hose connection to the high pressure service valve. Leave hose con-nection at low side service valve loose and cap or plug loosely the common hose connection. Crack both high and low pressure valves on the manifold. After a few seconds, tighten the hose connection at the low pressure service valve, and then tighten the cap or plug in the common connection. Close the valves on the manifold, crack the low pressure service valve, and the manifold is then connected to the system ready for use.

PUrGING NON-CONDENSaBLES

A leak in the low pressure side of an operating system frequently results in the entrance of air. In some cases it may be impractical to remove the refrigerant charge and evacuate the system, yet the air must be removed to prevent damaging chemical reactions.

Air is non-condensable under the temperatures and pressures encountered in an air conditioning or commercial refrigeration system. The liquid seal at the outlet of the receiver and condenser will normally trap the air in the top of the receiver and condenser. The system condensing pressure will be increased by the pressure exerted by the trapped air, the amount of the increase in pressure being dependent on the quantity of air trapped. Before starting to purge, note the compressor operating discharge pressure, and compare with the tem-perature of the condensing medium.

Restart the compressor and check to see if the discharge pressure is still abnormally high. If so, operate the system for a few minutes and repeat the purging procedure. Normally purging 3 or 4 times will remove most of the non-condensables trapped in the top of the condenser and restore

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normal operating pressures. However, purging should be used only as a short term emergency measure. In order to insure satisfactory compres-sor operation the system should be evacuated as soon as practical.

SYStEM PUMPDOWN

For any service work requiring access to the compressor or the sealed part of the system, the refrigerant must first be removed. On small sys-tems without service valves, it may be necessary to remove the refrigerant charge prior to servicing the equipment, and then recharge the system when put back in service.

On any system with service valves, the refrigerant can be pumped into the condenser and receiver (if used) and isolated there. This operation is termed pumping the system down, and is accomplished by closing the valve at the outlet of the receiver or condenser while the compressor is operating. Since no further refrigerant can flow to the evaporator, the refrigerant is pumped out of the evaporator and into the condenser.

Check the operating pressures by means of a gauge manifold, (see Figure 166) and when the suction pressure reaches 1 to 2 inches of vacuum, stop the compressor. (Note: If the unit is equipped with a low pressure control having a higher setting, it will be necessary to bypass the low pressure control in order to keep the compressor operating while pumping the system down.) If the pressure rises rapidly, this is an indication that there is still residual refrigerant in the compressor crankcase. Start the compressor and again pump the suction pressure down to 1 to 2 inches of vacuum. If the pressure remains at that point or rises very slowly, close the compressor discharge service valve. In the event the pressure should remain in a vacu-um, disconnect power from the compressor, and crack the receiver valve momentarily to introduce sufficient refrigerant to obtain a slight positive pressure.

The liquid line, the low pressure side of the sys-tem, and the compressor should now be at a slight positive pressure, (approximately 1 psig) and that part of the system can be opened for service. The refrigerant pressure prevents the inrush of air into

the open system, and reduces contamination to a minimum.

Note that if it is necessary to remove or gain access to the discharge line, condenser, or receiver, pumping the system down is of no ben-efit, and the refrigerant charge must be removed unless there are valves to isolate the defective component.

Pumpdown control is also used as a means of isolat-ing the refrigerant and preventing migration to the compressor crankcase during periods of shipment, storage, or long non-operating off cycles.

rEFrIGEraNt LEaKS

Refrigeration systems must be absolutely gas tight for two reasons. First, any leakage will result in loss of the refrigerant charge. Second, leaks allow air and moisture to enter the system.

Leaks can occur not only from joints or fittings not properly made at the time of the original installa-tion, but from line breakage due to vibration, gasket failure, or other operating malfunctions. A recent study by a major user of commercial equipment revealed that of approximately 3,000 service calls made during a typical year’s operation, 1 out of 6 were required because of refrigerant leaks. Since leak detection is such a common service complaint, it is essential that the service engineer check the system carefully to insure that it is leak tight before charging with refrigerant.

There are three common means of pressure test-ing a system for leaks. The pressure text method involves pressurizing the system and checking for leakage outward.

WarNING — Never use oxygen for pressur-izing a system; an explosion may occur if oil is present in the system. always use a gauge equipped pressure regulator on the high pres-sure back-up gas, and never interconnect the refrigerant cylinder and the inert gas cylinder through a gauge manifold. Nitrogen and car-bon dioxide cylinder pressures can rupture a refrigerant cylinder.

The electronic leak detector is the most sensitive

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type available. These are available at reasonable cost, and can detect small leaks of a fraction of an ounce per year, often missed when using other test-ing methods. Because of their extreme sensitivity, electronic detectors can only be used in a clean atmosphere not contaminated by refrigerant vapor, smoke, vapor from carbon tetrachloride, or other solvents which may give a false reaction.

The leak detector most widely used for field service is the halide torch. It consists of a small portable propane or L. P. gas tank, a sniffer hose, and a special burner which contains a copper element. The gas feeds a small flame in the burner, pulling a slight vacuum on the sniffer hose. When the probe is passed near a leak, the refrigerant is drawn into the hose and injected into the burner below the copper element. A small amount of refrigerant burning in the presence of copper has a bright green color. A larger amount will burn with a violet colored flame. When testing for leaks with a torch, always watch the flame for the slightest changes in color. With experience, very small leaks can be detected.

To use the halide torch to find a leak, explore each joint and fitting in the system. Check all gasketed joints at the compressor. Some manufacturers use the halide torch as a final check on packaged systems which are shipped in cartons, by punch-ing a hole in the carton and checking inside the carton several hours after the unit is packaged. A very small leak will tend to build up in strength in an enclosed area, and can thus be detected.

The simplest and oldest method of leak detection is by means of soap bubbles. Swab a suspected leak with liquid soap or detergent, and bubbles will appear if a leak exists. Despite its simplicity, the soap bubble method can be extremely helpful in pinpointing a leak which is difficult to locate.

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This type of detector is ideally suited for field service of air conditioning and refrigeration equip-ment.

The oldest and probably most widely used leak detector for fluorinated refrigerants is the halide type. The one illustrated is made to attach to a small, portable gas cylinder. This makes a very compact, easy to use, leak detection device.


When a leak is located, it should be marked. When leak testing is completed and all leaks have been located and marked, vent the test pressure gas. If a leak requiring brazing is found in the high pressure side of a system containing a refrigerant charge, in a location that cannot be isolated, it will be necessary to remove the refrigerant in order to make repairs.

When pressure has been removed from the area where the leak is located, the leak can be repaired as necessary. It may be necessary to re-braze fit-tings, replace gaskets, repair flare connections, or merely tighten connections. When all leaks have been repaired, the system should again be pres-surized and the leak testing process repeated.

