Piping

Refrigerant Piping

One of the Fundamental Series

A publication of Trane, a business

of American Standard Companies

Preface

© 2002 American Standard Inc. All rights reserved

ii TRG-TRC006-EN

Trane believes that it is incumbent on manufacturers to serve the industry by

regularly disseminating information gathered through laboratory research,

testing programs, and field experience.

The Trane Air Conditioning Clinic series is one means of knowledge sharing. It

is intended to acquaint a technical audience with various fundamental aspects

of heating, ventilating, and air conditioning (HVAC). We have taken special care

to make the clinic as uncommercial and straightforward as possible.

Illustrations of Trane products only appear in cases where they help convey the

message contained in the accompanying text.

This particular clinic introduces the reader to refrigerant piping.

Refrigerant Piping

notes

period one

Refrigerant Piping Requirements

The focus of this clinic is on the design and installation of the interconnecting

piping for vapor-compression refrigeration systems. Reviewing the physical

changes that the refrigerant undergoes within the refrigeration cycle will help

demonstrate certain demands that the piping design must meet.

This clinic focuses on systems that use Refrigerant-22 (R-22). While the general

requirements are the same for systems that use other refrigerants, velocities

and pressure drops will differ.

Figure 3 illustrates a basic vapor-compression refrigeration cycle. Refrigerant

enters the evaporator in the form of a cool, low-pressure mixture of liquid and

vapor (A). Heat is transferred to the refrigerant from the relatively warm air that

is being cooled, causing the liquid refrigerant to boil. The resulting refrigerant

vapor (B) is then pumped from the evaporator by the compressor, which

increases the pressure and temperature of the vapor.

period one

Refrigerant Piping Requirements

Refrigerant Piping Requirements

The resulting hot, high-pressure refrigerant vapor (C) enters the condenser

where heat is transferred to ambient air, which is at a lower temperature than

the refrigerant. Inside the condenser, the refrigerant vapor condenses into a

liquid and is subcooled. This liquid refrigerant (D) then flows from the

condenser to the expansion device. This device creates a pressure drop that

reduces the pressure of the refrigerant to that of the evaporator. At this low

pressure, a small portion of the refrigerant boils (or flashes), cooling the

remaining liquid refrigerant to the desired evaporator temperature. The cool

mixture of liquid and vapor refrigerant (A) enters the evaporator to repeat the

cycle.

The vapor-compression refrigeration cycle, and the four major components of

the refrigeration system (evaporator, compressor, condenser, and expansion

device), are discussed in more detail in separate clinics. Refer to the list of

references at the end of Period Six.

These individual components are connected by refrigerant piping. The suction

line connects the evaporator to the compressor, the discharge line connects the

compressor to the condenser, and the liquid line connects the condenser to the

expansion device. The expansion device is typically located at the end of the

liquid line, at the inlet to the evaporator.

There is more to the design of refrigerant piping than moving refrigerant from

one component to another. Regardless of the care exercised in selection and

application of the components of the refrigeration system, operational

problems may be encountered if the interconnecting piping is improperly

designed or installed.

Interconnecting Refrigerant Piping

compressor

condenser

evaporator

expansion

device

discharge

line

suction

line

liquid

line

Refrigerant Piping Requirements

notes

When a refrigeration system includes field-assembled refrigerant piping to

connect two or more of the components, the primary design goals are generally

to maximize system reliability and minimize installed cost. To accomplish these

two goals, the design of the interconnecting refrigerant piping must meet the

following requirements:

 Return oil to the compressor at the proper rate, at all operating conditions

 Ensure that only liquid refrigerant (no vapor) enters the expansion device

 Minimize system capacity loss that is caused by pressure drop through the

piping and accessories

 Minimize the total refrigerant charge in the system to improve reliability and

minimize installed cost

The first requirement is to ensure that oil is returned to the compressor at all

Refrigerant Piping Requirements

▲ Return oil to compressor

▲ Ensure that only liquid refrigerant enters

the expansion device

▲ Minimize system capacity loss

▲ Minimize refrigerant charge

Figure 5

Scroll Compressor

stationary

scroll

driven

scroll

intake

discharge

intake

discharge

port

motor

shaft

seal

Figure 6

4 TRG-TRC006-EN

notes

period one

Refrigerant Piping Requirements

operating conditions. Oil is used to lubricate and seal the moving parts of a

compressor. For example, the scroll compressor shown in Figure 6 on page 3

uses two scroll configurations, mated face-to-face, to compress the refrigerant

vapor. The tips of these scrolls are fitted with seals that, along with a thin layer

of oil, prevent the compressed refrigerant vapor from escaping through the

mating surfaces. Similarly, other types of compressors also rely on oil for

lubrication and for providing a seal when compressing the refrigerant vapor.

Characteristically, some of this lubricating oil is pumped along with the

refrigerant throughout the rest of the system. While this oil has no function

anywhere else in the system, the refrigerant piping must be designed and

installed so that this oil returns to the compressor at the proper rate, at all

operating conditions.

Returning to the system schematic, droplets of oil are pumped out of the

compressor along with the hot, high-pressure refrigerant vapor. The velocity of

the refrigerant inside the discharge line must be high enough to carry the small

oil droplets through the pipe to the condenser.

Inside the condenser, the refrigerant vapor condenses into a liquid. Liquid

refrigerant and oil have an affinity for each other, so the oil easily moves along

with the liquid refrigerant. From the condenser, this mixture of liquid refrigerant

and oil flows through the liquid line to the expansion device.

Next, the refrigerant–oil mixture is metered through the expansion device into

the evaporator, where the liquid refrigerant absorbs heat and vaporizes. Again,

the velocity of the refrigerant vapor inside the suction line must be high enough

to carry the droplets of oil through the pipe back to the compressor.

Without adequate velocity and proper pipe installation, oil may be trapped out

in the system. If this condition is severe enough, the reduced oil level in the

compressor could cause lubrication problems and, potentially, mechanical

failure.

Return Oil to Compressors

compressor

condenser

evaporator

expansion

device

discharge

line

suction

line

liquid

line

warm

liquid

hot

vapor

cool

vapor

Figure 7

TRG-TRC006-EN 5

period one

Refrigerant Piping Requirements

notes

The second requirement of the refrigerant piping design is to ensure that only

liquid refrigerant enters the expansion device. There are several types of

expansion devices, including expansion valves (thermostatic or electronic),

capillary tubes, and orifices.

In addition to maintaining the pressure difference between the high-pressure

(condenser) and low-pressure (evaporator) sides of the system, a

thermostatic expansion valve (TXV) also controls the quantity of liquid

refrigerant that enters the evaporator. This ensures that the refrigerant will be

completely vaporized within the evaporator, and maintains the proper amount

of superheat in the system.

Inside the condenser, after all of the refrigerant vapor has condensed into

liquid, the refrigerant is subcooled to further lower its temperature. This

subcooled liquid refrigerant leaves the condenser (A) and experiences a

pressure drop as it flows through the liquid line and accessories, such as a filter

Thermostatic Expansion Valve (TXV)

external equalizer

remote

bulb

evaporator

TXV

refrigerant

vapor

distributor

liquid

refrigerant

Figure 8

Subcooling

A B

pressure

enthalpy

subcooling

mixture of liquid

and vapor

saturated

vapor curve

saturated

liquid curve

expansion

device

condenser

evaporator

compressor

Figure 9

6 TRG-TRC006-EN

notes

period one

Refrigerant Piping Requirements

drier and solenoid valve, installed upstream of the TXV. On the pressureenthalpy chart, Figure 9 on page 5, this moves the condition of the refrigerant

toward the saturated liquid curve (B). If this pressure drop is high enough, or if

the refrigerant has not been subcooled enough by the condenser, a small

portion of the refrigerant may boil (or flash), resulting in a mixture of liquid and

vapor (C) entering the expansion device.

The presence of refrigerant vapor upstream of the expansion device is very

undesirable. Bubbles of vapor displace liquid in the port of the TXV, reducing

the flow rate of liquid through the valve, therefore substantially reducing the

capacity of the evaporator. This results in erratic valve operation.

The design of the piping system must ensure that only liquid refrigerant (no

vapor) enters the expansion device. This requires that the condenser provide

adequate subcooling at all system operating conditions, and that the pressure

drop through the liquid line and accessories not be high enough to cause

flashing. Subcooling allows the liquid refrigerant to experience some pressure

drop as it flows through the liquid line, without the risk of flashing.

The third requirement of the refrigerant piping design is to minimize system

capacity loss. To achieve the maximum capacity from the system, the

refrigerant must circulate through the system as efficiently as possible. This

involves minimizing any pressure drop through the piping and other system

components.

Whenever a fluid flows inside a pipe, a characteristic pressure drop is

experienced. Pressure drop is caused by friction between the moving liquid (or

vapor) and the inner walls of the pipe. The total pressure drop depends on the

pipe diameter and length, the number and type of fittings and accessories

installed in the line, and the mass flow rate, density, and viscosity of the

refrigerant.

As an example, the chart in Figure 10 demonstrates the impact of pressure

drop, through the suction line, on the capacity and efficiency of the system. For

this example system operating with Refrigerant-22, increasing the total

Pressure Drop in a Suction Line

period one

Refrigerant Piping Requirements

notes 

pressure drop in the suction line from 3 psi (20.7 kPa) to 6 psi (41.4 kPa)

decreases system capacity by about 2.5 percent and decreases system

efficiency by about 2 percent.

This reveals a compromise that the system designer must deal with. The

diameter of the suction line must be small enough that the resulting refrigerant

velocity is sufficiently high to carry oil droplets through the pipe. However, the

pipe diameter must not be so small that it creates an excessive pressure drop,

reducing system capacity too much.

The first three requirements have remained unchanged for many years.

However, years of observation and troubleshooting has revealed that the lower

the system refrigerant charge, the more reliably the system performs.

Therefore, a fourth requirement has been added for the design of refrigerant

piping: minimize the total amount of refrigerant in the system. To begin with,

this involves laying out the shortest, simplest, and most-direct pipe routing. It

also involves using the smallest pipe diameter possible, particularly for the

liquid line because, of the three lines, it impacts refrigerant charge the most.

The chart in Figure 11 shows that the liquid line is second only to the condenser

in the amount of refrigerant it contains.

This reveals another compromise for the system designer. The diameter of the

liquid line must be as small as possible to minimize the total refrigerant charge.

However, the pipe diameter cannot be small enough to create an excessive

pressure drop that results in flashing before the liquid refrigerant reaches the

expansion device.

Minimize Refrigerant Charge

liquid line suction line

condenser

evaporator

filter

drier

compressor

discharge

line

period one

Refrigerant Piping Requirements

This clinic discusses the processes for sizing the interconnecting piping in an

air-conditioning system. Some of the information required for selecting the

optimal line sizes is best known by the manufacturer. Therefore, if the

manufacturer of the refrigeration equipment provides recommended line sizes,

or tools for selecting the optimal line sizes, we recommend that you use those

line sizes.

If, however, line sizes are not provided by the manufacturer, the processes

outlined within this clinic could be used for selecting the sizes.

Before discussing the design and installation of the suction, discharge, and

liquid lines, there are some general requirements that apply to all of these lines.

First, copper tubing is typically used for refrigerant piping in air-conditioning

systems. This tubing is available in various standard diameters and wall

thicknesses. The nominal diameter of the tubing is expressed in terms of its

Involve the Manufacturer

▲ If provided, use refrigerant

line sizes recommended by

manufacturer

Figure 12

General Piping Requirements

▲ Use clean Type L copper tubing

◆ Copper-to-copper joints: BCuP-6 without flux

◆ Copper-to-steel (or brass) joints: BAg-28, non-acid flux

▲ Properly support piping to account for

expansion, vibration, and weight

▲ Avoid installing piping underground

▲ Test entire refrigerant circuit for leaks

Figure 13

TRG-TRC006-EN 9

period one

Refrigerant Piping Requirements

notes outside diameter. This tubing must be completely free from dirt, scale, and

oxide. New Type L or Type ACR tubing that has been cleaned by the

manufacturer and capped at both ends is recommended for air-conditioning

applications.

