technical paper #2 - iiarweb.iiar.org/membersonly/pdf/tc/t350.pdf · asme b31.3, process piping...

© IIAR 2005 33

Abstract

The paper examines issues faced by engineers, contractors and owners when designing, installing andoperating refrigeration piping. It reviews piping mechanical design by discussing applicable pipingcodes and the suitability of piping, castings, bolting and gasket materials as well as welding methodsfor low-, normal- and high-temperature service. It then analyzes pipe-sizing methods for a variety ofapplications, emphasizing the effects of undersizing or oversizing liquid piping, vapor lines and controlvalves. This evaluation is followed by a discussion on piping layout issues, in particular, layout ofcompressor suction mains, wet suction returns, elevated equipment and condensers. The paperconcludes with a presentation of special cases such as thermal expansion in liquid CO2 andsubcooled refrigerant lines, vapor condensation in hot gas and instrument tubing, liquid hammer,liquid carryover, internal and external corrosion, and a comparison of pumped and gravity-fedoverfeed systems.

2005 IIAR Ammonia Refrigeration Conference & Exhibition, Acapulco, Mexico

Technical Paper #2

Refrigeration Piping: A Simplified Guide to a Modern Approach

Paul DanilewiczPaul Orlando

Toromont Process SystemsHouston, Texas

ACKNOWLEDGEMENT

The success of the technical program of the 27th Annual Meeting of theInternational Institute of Ammonia Refrigeration is due to the quality of the technicalpapers in this volume. IIAR expresses its deep appreciation to the authors,reviewers, and editors for their contributions to the ammonia refrigeration industry.

Board of Directors, International Institute of Ammonia Refrigeration

ABOUT THIS VOLUME

IIAR Technical Papers are subjected to rigorous technical peer review.

The views expressed in the papers in this volume are those of the authors, not theInternational Institute of Ammonia Refrigeration. They are not official positions ofthe Institute and are not officially endorsed.

EDITORSM. Kent Anderson, President

Chris Combs, Project CoordinatorGene Troy, P.E., Technical Director

International Institute of Ammonia Refrigeration1110 North Glebe Road

Suite 250Arlington, VA 22201

+ 1-703-312-4200 (voice)+ 1-703-312-0065 (fax)

www.iiar.org

2005 Ammonia Refrigeration Conference & ExhibitionFairmont Acapulco Princess

Acapulco, Mexico

Technical Paper #2 © IIAR 2005 35

Introduction

Numerous articles, handbooks and manuals in our industry discuss good piping

practices for ammonia refrigeration systems. There is also a good understanding of

the major issues related to the subject among the designers, contractors, and end-

users of the refrigeration plants. The goal of this paper is not to repeat what has

already been written on the topic of ammonia piping. Instead, the paper shall

discuss the design issues that, in the authors’ opinion, are much ignored by the

industry but contribute significantly to the problems encountered in refrigeration

plants. The paper also presents modern piping design methods.

Mechanical Design Issues

This section covers the following mechanical design topics:

• Piping design for low temperatures

• Cast, malleable and nodular iron in refrigeration systems

• ASME piping classes

• ASHRAE 15 safety standard for refrigeration systems

Piping Design for Low Temperatures

Carbon steel is the material of choice for pipe, valves and other components used in

ammonia refrigeration systems and screw compressor lubricant service. At low

temperatures, however, carbon steel becomes brittle and loses flexibility and

toughness. For safety reasons, piping codes establish guidelines to ensure that the

materials used in piping systems provide adequate strength and flexibility at all

service temperatures.

The code most commonly used in industrial refrigeration piping is ASME B31.5,

Refrigeration Piping Code, although some end users, especially outside the food

Refrigeration Piping: A Simplified Guide to a Modern Approach — Paul Danilewicz and Paul Orlando

36 © IIAR 2005 Technical Paper #2

processing industry, require designers and installers to conform to the more stringent

ASME B31.3, Process Piping Code. (ASME B31.3, 2004; ASME, 2001)

Prior to the 2001 edition, the ASME B31.5 code stated that impact tests were not

required for ferrous material used in fabricating a piping system operating between

-20°F and -150°F [-29°C and -101°C], provided that the maximum circumferential or

longitudinal tensile stress, resulting pressure, thermal stresses, and bending stress

between the supports did not exceed 40% of the allowable stress of the material at

100°F [38°C].

The latest edition of ASME B31.5 brought the requirements more in line with the

methodology used in B31.3 and Section VIII of the ASME Boiler and Pressure Vessel

(B&PV) Code. (ASME B&PV, 2004) The new rules establish the minimum

temperature above which the material is exempt from impact test requirements

based on the grade of carbon steel and the thickness of the material. For operating

temperatures between -20°F and -50°F [-29°C and -46°C], if the system is to operate

below the nominal impact-test exemption temperature, the code allows the designer

to determine the additional reduction in minimum metal temperature for which

impact testing is not required. This reduction in minimum temperature is based on

the ratio of the design tensile stresses to the normal (room temperature) allowable

stress for a given material. (Figure 1) For metal temperatures below -50°F [-46°C],

the old 40% rule applies (i.e., impact testing is not required if this ratio is less than

40%.) Note that if the design temperature is below -55°F [-48°C] and the ratio is

less than 30%, ASME B31.3 requires impact testing.