Pressure leak testing is necessary to locate indi-vidual leaks. In order to determine if the system is free of all leaks, a vacuum test is helpful. After repairing all known leaks, draw a deep vacuum on the system with a good vacuum pump. The pressure should be reduced to 1 psia or less (the vacuum registered on the test gauge will vary with atmospheric pressure) and the system should be sealed and left for at least 12 hours. Any leakage of air into the system will cause the vacuum reading to decrease. (Some slight change in pressure may

be caused by changes in ambient temperature). If an air leak is indicated, the system should again be pressure leak tested, and the leaks located and repaired.

When all leaks have been repaired and the system satisfactorily passes the leak tests, it is ready for evacuation and charging.

EVaCUatION

Any time the compressor or system is exposed for prolonged periods to atmospheric air, or if the system becomes contaminated and removal of the refrigerant charge is necessary, the system should be evacuated in the same manner as at the original installation.

Liquid line filter-driers will effectively remove small amounts of moisture from a system, but the amount of moisture in an open system may be greater than a drier’s capacity. In both cases, evacuation is the only means of insuring a contaminant free system.

Under no conditions is the motor-compressor to be started or operated while the system is under a high vacuum. To do so may cause serious damage to the motor windings.

A small portable vacuum pump specifically built for refrigeration evacuation should be used. Do not use the refrigeration compressor as a vacuum pump. The serviceman who uses some discarded refrig-eration compressor as a vacuum pump is fooling himself and endangering the system.

The gauge manifold provides a convenient means of connecting the vacuum pump to a service valves on the compressor or in the system, and is adequate for field evacuation of relatively small systems with small displacement vacuum pumps. For larger systems and larger vacuum pumps, however, the pressure drop through the hose connections on the normal service gauge manifold is so high that evacuation is very slow, and gauge readings may be misleading. Copper tubing or high vacuum hoses of ¼ in. I.D. minimum size are recommended for high vacuum work.

Triple evacuation is strongly recommended for all field installed systems because of the greater

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Checking for refrigerant leaks with halide torch. Note sampling tube held adjacent to point of possible leak. Eye should be kept on flame to observe any color change.


degree of contamination that must be expected under actual operating conditions as opposed to laboratory or production line processing.

To evacuate a system with a small vacuum pump and a gauge manifold, attach the common con-nection on the gauge manifold to the suction con-nection on the vacuum pump. The high and low pressure connections on the gauge manifold should be securely connected to gauge ports on service valves on the high and low pressure sides of the system respectively.

With the valves on the gauge manifold closed (front-seated) open the service valves and adjust to a point approximately midway between the front-seat and back-seat position.

Start the vacuum pump and gradually open the gauge manifold valves. It may be necessary to restrict the vacuum pump suction pressure by means of the gauge manifold valves to avoid overloading the pump motor. Continue evacuation until the desired vacuum reading is obtained on both gauges.

When evacuation is complete, close the gauge manifold valves tightly, remove the line from the vacuum pump, and connect to a refrigerant cylinder of the same type refrigerant used in the system. Loosen the common hose connection at the gauge manifold, crack the refrigerant drum valve to purge the hose, and retighten the hose connection. Crack the valves on the gauge manifold until the system pressure rises to 2 psig. Close the refrigerant drum valve and the gauge manifold valves.

For triple evacuation, the above procedure should be repeated three times, evacuating twice to 1500 microns, and the last time to 500 microns, or to the limit of the vacuum pump’s ability.

When complete, the system is ready for charging. If it is not to be charged immediately, the system may be sealed by back-seating the service access valves, and plugging or capping all open gauge ports or connections.

CHarGING rEFrIGEraNt INtO a SYStEM

The proper performance of a refrigeration or air

conditioning system is dependent on the proper refrigerant charge. An under-charged system will starve the evaporator, resulting in excessively low compressor suction pressures, loss of capacity, and possible compressor overheating. Overcharg-ing can flood the condenser resulting in high dis-charge pressures, liquid refrigerant flooding, and potential compressor damage. Most systems have a reasonable area of tolerance for some variation in charge, although some small systems may ac-tually have a critical charge which is essential for proper operation.

Each system must be considered separately, since systems with the same capacity or horsepower rat-ing may not necessarily require the same refriger-ant or the same amount of charge. Therefore it is important to first determine the type of refrigerant required for the system, the unit nameplate nor-mally identifying both the type of refrigerant and the weight of refrigerant required.

Liquid Charging

Charging with liquid refrigerant is much faster than vapor charging, and because of this factor is almost always used on large field installed systems. Liquid charging requires either a charging valve in the liquid line, a process fitting in the high pressure side of the system, or a receiver outlet valve with a charg-ing port. It is recommended that liquid charging be done through a filter-drier to prevent any contami-nants from being inadvertently introduced into the system. Never charge liquid into the compressor suction or discharge service valve ports, since this can damage the compressor valves.

For original installations, the entire system should be pulled to a deep vacuum. Weigh the refrigerant drum, and attach the charging line from the refriger-ant drum to the charging valve. If the approximate weight of refrigerant required is known, or if the charge must be limited, the refrigerant drum should be placed on a scale so that the weight of refriger-ant can be checked frequently.

Purge the charging line and open the cylinder liquid valve and the charging valve. The vacuum in the system will cause liquid to flow through the charging connection until the system pressure is equalized with the pressure in the refrigerant cylinder.

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Close the receiver outlet valve and start the com-pressor. Liquid refrigerant will now feed from the refrigerant cylinder to the liquid line, and after passing through the evaporator will be collected in the condenser and receiver.

To determine if the charge is approaching the system requirement, open the receiver outlet valve, close the charging valve, and observe the system opera-tion. Continue charging until the proper charge has been introduced into the system. Again weigh the refrigerant drum, and make a record of the weight charged into the system.

Liquid-charging refrigerant through the charging valve on main liquid supply line. Note that cylin-der is safely held in inverted position on weighing scale. Liquid shut off valve on receiver would also be throttled to facilitate flow from cylinder.

Watch the discharge pressure gauge closely. A rapid rise in pressure indicates the condenser is filling with liquid, and the system pumpdown capacity has been exceeded. Stop charging from the cylinder immediately if this occurs, and open the receiver outlet valve.

On factory assembled package units utilizing welded compressors, charging is normally accom-plished by drawing a deep vacuum on the system, and introducing the proper charge by weight into the high pressure side of the system by means of a process connection which is later sealed and brazed closed. To field charge such systems, it may be necessary to install a special process fitting or charging valve, and weigh in the exact charge required.

Vapor Charging

Vapor charging is normally used when only small amounts of refrigerant are to be added to a sys-tem, possibly up to 25 pounds, although it can be more precisely controlled than liquid charging. Vapor charging is usually accomplished by means of a gauge manifold into the compressor suction service valve port. If no valve port is available—for example on welded compressors—it may be nec-essary to install a piercing valve or fitting in the suction line.

Vapor-charging refrigerant through compressor suction service valve. Gauges are connected to read both suction and discharge pressure. When adding refrigerant, discharge pressure should be observed to be sure system is not over-charged and refrigerant is not being added too rapidly. Higher than normal discharge pressure indicates either that condenser is filling with liquid or compressor is being over-loaded by too rapid charging. Charging

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manifold permits throttling of the vapor from the cylinder. Cylinder is mounted on scale to measure amount of refrigerant charged. Approved valve wrench is being used to operate cylinder valve.