The piping system is constructed by brazing copper tubes and fittings together.

When brazing copper-to-copper joints, use BCuP-6* without flux. For copper-tosteel or copper-to-brass joints, use BAg-28* with a non-acid flux.

The refrigerant piping must be properly supported to account for expansion,

vibration, and the total weight of the piping. When a pipe experiences a

temperature change, it is subject to a certain amount of expansion and

contraction. Because the refrigerant piping is connected to the compressor,

vibration forces are transmitted to the piping itself. Finally, the weight of the

refrigerant-filled pipe and fittings must be supported to prevent the pipes from

sagging, bending, or breaking.

Avoid installing refrigerant piping underground. It is very difficult to maintain

cleanliness during installation or to test for leaks. If underground installation is

unavoidable, each line must be insulated separately, and then the lines must be

waterproofed and protected with a hard casing (such as PVC).

After the piping has been installed, the entire refrigeration circuit must be

tested for leaks before it can be charged with refrigerant. This process typically

involves pressurizing the entire piping system with dry nitrogen to examine

each brazed joint for leaks.

Each of these issues is discussed in greater detail in the Trane Reciprocating

Refrigeration Manual.

* Based on the American Welding Society’s (AWS) Specification for Filler

Metals for Brazing and Braze Welding, publication A5.8–1992

10 TRG-TRC006-EN

notes

period two

Suction Line

The first line to be considered is the suction line. Again, this pipe conducts lowpressure refrigerant vapor from the evaporator to the compressor.

Requirements for Sizing and Routing

The diameter of the suction line must be small enough that the resulting

refrigerant velocity is sufficiently high to carry oil droplets, at all steps of

compressor unloading. If the velocity in the pipe is too high, however,

objectionable noise may result. Also, the pipe diameter should be as large as

possible to minimize pressure drop and thereby maximize system capacity and

efficiency.

period two

Suction Line

Refrigerant Piping

Figure 14

suction line

Requirements for Sizing and Routing

▲ Ensure adequate velocity to return oil to

compressor at all steps of unloading

▲ Avoid excessive noise

▲ Minimize system capacity and efficiency loss

Figure 15

TRG-TRC006-EN 11

period two

Suction Line

notes

It may be helpful to compare the old “rules” for selecting the diameter of the

suction line with the newer rules that result from changes in compressor

technology, recent research, and the additional requirement to minimize

system refrigerant charge.

In the past, many suction lines for systems operating with Refrigerant-22 were

sized to ensure that the minimum velocity in a vertical suction riser was more

than 1,000 fpm (5 m/s), and the minimum velocity in a horizontal section was

more than 500 fpm (2.5 m/s). Actually, the minimum allowable velocity in a

suction riser depends on the diameter of the pipe.

The minimum velocity required to carry oil droplets up a vertical riser is higher

for a larger diameter pipe than it is for a smaller diameter pipe. This is due to

the velocity profile of the refrigerant flowing inside the pipe. In a smaller

diameter pipe, the higher-velocity refrigerant is closer to the inner walls of the

pipe than it is in a larger-diameter pipe. For instance, while the minimum

allowable velocity in a 2 1/8 in. (54 mm)-diameter suction riser is approximately

1,000 fpm (5 m/s), the minimum velocity in a 1 1/8 in. (28 mm)-diameter riser is

only 700 fpm (3.6 m/s). While the old minimum-velocity limits were easy to

remember, they may lead to the unnecessary use of double suction risers.

The recommended maximum-velocity limit of 4,000 fpm (20 m/s) has not

changed. A higher velocity inside the suction line may cause objectionable

noise for those nearby.

Another common rule, in a system operating with R-22, was to limit the

pressure drop through the suction line to 3 psi (20 kPa). Although this was often

thought to be a maximum limit, this value was originally intended to be only a

recommendation, or guideline, for minimizing capacity and efficiency loss.

Today, architects and HVAC system design engineers are placing the

components of the refrigeration system farther apart, and this 3 psi (20 kPa)

limit is often overly restrictive. Longer line lengths and the associated higher

pressure drop can be tolerated, assuming that the loss of system capacity and

efficiency is acceptable for the given application. Of course, it is still good

suction line

Sizing “Rules”

minimum velocity

1,000 fpm (5 m/s) based on diameter of riser

for vertical risers

500 fpm (2.5 m/s) percentage of minimum

for horizontal sections riser velocity

maximum velocity

4,000 fpm (20 m/s) 4,000 fpm (20 m/s)

maximum pressure drop

3 psi (20 kPa) based on specific system

requirements

old rules for R-22 new rules for R-22

Figure 16

notes

period two

Suction Line

practice to minimize pressure drop, but an arbitrary limit places an unnecessary

restriction on the system designer.

Sizing Process

Following are the steps to follow when selecting the proper diameter of the

suction line:

1) Determine the total length of suction-line piping.

2) Calculate the refrigerant velocity at both maximum and minimum system

capacities.

3) Select the largest pipe diameter that will result in acceptable refrigerant

velocity at both maximum and minimum capacities.

4) Calculate the total “equivalent” length of piping by adding the actual length

of straight pipe to the equivalent length of any fittings to be installed in the

suction line.

5) Determine the pressure drop (based on the total equivalent length) due to

the straight pipe and fittings.

6) Add the pressure drop due to any accessories installed in the suction line.

To begin with, the refrigerant piping should be routed in the shortest and

simplest manner possible, minimizing the total length of piping. From the initial

layout, the total measured length of the suction line can be estimated.

suction line

Process for Sizing

1 Determine total length of piping

2 Calculate refrigerant velocity at maximum and

minimum capacities

3 Select largest pipe diameter that results in

acceptable velocity at both maximum and

minimum capacities

4 Calculate total equivalent length of straight pipe

and fittings

5 Determine pressure drop due to pipe and fittings

6 Add pressure drop due to accessories

period two

Suction Line

notes

If the system contains more than one independent refrigerant circuit, each

circuit requires its own set of refrigerant lines. Therefore, the capacity of each

individual circuit must be considered separately.

Some refrigeration circuits include only one compressor that cycles on and off.

This is very common in residential and light-commercial air-conditioning

systems. In this case, the refrigeration circuit only operates at one capacity—

fully on—so only the maximum system capacity needs to be considered.

However, if the circuit contains a compressor that is capable of unloading, such

as a single reciprocating compressor with cylinder unloaders, or if more than

one compressor is manifolded together on a single circuit, such as multiple

scroll compressors, then the minimum capacity of the circuit must also be

determined. When the circuit unloads, less refrigerant flows through the

system and the refrigerant velocity inside the piping is reduced.

Recall that the diameter of the suction line must provide adequate velocity at

both maximum and minimum capacities. At maximum capacity, the refrigerant

velocity through the suction line will be the highest. Therefore, maximum

capacity is important to ensure that the refrigerant velocity is below the upper

limit of 4,000 fpm (20 m/s). At minimum capacity, the refrigerant velocity will be

the lowest. Therefore, the minimum capacity of the refrigeration circuit is

critical for ensuring that the refrigerant velocity is high enough to properly

return oil to the compressor.

It is important to note that the diameter of a vertical riser does not necessarily

need to be the same as the diameter of the horizontal or vertical drop sections

of pipe. The horizontal or vertical drop sections can often be selected one

diameter larger than a vertical riser, reducing the overall pressure drop due to

the suction line. This will be demonstrated later in this period.

Unloading Refrigeration Circuits

reciprocating compressor

with cylinder unloaders

scroll compressors

manifolded on a

single circuit

notes

period two

Suction Line

The refrigerant velocity inside a pipe depends on the mass flow rate and

density of the refrigerant, and on the inside diameter of the pipe. The chart in

Figure 19 shows the velocity of R-22 inside pipes of various diameters at one

particular operating condition—40°F (4.4°C) saturated suction temperature,

125°F (51.7°C) saturated condensing temperature, 12°F (6.7°C) of superheat,

15°F (8.3°C) of subcooling, and 70°F (38.9°C) of compressor superheat. For an

example system with an evaporator capacity of 20 tons (70.3 kW), the

refrigerant velocity inside a 2 1/8 in. (54 mm)-diameter pipe at this condition is

about 1,850 fpm (9.4 m/s).

The easiest and most accurate method for determining refrigerant velocity is to

use a computer program that can calculate the velocity for various pipe sizes

based on actual conditions. However, if you do not have access to such a

program, a chart like this may be useful.

Assume that this example 20-ton (70.3-kW) system contains one refrigeration

velocity, fpm (m/s)

evaporator capacity, tons (kW)

R-22

suction line

Determine Refrigerant Velocity

pipe diameter, in. (mm)

Figure 19

suction line

Determine Refrigerant Velocity

pipe velocity, fpm (m/s)

diameter,

in. (mm)

20 tons

(70.3 kW)

10 tons

(35.2 kW)

1 1/8 (28) 7,000 (35.6) 3,500 (17.8)

1 3/8 (35) 4,600 (23.4) 2,300 (11.7)

1 5/8 (42) 3,250 (16.5) 1,625 (8.3)

2 1/8 (54) 1,850 (9.4) 925 (4.7)

2 5/8 (67) 1,200 (6.1) 600 (3.1)

3 1/8 (79) 850 (4.3) 425 (2.2)

period two

Suction Line

notes circuit with two steps of capacity. Maximum system (evaporator) capacity is 20 tons (70.3 kW) and the circuit can unload to 10 tons (35.2 kW) of capacity.

Using the chart in Figure 19 on page 14, the refrigerant velocity at both

maximum and minimum capacities is determined for several pipe diameters.

After these velocities have been determined, the largest acceptable pipe

diameter is selected to minimize the overall pressure drop due to the suction

line.

When this system operates at maximum capacity, use of either the 1 1/8 in.

(28 mm)- or the 1 3/8 in. (35 mm)-diameter pipes results in a refrigerant velocity

that is greater than the recommended upper limit of 4,000 fpm (20 m/s). Again,

these high velocities may cause objectionable noise, so these pipe sizes should

probably not be considered.

Figure 21 shows the minimum allowable refrigerant velocity, for both a vertical

suction riser and a horizontal (or vertical drop) section of suction line, for each

standard pipe diameter. As mentioned earlier in this period, the minimum

allowable velocity in a suction riser depends on the diameter of the pipe. The

minimum velocity for a horizontal, or vertical drop, section is 75 percent of the

minimum allowable velocity for a vertical riser of the same diameter.

The minimum velocities listed in this table assume a worst-case operating

condition of 20°F (-6.7°C) saturated suction temperature. This provides a safety

factor, because a system will probably operate at this type of condition at some

time in its life.

suction line

Minimum Allowable Velocities

Figure 21

16 TRG-TRC006-EN

notes

period two

Suction Line

Based on the minimum allowable velocities listed in Figure 21 on page 15, a

2 1/8 in. (54 mm) pipe will result in acceptable velocity, at both maximum and

minimum capacities, for the horizontal and vertical drop sections of this

example suction line. One size larger would result in a velocity that is too low

when the system operates at minimum capacity.

When the circuit unloads to 10 tons (35.2 kW), however, the velocity inside a

2 1/8 in. (54 mm) pipe—925 fpm (4.7 m/s)—will drop below the minimum

allowable velocity for a vertical riser of this diameter—980 fpm (5.0 m/s).

Downsizing all vertical suction risers to 1 5/8 in. (42 mm)-diameter pipe will

result in acceptable velocity at both maximum and minimum capacities. At

minimum capacity, the velocity—1,625 fpm (8.3 m/s)—is above the minimum

allowable velocity for a riser of this diameter—840 fpm (4.3 m/s). At maximum

capacity, the velocity—3,250 fpm (16.5 m/s)—is below the 4,000 fpm (20 m/s)

upper limit.

Select Suction Line Size

riser horiz/drop

20 (70.3) 1,850 (9.4)

10 (35.2) 925 (4.7)

circuit

capacity, minimum velocity for 2 1/8 in.