Typically, for higher operating temperatures, designers specify SA-106 grade B

seamless or SA-53 grade B ERW piping. For the pipe wall thicknesses typically

encountered in refrigeration piping, these materials have a maximum allowable

stress rated at temperatures above -20°F [-29°C]. For lower operating temperatures,

the specifications quite often call for SA-333 grade 6 material, which is impact-tested

by the producing mill at a minimum temperature of -55°F [-48°C].



Suppliers charge a substantial premium for impact-tested carbon steel pipe and

fittings. A recent informal survey showed that SA-333 grade 6 piping had a 10% to

20% higher cost than SA-106 grade B seamless piping, and a 40% to 50% higher

cost than SA-53 grade B ERW. The premium is even higher for large forgings used to

manufacture flanges, unions and other butt-welded fittings. Figure 2 details a price

comparison among the different types of material.

For packaged refrigeration systems, the cost of low-temperature carbon piping as a

percentage of total job costs is not as large as for field-installed systems. The piping

runs are typically short on packaged units and use of low-temperature specialty

piping material is usually limited to compressor suction lines. The cost becomes

significant for packaged units if the customer is concerned that the piping is suitable

for the normal boiling temperature (saturation temperature) of the refrigerant at

atmospheric pressure and specifies that all piping shall be rated for this temperature.

In the case of ammonia, the equivalent saturation temperature at sea-level

atmospheric pressure is -28.0°F [-33.3°C]. For R-22, the equivalent saturation

temperature is -41.3°F [-40.7°C] and for R-507, -52.1°F [-46.7°C]. In the packaged

refrigeration systems supplied to refineries, hydrocarbons such as propane and

propylene are quite often the refrigerants of choice; their saturation temperatures

are -38.4°F and -48.7°F [-39.1°C and -44.8°C] respectively. Some customers

require that the piping system be rated for these low temperatures at full design

working pressures.

The situation is dramatically different for ammonia refrigeration systems in typical

food processing installations, which have hundreds of feet of large diameter pipe,

valves and fittings operating at metal temperatures below the minimum -20°F

[-29°C]. At times, large overfeed systems have substantial wet suction lines or

liquid lines that operate at temperatures of -40°F [-40°C] or lower.

Valve manufacturers use either cast/ductile iron or impact-tested cast steel materials

to ensure that their products can operate at suitably low temperatures. This same



practice applies to manufacturers of pressure regulators, level switches, etc.

However, to meet piping code requirements, system suppliers must either install

expensive impact-tested piping or prove that the piping system will not be

overstressed at low temperatures considering the reduced maximum allowable

stresses of the piping material.

In refrigeration systems, operating pressures at low temperatures are typically very

low because the vapor pressure of the refrigerant falls as the temperature drops. For

example, ammonia vapor pressure at 100°F is 211.9 psia [38°C/14.41 bar]. At -35°F

[-37°C], it is only 12.1 psia [0.823 bar]. It is apparent that the stresses imposed on

the piping due to direct refrigerant pressure will be generally low. The majority of

the stress that the designer will have to consider is a thermal stress due to

expansion or contraction of the piping when its operating temperature fluctuates

between ambient and operating temperatures.

Currently, system designers may choose from several software packages to verify

the stresses imposed on piping. By using this software, the system designer can

prove the suitability of non-impact-tested, less-expensive piping materials for

operation at low temperatures. In most cases, a software analysis for large systems

proves to be beneficial because non-impact-tested piping costs far less than

impact-tested materials.

Many companies use three-dimensional modeling as their standard method for

laying out the piping in the plant, which makes the analysis quite simple. The

models are used to estimate the labor required for the installation, to prepare bills of

materials, and to create isometric drawings for pipe spools. Once the model has

been created, the pipe runs can be evaluated using the stress analysis software. If

the initial analysis shows that the allowable stresses would be exceeded, the

designer can reduce them by modifying the proposed layout and location of the

anchor points. Installations designed in such a manner are safe, less expensive and

meet the piping code requirements. This type of design process does not just apply



to low-temperature applications; designers may check stresses in other types of

piping systems as well, including hot gas lines or compressor discharge lines.

Figures 3 and 4 illustrate a sample analysis of a -50°F [-46°C] wet suction line and

120°F [49°C] hot gas line. In the example, the suction main and hot gas runs were

first drawn using computer-aided design (CAD) software based on a typical roof

routing of the piping mains. Once the mains had been designed, the pipe supports

were added to the drawing, and the system was then evaluated using piping-stress-

analysis software. The analysis showed that both SA106 grade B and SA 333 grade 6

piping materials were suitable for the installation, based on the pipe size and

thickness, ambient temperature range, operating pressures and temperatures, and

fluid density.

Process requirements define operating temperatures. While the designer cannot

change this constraint, he or she can lay out the plant pipe in such a manner to

limit the imposed thermal stresses. When laying out system pipe runs, no single rule

assures that the proposed layout will not exceed the maximum allowable stresses for

the material. Instead, a common sense approach that considers the behavior of pipe

subjected to temperature fluctuations should guide the design process. The system

design must allow the pipe to move without breaking the welds and prevent it from

becoming overly stressed while it experiences temperature reductions or increases.

One good method to mitigate such stresses in a pipe run is to add a few elbows to

create a loop that can flex when the pipe run expands or contracts. Also, vessels and

attached pipe can be installed such that thermal expansion and contraction will not

cause forces acting on the nozzles to exceed the maximum allowable limits.

Like pipe, all other components must be suitable for operation at the design

operating temperature. The ASME B31.5 piping code does not require the use of

impact-tested carbon steel bolting for temperatures above -50°F [-46°C] as long as

the bolting used is high strength A193 grade B7. The designer also needs to confirm

that the elastomers used for gaskets and seals are not only compatible with the



refrigerant, but that they will maintain sealing capabilities at low temperatures.