Weigh the refrigerant cylinder prior to charging. Connect the gauge manifold to both suction and discharge service valves, with the common con-nection to the refrigerant cylinder. Purge the lines, open the refrigerant cylinder vapor valve, start the compressor, and open the suction connection on the gauge manifold. Modulate the rate of charging with the gauge manifold valve.

The refrigerant cylinder must remain upright with refrigerant withdrawn only through the vapor valve to insure vapor only reaching the compressor. The vaporizing of the liquid refrigerant in the cylinder will chill the liquid remaining and reduce the cyl-inder pressure. To maintain cylinder pressure and expedite charging, warm the cylinder by placing it in warm water or by using a heat lamp. Do not apply heat with a torch.

To determine if sufficient charge has been intro-duced, close the refrigerant cylinder valve and observe the system operation. Continue charging until the proper charge has been added. Again weigh the refrigerant drum and make a record of the weight charged into the system.

Watch the discharge pressure closely during the charging operation to be certain that the system is not overcharged.

How to Determine the Proper Charge

1. Weighing the Charge.

The most accurate charging procedure is to actu-ally weigh the refrigerant charged into the system. This can only be done when the system requires a full charge and the amount of charge is known. Normally such data is available on packaged unitary equipment. If the charge is small, it is common prac-tice to vent the system charge to the atmosphere if repairs are required, and add a complete new charge after repairs are complete.

2. Using A Sight Glass

The most common method of determining the proper system charge is by means of a sight glass in the liquid line. Since a solid head of liquid refrigerant is essential for proper expansion valve control, the system can be considered properly charged when a clear stream of liquid refrigerant is visible. Bubbles or flashing usually indicate a shortage of refrigerant. Bear in mind that if there is vapor and no liquid in the sight glass, it will also appear clear.

However, if the service engineer should be aware of the fact that at times the sight glass may show bubbles or flash gas even when the system is fully charged. A restriction in the liquid line ahead of the sight glass may cause sufficient pressure drop to cause flashing of the refrigerant. If the expansion valve feed is erratic or surging, the increased flow when the expansion valve is wide open can create sufficient pressure drop to create flashing at the receiver outlet. Rapid fluctuations in condensing pressure can be a source of flashing. For example, in a temperature controlled room, the sudden open-ing of shutters or the cycling of a fan can easily cause a change in condensing temperature of 10°F. to 15°F. Any liquid in the receiver may then be at a temperature higher than the saturated temperature equivalent to the changed condensing pressure, and flashing will occur until the liquid temperature is again below the saturation temperature.

Some systems may have different charge require-ments under different operating conditions. Low ambient head pressure control systems for air cooled applications normally depend on partial flooding of the condenser to reduce the effective surface area. Under such conditions a system operating with a clear sight glass under summer conditions may require a refrigerant charge twice as large for proper operation under low ambient conditions.

While the sight glass can be a valuable aid in determining the proper charge, the system per-formance must be carefully analyzed before plac-ing full reliance on it as a positive indicator of the system charge.

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3. Using A Liquid Level Indicator

On some systems, a liquid level test port may be provided on the receiver. The proper charge can then be determined by charging until liquid refriger-ant is available when the test port is cracked. With less than a full charge, only vapor will be available at the test port.

Larger receiver tanks may be equipped with a float indicator to show the level of liquid in the receiver much in the same manner as a gasoline tank gauge on an automobile.

4. Checking Liquid Subcooling

On small systems, if no other easy means of check-ing the refrigerant charge is available, a determina-tion of the liquid subcooling at the condenser outlet can be used. With the unit running under stabilized conditions, compare the temperature of the liquid line leaving the condenser with the saturation tem-perature equivalent to the condensing pressure. This provides an approximate comparison of the condensing temperature and the liquid temperature leaving the condenser. Continue charging until the liquid line temperature is approximately 5°F. below the condensing temperature under maximum load conditions. This type of charging is normally used only on factory packaged systems, but it does pro-vide a means of emergency field checking which should indicate proper system operation.

5. Charging By Superheat.

On small unitary systems equipped with capillary tubes, the operating superheat may be used to determine the proper charge.

If a service port is available so that the suction pressure can be determined, the superheat may be calculated by determining the difference be tween the temperature of the suction line approximately 6 inches from the compressor and the saturation temperature equivalent of the suction pressure. If no means of determining pressure is available, then the superheat can be taken as the difference between the suction line temperature 6 inches from the compressor and a temperature reading on an evaporator tube (not a fin) at the midpoint of the evaporator.

With the unit running at its normal operating con-dition, continue charging until the superheat as determined above is approximately 20° to 30°. A superheat approaching 10° indicates an over charged condition, a superheat approaching 40° indicates an undercharge.

6. Charging by Manufacturer’s Charging Charts.

Some manufacturers of unitary equipment have charging charts available so that the proper charge may be determined by observing the system oper-ating pressures. Follow the manufacturer’s direc-tions for determining proper charge if the unit is to be charged in this fashion.

rEMOVING rEFrIGEraNt FrOM a SYStEM

Occasionally it will be necessary to remove re-frigerant from a system. To properly remove the refrigerant, the individual servicing the unit must abide by the following guidelines. 1. Complying with Law and Regulation

During the recovery, recycle and reuse of any and all refrigerants it is imperative that one complies with current laws and regulations. It is the respon-sibility of the individual servicing the refrigeration unit to follow all current local, state, and federal laws, regulations, and ordinances. It is also their responsibility to follow any directions or guidelines that are set forth by the recovery unit equipment manufacturer.

2. Using the System Compressor

Connect the gauge manifold from the compres-sor discharge valve service port to the refrigerant container and purge the lines. Note the maximum allowable refrigerant container weight.

Place the refrigerant container in ice. Place the compressor in normal system operation. Turn the discharge service valve in a few turns to open the service port, open the refrigerant container valve and the gauge manifold so that discharge gas can enter the cold container, with the discharge pressure registering on the manifold high pressure gauge. WarNING. Do not close off the discharge valve to the condenser. A portion of the discharge

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gas will now enter the container and condense. Weigh the container frequently to check the prog-ress in filling. Continue bypassing a portion of the discharge gas into the refrigerant container until it is filled to its weight capacity. Do Not Overfill. Use additional containers as necessary.

When a major portion of the refrigerant has been removed, system pressures may fall so low that re-frigerant can no longer be efficiently transferred.

3. Using a Transfer Condensing Unit

A small, air cooled condensing unit equipped with an oil separator may be used as a scavenging or transfer pump to transfer refrigerant to storage containers. By means of a gauge manifold connect system discharge and suction service ports to the transfer pump, and connect the transfer unit liquid outlet connection to the refrigerant container.