(54 mm) pipe, fpm (m/s) velocity inside

2 1/8 in. (54 mm)

tons (kW)

pipe, fpm (m/s)

980 (5.0) 735 (3.7)

riser horiz/drop

20 (70.3) 3,250 (16.5)

10 (35.2) 1,625 (8.3)

circuit

capacity, minimum velocity for 1 5/8 in.

(42 mm) pipe, fpm (m/s) velocity inside

1 5/8 in. (42 mm)

tons (kW)

pipe, fpm (m/s)

840 (4.3) 630 (3.2)

Figure 22

TRG-TRC006-EN 17

period two

Suction Line

notes

For this example system, a 2 1/8 in. (54 mm)-diameter pipe can be used for the

horizontal and vertical drop sections of the suction line, and a 1 5/8 in. (42 mm)-

diameter pipe should be used for the vertical riser. This combination of pipe

sizes will ensure proper oil movement at both maximum and minimum system

capacities, and minimize the overall pressure drop due to the suction line.

This process suggests selecting the largest-possible pipe diameter to minimize

pressure drop. For some applications, however, using a smaller pipe diameter

may be possible and can reduce the installed cost of the system. Of course, the

smaller pipe diameter will result in a higher pressure drop and must meet the

minimum and maximum velocity requirements.

For this example system, for the vertical riser, a single pipe diameter was

selected that provides adequate velocity at minimum capacity without

exceeding the upper velocity limit at maximum capacity. This is generally

possible when the minimum capacity of the refrigeration circuit is not less than

30 to 37 percent of maximum capacity.

Selected Suction Line Sizes

compressor

evaporator

2 1/8 in.

(54 mm)

1 5/8 in. (42 mm)

riser

2 1/8 in.

(54 mm)

drop

Figure 23

18 TRG-TRC006-EN

notes

period two

Suction Line

If the refrigeration circuit unloads to less than 30 to 37 percent of maximum

capacity, a double suction riser may be required.

Using the same example 20-ton (70.3-kW) system, assume that the refrigeration

circuit can, instead, unload to 5 tons (17.6 kW). At this minimum capacity, the

velocity inside the 1 5/8 in. (42 mm) pipe—810 fpm (4.1 m/s)—will drop below

the minimum allowable velocity for a riser of this diameter—840 fpm (4.3 m/s).

Selecting the smaller 1 3/8 in. (35 mm)-diameter pipe for the vertical riser would

result in a velocity that is too high—4,600 fpm (23.4 m/s)—at maximum

capacity.

A double suction riser is constructed of a larger-diameter riser with a trap at the

base and a smaller-diameter riser in parallel. At maximum capacity, the

refrigerant vapor flows up both risers at velocities that are adequate to carry the

oil droplets.

Double Suction Riser

20 (70.3) 3,250 (16.5)

5 (17.6) 810 (4.1)

circuit

capacity,

tons (kW)

minimum velocity for

1 5/8 in. (42 mm) riser,

fpm (m/s)

velocity inside

1 5/8 in. (42 mm)

pipe, fpm (m/s)

840 (4.3)

20 (70.3) 4,600 (23.4)

5 (17.6) 1,150 (5.9)

circuit

capacity,

tons (kW)

minimum velocity for

1 3/8 in. (35 mm) riser,

fpm (m/s)

velocity inside

1 3/8 in. (35 mm)

pipe, fpm (m/s)

780 (4.0)

Figure 24

Double Suction Riser

larger

riser

smaller

riser

trap

period two

Suction Line

notes At minimum capacity, the refrigerant velocity in the two risers becomes too low

to carry the oil droplets. The oil from both risers therefore drains down, filling

the trap at the base of the larger riser. When this trap becomes completely filled

with oil, it prevents refrigerant vapor from flowing up the larger riser, and

diverts all the refrigerant up the smaller riser. This smaller riser is constructed

of a pipe with a diameter that is small enough to maintain adequate velocity at

minimum capacity. When system capacity is increased again, the higher

refrigerant velocity clears the trap of oil, and refrigerant vapor again flows up

both risers.

Notice the inverted trap at the top of the larger riser. When the trap at the base

of the larger riser is filled with oil, and refrigerant flows up only the smaller

riser, this inverted trap prevents oil from draining back into the larger riser. This

configuration minimizes the amount of oil trapped in the double riser under this

condition, therefore maximizing the amount of oil that is returned to the

compressor.

When sizing a double suction riser, the diameter of the smaller riser should be

selected to provide the minimum allowable velocity at minimum capacity. The

diameter of the larger riser should then be selected so that, at maximum

capacity, the velocities in both risers are greater than the minimum allowable

velocity and less than 4,000 fpm (20 m/s).

For this example system, the diameter of the smaller riser is 1 3/8 in. (35 mm).

When the system operates at its minimum capacity of 5 tons (17.6 kW), the

velocity inside the 1 3/8 in. (35 mm) riser—1,150 fpm (5.8 m/s)—is above the

minimum allowable velocity for a riser of this diameter.

At the maximum system capacity of 20 tons (70.3 kW), the smaller riser is

assumed to handle the same 5 tons (17.6 kW) of capacity, and the larger riser

must handle the remaining capacity. At 15 tons (52.8 kW) of capacity, the

refrigerant velocity in a 1 5/8 in. (42 mm) riser—2,450 fpm (12.4 m/s)—is

between the maximum and minimum allowable velocities for a riser of this

diameter.

Sizing a Double Suction Riser

▲ Sizing the smaller riser

▲ Sizing the larger riser

5 (17.6) 1,150 (5.8)

circuit

capacity,

tons (kW)

minimum velocity

for 1 3/8 in. (35 mm)

riser, fpm (m/s)

velocity inside

1 3/8 in. (35 mm)

pipe, fpm (m/s)

780 (4.0)

15 (52.8) 2,450 (12.4)

circuit

capacity,

tons (kW)

minimum velocity

for 1 5/8 in. (42 mm)

riser, fpm (m/s)

velocity inside

1 5/8 in. (42 mm)

pipe, fpm (m/s)

840 (4.3)

notes

period two

Suction Line

For this example system, a 2 1/8 in. (54 mm)-diameter pipe can be used for the

horizontal or vertical drop sections, and the double suction riser should be

constructed using 1 5/8 in. (42 mm)-diameter pipe for the larger riser and

1 3/8 in. (35 mm)-diameter pipe for the smaller riser.

Do not use a double suction riser unless it is absolutely necessary! The double

riser stores a large amount of oil—oil that is preferably left in the crankcase of

the compressor. If a double riser is required, check with the equipment

manufacturer to determine if more oil must be added to the system.

After the pipe diameters are selected, the overall pressure drop due to the

suction line can be calculated. The next step is to calculate the total

“equivalent” length of the suction line. The total equivalent length is the actual

measured length of straight pipe plus the pressure drop through any fittings,

such as elbows and tees. The pressure drop through the fittings is expressed in

terms of the length of straight pipe that would produce the same pressure drop.

To determine the equivalent length of various copper fittings, Trane now uses

the results from two decades of research conducted for the American Society of

Mechanical Engineers (ASME). This research produced a method for predicting

the pressure drop, in terms of equivalent length (Leq), through various types of

copper fittings by using the radius of curvature (r) of the fitting divided by its

diameter (d), or r/d. Table 1 on page 21 includes this data.

Calculate Total Equivalent Length

L

eq is based on r/d

radius of curvature (r)

diameter (d)

period two

Suction Line

notes

The research findings show that the equivalent lengths of copper fittings are

much smaller than the data used in the past. The table in Figure 28 compares

some of the equivalent-length data from this research to the values commonly

used in the past. While this data is for long-radius elbows, the same method is

used to predict the equivalent length of short-radius elbows and tees.

The suction line for this example 20-ton (70.3-kW) system, shown in Figure 23

on page 17, is 70 ft (21.4 m) long. The horizontal and vertical drop sections total

55 ft (16.8 m) in length, and are constructed of 2 1/8 in. (54 mm)-diameter pipe

with three long-radius elbows and three short-radius elbows. The single vertical

riser is 15 ft (4.6 m) long, and is constructed of 1 5/8 in. (42 mm)-diameter pipe

with no fittings.

suction line

Determine Pressure Drop

p2 1/8 in. = = 66.1 ft 1.4 psi/100 ft 0.93 psi

p54 mm = = 20.2 m 3.2 kPa/10 m 6.4 kPa

p1 5/8 in. = = 15 ft 6 psi/100 ft 0.9 psi

p42 mm = = 4.6 m 13.5 kPa/10 m 6.2 kPa

period two

Suction Line

notes Therefore, the total pressure drop through the straight pipe and fittings for this example suction line is 1.83 psi (12.6 kPa).

The final step toward calculating the total pressure drop is to add the pressure

drop due to any accessories that are to be installed in the suction line. The most

common accessories installed in a suction line are a suction filter and manual

ball valves that can be used to isolate this filter when it needs to be replaced.

Assuming that this example system includes one suction filter and two manual

ball valves installed in the 2 1/8 in. (54 mm) line, the total pressure drop due to

the suction line is 3.86 psi (26.6 kPa).

As mentioned earlier, a common rule in the past was to limit the pressure drop

through the suction line to 3 psi (20 kPa) for a system operating with R-22.

However, long line lengths, and the associated higher pressure drops, can be

tolerated assuming that the impact on system capacity and efficiency is

acceptable for the given application.

∆p = = 0.93 psi 0.9 psi + 1.83 psi

∆p = = 6.4 kPa 6.2 kPa + 12.6 kPa

suction line

Add Pressure Drop of Accessories

▲ Suction filter = 2 psi (13.8 kPa)

▲ Angle valve = 1 psi (6.9 kPa)

▲ Ball valve = 1 equivalent ft (0.3 m)

suction line filter

p 1.83 psi 2 psi 2 1 ft = = + + 1.4 psi/10 ft 3.86 psi

p 12.6 kPa 13.8 kPa 2 0.3 m = = + + 3.2 kPa/10 m 26.6 kPa

notes

period two

Suction Line

Other Considerations

Certain precautions should be taken when routing the suction line. First, proper

location and attachment of the remote expansion-valve bulb is very

important. This remote bulb, which measures the temperature of the refrigerant

leaving the evaporator, should be firmly attached to a straight, well-drained,

horizontal section of the suction line. The external equalizer line should be

inserted downstream of the remote bulb, in order to prevent influencing the

temperature measured by the bulb due to any leakage of liquid refrigerant

through the equalizer line.

The section of pipe at the outlet of the evaporator must be long enough to

permit the attachment of the remote bulb and the insertion of the external

equalizer line. A good rule is to allow 12 in. (300 mm) of straight horizontal pipe

for these two connections. Under no circumstances should the remote bulb be

located on a section of pipe where oil or liquid refrigerant could be trapped. The

presence of oil or liquid refrigerant can cause false temperature readings.

Slightly pitch this short horizontal section of pipe downward from the suction

header—1 in./10 ft (10 mm/3 m) in the direction of flow. For best temperature

sensing, the remote bulb should be firmly attached to the top of this section of

pipe and well insulated.

In addition, an access port should be located near the external equalizer line

connection. This port provides a point for accurate pressure measurement

when checking or adjusting the superheat setting of the TXV.

TXV Installation

external equalizer line

remote

bulb

evaporator

TXV

refrigerant

vapor

distributor

liquid

refrigerant

period two

Suction Line

notes

After this short horizontal section, the suction line should drop vertically

downward to allow the evaporator, and the section of pipe with the TXV bulb

attached, to drain freely when the system is operating.

If the suction line leaves the evaporator and then must rise immediately, this

can be accomplished by using a small trap at the end of the horizontal section

of pipe, just before the suction line rises. The purpose of this trap is to provide

free drainage from the evaporator and the section of pipe to which the TXV

bulb is attached. This ensures that the TXV bulb is not the “low spot” in the

piping where oil or liquid refrigerant could be trapped. The purpose of this trap

is not to drain the suction riser.