Neoprene, which is the most common elastomer used in ammonia systems, is

usually rated for temperatures above -40°F [-40°C]. For temperatures below this

threshold, the designer may have to consider either PTFE gaskets or spiral wound

stainless steel gaskets that are suitable for very low operating temperatures.

Also, the welding procedures used in fabricating piping systems should be rated for

the minimum metal design temperature. A suitable Welding Procedure Specification

(WPS) and Procedure Qualification Record (PQR) must be prepared in accordance

with Section IX of the B&PV Code, and a set of pipe coupons must be impact-tested

at the minimum design metal temperature to be used for the procedure. The main

reason for the impact-testing requirement is to establish the heat input required to

achieve the desired weld toughness at the design temperature. Although the B&PV

Code does not specifically require that the welding consumables be impact-tested

for the minimum metal temperature, it may be beneficial to utilize impact-tested or

low-temperature-rated consumables to assure that the coupon passes the impact test.

Also, the welders must be currently qualified to perform the approved WPS.

It is worth pointing out that common stainless steel materials such as type 304, 304L,

316 or 316L do not require impact testing, nor does aluminum. Stainless steel and

aluminum are commonly used in the construction of evaporator coils in air units.

Cast, Malleable and Nodular Iron in Refrigeration Systems

Valve manufacturers commonly use cast, malleable and nodular types of iron. Most

of the relief valves and regulators used in ammonia refrigeration are made of gray

cast iron or ductile iron materials. The ASME B31.5 code allows cast and malleable

types of iron as long as they are not used for hydrocarbon or flammable fluid service

at temperatures above 300°F [149°C] or operating pressures above 300 psig

[20.4 barg]. Restrictions against nodular iron are somewhat different; the code

prohibits the use of nodular iron above 1000 psig [20.4 barg], which obviously



would not be a common case with ammonia refrigeration. Notably, the code allows

the usage of cast, malleable and nodular iron for temperatures above -150°F

[-101°C] without impact testing.

ANSI/ASME Piping Classes

The ASME B16 Committee developed several standards to ensure consistency in the

design and application of valves, flanges, fittings and gaskets. The committee

defined several pressure-temperature classes for a variety of cast and forged

materials. The two main steel flange classes defined by the ASME B16.5 standard

that are commonly used in ammonia refrigeration piping are Class 150 and 300.

(ASME, 2003) The similar classes for cast and nodular irons are defined as Class 125

and 250. The class defines the maximum allowable working pressure for which a

given component can be used depending on the system operating temperature.

Figure 5 details the basic pressure ratings for both classes and common carbon and

stainless steel materials. Class 150 and 300 represent nominal 150-psig [10.2-barg]

and 300-psig [20.4-barg] designs, respectively. For carbon and stainless steel fittings,

the maximum allowable working pressure decreases as temperature increases. For

example, for Class 150 carbon steel flanges, the maximum allowable working

pressure at 100°F [38°C] is 284 psig [19.3 barg] and falls to 200 psig [13.6 barg] at

392°F [200°C]. The examination of the pressure-temperature charts for each class

shows that if the refrigeration system is designed entirely for 300-psig [20.4 barg]

maximum allowable working pressure, Class 150 valves, flanges or other

components should not be used. As indicated above, for Class 150, the maximum

pressure rating at any temperature is less than 300 psig [20.4 barg]. This restriction

is also applicable to compressor suction flanges or flanges supplied with pressure

vessels, heat exchangers or pumps. Some compressor and pump manufactures

utilize Class 150 flanges at compressor or pump suction connections; thus, they

should not be used if the low side of the system is protected with a relief valve set

for 300 psig [20.4 barg].



ANSI/ASHRAE 15-2004 and Mechanical Design

ASHRAE 15-2004, a model refrigeration safety code, provides recommendations for

the design pressures for refrigeration systems. The code states that all equipment

should be designed for a 29” Hg vacuum [23 Torr]. For the low side of ammonia

systems, the code requires 138 psig [9.39 barg] minimum design pressure, which is

equivalent to 80°F [27°C] saturation temperature. For the high side, with an

evaporative or water-cooled condenser, the code recommends a 210-psig [14.3-barg]

design, which corresponds to 104°F [40°C] saturation temperature. If air-cooled

condensers are used, then the minimum high side design pressure should be

280 psig [19.0 barg], equivalent to a 122°F [50°C] saturation temperature.

If a designer expects ambient or inside temperatures hotter than 80°F [27°C],

he/she should then design the system for the pressure equivalent to the maximum

expected temperature, with at least a 15% margin of safety to account for tolerance

in the relief valve set pressure. For example, for a 95°F [35°C] ambient, which is a

common temperature experienced in the summer in the southern U.S., the

equivalent saturation gauge pressure is 181 psig [12.3 barg]. At 100°F [38°C], the

ammonia vapor pressure is 197 psig [13.4 barg]. For these temperatures, even a

200-psig [13.6-barg] low side design appears to be insufficient.

Designing the low side for 250 psig [17.0 barg] can prevent the relief valves from

opening and releasing ammonia during elevated ambient temperatures. For systems

designed for 250 psig [17.0 barg] and operating temperatures below 212°F [100°C],

Class 150 flanges and fittings can still be used on the compressors, pumps, vessels

and heat exchangers.

At times, designers specify the high side of the system for 300 psig [20.4 barg].