Purge lines as previously outlined, start the trans-fer pump, and modulate the suction pressure as necessary with the gauge manifold to prevent overloading.

WarNING. Watch refrigerant cylinder weight closely. Do no overfill.

4. Charge Migration.

In the absence of a transfer condensing unit, and when the system compressor is inoperative, re-frigerant may be transferred to a storage container by migration. Evacuate the container if possible, and connect to the system by means of the gauge manifold.

Chill the refrigerant container to the lowest possible temperature. Pack in ice or dry ice if available. Open the valves so that the refrigerant can migrate from the warm and therefore higher pressure sys-tem to the cold and lower pressure cylinder. Do Not overfill.

Migration will continue until the system pressure is the equivalent of the saturated pressure of the re-frigerant at the cylinder temperature. For example, if the cylinder is 40°F. and the refrigerants R-12, migration will continue until the system pressure is approximately 37 psig.

A disadvantage of this system is the length of time required for the transfer.

HaNDLING rEFrIGEratION OIL

Oil processed for use in refrigeration compres-sors is highly refined, dewaxed, and dehydrated. In order top protect its quality, refrigeration oil is shipped in tightly sealed containers. Exposure to air and moisture for extended periods will result in contamination of the oil, and can cause harmful reactions in the compressor.

Refrigeration oils are available in sealed containers in various sizes, but should be purchased only in the sized container needed for the immediate ap-plication. It is highly recommended that oil added to a compressor be taken only from sealed containers opened at the time of use. Do not transfer oil from one container to another, and do not store in open containers. Buying oil in large containers to obtain a better price is false economy. In the long run, it will be far more costly in terms of compressor damage and customer ill will.

Compressors leaving the Emerson Climate Tech-nologies, Inc. factory are charged with Suniso 3G or 3GS, 150 SUS viscosity refrigeration oil, and the use of any other oil must be specifically cleared with the Emerson Climate Technologies, Inc. Ap-plication Engineering Department.

DEtErMINING tHE OIL LEVEL

All service compressors are shipped with a charge of the proper refrigeration oil. Normally the factory oil charge in the compressor is somewhat greater than the normal oil level required for adequate lubrication, in order to provide some allowance for oil which will be circulating in the system during operation. Depending on the system design, the amount of oil in the system at the time of compres-sor installation, oil lost due to leakage, etc., it may be necessary either to add or remove oil from a system any time it is first placed in operation with a different compressor.

On Copeland® brand compressors equipped with crankcase sightglasses, the oil level should be maintained at or slightly above the center of the sight glass while operating. An abnormally low oil

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level may result in a loss of lubrication; while an excessively high oil level may result in oil slugging and possible damage to the compressor valves or excessive oil circulation. The oil level may vary considerably on initial start-up if liquid refrigerant is present in the crankcase, and the oil level should be checked with the compressor running after having reached a stabilized condition.

Most welded hermetic compressors have no means of determining the oil level. This type of compres-sor is primarily designed for installation in factory designed, assembled, and charged systems where the oil charge can be accurately measured into the system at the time of original assembly. In the case of a leak, if the amount of oil lost is small and can be reasonably calculated, this amount should be added to the compressor. If however, there is a major loss of oil, the serviceman must remove the compressor, drain the oil, and add the correct measured charge before placing the compressor in operation.

aDDING OIL tO a COMPrESSOr

1. Open System Method

If the compressor is equipped with an oil fill hole in the crankcase, the simplest means of adding oil is to isolate the compressor crankcase, and pour or pump the necessary oil in. If the system contains no refrigerant, or the compressor is open for repairs, no special precautions are necessary other than the normal measures of keeping the oil clean and dry, since the system should be evacuated prior to start-up.

If the system contains a charge of refrigerant, close the compressor suction valve and reduce the crankcase pressure to approximately 1 to 2 psig. Stop the compressor and close the compressor discharge valve.

Remove the oil fill plug and add the required amount of oil. The residual refrigerant in the crankcase will generate a slight continuing pressure and outflow of refrigerant vapor during the period when the compressor is exposed to the atmosphere, prevent-ing the entrance of serious amounts of either air or moisture. Purge the crankcase by cracking the suction service valve off its seat for 1 or 2 seconds.

Replace the oil fill plug, open the compressor valves, and restore the system to operation.

In the case of welded compressors installed in systems without service fittings, the only means of adding oil to the compressor may be by cutting the refrigerant lines so that oil can be poured directly into the suction line since the suction connection on a welded compressor opens directly into the shell.

2. Oil Pump Method

Many servicemen have either fabricated or pur-chased a small oil pump for adding oil to compres-sors. The pump is quite similar to a small bicycle tire pump, and allows the addition of oil to an operating compressor through the service port if necessary, or can be used to add oil directly to the crankcase where space may not permit a gravity feed. When the compressor is in operation, the pump check valve prevents the loss of refrigerant, while allow-ing the serviceman to develop sufficient pressure to overcome the operating suction pressure and add oil as necessary.

3. Close System Method

In an emergency where an oil pump is not avail-able to the compressor is inaccessible, oil may be drawn into the compressor through the suction service valve.

Connect the suction connection of the gauge manifold to the compressor suction service valve, and immerse the common connection of the gauge manifold in an open container of refrigeration oil. Close the manifold valve and the compressor suction service valve and pull a vacuum in the compressor crankcase. Then open the manifold valve, drawing oil into the compressor through the suction service valve.

WarNING. Extreme care must be taken to insure the manifold common connection remains immersed in oil at all times. Otherwise air will be drawn into the compressor. On smaller horsepower or older style compressors where the suction vapor and oil are returned directly into the suction chamber, oil must be added very slowly since drainage to the crankcase may be quite slow.

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Continue as necessary until the proper amount of oil has been drawn into the compressor.

rEMOVING OIL FrOM a COMPrESSOr

Occasionally problems in line sizing or system operation may cause oil to trap in the evapora-tor or suction line, and large amounts of oil may be added to the system in an effort to maintain a satisfactory oil level in the compressor. When the basic oil logging problem is corrected, the excess oil will return to the compressor crankcase, and unless removed from the system, can cause oil slugging, excessive oil pumping, and possible compressor damage. Also in cases where the system has been contaminated, for example by a broken water tube in a water cooled condenser, or in cleaning a system after a bad motor burn, it may be necessary to completely remove the oil from the compressor crankcase.

To some extent the choice of a method for removing oil depends on the degree of system contamination. For removing excess oil or on systems with only slight contamination, almost any method is accept-able. However if the system is badly contaminated, it may be advisable to remove the compressor bottom plate and thoroughly clean the interior of the crankcase.

1. Removing by Oil Drain Plug

Some compressors are equipped with oil drain plugs. If so, this provides an easy method for removing oil.

Close the suction service valve, and operate the compressor until the crankcase pressure is reduced to approximately 1 to 2 psig. Stop the compressor and isolate the crankcase by closing the discharge service valve. Carefully loosen the oil drain plug, allowing any pressure to bleed off before the threads are completely disengaged. Drain oil to the desired level by seepage around the threads without removing the plug.