The suction line must then rise above the height of the evaporator coil. This

prevents refrigerant and oil inside the evaporator from free draining into the

suction line, and toward the compressor, when the system is off.

The configuration in Figure 32 is typical for a system with an indoor air handler,

where the refrigerant piping is routed along the ceiling and must drop down to

the evaporator.

single distributor on circuit

Evaporator Coil Connection

TXV

bulb

must rise above

height of evaporator

must drop below

header outlet

suction header

Figure 32

26 TRG-TRC006-EN

notes

period two

Suction Line

If the suction line leaves the evaporator and then must drop vertically, an

inverted trap is used to prevent refrigerant and oil inside the evaporator from

free draining into the suction line, and toward the compressor, when the

system is off.

This involves piping the suction line to rise above the height of the evaporator

coil, installing an inverted trap, and then allowing the line to drop back down

below the evaporator.

When an evaporator has more than one distributor connected to a single

refrigeration circuit, there are some additional considerations for manifolding

the suction headers into a single suction line.

First, arrange the suction line so that the refrigerant vapor leaving the top

suction header flows downward, connecting to the pipes leaving the other

suction headers. This manifold pipe should drop below the lowest suction

single distributor on circuit

Evaporator Coil Connection

TXV

bulb

must rise above

height of evaporator

must drop below

header outlet

suction header

Figure 33

multiple distributors on circuit

Evaporator Coil Connections

TXV

bulb

must drop below

lowest header outlet double-elbow

configuration

must rise above

height of evaporator

Figure 34

TRG-TRC006-EN 27

period two

Suction Line

notes header outlet before being allowed to turn upward. This allows all of the oil to

drain out of the evaporator and congregate before it must be carried up a

vertical riser.

Second, the remote TXV bulbs and external equalizing line connections must

be located so that the conditions affecting one valve cannot influence the other

valves. The double-elbow configuration allows the corresponding section of the

evaporator coil to drain freely while also isolating the remote bulb and

equalizer line from the suction pressure and temperature of the above coil

sections.

The single suction line then rises above the height of the evaporator coil,

preventing refrigerant and oil inside the evaporator from free draining into the

suction line, and toward the compressor, when the system is off.

As mentioned earlier, the trap shown in Figure 32 on page 25, and Figure 33 on

page 26, is used to provide free drainage from the evaporator and section of

pipe to which the TXV bulb is attached. Its purpose is not to drain the suction

riser.

A common practice in the past was to install a trap at the base of any vertical

suction riser, and often part of the way up the riser. It was believed that these

traps would cause turbulence at the base of the riser section so that oil

accumulating in the trap would be more easily “broken up” into a mist, which

could then be carried to the top of the riser with the refrigerant vapor.

Today, we understand that oil droplets are moved inside a pipe by the force of

mass flow, not by turbulence. It is possible to size a vertical suction riser, which

will allow the refrigerant to carry oil droplets at minimum capacity, without the

need for a trap. If a suction riser is sized properly, oil will return to the

compressor regardless of whether a trap is present at the base of the riser. If a

suction riser is oversized, however, adding such a trap will not restore proper

oil movement.

suction line

Riser Traps Are Not Required

notes

period two

Suction Line

Horizontal sections of the suction line should be pitched slightly—1 in./10 ft

(10 mm/3 m)—so that the refrigerant drains back toward the evaporator. This

prevents any refrigerant that condenses in the suction line from flowing into the

compressor when the circuit is off.

The refrigerant temperature inside the suction line is generally cooler than the

surrounding air. Therefore, it is always good practice to insulate the entire

suction line to prevent condensation and loss of capacity due to heat gain.

Without insulation, moisture may condense on the outside surface of the pipe

and drip onto the floor or suspended ceiling below. Also, any heat gained by

the suction line places an additional load on the system that reduces cooling

capacity, and also may result in improper motor cooling with some compressor

designs.

The suction-line filter should be located as close as possible to the

compressor—its purpose is to remove foreign matter from the system and

protect the compressor. This filter should be accompanied by manual shutoff

valves on both sides, allowing it to be isolated when the filter core needs to be

replaced.

Additionally, the suction line typically includes two access ports. One is

installed near the compressor and is used to measure suction pressure. The

other is located near the external equalizer line connection for the TXV, and is

used to measure superheat when checking or adjusting the TXV.

suction line

Other Considerations

▲ Do not use suction riser traps: they are not

required

▲ Pitch horizontal sections to drain toward

evaporator

▲ Insulate entire suction line

◆ Prevents condensation

◆ Minimizes loss of capacity due to heat gain

▲ Install suction-line filter close to compressor

◆ Manual shutoff valves allow isolation for replacement

▲ Install access ports to measure suction

pressure and superheat

period two

Suction Line

notes

The compressor is designed to compress refrigerant vapor only. A suctionline accumulator is a device that attempts to prevent a slug of liquid

refrigerant or oil from causing damage to the compressor. The accumulator

allows liquid refrigerant and oil to separate from the refrigerant vapor, and then

be drawn into the compressor at a rate that will not cause damage.

Oversizing the accumulator, however, can cause inadequate oil return due to

low refrigerant velocity through the accumulator. An accumulator also

increases the refrigerant charge of the system and increases the pressure drop.

Some refrigeration systems that use a flooded evaporator may require a

suction-line accumulator for freeze protection or other functions. Check with the

equipment manufacturer to determine if a suction-line accumulator is required,

recommended, or discouraged.

Suction-Line Accumulator

▲ Check with

equipment

manufacturer to

determine if required,

recommended, or

discouraged

notes

period two

Suction Line

Finally, when the system includes a reciprocating compressor, an anchored,

45-degree canted loop should be constructed very close to the location at which

the suction line connects to the compressor. This loop absorbs vibration

generated by the reciprocating motion of the compressor, minimizing the

amount of vibration transmitted to the piping system.

The loop is constructed of four straight sections of pipe, each one at least ten

pipe diameters in length, and four elbows. The pipe is anchored a short

distance downstream from the loop, allowing the loop to absorb vibration in

both the vertical and horizontal planes. Note that the loop is pitched downward

to provide free drainage away from the compressor.

suction line

Reciprocating Compressors

pipe

anchor

pitch

notes

period three

Discharge Line

The next line to be considered is the discharge line. This section of pipe

conducts hot, high-pressure refrigerant vapor from the compressor to the

condenser.

The design of the discharge line is probably less critical than that of the suction

line because the refrigerant vapor is at a higher temperature, allowing the oil to

be carried along more easily than in the cooler suction line. Even so, if the

discharge line is not properly sized and installed, reliability or performance

problems can result.

Requirements for Sizing and Routing

Similar to the suction line, the diameter of the discharge line must be small

enough that the resulting refrigerant velocity is sufficiently high to carry oil

droplets, at all steps of compressor unloading. However, if the velocity in the

pipe is too high, it may cause objectionable noise.

period three

Discharge Line

Refrigerant Piping

Figure 39

discharge line

Requirements for Sizing and Routing

▲ Ensure adequate velocity to return oil to

compressor at all steps of unloading

▲ Avoid excessive noise

▲ Minimize efficiency loss

notes

period three

Discharge Line

Finally, anything that causes the compressor discharge pressure to rise causes

the compressor to work harder. Therefore, the pipe diameter should be as large

as possible to minimize pressure drop and maximize compressor efficiency.

As with the suction line, it may be helpful to compare the old “rules” for

selecting the diameter of the discharge line with the newer rules that result

from changes in compressor technology, recent research, and the additional

requirement to minimize system refrigerant charge.

In the past, many discharge lines for systems operating with R-22 were sized

using the same minimum-velocity values that were used for sizing suction

lines. As explained in Period Two, the minimum allowable velocity at a specific

operating condition depends on the diameter of the pipe. The minimum

allowable velocity values for discharge lines are different than for suction lines,

because the refrigerant vapor is at a higher temperature, and oil moves more

easily than in the cooler suction line.

The recommended maximum velocity limit for discharge lines is 3,500 fpm

(17.5 m/s). Again, a higher velocity inside the pipe may cause objectionable

noise.

Another “old” rule, for a system operating with R-22, was to limit the pressure

drop through the discharge line to 6 psi (41 kPa). While it is still good practice to

minimize pressure drop, higher pressure drops can be tolerated, assuming that

the impact on system efficiency is acceptable for the given application.

discharge line

Sizing “Rules”

minimum velocity

1,000 fpm (5 m/s) based on diameter of riser

for risers

500 fpm (2.5 m/s) percentage of minimum

for horizontal sections riser velocity

maximum velocity

3,500 fpm (17.5 m/s) 3,500 fpm (17.5 m/s)

maximum pressure drop

6 psi (41 kPa) based on specific system

requirements

old rules for R-22 new rules for R-22

period three

Discharge Line

notes

Sizing Process

The steps for selecting the proper diameter of the discharge line are identical to

the steps for sizing the suction line:

1) Determine the total length of discharge-line piping.

2) Calculate the refrigerant velocity at both maximum and minimum system

capacities.

3) Select the largest pipe diameter that will result in acceptable refrigerant

velocity at both maximum and minimum capacities.

4) Calculate the total “equivalent” length of piping by adding the actual length

of straight pipe to the equivalent length of any fittings to be installed in the

discharge line.

5) Determine the pressure drop (based on the total equivalent length) due to

the straight pipe and fittings.

6) Add the pressure drop due to any accessories installed in the discharge

line.

Again, the total measured length of the discharge line can be estimated from

the initial layout.

discharge line

Process for Sizing

1 Determine total length of piping

2 Calculate refrigerant velocity at maximum and

minimum capacities

3 Select largest pipe diameter that results in

acceptable velocity at both maximum and

minimum capacity

4 Calculate total equivalent length of straight pipe

and fittings

5 Determine pressure drop due to pipe and fittings

6 Add pressure drop due to accessories

notes

period three

Discharge Line

The chart in Figure 43 shows the velocity of R-22 inside discharge pipes of

various diameters at one particular operating condition—40°F (4.4°C) saturated

suction temperature, 125°F (51.7°C) saturated condensing temperature, 12°F

(6.7°C) of superheat, 15°F (8.3°C) of subcooling, and 70°F (38.9°C) of

compressor superheat.

For the same example 20-ton (70.3-kW) system, the refrigerant velocity inside a

1 5/8 in. (42 mm)-diameter pipe is about 1,100 fpm (5.6 m/s) at this condition.

Again, a computer program can be used to calculate the velocity based on

actual conditions, but without a program, a chart like this may be useful.

This example 20-ton (70.3-kW) system contains one refrigeration circuit with

two steps of capacity. Maximum capacity is 20 tons (70.3 kW) and minimum

capacity is 10 tons (35.2 kW). Using the chart in Figure 43, the refrigerant

discharge line

Determine Refrigerant Velocity

velocity, fpm (m/s)

evaporator capacity, tons (kW)

pipe diameter, in. (mm)

discharge line

Determine Refrigerant Velocity

7/8 (22) 4,050 (20.6) 2,025 (10.3)

1 1/8 (28) 2,400 (12.2) 1,200 ( 6.1)

1 3/8 (35) 1,550 ( 7.9) 775 ( 3.9)

1 5/8 (42) 1,100 ( 5.6) 550 ( 2.8)

2 1/8 (54) 650 ( 3.3) 325 ( 1.7)

pipe velocity, fpm (m/s)

diameter,

in. (mm)

20 tons

(70.3 kW)

10 tons

(35.2 kW)

period three

Discharge Line

notes velocity at both maximum and minimum capacities is determined for several

pipe diameters.

After these velocities have been determined, the largest acceptable pipe

diameter is selected to minimize the overall pressure drop due to the discharge

line.

When this system operates at maximum capacity, use of the 7/8 in. (22 mm)-

diameter pipe results in a refrigerant velocity that is greater than the

recommended upper limit of 3,500 fpm (17.5 m/s) for discharge lines. This high

velocity may cause objectionable noise, so this pipe size should probably not

be considered.