Many ammonia systems use evaporative condensers operating at 181 psig

[12.3 barg], with occasional excursions to 210 from 240 psig [14.3 to 16.3 barg]. The

differential between these pressures and the 300-psig [20.4-barg] relief valve set



point permits the valve to remain securely closed, preventing weeping that can occur

when the differential is less than 25 psi [1.7 barg], as would be the case had the

design pressure been chosen at 250 psig [17.0 barg].

Sizing of Piping Runs

New Tools for Selecting Refrigerant Pipe Sizes

Design engineers may choose among a wide variety of selection charts for sizing

refrigerant piping. These charts are generally based on the Darcy equation, which is

suitable for single-phase flow of both liquids and gases as long as the average flow

velocity does not exceed approximately 0.3 Mach. For vapor flow, the equation can

be used only when the pressure losses in the pipe run are relatively low, which is

the case for a majority of pipe runs. This limitation is due to the fact that for vapors,

as the pressure decreases, the specific volume increases. Thus, for the same mass

flow rate the average flow velocity increases. However, because the Darcy equation

assumes a constant average flow velocity for calculation of the pressure drop, it

becomes less accurate as vapor pressure drops increase in size.

This Darcy-based methodology does not address two-phase flow situations, common

for all overfeed systems. The ammonia refrigeration industry has circumvented this

problem by the practice of selecting wet suction lines one size larger than the dry

suction line required for the given duty. Also, some refrigeration system designers

and pipe sizing software developers have adopted Beattie’s simplistic method of

calculating pressure drop in two-phase flow. (Beattie, 1982)

For some time, the process industry has used simulation software to aid engineers in

designing process equipment, such as absorbers, strippers, mixers, etc. Industries

that compress gases also use this software, especially for compression of wet mixed

hydrocarbon gases. Further, the software aids in the design of carbon dioxide



purification and liquefaction plants. The software output is essential engineering

data that is then used for selecting all heat exchangers, pressure vessels and piping

for single- and two-phase flows in the plant. A typical Process Flow Diagram, a

software output report, and an example of pipe sizing are included in Figures 6

through 8.

There are many advantages to using simulation software for plant design. The

software allows for a detailed analysis of the plant performance, including predicting

system performance under different operating situations. It eliminates guessing in

the design process by rigorously calculating all relevant flow data. It assures that all

pipe sizes are correct and designed for the actual plant requirements. The software

output can also be used to confirm that compressors, pumps and other equipment

selected have enough capacity. The software allows for optimizing plant

performance by allowing the designer to evaluate different what-if scenarios.

Obviously, the software can also be a handy tool for analyzing existing installations.

The Effects of Undersizing or Oversizing Piping Systems

Liquid Lines: It is essentially impossible to oversize liquid lines from a flow and

pressure drop point of view. The designer has to be aware that the larger-than-

necessary liquid lines will contribute to overall plant liquid inventory, which in turn

will require a larger receiver volume, adding cost to the installation. Undersized

liquid piping causes liquid pressure to drop, which in turn will cause flash gas to

form during flow. Flash gas has an adverse impact on the performance of expansion

valves. This problem is especially acute for liquid lines containing saturated liquid.

For pumped liquid overfeed or pressure-fed liquid systems, excessive pressure drop

in liquid piping can result in insufficient liquid pressure at the expansion valve to

overcome the pressure drop at the required flow through the evaporator.

When analyzing liquid lines, the designer has to be aware of the consequences of

installing larger-than-necessary liquid control valves. In the case of thermostatic



expansion valves, if the valve is too big, it can become unstable, which in turn can

result in poor control and eventual flooding of the evaporator. The valve seats of

automated or hand-operated expansion valves for flooded evaporators will wear out

faster if forced to operate barely open for long periods. On the other hand, if the

control or expansion valve is too small, the evaporator will starve. Also, some

solenoid valves require a minimum pressure differential to operate properly. If liquid

line solenoid valves are oversized, there may not be enough flow to create a

pressure drop adequate to hold the valve open.

Designers should also take advantage of the fact that the plant can operate more

efficiently at lower-than-design ambient dry- or wet-bulb temperatures, which occur

the majority of the time. At low ambient conditions, the differential pressure

between the high and low side of the system can be substantially less than the

design differential pressure, which reduces the required compressor horsepower.

Therefore, both liquid and hot gas lines should be sized for flow at the reduced

differential pressure to take advantage of these conditions. Otherwise, the plant will

pay an expensive penalty due to the excessive power consumption.

Suction Lines: Special care should be taken when sizing suction lines. The lower the

design suction temperature, the greater the effect of undersizing. Undersized low

temperature suction piping and components will result in increased compressor

horsepower requirements for the same refrigeration duty. As the pressure decreases,

the gas specific volume and the horsepower-per-ton increase (assuming constant

condensing pressure). The lower suction pressure requires a greater compression

ratio to compress the refrigerant vapor to condensing pressure.

As the suction pressure decreases, the required compressor volumetric flow

increases. For the same duty, this may result in having to provide larger compressor

capacity, thus increasing equipment costs. A good approach to avoid excessive

capital costs is to establish the optimum economical velocity based on a comparison

of initial and operating costs. (Richards, 1984) Based on the historical relationship



between these costs, several design recommendations for the compressor suction

lines have been developed. One contractor proposed the following recommendations

in a form of overall pressure drop per 100 feet of equivalent length of pipe:

-40ºF — 0.15 to 0.20 psi

-20ºF — 0.25 psi

+15ºF — 0.5 psi

+35ºF — 1 psi

Another economics-based method is described in IIAR’s Ammonia Refrigeration

Piping Handbook. (IIAR, 2004) The Handbook includes several tables for both

suction and discharge lines.