When draining is complete, tighten the drain plug, open the compressor valves, and restore the compressor to operation. The oil seal at the drain hole and the residual refrigerant pressure in the crankcase will effectively block the entrance of

any measurable quantities of air or moisture into the system.

2. Removing by Oil Fill Hole

If a drain plug is not convenient or is not furnished on the compressor, oil may be removed by means of the oil fill hole.

Close the compressor suction service valve, reduce the crankcase pressure to 1 to 2 psig, and isolate by closing the discharge service valve.

Carefully loosen the oil fill plug, allowing any pres-sure to bleed off before the threads are completely disengaged. Remove the oil fill plug, and insert a ¼ in. O. D. copper tube so that the end is at or near the bottom of the crankcase. If possible use a tube of sufficient length so that the external end can be bent down below the crankcase, thus forming a syphon arrangement. Wrap a waste rag tightly around the oil fill opening, and crack the suction service valve, pressurizing the crankcase to ap-proximately 5 psig, and then close the valve.

Oil will be forced out the drain line, and will continue to drain by the syphon effect until the crankcase is emptied. If the syphon arrangement is not pos-sible, repressurize the crankcase as necessary to remove the desired amount of oil.

The residual refrigerant pressure in the crankcase

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will prevent the entrance of any serious amounts o of moisture or air into the system. Purge the crankcase by cracking the suction service valve off its seat for 1 to 2 seconds. Reinstall the fill plug, tighten, open the compressor valves, and restore the compressor to operation.

In large systems where a large amount of excess oil must be removed, or where oil must be removed at intervals over a prolonged period, considerable time can be saved by brazing a dip tube in a valve so that oil can be removed as desired as long as the crankcase pressure is above 0 psig. (See Figure 175.) To speed up separation, the oil should be removed to a ¼ in. sight glass level. After oil removal is complete, the oil level may then be raised to the normal operating level.

3. Removal by Means of Baseplate

On accessible compressors, it may be necessary to remove the base plate if complete crankcase cleaning is necessary.

Pump the system down to isolate the compressor, remove the base plate, clean as necessary, and reinstall with new gasket. Since both air and mois-ture can enter the crankcase during this operation, the crankcase should be evacuated with a vacuum pump before restoring to operation. In an emer-gency, the crankcase may be purged by cracking the suction service valve and venting through the oil fill hole and the discharge service port. Replace the plug in the oil fill hole and jog the compressor a few times by starting and stopping, discharging through the discharge service port. Cap the dis-charge service port, open the discharge valve, and the compressor can be restored to operation.

4. Removing Oil From Welded Compressors

If the oil must be removed from a welded com-pressor, for example to recharge with a measured amount of oil, the compressor must be removed from the system, and the oil drained out the suction line stub by tilting the compressor.

After the compressor is reinstalled the system must then be evacuated by means of an access valve or the process tube before recharging with refrigerant and restoring to operation.

HaNDLING FILtEr-DrIErS

Regardless of the precautions of care taken, any time a system is opened for repair or maintenance, some amount of moisture and air enters. In order to avoid freezing of the moisture at the expan-sion valve or capillary tube, and to prevent acid formation and other detrimental system effects, the moisture level in the system must be kept at a minimum. Therefore every system opened for repair or installed in the field must have a liquid line filter-drier.

Self-contained filter-driers or replaceable drier ele-ments are factory sealed for protection. If the seal is broken and the drier is exposed to the atmosphere for more than a few minutes, the drier will pick up moisture from the atmosphere and will quickly lose much of its moisture removal ability.

The system must be sealed and evacuated within a few minutes of the installation of the drier. Leaving a system open overnight after installation of a drier may completely destroy the drier’s value.

COMPrESSOr BUrNOUtS - WHat tO DO

(Excerpts from a speech by Raymond G. Mozley, Vice President, Emerson Climate Technologies, Inc. Application Engineering)

Sometimes we get so involved in the technical details of how to solve a problem that we lose sight of the ultimate objective—how to get rid of the problem. As the old saying goes, “We can’t see the forest for the trees.”

Our objective in any refrigeration or air conditioning application is a satisfactory trouble free system. And, viewed from that standpoint, our answer to the question of compressor burnouts is at once simple and logical—prevent them before they occur. Our ultimate objective is to prevent the occurrence of a burnout, and this can only be done before a burnout occurs, not afterward.

It is true that occasionally a fault in the motor insu-lation may result in a motor burn, but in a system with proper design, manufacture, application, and installation, burnouts rarely occur. Of those that do occur, most are the result of mechanical or

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lubrication failures, resulting in the burnout as a secondary result.

If the problem is detected and corrected I time, a large percentage of compressor failures can be prevented. Periodic maintenance inspections by an alert serviceman on the lookout for abnormal operation can be a major factor in maintenance cost reduction. It is far easier, far less costly, and far more satisfactory to all parties concerned to take the few simple steps necessary to insure proper system operation than it is to allow a compressor failure to occur that could have been prevented, and then have to restore the system to satisfac-tory condition.

Probably no single type of failure has been more publicized, more studied, more debated, and more blamed for compressor failures than burnouts. As a result of this widespread publicity, the burnout problem has been the whipping boy for other sys-tem problems of far more serious proportions, and in many cases has been blown up out of proper perspective for competitive reasons. At one time several years ago, motor burnouts were a serious problem in hermetic compressors, and even today, many service engineers feel that burnouts are a major source of compressor failures. Our experi-ence certainly indicates that due to the tremendous improvements in compressor design and system practices over the past years, burnouts as a cause of system failures ceased to be a major factor several years ago.

Motor failures do occasionally occur as a result of other malfunctions in the system, most often as an after effect of a lubrication failure. This is one of the major factors which contribute to frequent field misunderstanding of the type of compressor failure which may have actually occurred. As a result many service personnel mistakenly attribute the cause of recurring failures to motor burns, whereas in reality the motor failure has been an after effect of other system difficulties.

If the service engineer is to help in eliminating un-necessary compressor failures, he must thoroughly understand both the operation of the system and possible causes of failure that might occur, and he must be on the alert for any signs of system malfunction. On the return material cards which are

attached to the compressors returned to our factory, a space is provided to note the cause of failure if known. On a great majority of the cards, there is a notation of bad compressor, compressor won’t run, motor burn, or compressor locked up, and we suspect that the majority of these are classified in the service engineer’s mind as compressor burnouts. The truth of the matter is that of the compressors returned to our factory during the warranty period, probably 65% to 75% have failed due to lack of lu-brication or damage from liquid refrigerant. Seldom do we see any notation to this effect on a return material card. We suspect that in the great major-ity of cases the serviceman did not know what the cause of failure was, and installed a replacement compressor without determining whether he had corrected the basic cause of failure.