The table in Figure 45 shows the minimum allowable refrigerant velocity, for

both a vertical riser and a horizontal (or vertical drop) section of discharge line,

for each standard pipe diameter. As with suction lines, the minimum velocity

for a horizontal, or vertical drop, section of discharge line is 75 percent of the

minimum allowable velocity for a vertical riser of the same diameter.

The minimum velocities listed in this table assume a worst-case operating

condition of 80°F (26.7°C) saturated condensing temperature.

discharge line

Minimum Allowable Velocities

period three

Discharge Line

Based on the minimum allowable velocities listed in Figure 45 on page 35,

1 5/8 in. (42 mm) is the largest diameter of pipe that will result in acceptable

velocity, at both maximum and minimum capacities, within horizontal and

vertical drop sections.

When the circuit unloads to 10 tons (35.2 kW), the velocity inside a 1 5/8 in.

(42 mm) pipe—550 fpm (2.8 m/s)—remains above the minimum allowable

velocity for a riser of this diameter—520 fpm (2.6 m/s). One size larger (2 1/8 in.

[54 mm]) would result in a velocity that is too low when the system operates at

minimum capacity.

For this example system, a 1 5/8 in. (42 mm)-diameter pipe can be used for all

horizontal and vertical sections of the discharge line, including vertical risers.

This pipe size will ensure proper oil movement at both maximum and minimum

system capacities, and minimize the overall pressure drop due to the discharge

line.

Select Discharge Line Size

riser horiz/drop

20 (70.3) 1,100 (5.6)

10 (35.2) 550 (2.8)

tons (kW)

pipe, fpm (m/s)

520 (2.6) 390 (2.0)

Selected Discharge Line Size

1 5/8 in.

(42 mm)

compressor

condenser

1 5/8 in. (42 mm)

riser

1 5/8 in.

(42 mm)

period three

Discharge Line

notes For this example, a single pipe diameter can be selected that provides adequate

velocity at minimum capacity, without exceeding the maximum velocity limit at

maximum capacity. If the maximum and minimum velocity requirements

cannot be achieved with a single pipe diameter, a double riser may be required.

After the pipe diameter is selected, the overall pressure drop due to the

discharge line can be calculated. The next step is to calculate the total

“equivalent” length of the straight pipe and fittings. The equivalent length data

for copper fittings used in the discharge line is the same as for the suction line.

The discharge line for this example 20-ton (70.3-kW) system, shown in

Figure 47 on page 36, is a total of 50 ft (15.2 m) long. Both the horizontal

sections and the vertical riser are constructed of 1 5/8 in. (42 mm)-diameter pipe

and contain a total of two long-radius elbows. Based on the “new” data from

Table 1 on page 21, the equivalent length of each long-radius elbow of this

diameter is 1.0 ft (0.30 m), so the total equivalent length of the discharge line is

52 ft (15.8 m).

The next step is to determine the pressure drop due to the straight pipe and

fittings, based on this calculated total equivalent length. The chart in Figure 48

shows the pressure drop, per unit of equivalent length, for different pipe

diameters and various capacities. At the maximum capacity of 20 tons

(70.3 kW), the pressure drop due to the straight, 1 5/8 in. (42 mm)-diameter pipe

and fittings is 2.1 psi per 100 ft of equivalent length (4.8 kPa/10 m), or 1.1 psi

(7.6 kPa).

discharge line

Determine Pressure Drop

period three

Discharge Line

The final step is to add the pressure drop due to any accessories that will be

installed in the discharge line, and calculate the total pressure drop. The most

common accessories installed in a discharge line are manual ball valves that

are used to isolate portions of the system during maintenance.

Assuming that this example system includes one manual angle valve, the total

pressure drop due to the discharge line is 2.1 psi (14.5 kPa).

As mentioned earlier, in the past, a common rule was to limit the pressure drop

through the discharge line to 6 psi (41 kPa) for a system operating with R-22.

However, long line lengths, and the associated higher pressure drops, can be

tolerated assuming that the impact on system efficiency is acceptable for the

given application.

discharge line

Add Pressure Drop of Accessories

▲ Angle valve = 1 psi (6.9 kPa)

▲ Ball valve = 1 equivalent ft (0.3 m)

period three

Discharge Line

Other Considerations

As with the suction line, riser traps are not required in a discharge line,

assuming that the riser is properly sized to return oil at minimum capacity.

Horizontal sections of the discharge line should be pitched slightly—1 in./10 ft

(10 mm/3 m)—so that the refrigerant drains toward the condenser. This

prevents any refrigerant that may condense in the discharge line from flowing

back into the compressor when the system is off.

In most cases, the discharge line does not need to be insulated. The refrigerant

temperature inside the line is generally warmer than the surrounding air. Any

heat loss reduces the heat-rejection load of the condenser, improving system

efficiency. However, the outer surface of the discharge line can be hot. Some

sections of the discharge line may be insulated to prevent injury to someone

who may come in contact with the piping. Also, if the system has long

refrigerant lines and is expected to operate when it is cold outdoors, the

discharge line should be insulated to prevent refrigerant from condensing

inside the line before it reaches the condenser.

discharge line

Other Considerations

▲ Do not use discharge riser traps: they are not

required

▲ Pitch horizontal sections to drain toward

condenser

▲ Insulate only to prevent accidental burns or

when system must operate in cold ambient

conditions

period three

Discharge Line

Finally, when the system includes a reciprocating compressor, an anchored,

45-degree canted loop should be constructed very close to the location at which

the discharge line connects to the compressor. This loop absorbs vibration

generated by the reciprocating motion of the compressor. Note that the loop is

pitched downward to provide free drainage away from the compressor.

discharge line

Reciprocating Compressors

pipe

anchor

pitch

period four

Liquid Line

The third line to be considered is the liquid line. This section of pipe conducts

warm, high-pressure liquid refrigerant from the condenser to the expansion

device and evaporator.

Requirements for Sizing and Routing

The liquid line must be designed and installed to ensure that only liquid

refrigerant (no vapor) enters the expansion device. The presence of refrigerant

vapor upstream of a TXV can result in erratic valve operation and reduced

system capacity. In order to meet this requirement, the condenser must provide

adequate subcooling, and the pressure drop through the liquid line and

accessories must not be high enough to cause flashing upstream of the

expansion device. Subcooling allows the liquid refrigerant to experience some

pressure drop as it flows through the liquid line, without the risk of flashing.

period four

Liquid Line

Refrigerant Piping

Figure 52

liquid line

Requirements for Sizing and Routing

▲ Ensure that only liquid refrigerant enters

expansion device

▲ Minimize refrigerant charge

▲ Avoid excessive noise and pipe erosion

period four

Liquid Line

Oil and liquid refrigerant mix readily, so oil movement within the liquid line is

not a concern. However, the design of the liquid line is the most critical when it

comes to minimizing the system refrigerant charge. This is because, of the

three lines, it has the greatest impact on the quantity of refrigerant required to

charge the system. The diameter of the liquid line must be as small as possible

to minimize the refrigerant charge, therefore improving reliability and

minimizing installed cost. However, if the pipe is too small, the increased

pressure drop may cause flashing upstream of the expansion device.

The final requirement is to limit the refrigerant velocity in order to prevent

objectionable noise or erosion of the inner surfaces of the piping.

In this example system, the refrigerant (R-22) is condensed at a temperature of

125°F (51.7°C), which corresponds to a pressure of 293 psia (2,020 kPa). Inside

the subcooling tubes, the temperature of the liquid refrigerant is reduced by

15°F (8.3°C) to 110°F (43.3°C). Ignoring the pressure drop due to the condenser

and subcooler tubes, the pressure of this subcooled liquid refrigerant is the

same, 293 psia (2,020 kPa).

At 110°F (43.3°C), the pressure of the refrigerant can be reduced to 241 psia

(1,662 kPa) before it reaches the saturated liquid condition, where flashing

occurs. In other words, the liquid refrigerant could experience a pressure drop

of up to 52 psi (358 kPa) before flashing occurs.

In years past, the liquid line was considered an independent component of the

system and was sized based solely on the physics of fluid flow through the

pipe. It was believed that increasing the diameter of the liquid line, which would

reduce the pressure drop through the piping, would increase the amount of

subcooling remaining at the inlet of the expansion device.

The internal volume of the liquid line, however, has a significant impact on the

ability of the condenser to provide subcooling at part-load operating

conditions. When this interaction between the liquid line and condenser is

considered, we find that the amount of subcooling available at the expansion

device is as dependent on the ability of the condenser to provide subcooling as

Subcooling

enthalpy, Btu/lb (kJ/kg)

period four

Liquid Line

it is on the pressure drop due to the liquid-line piping. In fact, if the liquid line is

too large, it will actually reduce the subcooling at the expansion device under

cooler ambient conditions.

Therefore, proper sizing of the liquid line requires an understanding of the

entire refrigeration system, not just the liquid-line piping.

Take a look at an example 25-ton (87.9-kW) system with 150 ft (45.7 m) of liquidline piping to illustrate this point. System A uses a 7/8 in. (22 mm)-diameter

pipe for the liquid line, while system B uses a larger 1 1/8 in. (28 mm)-diameter

pipe. Both systems are charged so that the condenser provides 18°F (10°C) of

subcooling at the same high outdoor temperature. Even though the rest of the

system components are identical, system B requires 25 lb (11.3 kg) more

refrigerant to charge the system because of the larger liquid line.

As expected, the pressure drop due to the piping is less for system B with the

larger diameter pipe. When the outdoor temperature is high, this reduced

pressure drop results in more subcooling remaining at the expansion device—

14.2°F (7.9°C) compared to 12.3°F (6.8°C) for system A.

However, one should examine what happens when the system operates at a

cooler outdoor temperature. As the outdoor temperature decreases, the

temperature of the refrigerant inside the condenser and liquid line also

decreases. This lower temperature causes the density of the refrigerant to

increase, requiring more mass of refrigerant in order to fill up the internal

volume of the liquid line. The result is less refrigerant remaining inside the

condenser, and therefore less subcooling provided by the condenser.

This change in density has a more severe impact on system B with the largerdiameter liquid line, because the internal volume of the liquid line is greater. For

these two example systems, when the outdoor temperature is cooler, the

amount of subcooling provided by the same condenser is 12.9°F (7.2°C) for

system A, but only 7.2°F (4°C) for system B.

Even though the pressure drop is still less through the larger-diameter line—

5.2°F (2.9°C) loss in subcooling due to pressure drop in system B versus 7.4°F

Impact of Refrigerant Density Change

A 7/8 (22) high 18.0 (10) - 5.7 (3.2) = 12.3 (6.8)

B 1 1/8 (28) high 18.0 (10) - 3.8 (2.1) = 14.2 (7.9)

A 7/8 (22) low 12.9 (7.2) - 7.4 (4.1) = 5.5 (3.1)

B 1 1/8 (28) low 7.2 (4.0) - 5.2 (2.9) = 2.0 (1.1)

available subcooling, °F (°C)

liquid line

diameter,

in. (mm)

period four

Liquid Line

(4.1°C) in system A—the impact of the change in refrigerant density has a

greater impact on the overall system. At this cooler outdoor condition, only

2.0°F (1.1°C) of subcooling is available at the TXV in system B, while 5.5°F

(3.1°C) of subcooling is available at the TXV in system A.

Again, proper sizing of the liquid line requires an understanding of the entire

refrigeration system, not just the liquid-line piping. The diameter of the liquid

line should be as small as possible, to minimize the impact of the change in

refrigerant density. If the liquid line is too large, it may result in too little

subcooling at some part-load operating conditions.

Figure 56 compares the old “rules” for selecting the diameter of the liquid line

with the newer rules that result from this understanding of the impact of

refrigerant density change, recent research, and the additional requirement to

minimize system refrigerant charge.

In the past, many liquid lines in systems operating with R-22 were sized to limit

the pressure drop through the piping and accessories to 6 psi (41 kPa).