Again, the overall pressure loss must be considered when selecting isolation and

control valves. For isolation valves, angle or butterfly designs are recommended

because they require less pressure drop than globe valves. Control valves should be

selected and sized to allow proper operation of the valves at the lowest pressure

drop that is economically feasible. However, if the system design requires a greater

pressure drop for a certain piece of equipment (e.g., multiple evaporators operating

at different evaporating temperatures connected to a common suction header), then

the designer should disregard this guideline.

Check valves should also be specified taking into account the overall pressure loss.

When selecting check valves, both the operating and cracking pressure drop required

to lift or open the valve should be considered. Cracking pressure is the pressure drop

required for the valve to open and overcome valve spring pressure and/or disk or

piston weight. Oversized piston-type check valves will cause the piston to operate

too close to the seat, which, in the event of a minor upset, will cause the piston to

bounce off the seat. Also, when excessive forces are required to open check valves,

the resulting pressure drop fluctuation across the valve may lead to bouncing or

chattering and eventually damage the valve. The use of split-wafer check valves

allows minimum cracking force and low operating pressure drop if the valve is



properly sized and installed. These valves are available with low-torque springs that

have very low valve cracking pressures.

In summary, when rating and selecting compressors, the designer must consider all

combined pressure drops between the evaporator and the compressor flange.

Hot Gas Lines: During hot gas defrost, the required hot gas duty is about twice the

refrigeration capacity of the evaporator that operates at a 10ºF [5.6K] temperature

difference. Hot gas piping should be sized to accommodate the volume of hot gas

required to condense this amount of vapor. Hot gas headers should be sized to

accommodate all of the evaporators scheduled to defrost at the same time. The

volume of hot gas flow is determined by the difference between the compressor

discharge pressure and the highest defrost pressure regulator setting on any of the

evaporators in the system. Piping should be designed to permit adequate flow even

when the condensing pressure is relatively low, such as during winter months. At

lower-than-design ambient conditions, undersized hot gas lines waste energy

unnecessarily and prolong defrost cycles. Extended defrost cycles give the

evaporators less time to recover between cycles, potentially compromising room

temperatures or causing product quality issues.

Piping Layout and Special Issues

Thermal Expansion of Subcooled Liquid

Subcooling: Subcooled liquid is frequently used in ammonia refrigeration plants and

is produced by cooling liquid below its saturation temperature at the prevailing

pressure. Subcooling may increase the efficiency of the refrigeration cycle and

prevent the formation of flash gas in downstream piping. In a two-stage system, for

example, by exchanging heat between liquid supply A (which feeds low-temperature

loads) and liquid supply B (which flashes off and is directed to the high-stage



compressor), liquid supply A would become subcooled and its refrigerating effect

would increase. While the flash gas would increase the volume normally handled by

the high-stage compressor, the subcooling effect would reduce the amount of gas

handled by the low-stage compressor. However, because the high stage compressor

operates more efficiently in terms of Btuh/BHP than the low stage, the process

yields a net gain in efficiency.

Subcooling is required in piping that suffers excessive frictional pressure losses or

that must be delivered to elevated evaporators. In the case of elevated evaporators

fed with saturated ammonia at 95°F [35°C], for every 3.92 ft [1.19 m] of elevation

increase, the liquid vapor pressure falls by 1 psi [0.068 bar], causing subsequent

flashing of liquid.

Preventing flash gas is also important for sizing liquid expansion valves. Most of the

selection charts and software assume that pure liquid exists upstream of the

expansion device. If this is not the case, the device will either not operate properly,

or will be too small for the duty, and may eventually fail.

Thermal expansion of cold liquid: Consider a case in which we have trapped

subcooled liquid ammonia in a 200’ length of 2” schedule 40 pipe. If the subcooled

liquid were allowed to warm up from 10°F to 95°F [-12°C to 35°C], for example,

during system shutdown, it would expand, but how much?

The density of subcooled liquid ammonia at 10°F [-12°C] is approximately

40.9 lb/ft3 [656 kg/m3]. Therefore, the specific volume of the liquid is

0.0244 ft3/lb [0.00152 m3/kg]. At 95°F [35°C], its density would be 36.7 lb/ft3

[589 kg/m3]; specific volume would be 0.0272 ft3/lb [0.00170 m3/kg]. Thus, each

pound of trapped liquid would increase in volume by 0.0028 cubic feet [0.18 liters].



The volume of the pipe in our example, 200’ of 2” schedule 40, is calculated:

200’ x 0.0233 ft3/linear foot = 4.66 ft3 [0.132 m3].

The mass of this volume of ammonia at 10°F [86.6 kg at -12°C] is:

4.66 ft3 x 40.9 lbs/ft3 = 190 lbs.

After warming to 95°F [35°C], the ammonia will expand to a volume of:

190 x 0.0272 = 5.17 ft3 [0.147 m3]

The net increase in liquid volume, then, would be:

5.17 – 4.66 = 0.51 ft3 [0.015 m3].

However, liquid ammonia cannot be compressed. Therefore, as trapped liquid warms,

the pipe will at first expand and stretch to accommodate the larger volume and rising

pressure. As the pressure exceeds the mechanical limits of the pipe, it will fail.