System malfunctions rarely originate from normal operation. They may be caused by some quirk in system design, by contaminants left in the system at the time of installation, by refrigerant leaks, by the improper operation of some electrical or refrigerant control components, or from a dozen other possible causes. In many cases, it may be a long period of time before the effects of some system fault begin to affect the compressor operation. In practically every case, indications of a system malfunction are clearly evident prior to the compressor failure.

When a Burnout Occurs

But suppose, despite your best precautions, a motor burn does occur. What can you do? Hap-pily, today as a result of many years’ experience, techniques are available which make system cleaning simple, effective, relatively inexpensive, and dependable.

The type of failure which has created the most hardship on the user, and the one which has received the most publicity in recent years has been the repeat burnout type of failure, where the initial burnout has triggered a series of failures on the same system—each after decreasing period of operation. It was recognized at an early stage that contamination resulting from previous burnout, remaining in the system, was the source of the succeeding failure, but developing a dependable cure for the system was not an easy task.

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Our company has been interested in cleaning meth-ods for a long time, and because of the problems involved in the flushing process, we felt a simpler, less expensive procedure was badly needed. Field experience in removing moisture from systems had indicated that filter-driers might be the answer.

In the early 1960s in cooperation with major air conditioning and filter-drier manufacturers, we launched an intensive field trial of the filter-drier cleanout method. Basically this involved the use of approved filter-driers, incorporating an adequate desiccant in both the liquid and suction lines.

A great deal of corporate money and prestige was risked in early field tests, but it paid off not only in successfully cleaned systems, but in developing a background of test data so that we could safely recommend field proven components and know they would do the job. Several manufacturers are now producing suitable filter-driers, and many of these have been proven by field experience to be of equal value in successfully cleaning systems.

Due to its simplicity, the cost of the filter-drier cleanout system is quite inexpensive. There is no need for large quantities of refrigerant for flushing the system, no waste of man hours in laboriously cleaning each circuit, no long periods of down time. In most cases even the refrigerant in the system can be saved. This is the only practical method which can assure proper cleaning especially where long lines and multiple evaporators and circuits are involved.

This procedure has been used in thousands of installations during the past few years, and where properly used we do not know of a single instance of a second burnout due to improper cleaning. We do not feel there is any excuse today for repeat failures due to improperly cleaned systems.

We do still get occasional reports of repeat failures on systems which have experienced a motor burn. In practically every case we find the system has been improperly cleaned, in many instances because the serviceman felt the system could be cleaned merely by purging with refrigerant or by failure to use the recommended suction line filter-drier.

We feel there is no such thing as a mild burnout.

The only safe procedure is to treat every motor burn as a serious one. We do not agree with those people in the industry who feel a system which has experienced a so-called mild burn can be safely cleaned by use of filters only in the suction line in conjunction with a liquid line filter-drier. In our opinion the risks are far too great to gamble on a half-way cleaning job when the stakes are possible future costly trouble on the system, and the pos-sible savings are nominal at best.

Cleanout Procedure

The actual cleanout procedure with the filter-drier system is quite simple.

A. Save refrigerant

On any system the refrigerant charge should be saved if the volume is large enough to be worthwhile. If the compressor has service valves, there may be no need to even remove the refrigerant charge. If a separate condensing unit or transfer pump is available, flange adapters may be used to pump the system down or pump the system charge into an empty drum. If a separate condensing unit is not available, in an emergency the replacement compressor can be installed to pump the system down prior to the cleanout. Although some contami-nants will be returned to the compressor during the pumpdown procedure, the compressor will not be harmed by the short period of operation required, and the contaminants will be safely removed as they are circulated through the system after installation of the system cleaner.

B. remove Old Driers

All filter-driers previously installed in the system must be replaced, all filters or strainers cleaned or replaced, and in the event of a bad burn, refriger-ant control devices such as expansion valves and solenoid valves should be checked and cleaned if necessary.

C. Install New Filter-Driers

Adequate filter-driers of the proper size must be installed in both the liquid and suction lines. The suction line filter-drier is most important, since contaminants may not be effectively removed by a

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liquid line filter-drier alone. A pressure fitting should be provided ahead of the suction line filter-drier, preferably in the shell, to facilitate checking the pressure drop across the filter-drier.

D. Check Electrical Circuits

All electrical connections should be checked to be sure they are tight and properly made. The contactor should be examined, and any worn or pitted con-tacts should be cleaned or replaced. On externally protected motors standard service replacement compressors are normally supplied with motor protectors mounted in the terminal box. No attempt should be made to salvage the external inherent protector or external supplementary protectors mounted in the terminal box of the compressor be-ing replaced, as these might have been damaged and could contribute to another failure.

If an electrical problem was responsible for the original motor burn, and is not corrected, it can result in the loss of the replacement compressor.

E. Place In Operation

The system may then be placed in operation, but should be closely watched for at least two hours after start-up. As the contaminants in the system are filtered out, the pressure drop across filter-driers will increase. Check the pressure drop across the suction line filter-drier frequently. If the pressure drop increases to the point where it exceeds the manufacturer’s recommended maximum limit, the filter-drier should be replaced.

F. 48 Hour Check

The system may then be allowed to operate for 48 hours, at which time the color and odor of the oil should be checked. Normally the system will be adequately cleaned by this time. However, if an acid content is present, if the oil is still discolored, or has an acid odor, the filter-driers should be changed, and in the case of bad burns, the compressor oil should be changed. After an additional 48 hours of operation, the oil should be checked again, and the filter-drier change repeated until the oil remains clean, odor free, and the color approaches that of new oil. The suction line filter-drier may then be removed, preferably replaced with a permanent

suction line filter, the liquid line filter-drier should be changed, and the system can be returned to normal operation.

acid Check

Acid test kits are available from several manu-facturers for measuring the acid level in the oil. These are capable of making quite accurate acid measurements, but if they are not available, a check of the oil by sight and smell can give a quick indication if contamination remains in the system. Since refrigeration oil varies in color, a sample of the new oil in the replacement compressor should be removed prior to installation and sealed in a small glass bottle for comparison purposes. Suitable 2 ounce bottles are obtainable at any drug store. If the oil has been exposed to refrigerant, the bottle should not be tightly capped, since the residual refrigerant may create a high pressure if tightly sealed and exposed to high temperature.

Conclusion

In conclusion, let me stress again our answer to the question of compressor burnouts. Prevent them before they occur. It is impossible to view burnouts as a separate item, apart from the rest of the system. If you follow good refrigeration practice, apply the compressor properly, keep the system free of contamination, and stay on the alert for any system malfunctions, the burnout problem will take care of itself.

COMPrESSOr FaILUrES tHat COULD HaVE BEEN PrEVENtED

To enable the user and service engineer to better understand the type of damage that can occur in the compressor from improper system control or external system malfunctions, the following typical examples illustrate mechanical damage from com-pressor failures that could have been prevented.