However, as we just explained, limiting the pressure drop through the liquid

line does not necessarily ensure that only liquid refrigerant (no vapor) enters

the expansion device. This rule has been rewritten to specify that adequate

subcooling must be available at the inlet of the expansion device at all

operating conditions, acknowledging this interaction between the liquid line

and condenser. The condenser must provide adequate subcooling, and the

pressure drop through the liquid line and accessories must not be high enough

to cause flashing upstream of the expansion device. A margin of safety, such as

5°F (2.8°C), may be used to account for unexpected variations.

In the past, the recommended maximum velocity limit for a liquid line was

360 fpm (1.8 m/s). Velocities in excess of this limit caused concern for liquidhammer effect when the flow of refrigerant was suddenly stopped, such as

when a solenoid valve closed. However, most solenoid valves used in airconditioning systems today are not quick-closing valves. That is, they do not

close rapidly enough to cause a serious hammer effect.

liquid line

Sizing “Rules”

maximum pressure drop

6 psi (41 kPa) 5°F (2.8°C) subcooling

remaining at inlet to expansion

device at all operating conditions

maximum velocity

360 fpm (1.8 m/s) 600 fpm (3 m/s)

old rules for R-22 new rules for R-22

period four

Liquid Line

The new maximum-velocity limit of 600 fpm (3 m/s) is based on preventing

erosion of the inner surfaces of the pipe.

Sizing Process

Following are the steps to follow when selecting the proper diameter of the

liquid line:

1) Determine the total length of liquid-line piping.

2) Obtain the amount of subcooling provided by the condenser from the

manufacturer.

3) Determine the refrigerant velocity at maximum system capacity.

4) Select the smallest pipe diameter that will result in acceptable refrigerant

velocity at maximum system capacity.

5) Calculate the total equivalent length of piping by adding the actual

measured length of straight pipe to the equivalent length of any fittings to

be installed in the liquid line.

6) Determine the pressure drop (based on the total equivalent length) due to

the straight pipe and fittings.

7) Add the pressure drop due to any accessories installed in the liquid line.

8) Calculate the loss of subcooling due to pressure drop and to any change in

elevation, to verify adequate subcooling at the inlet to the expansion

device.

From the initial layout, the total measured length of the liquid line can be

estimated. Again, the refrigerant piping should be routed in the shortest and

simplest manner possible, minimizing the total length of piping.

liquid line

Process for Sizing

1 Determine total length of piping

2 Obtain subcooling provided by condenser

3 Determine refrigerant velocity

4 Select smallest pipe diameter that results in

acceptable velocity at maximum capacity

5 Calculate total equivalent length of pipe and fittings

6 Determine pressure drop due to pipe and fittings

7 Add pressure drop due to accessories

8 Calculate loss of subcooling due to pressure drop

and elevation change 

period four

Liquid Line

The next step is to determine how much subcooling is provided by the

condenser. This is difficult because the amount of subcooling provided by an

air-cooled condenser depends on many factors, including the total amount of

heat that must be rejected, the outdoor dry-bulb temperature, the internal

volume of the liquid line, and the system refrigerant charge.

To demonstrate, look at an example operating envelope for a system comprised

of an air-cooled condensing unit (compressors and condenser packaged

together) and a cooling coil (evaporator). The operating envelope is the entire

range of cooling loads for the system over the entire range of outdoor

temperatures at which the system will be operating. Identifying the extreme

corners of this system operating envelope helps us to understand how the

amount of subcooling provided by the condenser varies.

In Figure 58, two of the lines depict the performance of the condensing unit

operating over a range of cooling loads. One line assumes that the unit

operates at a constant high outdoor-air (outdoor or OA) temperature, while the

other line assumes that the same unit operates at a constant low outdoor

temperature.

The other two lines depict the performance of the evaporator coil operating

over a range of outdoor-air temperatures. One line assumes that the coil

operates with a constant high cooling load, while the other line assumes that

the same coil operates with a constant low cooling load.

System Operating Envelope

saturated suction temperature

condensing unit,

high OA temperature

condensing unit,

low OA temperature

evaporator,

high cooling load

evaporator,

low cooling load

system cooling capacity

period four

Liquid Line

This operating envelope defines four extreme corners. Condition A depicts the

system operating with a high cooling load at a high outdoor temperature, and

condition B depicts a low cooling load at the same high outdoor temperature.

Condition C depicts the system operating with a low cooling load at a low

outdoor temperature, and condition D depicts a high cooling load at the same

low outdoor temperature.

A common belief is that the amount of subcooling provided by the condenser

increases when the outdoor temperature is cooler. But this is not necessarily

true in systems that use a TXV for the expansion device. Figure 60 shows how

subcooling changes throughout the operating envelope of an example system.

When the system operates with a high cooling load and a high outdoor

temperature (A), the condenser provides 18°F (10°C) of subcooling. When

operating with the low cooling load at the same high outdoor temperature (B),

less heat must be rejected by the condenser, allowing it to provide more

subcooling—21°F (11.7°C).

System Operating Envelope

high cooling load,

high OA temperature

A

B

D

C

high cooling load,

low OA temperature

low cooling load,

high OA temperature

low cooling load,

low OA temperature

saturated suction temperature

system cooling capacity

Figure 59

effect of cooling load and outdoor temperature

Subcooling Leaving Condenser

A high high 18 (10.0)

B low high 21 (11.7)

C low low 13 (7.2)

D high low 5 (2.8)

subcooling leaving

condenser*, °F (°C)

cooling

load

outdoor

temperature

* This example system uses a TXV for the expansion device.

Figure 60

48 TRG-TRC006-EN

notes

period four

Liquid Line

When the system operates with the low cooling load at a low outdoor

temperature (C), the impact of the refrigerant density change reduces the

subcooling provided by the condenser to 13°F (7.2°C). Finally, when the system

operates with the high cooling load at the same low outdoor temperature (D),

subcooling drops to 5°F (2.8°C). At condition D, the mass flow rate of

refrigerant is high due to the high cooling load, and the refrigerant density is

high due to the low outdoor temperature. This results in very little liquid

refrigerant remaining in the condenser for subcooling.

The key points here are that the amount of subcooling provided by the

condenser changes within the operating envelope, and this subcooling is

actually reduced at cooler outdoor temperatures due to the change in

refrigerant density. This example demonstrates that the condenser provides the

smallest amount of subcooling when the system is operating with a high

evaporator load at a low outdoor temperature.

Defining the operating envelope for a particular system is no easy task, and

therefore is generally not done for each project. Determining the amount of

subcooling provided by the condenser throughout an operating envelope is

also quite complicated. It requires a model of the entire refrigeration system

and help from the manufacturer of the condenser. Therefore, if the

manufacturer of the refrigeration equipment provides recommended line sizes,

or tools for selecting the optimal line sizes, we suggest that you use them.

If recommended line sizes are not provided, however, consult the manufacturer

to determine the amount of subcooling provided by the condenser. For this

example, we will assume 15°F (8.3°C) of subcooling is provided by the

condenser.

Involve the Manufacturer

▲ If provided, use refrigerant

line sizes recommended by

manufacturer

Figure 61

TRG-TRC006-EN 49

period four

Liquid Line

notes

The chart in Figure 62 shows the velocity of liquid R-22 inside liquid lines of

various diameters at one particular operating condition—40°F (4.4°C) saturated

suction temperature, 125°F (51.7°C) saturated condensing temperature, 12°F

(6.7°C) of superheat, 15°F (8.3°C) of subcooling, and 70°F (38.9°C) of

compressor superheat.

For the same example 20-ton (70.3-kW) system, the refrigerant velocity inside a

5/8 in. (15 mm)-diameter pipe is about 520 fpm (2.6 m/s) at this condition.

Again, a computer program can be used to calculate the velocity based on

actual conditions, but without a program, a chart like this may be useful.

Using the chart in Figure 62, the refrigerant velocity at maximum system

capacity—20 tons (70.3 kW), for this example—is determined for various pipe

diameters. Use of the 1/2 in. (12 mm)-diameter pipe results in a refrigerant

liquid line

period four

Liquid Line

velocity that is greater than the recommended upper limit of 600 fpm (3 m/s) for

liquid lines, so it should probably not be considered.

After the velocity inside pipes of various diameters has been determined, the

smallest acceptable pipe diameter is selected to minimize the system

refrigerant charge. The table in Figure 63 on page 49 shows that 5/8 in. (15 mm)

is the smallest diameter of pipe that will result in a velocity below the

recommended upper limit of 600 fpm (3 m/s).

The next step is to calculate the total equivalent length by adding the actual

measured length of straight pipe to the equivalent length of any fittings to be

installed in the liquid line. The equivalent length data for copper fittings used in

the liquid line is the same as for the suction and discharge lines.

The liquid line for this example 20-ton (70.3-kW) system, shown in Figure 64, is

100 ft (30.5 m) long. It is constructed entirely of 5/8 in. (15 mm)-diameter pipe

and contains a total of four long-radius elbows. Based on the “new” equivalent

length data from Table 1 on page 21, the equivalent length of each elbow of this

diameter is 0.4 ft (0.12 m). The total equivalent length of this example liquid line

is calculated to be 101.6 ft (31.0 m).

liquid line

Calculate Total Equivalent Length

evaporator

condenser

period four

Liquid Line

The next step is to determine the pressure drop due to the straight pipe and

fittings, based on the calculated total equivalent length. The chart in Figure 65

shows the pressure drop, per unit of equivalent length, for different pipe

diameters and capacities. This chart is based on one specific operating

condition, but is likely to be representative of the condition when the system

operates at maximum capacity. Alternatively, a computer program can be used

to determine pressure drop at specific conditions.

At the maximum capacity of 20 tons (70.3 kW), the pressure drop due to the

straight pipe and fittings of the liquid line is 17 psi per 100 ft of equivalent

length (72.2 kPa/10 m), or 17.3 psi (119 kPa).

liquid line

Determine Pressure Drop

period four

Liquid Line

The next step is to add the pressure drop due to any accessories that will be

installed in the liquid line, and calculate the total pressure drop. The most

common accessories installed in a liquid line are a filter-drier, a sight glass, a

solenoid valve, and manual ball valves that are used to isolate portions of the

system during maintenance, such as when replacing the core of the filter-drier.

The pressure drop for the filter-drier, shown in Figure 66, assumes that it is

properly sized but dirty.

Assuming that this example system includes a filter-drier, solenoid valve, sight

glass, and two ball valves, the total pressure drop due to the liquid line is

27.8 psi (191 kPa).

liquid line

Add Pressure Drop of Accessories

▲ Filter-drier = 6 psi (41.4 kPa)

▲ Solenoid valve = 4 psi (27.6 kPa)

▲ Sight glass = 1 equivalent ft (0.3 m)

▲ Angle valve = 1 psi (6.9 kPa)

▲ Ball valve = 1 equivalent ft (0.3 m)

solenoid

valve

period four

Liquid Line

The next step is to calculate the loss of subcooling due to the liquid-line

pressure drop and any elevation change.

When the condenser is at a lower elevation than the evaporator, the refrigerant

must overcome the pressure due to a column of liquid refrigerant when flowing

up to the evaporator. This liquid column causes a pressure drop that results in a

decrease in the amount of subcooling available at the expansion device. When

the condenser is at a higher elevation than the evaporator, however, the liquid

column causes a pressure gain that results in an increase in subcooling.

Most systems will include some vertical risers and some vertical drops in the

liquid line, but the critical distance is net elevation difference. For example, the

liquid line in this example system (Figure 67) rises 15 ft (4.6 m) and then later

drops 11 ft (3.4 m). The vertical rise will cause a pressure drop and the vertical

drop will cause a pressure gain, but the net effect of these two vertical sections

will be 4 ft (1.2 m) of rise.

liquid line

Pressure Drop Due to Elevation

evaporator

condenser

15 ft (4.6 m)

vertical rise

creates

pressure drop

11 ft (3.4 m)

vertical drop

creates

pressure gain

period four

Liquid Line

The chart in Figure 68 shows that the magnitude of the associated pressure

drop (or gain) depends primarily on the temperature (density) of the liquid

refrigerant. The greatest pressure drop (or gain) due to this vertical column of

liquid occurs when the temperature of the liquid refrigerant is the lowest.