Fortunately, the designer of the system can address this problem by providing the

liquid a place in which to expand, such as an upstream or downstream vessel or

heat exchanger. Isolation valves at both ends of a liquid line should also be tagged

or labeled, advising the operator not to close both valves simultaneously. In

addition, liquid relief valves can be installed and piped either to a relief header or a

low side pressure vessel. In carbon dioxide installations, it is a common practice to

protect subcooled and saturated liquid lines with “pop-up” hydrostatic relief valves.

These relatively small and inexpensive devices prevent the rising of hydrostatic

pressure in the piping system by relieving expanding liquid to the atmosphere.

A similar problem is encountered in low temperature cascade systems. These

systems utilize various refrigerants, including carbon dioxide. Because the critical

temperature of carbon dioxide is below typical ambient temperatures, it will develop

a high pressure at relatively moderate temperatures. For example, saturated carbon

dioxide’s vapor pressure at 50°F [10°C] is approximately 652 psia [44.4 bar]. In the



case of R-23, another commonly used low temperature refrigerant, the vapor

pressure at the same temperature is 472 psia [32.1 bar]. In those installations, any

trapped saturated liquid must be vented to prevent the system pressure from rising

above its maximum allowable design pressure. Hydrostatic relief valves can be used

for this purpose.

Vapor Condensation in High Pressure Lines

One problem commonly overlooked is the effect of low ambient temperatures on

high-pressure refrigerant vapor lines, encountered when routing compressor

discharge lines or high-pressure vent lines through air-conditioned spaces or when

high-pressure pipes are installed outside the building.

At 95°F [35°C] condensing temperature, the ammonia vapor pressure is about

195.7 psia [13.31 bar]. If a system is shut down under this condition and the pipe

pressure is maintained at the pressure equivalent to the high ambient temperature,

then once the pipe wall temperature starts falling, the vapor refrigerant in the pipe

will begin to condense.

Condensing liquid may flow back down the compressor discharge lines and

accumulate on top of the compressor discharge check valve. If the check valve is not

fully closed, liquid will drain into the compressor’s oil separator. If the discharge line

is large enough and the oil separator does not have a heat source sufficiently great

to boil off all of the condensed liquid, the refrigerant will increasingly dilute the

compressor oil. At high levels of oil dilution, the oil viscosity can fall so far that the

oil will not provide sufficient bearing lubrication, causing compressor failure once

the compressor is restarted.

An easy and relatively inexpensive solution to this problem is to use inverted traps

for individual compressor discharge risers at the discharge header, or trap discharge

lines in which any condensed liquid can collect. The trap can be either drained with



a liquid drainer to a high or low side vessel, or liquid can be allowed to boil off once

the system is restarted. Additionally, heat tracing and insulation may be applied to

maintain a warm pipe surface temperature to prevent liquid from condensing.

Differential pressure transmitters are not commonly used in ammonia refrigeration.

However, where they are used to measure liquid level in the high-pressure vessel, it

is important to heat-trace and insulate the vapor leg of the device. Otherwise, on a

cold day, liquid will condense in the tubing and stack up on the low side of the

transmitter. This condition may cause the transmitter to falsely indicate low liquid

levels in the vessels. Likewise, the tubing runs to the pressure transmitters and

pressure transducers on the high side of the system should not be trapped. They

should be installed to allow ammonia condensate to drain freely into the pipe.

Layout of Suction Piping and Liquid Carry-over

It is very important to prevent overfed liquid from reaching refrigerant compressors.

In a liquid-vapor separator vessel, the liquid-vapor flow velocity is reduced below a

certain critical velocity, and the vapor can no longer fully entrain liquid droplets.

The liquid droplets fall out of the vapor stream and settle into the liquid pool. The

separated vapor is saturated (but not wet), and is collected to flow into the

compressor. The key distinction between saturated and wet vapor is that saturated

vapor does not carry any free liquid droplets, while wet vapor includes liquid. All

liquid overfeed systems have wet return lines. The compressor suction lines should

be dry.

The maximum allowable flow velocity is usually determined from a well-known

industry equation:

Vmax = k


ρl – ρv

ρv


where:

Vmax = maximum vapor flow velocity (ft/sec)

k = flow coefficient

ρl = liquid density (lb/ft3)

ρv = vapor density (lb/ft3)

Separation vessel manufacturers provide the maximum vessel capacity, expressed in

most cases in terms of tons of refrigeration. The capacity depends on the vessel size,

operating temperature, maximum allowable vessel flow velocity, vapor density, and

mass flow rate.

For example, consider a load of 80 tons of refrigeration (TR) [280 kW] at -40°F

[-40°C] with a 90°F [32°C] liquid supply. We wish to calculate the mass flow. First,

we determine the refrigerating effect of each pound of ammonia by subtracting the

enthalpy of the liquid (143 Btu/lb) from the vapor enthalpy (597 Btu/lb). The result,

454 BTU/lb [293 W-hr/kg], is the refrigerating effect.

We convert the TR units into Btu/hr and multiply by the refrigerating effect to

obtain the mass flow. The estimated mass flow is:

80TR x x =2,115 lb/hr [961 kg/hr]

Ammonia vapor density at -40°F [-40°C] is 0.04 lb/ft3 [0.64 kg/m3]. We divide the

mass flow by the density to obtain the volumetric flow rate:

x x = 881 ft3/min [25.0 m3/min]

Therefore, for a given vessel internal diameter, the flow velocity inside the vessel

can be easily established by dividing the volumetric flow rate by the vessel cross-

sectional area and comparing it with the maximum allowable velocity.