Liquid refrigerant or Oil Slugging

Figures 176 and 177 are different views of a typical valve plate from a smaller horsepower compressor. A discharge valve and discharge valve backer as-sembly are shown in the foreground in new condition for comparison. Note how the valve backers on both

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discharge valves have been bent from the original shape. The backers are made of steel, and only the force generated by a slug of liquid refrigerant or oil would have sufficient impact to cause this distortion. Once the valve backer is bent, it is only a matter of time before the reed is broken, since the reed is then subjected to stresses beyond its designed strength.

The solution to this problem lies in proper control of liquid refrigerant, and may require the use of a suction accumulator, crankcase heaters, or a pumpdown cycle in a system with an excessively large refrigerant charge.

Carbon Formation From Heat and Contaminants

Figure 178 is a similar type valve plate showing extreme carbon formation. This occurs due to oil breakdown, and can be caused by contaminants such as moisture and air in the system, or by ex-cessively high discharge temperatures. Contrary to popular belief, carbon formation such as this can occur without a motor failure.

This will not occur in a system that is properly cleaned, dehydrated, and evacuated, with motor and discharge temperatures maintained within recommended operating limits.

Broken Discharge reed

Figure 179 shows a discharge reed which has been damaged by excessively high head pressures. Note the round hole which has been broken out of the reed, and which actually has been forced down into the discharge port. This occurs when the discharge pressure builds up to a point at which there is sufficient pressure to actually shear the steel discharge reed against the sides of the discharge port in the valve plate on the piston suction stroke. Slugging is not necessarily connected with this type of damage, and the valve backer shows no signs of distortion. In order for pressure of this magnitude to build up it is probable that either a restriction in the discharge line or liquid line—possibly in the

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refrigerant control device—may have created a hydraulic pressure condition in the compressor discharge chamber.

Head pressures must be kept under proper control for the proper operation of any system. The pres-sures required to shear a reed in this manner are far in excess of established compressor limits.

ruptured Discharge Chamber

A discharge chamber from a welded compressor which has been ruptured by excessive liquid pres-sure is shown in Figure 180. Note the relief valve on the end of the discharge chamber which will relieve pressures between discharge and suction pressures to the compressor crankcase when the difference between discharge and suction pres-sures exceeds 550 ± 50 psig. This will effectively prevent gas pressures from exceeding the relief valve setting, but liquid cannot be forced through the valve quickly enough to prevent excessive pressures. This type of damage will occur only when excessive liquid slugging is taking place, or when excessive liquid in some way enters the discharge chamber, and is normally encountered only when the system refrigerant charge exceeds the compressor charge limitation. Peak pressures in excess of 2500 psig must have been experienced to cause this failure.

To avoid this type of damage, either the compressor charge must be kept within allowable limitations, or adequate safety provisions must be provided in system design. Automatic pumpdown system control, suction line accumulators, or crankcase heaters may be required.

Connecting rod Pinhole Wear

Figures 181 and 182 illustrate progressive stages of connecting rod damage from wear of the con-necting rod pinhole. This condition occurs when the discharge valve is broken, and this piston is subjected to discharge pressure on both discharge and suction strokes. As a result, the bottom side of the pinhole is always under pressure and receives no lubrication. As the pinhole elongates, excessive play develops, and the connecting rod starts hitting the underside of the piston. Eventual rod breakage results, either at the pinhole, or at the connecting rod shaft.

This type of failure normally is caused by discharge valve breakage. Usually it originates in liquid slug-ging or excessively high discharge pressures or temperatures causing the original valve damage.

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Piston Damage Due to Lack of Lubrication

Figure 183 illustrates the condition of a piston after operating without adequate lubrication for pro-longed periods. This is most frequently encountered on low temperature applications. This can occur from excessive cylinder wall temperatures result-ing from high compression ratios and low suction pressures, or from inadequate air flow over the compressor head and body. This same condition can also result in any air conditioning or refrigera-

tion application from the continuous return of liquid refrigerant, or an excessively wet mixture of liquid and vapor returning to the crankcase, which can wash lubricant from the cylinder walls.

The oval shape resulting from piston wear is shown in Figure 184 and the eventual condition of the piston after failure due to contact with metal frag-ments is shown in Figure 185.

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Crankshaft Damage Due to Lack of Lubrication

Scoring and wear of a crankshaft due to lack of adequate lubrication is shown in Figure 186. The ridge on the one throw is due to wear from a grooved connecting rod. The heat generated in the rods and bearings can cause eventual seizure of either rods or bearings, and possible connecting rod breakage.

To avoid damage to bearings, crankshaft, pistons, and connecting rods, continuous lubrication must be maintained at all times. Repeated short periods of

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operation or prolonged periods of operation without adequate lubrication are almost certain to result in compressor failure.

Connecting rod Damage Due to Liquid Slugging

While most Copeland® brand compressors have aluminum connecting rods which usually will break rather than bend when subjected to excessive stress, some rods in older models of belt-drive compressors are made with forged steel. Figures


187 and 188 illustrate the distortion and bending of the steel rods caused by liquid slugging. This is an excellent example of the tremendous force generated by hydraulic compression when liquid refrigerant enters the cylinders.

A refrigeration compressor is designed to pump vapor only. While small amounts of liquid can be tolerated, large amounts of liquid returning to the compressor crankcase can cause major damage. Systems must be designed and applied so the compressor is not subjected to such abuse.

System Cleaning with Filer-Drier

Emerson Climate Technologies, Inc. recommends only the filter-drier cleaning procedure in the case of a motor burn. Basically this involves the use of approved filter-driers incorporating an adequate desiccant (not a filter only) in both the liquid and suction lines.

To illustrate the effectiveness of a suction line filter-drier, 26 ounces of badly contaminated oil were put in an operating system with a 10 H. P. compressor. Bottle #1 shows a sample of oil removed from the system prior to the test. Bottle #2 is a sample of the contaminated oil introduced into the system, which was then allowed to operate for 24 hours without a filter-drier. The system was then equipped with a suction line filter-drier, and a sample of the oil taken from the compressor at this time is shown in Bottle #3. Sample #4 was taken from the crankcase one hour after the filter-drier was installed, and Sample #5 was taken from the crankcase 72 hours after the filter-drier was installed. The oil has been ef-fectively cleaned so that the color and appearance are equal to the original oil.

This system cleaning procedure has been used in thousands of installations during the past few years, and when this procedure has been prop-erly followed, we do not know of a single instance where a second failure has resulted because of improper cleaning.

Discharge Valves Damaged From Slugging

Figure 190 shows a comparison of a new valve plate of the type used on larger compressors with a valve plate on which the discharge reeds and

discharge reed backstops have been damaged from liquid refrigerant or oil slugging. The backstops are made of ⅛” hardened steel, and the distortion clearly illustrates the tremendous force exerted by the liquid or oil slug.

Note that on the valves on which the backstops have been badly bent, the discharge reed has been broken, due to the resulting excessive stress on the reed.

PrEVENtIVE MaINtENaNCE

The question is frequently asked as to how long it takes a compressor to wear out. It is almost im-possible to answer that question because seldom if ever does a compressor fail from wear due to

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normal operation. Almost invariably a compres-sor failure results from malfunctions in either the refrigerant or electrical system, or from system operating conditions beyond the limitations of the system design.