Using the example liquid line in Figure 67 on page 53, and assuming a liquidrefrigerant temperature of 110°F (43.3°C), the pressure drop due to the 4 ft (1.2

m) net liquid-line rise is 1.9 psi (13.4 kPa).

As you can see, the effect of this liquid column of refrigerant can dramatically

affect the amount of subcooling available at the expansion device, especially at

cooler ambient conditions when the condenser is already providing less

subcooling than it could under design conditions. In fact, it may even use up all

of the subcooling provided by the condenser, resulting in no subcooling

remaining at the expansion device.

Historically, the rule-of-thumb has been to assume a pressure drop of 0.5 psi for

each foot (11.3 kPa/m) of net liquid-line rise.

liquid line

Pressure Drop Due to Elevation

period four

Liquid Line

Next, the pressure drop due to the elevation change is added to the frictional

pressure drop due to the straight pipe, fittings, and accessories. For our

example 20-ton (70.3-kW) system, the total pressure drop due to all of these

effects is 29.7 psi (204.4 kPa).

Then, using the chart in Figure 69, this total pressure drop is converted to an

equivalent loss in subcooling. Assuming a liquid refrigerant temperature of

110°F (43.3°C), the loss of subcooling is 3 psi/°F (37.3 kPa/°C), or 9.9°F (5.5°C).

liquid line

period four

Liquid Line

The last step is to verify that adequate subcooling is available at the inlet to the

expansion device. This involves subtracting the impact of pressure drop and

elevation difference from the amount of subcooling provided by the condenser.

At this condition, 5.1°F (2.8°C) of subcooling is remaining at the inlet to the

expansion device. This is above the minimum limit of 5°F (2.8°C).

As mentioned earlier, in the past, a common rule was to limit the pressure drop

through the liquid line to 6 psi (41 kPa) for a system operating with R-22.

However, as this example demonstrates, in many cases a much higher pressure

drop can be tolerated and allows the liquid line to be smaller. This reduces the

refrigerant charge, which improves system reliability and lowers the installed

cost.

Subcooling Remaining at

Expansion Device

subcooling remaining at expansion device 15 = = F 9.9 – F 5.1F

subcooling remaining at expansion device 8.3 = = C 5.5 – C 2.8C

period four

Liquid Line

Other Considerations

When an evaporator has more than one distributor connected to a single

refrigeration circuit, the liquid line should be designed and installed to provide

equal inlet pressures to the distributors.

In this example, the liquid line is split and the same sizes and lengths of piping

are used to make the connections to the expansion devices and distributors.

This equalizes the pressure drops of the two connections and provides the

most-balanced expansion-device performance.

For proper refrigerant distribution into the evaporator coil, one TXV is required

for each distributor.

A sight glass is installed as close as practical, upstream of the expansion

device. This is a fitting that contains a glass window through which the

refrigerant entering the expansion device may be viewed and its state

liquid line

Evaporator Coil Connections

TXV

distributor

evaporator

solenoid

valve

liquid line

Refrigerant Accessories

sight

glass

TXV

liquid line

solenoid

valve

filter-drier

period four

Liquid Line

determined. As mentioned earlier, flashing of refrigerant upstream of the

expansion device is undesirable. A sight glass can be used to indicate that this

problem exists.

A solenoid-operated shutoff valve is typically installed in the liquid line to

facilitate system pumpdown or to prevent refrigerant migration when the

system is off. Solenoid valves may also be used to control the flow of liquid

refrigerant to multiple sections of the evaporator. In this application, a valve

would be installed upstream of the TXV for each individually controlled section

of the evaporator coil.

A liquid line filter-drier should be installed upstream of, and as close as

possible to, the solenoid valve and TXV. The filter-drier prevents moisture and

foreign matter, introduced during the installation process, from damaging the

solenoid valve or TXV. Manual ball valves should be installed for isolating the

filter-drier when the core needs to be replaced. It is important to realize that the

filter-drier is not a cure-all for poor installation practice. The piping must be

clean, and any air or moisture must be thoroughly evacuated from the system.

Finally, an access port is typically installed in the liquid line in a convenient

location and is used to charge the system with liquid refrigerant. It is also used

to measure the amount of subcooling in the system.

How the liquid line is routed can be important, particularly when the condenser

is located below the evaporator. As an example, the upper system in Figure 73

routes the liquid-line riser before a long horizontal run. The lower system

routes the riser after a long horizontal run.

In both cases, when the refrigeration circuit shuts off, the liquid refrigerant

inside the vertical riser will drain back into the condenser. In the upper system,

however, some of the liquid refrigerant from inside the horizontal run will also

drain back into the condenser. If this quantity of refrigerant is large enough, it

will fill the condenser and overflow into the discharge line and then into the

compressor. This can result in compressor lubrication problems, system

shutdown due to activation of a protective safety device, or mechanical failure.

Liquid Line Routing

period four

Liquid Line

Alternatively, a check valve can be installed in the liquid line, near the outlet of

the condenser, whenever the liquid line rises above the height of the condenser.

This check valve will prevent liquid refrigerant from draining back into the

condenser when the system is off.

A pressure-relief valve is required if both a check valve and a solenoid valve are

installed in the same liquid line. When the refrigeration circuit is off, these two

valves trap liquid refrigerant between them. The pressure-relief valve is

required to relieve the increased pressure that can result when the temperature

of this trapped refrigerant increases.

Generally, horizontal sections of the liquid line should be pitched to run

alongside of the accompanying suction or discharge line. This simplifies the

installation of the line set. Sloping the liquid line in one direction will not always

prevent the flow of refrigerant in the opposite direction because it may move

due to the siphon effect.

The refrigerant temperature inside the liquid line is generally higher than that of

the surrounding air. Therefore, insulation is not typically required. In fact, heat

loss may be desirable because it provides additional subcooling. If the liquid

line is routed through areas that are significantly warmer than the liquid

refrigerant (such as unconditioned attic spaces or boiler rooms), however, the

sections of the liquid line that are exposed to the warmer temperatures should

be insulated to prevent loss of subcooling.

liquid line

Other Considerations

▲ Pitch horizontal sections to run alongside

suction or discharge line

▲ Insulate lines that are routed through very

warm spaces

period five

Hot-Gas Bypass Line

The last refrigerant line to be considered is the hot-gas bypass line. The hot-gas

bypass line diverts hot, high-pressure refrigerant vapor from the discharge line

to the low-pressure side of the refrigeration system. This adds a “false load” to

the system that can help stabilize the suction pressure and temperature.

Hot-gas bypass is sometimes used to help prevent evaporator frosting in directexpansion (DX) applications, or in applications where the compressor cannot

be allowed to cycle on and off, such as a process-cooling application.

In general, however, hot-gas bypass should be avoided for most comfortcooling applications. It increases the refrigerant charge in the system, adds

more paths for potential refrigerant leaks, and increases the likelihood of

refrigerant-distribution and oil-return problems. It also increases the first cost of

the system, and the operating cost, because it prevents compressors from

cycling off with fluctuating loads.

period five

Hot-Gas Bypass Line

Refrigerant Piping

period five

Hot-Gas Bypass Line

Similar to the discharge line, the refrigerant flowing through the hot-gas bypass

line is in the form of a hot, high-pressure vapor. However, unlike the discharge

line, the minimum velocity is difficult to determine. A typical hot-gas bypass

valve is a modulating valve, so the refrigerant velocity through the valve and

piping can range all the way from zero flow to full flow. At times, the refrigerant

velocity inside a given pipe diameter will fall below the minimum velocity

required to carry oil droplets.

Therefore, it is not possible to select a pipe diameter that is small enough to

move oil at all valve positions. The hot-gas bypass line should be sized as small

as possible, without the maximum velocity being high enough to cause

objectionable noise. Because the hot-gas bypass line is basically an extension

of the discharge line, the recommended maximum velocity is 3,500 fpm

(17.5 m/s). Selecting the smallest possible line size also minimizes the amount

of oil and condensed refrigerant that could be trapped in the hot-gas bypass

line.

Also, the hot-gas bypass line must be installed so that oil and condensed

refrigerant can freely drain out of all sections of the line, at all operating

conditions.

hot-gas bypass line

Requirements for Sizing and Routing

▲ Ensure that oil and refrigerant are not trapped

in piping

▲ Avoid excessive noise

▲ Proper routing is more critical than pipe

diameter

period five

Hot-Gas Bypass Line

Hot-Gas Bypass to Evaporator Inlet

In a DX application, there are two bypass methods that are commonly used.

The first, and preferred, method bypasses refrigerant vapor from the

compressor discharge line to the inlet of the evaporator. Sensing a decrease in

suction pressure, the hot-gas bypass valve opens to bypass hot refrigerant

vapor from the compressor discharge line to the inlet of the evaporator,

between the TXV and the distributor. This adds a “false load” to the evaporator

and increases the suction pressure (and temperature).

The principal advantage of hot-gas bypass to the evaporator inlet is that the

refrigerant velocity inside the evaporator and the suction line is higher at low

loads. This promotes a uniform movement of oil through the evaporator and

the suction line.

Due to the widely varying velocity in the hot-gas bypass line, the selected

diameter of the pipe is not as critical as its routing. Install the hot-gas bypass

Hot-Gas Bypass to Evaporator Inlet

evaporator

condenser

compressor

hot-gas

bypass valve

TXV

distributor

suction line

liquid line

Figure 77

Hot-Gas Bypass Valve Installation

condenser

compressor

hot-gas

bypass valve

pitch back toward

discharge line

pilot

line

suction

line

discharge

line

pitch away

from hot-gas

bypass valve

hot-gas

bypass line

period five

Hot-Gas Bypass Line

valve above the discharge line and close to the compressor. Pitch the section of

pipe between the discharge line and the hot-gas bypass valve so oil will drain

back into the discharge line.

When using a pressure-actuated hot-gas bypass valve, tap the pilot line for the

valve close to the compressor inlet. This pilot line senses suction pressure, to

which the valve responds by modulating open or closed. Pitch the piping

downstream of the hot-gas bypass valve so that oil or condensed refrigerant

will drain away from the valve, toward the evaporator.

If the hot-gas bypass line includes a vertical riser of any height, or if the line

does not allow free drainage from the valve to the evaporator, the retention of

oil or condensed refrigerant within the riser must be considered. As mentioned,

because the refrigerant velocity within the hot-gas bypass line varies over a

wide range, no pipe, regardless of its diameter, can ensure adequate velocity to

carry oil up a riser of any height. Oil will collect at the base of the riser when the

hot-gas bypass valve throttles to lower flow rates.

This problem is commonly addressed by installing a small oil-return line

between the base of the riser and the suction line. First, a drain leg is provided

at the base of the riser to collect the oil and condensed refrigerant. This short

section of pipe should be about 12 in (300 mm) long and be the same diameter

as the riser.

Next, the oil-return line is connected to the side of the drain leg to prevent any

foreign material from plugging the tube. This oil-return line is constructed of

1/8 in. (3 mm)-diameter copper tubing and should be at least 5 ft (1.5 m) in

length. This length is required to create the necessary pressure drop between

the high-pressure and low-pressure sides of the system, yet this tubing is large

enough to drain oil and liquid refrigerant from the drain leg. If the oil-return line

must be longer than 5 ft (1.5 m), 1/4 in. (6 mm)-diameter tubing can be used for

the rest of the distance.

Hot-Gas Bypass Risers

hot-gas

bypass riser

suction

line

compressor

drain leg

oil-return

line

condenser

hot-gas

bypass valve

period five

Hot-Gas Bypass Line

The hot-gas bypass line is then piped to the evaporator inlet. It must be

connected to the top of the pipe, between the TXV and the distributor. Any hotgas bypass piping upstream of this connection should be pitched so that oil or

condensed refrigerant will drain toward the evaporator.