12,000 Btu

Ton – hr

2,115 lb

hr

1 ft3

0.04 lb

1 hr

60 min

1 lb

454 Btu


The plant designer can depend on experienced refrigeration vessel manufacturers to

provide a vessel with a suitable capacity for the design conditions or follow the

maximum separating velocity guidelines published by ASHRAE.

Several additional factors must be considered to assure that the objective of liquid

separation is achieved. First, the designer must consider off-design conditions (e.g.,

system pull-down, heavy evaporator defrost duty) during which the actual flow may

be well above the steady state flow conditions. Second, all the wet and dry suction

vapor piping should be sloped back to the separation vessel to make sure that any

liquid condensed or entrained in the stream is returned into the separation vessel.

Third, all individual compressor suction lines should be piped to the top of the main

header in lieu of the bottom or side to prevent liquid refrigerant from draining back

down the suction riser during the off-cycle.

Trapped lines, obviously, does not slope back to the vessel, so individual suction lines

and mains should not be trapped. Traps will accumulate liquid until it is pushed back

to the separator, which in turn can overload the separator and create overall system

instability. If it is not possible to drain the liquid back to the vessel, then a means to

return liquid must be provided (e.g., collection pots, a double riser return).

Another issue that needs to be considered while designing liquid-vapor separation

systems is how to return the collected liquid back into the system. If the separator

operates at an intermediate pressure, then liquid can be drained back to the low side

of the system utilizing the differential pressure between the intermediate and low

sides. If the separator also acts as an intercooler, collected liquid can vaporize during

desuperheating of the boosters discharge vapor. The low side separators, however,

require a source of heat to boil off the refrigerant. In ammonia systems, this is

typically achieved with a boil-off coil installed in the bottom of the vessel through

which warm liquid refrigerant is circulated. Another suitable, but much less efficient

way, is to use an electric heater or vessel heat tracing. It is also possible to collect

liquid in a liquid trap and return it by pushing it with discharge gas, or a pump, into

an intermediate pressure vessel.



Risers for flooded shell-and-tube evaporators operate below their flooding velocities

much of the time and must be large enough to allow separated liquid to drain back

while allowing vapor to flow out of the exchanger.

Pumped and Gravity-fed Evaporators

Pumped systems are common for liquid overfeed applications. Gravity-fed (also

called thermosyphon) systems are also frequently used for evaporators and for

cooling compressor oil.

Piping of pumped systems is quite simple. The refrigerant pump develops sufficient

head, normally 30-40 psi, to deliver liquid to even the most remote evaporators in

the plant. The piping from the pump to the evaporators can be routed in any

manner and the system will operate properly as long as the pump flow is large

enough for the duty and pump head is greater than the pressure drop between the

pump and the evaporator.

In the case of gravity fed evaporators, the driving force to supply liquid to the

evaporator is the elevation difference between the liquid level above the evaporator

and the evaporator itself, plus the difference between the density of liquid in the

supply line and the liquid-vapor mix in the return piping. This force must overcome

liquid line pressure losses, pressure drop across the evaporator and the static head of

the liquid-vapor mix in the return line.

It should be noted that for gravity systems, the amount of liquid supplied to the

evaporator would depend on the evaporator heat load. As the heat load increases,

the amount of liquid vaporized increases, causing the density of vapor liquid-vapor

mix leaving the evaporator to decrease. Thus, as the heat load increases, the amount

of liquid flowing into the evaporator also increases, which increases the pressure

losses (which are proportional to the square of the flow velocity).



If the system is designed for a 3 to 1 overfeed ratio at the given heat load, then the

liquid supply line will carry 3 times as much liquid as necessary for vaporization.

Thus, the return line will carry two times more liquid than vapor in terms of mass

flow. If the piping is undersized or pressure losses through the evaporator are too

high, the system will self-compensate by moving less than design overfeed ratio

through the evaporator because the flow velocity will decrease and the density

difference between supply and return will increase.

In the case of gravity-fed systems, it is important that the evaporator’s surge drum is

sized properly and capable of storing fluctuating and transient liquid levels without

tripping the high-liquid-level switch installed on the drum. Other important issues

requiring attention in gravity-fed systems are oil management and prevention of oil

and water accumulation in the evaporator coil. The most efficient way to prevent oil

accumulation in the evaporator is to avoid drawing liquid from the bottom of the

surge drum, which permits oil to accumulate on the bottom, and to provide for

draining oil from the bottom of the surge drum and the coil.

Both pumped and gravity systems can be utilized for a variety of evaporators and

refrigerants. In the process refrigeration industry, thermosyphon arrangements are

popular for heat exchangers that cool viscous liquids, such as ethylene glycol. In

those cases, the viscous brine flows through the shell side of the evaporator while

liquid refrigerant is supplied to the tube side of the exchanger. For the same reason,

oil flows on the shell side of a typical shell-and-tube thermosyphon oil cooler.

Not only ammonia, but also a variety of other refrigerants are used in pumped or

gravity systems in process refrigeration. The installations utilizing the other

refrigerants are designed in a similar fashion. However, for most of these, the

overfeed ratio is usually lower, because the amount of mass required for

vaporization is much greater. As a result, the velocity through the exchangers is high

enough for a reasonable heat transfer coefficient even at low overfeed ratios.