Just what does this mean in terms of preventive maintenance? In practically every case, indications of a system malfunction are clearly evident prior to the compressor failure. If the problem is detected and corrected in time, a large percentage of com-pressor failures can be prevented. If the inspector is alert and on the lookout for any indication that operation of the system is in any way abnormal, a periodic maintenance inspection can be a major factor in maintenance cost reduction. Inspections should be made at least three times per year, and more frequent inspection is recommended during heavy usage periods.

Following is a summary of the major items to be checked.

Check With Operating Personnel

Always check with the operating personnel who are using the equipment to see if there have been any reports of abnormal or erratic operation. Fre-quently indications of abnormal operation may be observed by operating people who do not realize their significance, and this information may never be given to the service personnel unless brought out by specific questions concerning system opera-tion. Ask particularly about trips of the oil pressure safety control, or other safety devices.

Operating Pressures and temperatures

If permanent gauges are available, check the com-pressor suction and discharge pressures to be sure they are within the normal range for the application and the temperature of the condensing medium. If there are any indications of abnormal operation such as short cycling on pressure controls or excessive compressor temperatures, use a gauge manifold to check the operating pressures on systems without permanently installed gauges.

If abnormal operating pressures are found, the cause must be found and the malfunction cor-rected.

Check the compressor head temperature by touch. An abnormally cool head can indicate a broken valve, a broken connecting rod, or excessive liquid refrigerant flood back. An abnormally hot head can indicate a broken discharge valve, a blown or improper head gasket, or inadequate compressor cooling.

Oil Level and Condition

On Copeland® brand compressors, the oil level should be at or slightly above the center of the sight glass. It should be kept in mind that some slight fluctuation in oil level may occur during an operating cycle—particularly before and after defrost periods. So long as the oil level is maintained well within the sight glass such fluctuations are not harmful.

If the oil is black in color, the crankcase should be drained and the oil replaced. If there has been a recent compressor failure on the system and the oil has an acid odor, a fresh filter-drier should be installed in the suction line and left in the line for a period of 48 hours. If the oil is still discolored, the suction line filter-drier is still discolored, the suction line filter-drier element should again be changed. This procedure should be continued until the oil remains clean, odor free, and the color approaches that of new oil. The filter-drier element may then be replaced with a permanent type suction line filter.

It is recommended that only Suniso 3G or 3GS oil be used in Copeland® brand compressors. Unless there are reasons as outlined above for changing the oil, the refrigeration oil should not be changed. It does not deteriorate or wear out with normal usage.

System refrigerant Charge

All systems must have a full head of liquid refrigerant at the expansion valve to insure proper operation. A clear sight glass indicates an adequate charge. Bubbles or flashing in the sight glass may indicate a shortage of refrigerant, but flashing can also be caused by a liquid line restriction, hunting expansion valves, sudden changes in condensing pressure, and rapid changes in the refrigeration load. If there is a doubt as to the refrigerant charge, check the liquid level in the receiver. If no test cock is available, pass a torch flame momentarily back and forth on

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the receiver. If the metal remains relatively cool, a liquid level is indicated, but if the metal heats up rapidly, vapor is indicated. The liquid level can be determined by the point where the temperature change occurs.

Units with roof mounted condensers equipped with low ambient head pressure controls will require a great deal more refrigerant for low ambient condi-tions, since the head pressure is normally main-tained by partially flooding the condenser.

Too little refrigerant can result in lack of refrigeration, loss of oil in the evaporator, and overheating of the compressor. Too much refrigerant can contribute to high discharge pressures, liquid refrigerant flooding, liquid slugging, with resulting compressor lubrication problems.

Special care should be taken in finding and repair-ing any leaks if a loss of refrigerant occurs.

System Control Settings

If there is any question as to the proper operation of the low pressure control, high pressure control, or oil pressure safety control, the pressure setting should be checked. The accuracy of indicating scales on refrigeration pressure controls is not dependable, and if operation is questionable, the control should be checked with serviceman’s gauges.

Do not set the low pressure control below the rec-ommended operating limits of the compressor.

The cause of any short cycling condition must be corrected.

If the operation of an oil pressure safety controls is questionable, it should be checked by running a jumper connection across the pressure contacts to determine if the control will trip.

Liquid Line Filter-Drier

Check the color code of the moisture indicator. A positive moisture indication indicates the filter-drier should be replaced.

If the drier is frosted or if there is a perceptible temperature change between the liquid line enter-

ing and leaving the drier, an excessive pressure drop in the drier is indicated, and the drier or drier element should be replaced.

Vibration Eliminators

If the wire braid cover on a metal vibration elimina-tor is starting to pull out of the brazed end connec-tors, the vibration eliminator should be replaced to prevent possible rupture, loss of the refrigerant charge, and potential personal injury.

Capillary tubes and refrigerant Lines

Check all capillary lines for wear and vibration. Tape or support as necessary. Check refrigerant line supports and braces to make certain they are not wearing or cutting the refrigerant lines.

Oil traces at flare nuts or valve connections indicate the possibility of a refrigerant leak. Wipe clean and tighten the flare nut.

Liquid refrigerant Control

Check for any indications of liquid refrigerant flood-ing such as sweating or frosting of the compressor, rust on the suction service valve or compressor body, tripping of the oil pressure safety control, audible slugging, or excessive foaming in the crankcase. If there is any question as to liquid control, the operation of the system immediately after a defrost cycle should be observed. Exces-sive sweating or frosting of the suction line and/or compressor body must be corrected.

If the refrigerant cannot be properly controlled with the existing system controls, a suction line accumulator may be required.

Suction Line Filter

Check pressure drop across suction line filter, and replace element if pressure drop exceeds manufacturer’s recommended maximum.

Electrical Control Panel

Check the electrical control panel to see that heaters or motor protectors are not jumpered. Look for burn marks on the cabinet that might indicate possible

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shorts, and check the contacts on any contactor on which there is any question.

air Cooled Machine room

Check the exhaust fan and fan motor, and lubricate if necessary. Check operation of dampers and lou-vers, and lubricate as necessary. Run fan through on and off cycle by means of thermostat.

remote air Cooled Condenser

Check belt condition, and lubricate motor and shaft bearings. Clean condenser face if necessary. In-spect all line supports for vibration and line wear.

Walk-in Coolers, Freezers, refrigerated Fix-tures

Check coils for ice build-up and cleanliness. Check temperatures being maintained in refrigerated space. Inspect door latches, gaskets, moulding, etc.

air Conditioning air Handlers

Check filters and change if necessary. Check belt condition and lubricate motor and shaft bearings.

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SECtION 27USEFUL ENGINEErING Data

The following reference tables and charges cover miscellaneous data and conversion factors frequently required in engineering calculations. Data specifically pertaining to refrigeration has been included where appropriate in previous sections.


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