An orifice-type distributor is typically not able to permit adequate flow of

refrigerant vapor through the orifice. If this type of distributor is used, the

orifice must be relocated upstream of the point at which the hot-gas bypass line

connects to the distributor.

When the evaporator contains multiple distributors on a single refrigeration

circuit, the hot-gas bypass line should be piped to each distributor that is

“active” when the hot-gas bypass valve opens. The evaporator shown in

Figure 80 contains two distributors on the refrigeration circuit, each equipped

with a solenoid valve. In this example, the top solenoid valve closes when the

compressor unloads. Only the bottom distributor is “active” at minimum

capacity when the hot-gas bypass valve opens. Therefore, the hot-gas bypass

pipe only connects to the bottom distributor.

hot-gas bypass line

Connection to Evaporator

evaporator

distributor

liquid

line

hot-gas

bypass line

solenoid

valve

TXV

period five

Hot-Gas Bypass Line

The evaporator shown in Figure 81 also contains two distributors on a single

refrigeration circuit, but it only contains one solenoid valve. In this example, the

solenoid valve closes only after the last compressor on the circuit shuts off.

Both distributors are “active” at all times and the hot-gas bypass pipe must be

connected to both distributors.

When the hot-gas bypass line is piped to more than one distributor, check

valves should be installed in each section of pipe that feeds a distributor.

Without these check valves, a path is created between coil circuits whenever

the hot-gas bypass valve is closed. The coil circuits will become unequally

loaded and refrigerant will travel from the outlet of one TXV into the section of

coil that is served by a different TXV. This will result in the loss of superheat

control and potential compressor damage.

hot-gas bypass line

Connection to Evaporator

evaporator

distributor

hot-gas

bypass line

solenoid

valve

check

valves

TXV

period five

Hot-Gas Bypass Line

Hot-Gas Bypass to the Suction Line

The second bypass method conducts refrigerant vapor from the compressor

discharge line to the suction line. This method requires an additional expansion

valve, called a liquid-injection valve. The remote bulb of this liquid-injection

valve is attached to the suction line near the inlet to the compressor. When

reduced suction pressure causes the hot-gas bypass valve to open, the liquidinjection valve senses the resulting rise in suction temperature and opens to

mix liquid refrigerant with the bypassed refrigerant vapor. The hot refrigerant

vapor causes the liquid refrigerant to vaporize, cooling the mixture and

stabilizing the suction pressure.

The advantage of this method is that it requires less actual piping than

bypassing hot gas to the evaporator inlet. However, the disadvantage is that the

refrigerant velocity inside the evaporator and suction line (upstream of the hotgas bypass connection) can drop very low when the hot-gas bypass valve is

open. This may result in oil being trapped in the evaporator and the suction

line. For this reason, this method is only acceptable in applications where the

evaporator and the suction line are able to freely drain to the point at which the

hot-gas bypass line connects into the suction line. Therefore, hot-gas bypass to

the evaporator inlet is the preferred method.

Hot-Gas Bypass to Suction Line

evaporator

condenser

compressor

hot-gas

bypass valve

suction line

liquid line

liquid-injection

valve

TXV

period five

Hot-Gas Bypass Line

Again, routing of the hot-gas bypass piping is critical. Install the hot-gas bypass

valve above the compressor discharge and close to the compressor. Pitch the

section of pipe between the discharge line and the hot-gas bypass valve so that

oil will drain back into the discharge line. Tap the pilot line for the hot-gas

bypass valve close to the compressor inlet, and pitch the piping downstream of

the hot-gas bypass valve so that oil will drain away from the valve, toward the

suction line.

A separate pipe routes liquid refrigerant from the liquid line to mix with the hot,

bypassed refrigerant vapor. This mixture is then piped into the suction line at

an angle, at least 5 ft (1.5 m) upstream of the compressor inlet, and upstream of

the pilot line tap for the hot-gas bypass valve. This angled connection promotes

mixing, and the minimum distance helps ensure complete vaporization of the

liquid refrigerant before it reaches the compressor.

The liquid-injection line contains a thermal expansion valve, called a liquidinjection valve. The remote bulb for the liquid-injection valve is attached to the

suction line, downstream of the point at which the hot gas enters the suction

line. This valve is adjusted to maintain between 30°F and 35°F (-1°C and 1.5°C)

of superheat at the inlet to the compressor. A solenoid valve is installed

upstream of the liquid-injection valve to shut off the flow of liquid refrigerant

whenever the hot-gas bypass valve is closed.

hot-gas bypass line

Connection to Suction Line

condenser

compressor suction line

liquid

line

hot-gas

bypass valve

liquid-injection

valve

solenoid

valve

period five

Hot-Gas Bypass Line

Other Considerations

The hot-gas bypass valve, and the diameter of the hot-gas bypass line, are

selected based on the maximum amount of refrigerant vapor that will be

bypassed. This depends on the application and the purpose of the hot-gas

bypass system.

For a traditional comfort-cooling application, the hot-gas bypass valve is used

to modulate the amount of refrigerant vapor bypassed to maintain the suction

temperature at a setpoint, such as 28°F (-2.2°C). This typically requires the valve

to only open when the cooling load on the system drops below the minimum

compressor capacity.

For example, consider the same 20-ton (70.3-kW), single-circuit system that can

unload to a minimum capacity of 10 tons (35.2 kW). The hot-gas bypass valve

and line would probably be sized for 10 tons (35.2 kW) of capacity or less.

Assuming a maximum bypass capacity of 10 tons (35.2 kW), the refrigerant

velocity is determined for various pipe diameters using the chart in Figure 43

on page 34.

As explained at the beginning of this period, the diameter of the hot-gas bypass

line should be as small as possible while preventing the maximum velocity

from exceeding the recommended upper limit of 3,500 fpm (17.5 m/s). For this

example system, a 3/4 in. (18 mm)-diameter pipe is selected. If the overall

length of the hot-gas bypass line exceeds 75 ft (23 m), special precautions,

including increasing the line size, may be required.

period five

Hot-Gas Bypass Line

The full length of the hot-gas bypass line must be insulated to prevent the hot

refrigerant vapor from condensing before it reaches the evaporator.

As the length of the hot-gas bypass line increases, the risk of refrigerant vapor

condensing inside the line increases, even if the line is insulated. This can cause

a problem when the hot-gas bypass valve opens, because the refrigerant vapor

could force a slug of liquid refrigerant into the evaporator. This over-feeding of

the evaporator may cause liquid refrigerant to enter the suction line, possibly

slugging the compressor. For this reason, the overall length of the hot-gas

bypass line should be as short as possible.

If the refrigeration system includes a pump-down cycle, then a solenoid valve

must be installed in the hot-gas bypass line. A pump-down cycle closes the

liquid-line solenoid valve to stop the flow of refrigerant into the evaporator, and

allows the compressor to run for a short period of time. This pumps refrigerant

from the low-pressure side of the system (evaporator and suction line) to the

high-pressure side of the system (discharge line, condenser, and liquid line).

The solenoid valve in the hot-gas bypass line must also close when the system

begins the pump-down cycle, to prevent the high-pressure refrigerant vapor

from migrating to the low-pressure side of the system. Some hot-gas bypass

valves are equipped with an integral solenoid that can be used to hold the valve

closed for this purpose.

hot-gas bypass line

Other Considerations

▲ Insulate entire line to prevent refrigerant vapor

from condensing

▲ Minimize overall length of the line

▲ Install solenoid valve if system includes a

pump-down cycle

We will now review the main concepts that were covered in this clinic regarding

the design and installation of the interconnecting piping for vapor-compression

air-conditioning systems.

As discussed in Period One, when a refrigeration system includes fieldassembled refrigerant piping, the primary design goals are generally to

maximize system reliability and minimize installed cost. When the refrigerant

piping is designed and installed, it is imperative that the following requirements

are met:

 Return oil to the compressor at all operating conditions

 Ensure that only liquid refrigerant (no vapor) enters the expansion device

 Minimize system capacity loss that is caused by pressure drop through the

piping

period six

Review

Refrigerant Piping

Figure 86

Review—Period One

▲ Refrigerant Piping Requirements

◆ Return oil to compressor

◆ Ensure that only liquid refrigerant (no vapor) enters

the expansion device

◆ Minimize system capacity loss

◆ Minimize refrigerant charge

▲ If provided, use line sizes recommended

by the manufacturer

period six

Review

period six

Review  Minimize the total refrigerant charge in the system to improve reliability and

minimize installed cost

If the manufacturer of the refrigeration equipment provides recommended line

sizes, or tools for selecting the proper line sizes, use those line sizes.

Period Two reviewed the design of the suction line. The diameter of the suction

line should be selected small enough so the refrigerant velocity is high enough

to move oil, at all steps of compressor unloading. However, the pipe diameter

should be as large as possible to minimize pressure drop, maximize system

capacity and efficiency, and avoid causing objectionable noise.

The minimum allowable velocity in a suction line varies based on the diameter

of the pipe. The smaller the pipe diameter, the less velocity is required to move

the oil. Long line lengths and the associated higher pressure drop can be

tolerated, assuming that the loss of system capacity and efficiency is acceptable

for the given application. Of course, it is still good practice to minimize pressure

drop.

Review—Period Two

▲ Suction line

◆ Ensure adequate velocity to return oil to the

compressor at all steps of unloading

◆ Avoid excessive noise

◆ Minimize system capacity and efficiency loss

72 TRG-TRC006-EN

period six

Review

Period Three discussed the discharge line. Similar to the suction line, the

diameter of the discharge line must be small enough that the refrigerant

velocity is high enough to move oil, at all steps of compressor unloading.

However, the pipe diameter should be as large as possible to minimize

pressure drop, maximize system efficiency, and avoid causing objectionable

noise.

The liquid line was the focus of Period Four. The liquid line must be designed

and installed to ensure that only liquid refrigerant (no vapor) enters the

expansion device. This requires the condenser to provide adequate subcooling,

and the pressure drop through the liquid line and accessories must not be high

enough to cause flashing upstream of the expansion device.

An understanding of the entire refrigeration system, not just the piping, is

required to select the optimal liquid line size. The diameter of the liquid line

must be as small as possible to minimize the quantity of refrigerant in the

Review—Period Three

▲ Discharge line

◆ Ensure adequate velocity to return oil to the

compressor at all steps of unloading

◆ Avoid excessive noise

◆ Minimize efficiency loss

Review—Period Four

▲ Liquid line

◆ Ensure that only liquid refrigerant enters the

expansion device

◆ Minimize refrigerant charge

◆ Avoid excessive noise and pipe erosion

period six

Review

system and minimize the impact of changes in refrigerant density on the ability

of the condenser to provide subcooling. This improves system reliability and

minimizes the installed cost. However, if the pipe is too small, the increased

pressure drop may cause flashing upstream of the expansion device, and the

high velocity may result in objectionable noise or erosion to the inside surfaces

of the piping.

Period Five discussed the importance of proper routing when hot-gas bypass is

to be used. The two most common methods are hot-gas bypass to the

evaporator inlet and hot-gas bypass to the suction line. Hot-gas bypass to the

evaporator inlet is preferred because the refrigerant velocity in the evaporator

and the suction line is higher at low loads, resulting in improved oil return.

Because the hot-gas bypass valve modulates from full flow down to zero flow,

at times the refrigerant velocity will fall below the minimum velocity required to

carry oil droplets. Therefore, it is not possible to select a pipe diameter that is

small enough to move oil at all valve positions. The hot-gas bypass line must

be designed and installed so that oil and condensed refrigerant can freely drain

out of all sections of the line, at all operating conditions.

When used, the overall length of the hot-gas bypass line should be as short as

possible. In general, however, hot-gas bypass should be avoided for most

comfort-cooling applications.

Review—Period Five

▲ Hot-gas bypass line

◆ Ensure that oil and refrigerant are not trapped in the

piping

◆ Avoid excessive noise

◆ Proper routing is more critical than pipe diameter

period six

From*book-Trane Refrigerant Piping