Avoiding Liquid Hammer

Special precautions must be taken in the design and layout of refrigeration piping to

avoid the conditions that allow for the sudden deceleration of liquid flow, which can

result in failure of components and/or piping. This phenomenon is commonly

known as liquid hammer. It can occur when high-pressure vapor is suddenly

introduced in a low-pressure line partially filled with liquid. The liquid becomes

entrained, forming a slug that grows and accelerates until it seals the pipe off

completely and is propelled with the full velocity of gas flowing in the pipe. Liquid

hammer can occur during defrost either when pressure is released too quickly from

an evaporator at the end of the cycle, or, when condensate that forms in a hot gas

line is propelled with the hot gas when released into an evaporator.

Because hot gas velocities can easily reach 100 feet per second, the pressure of the

liquid slug upon impact can exceed 3,000 psig [200 bar] (IIAR, 1992) Soft gas defrost

cycles may be used to prevent this situation and piping should be designed to

prevent liquid slugs from hitting a dead end in the piping.

Cleanliness

With carbon steel piping, particulates and moisture potentially may remain in the

system after assembly. These contaminants can wreak havoc with components such

as solenoids, check valves and regulators, plug strainers and orifices, and worst of all,

damage compressors. After assembly, the piping system should be cleaned and dried.

Steps should be taken to reduce the amount of contaminants initially introduced into

the system:

• To reduce the amount of rust in the system, piping should be kept capped at all

times until it is welded to the system.

• Welding slag should be removed or procedures used that reduce the amount of

slag introduced into the system.



• Procedures should be used to keep foreign materials from being placed or

forgotten in the system such as rags, welding rods, soap stones and other items

and checks performed to insure such items are not left in the system before

sealing.

• Prior to system start up, socks should be installed in strainers to collect the initial

mass of particulate.

Water Contamination

The introduction of moisture in the system should be minimized. Purchasing

ammonia from a reliable vendor who provides refrigerant grade ammonia with

certification is one step to help reduce moisture content. The sometimes-overlooked

situation during installation is when a pipe penetration, through the building wall,

for example, is sealed on the cold side but not sealed on the warm side. When the

cold side is subject to temperatures below the dew point of the air on the warm

side, water will condense inside the pipe even in dry climates. If the water is not

removed, it will introduce a large volume of moisture into the system.

The refrigeration system should be initially dried out to an acceptably low moisture

content and should be periodically checked. The effects of moisture contamination

can be found in IIAR Bulletin No. 108. (IIAR, 1986)

Purgers should be used to remove air in the systems that operate in a vacuum. A

method of removing air-borne moisture should also be provided. An ammonia/water

distiller is a good way to periodically remove moisture.

Excessive oil is also considered to be a contaminant in the system. Equipment

should be in place to reduce the amount of oil introduced into the system by having

properly sized oil separators. The oil pots and other oil accumulation point drains to

remove accumulated oil from the system should also be provided. Where applicable,

systems should be in place to return or recycle oil back to the compressors.



Conclusion

The paper reviewed only the most common mechanical, sizing and layout challenges

experienced designing ammonia refrigeration systems. The IIAR Ammonia

Refrigeration Piping Handbook should be consulted for detailed design

recommendations and practices. New design tools, used in the chemical process

industry, have also been presented to show the readers the direction in which the

refrigeration industry may turn in the near future.



References

ASME. B31.5: Refrigeration Piping and Heat Transfer Components. American Society

of Mechanical Engineers (ASME). 2001.

ASME. B16.5: Pipe Flanges and Flanged Fittings. American Society of Mechanical

Engineers (ASME). 2003.

ASME B31.3. B31.3: Process Piping. American Society of Mechanical Engineers

(ASME). 2004.

ASME B&PV. Boiler and Pressure Vessel Code: Section VIII, Pressure Vessels.

American Society of Mechanical Engineers (ASME). 2004.

Beattie, W. International Journal of Multiphase Flow. Vol. 8, No 1, pp 83-87. 1982.

IIAR. Bulletin No. 108: Guidelines for: Water Contamination in Ammonia

Refrigeration Systems. International Institute of Ammonia Refrigeration (IIAR). 1986.

IIAR. Bulletin No. 116: Avoiding Component Failure in Industrial Refrigeration

Systems Caused by Abnormal Pressure or Shock. International Institute of Ammonia

Refrigeration (IIAR). 1992.

IIAR. Ammonia Refrigeration Piping Handbook. International Institute of Ammonia

Refrigeration (IIAR). 2004.

Richards, W.V. Practical Pipe Sizing for Refrigerant Vapor Lines. Proceedings of the

IIAR Ammonia Refrigeration Conference & Exhibition. 1984.



Figure 1: Impact Exemption Criteria from ASME B31.5-2001


Reprinted by permission from ASME



Figure 2: Cost Escalation Factors for Low-Temperature Materials

960"

1200"

2400"

2400"

960"

960"960"

1218"

24"

24"

72"

72"

72"

72"

240

"

240

"

8" SUCTION LINE

3" HOT GAS LINE

Figure 3: Suction and Hot Gas Mains in Pipe Stress Analysis



Figure 4: Sample of Suction Main Stress Calculations



Figure 5: Pressure Rating for Different Piping Classes



Fig

ure

6:

Sam

ple

So

ftw

are

Ou

tpu

t -

Pro

cess

Flo

w D

iag

ram



Figure 7: Sample Software Output - System Mass Balance



Figure 8: Sample Software Output - Pipe Sizing Calculations


Notes:


Notes:

technical paper #2 - iiarweb.iiar.org/membersonly/pdf/tc/t350.pdf · asme b31.3, process piping...

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