ARNOLD, K. (1999). Design of Gas-Handling Systems and Facilities (2nd ed.) Episode 2 Part 4 pdf

Single-acting cylinder, crank endDouble-acting cylinder where RLC = rod load in compression, ib RL, = rod load in tension, Ib ap = cross-sectional area of piston, in.2 Pd = discharge pre

Single-acting cylinder, crank end

Double-acting cylinder

where RLC = rod load in compression, ib

RL, = rod load in tension, Ib

ap = cross-sectional area of piston, in.2

Pd = discharge pressure, psia

Ps = suction pressure, psia

Pu = pressure in unloaded area, psia

ar = cross-sectional area of rod, in,2

The calculations shown above provide the gas load imposed on the rod(and crosshead bushing) by the compressor cylinder piston To provide areasonable crosshead pin bushing life, the rod loading at the crossheadbushing must change from compression to tension during each revolu-tion This is commonly referred to as "rod reversal" and allows oil tolubricate and cool one side of the bushing while load is being applied tothe other side of the bushing

A single-acting, head end cylinder will not have load reversal if tion pressure is applied to the crank end Similarly, if discharge pressure

suc-is applied to the head end of a single-acting, crank end cylinder, loadreversal will not occur

In addition to the gas load, the rod and crosshead pin bushing is ject to the inertia forces created by the acceleration and deceleration ofthe compressor reciprocating mass The inertia load is a direct function

sub-of crank radius, the reciprocating weight, and speed squared The totalload imposed on the crosshead pin and bushing is the sum of the gas loadand the inertia load and is referred to as the "combined rod load."

The combined rod load should be checked anytime the gas loads areapproaching the maximum rating of the compressor frame or anytime rodreversal is marginal or questionable

COOLING AND LUBRICATION SYSTEMS

Compressor Cylinder Cooling

Traditional compressor cylinder designs require cooling water jackets

to promote uniform distribution of heat created by gas compression andfriction Some of the perceived advantages of water-cooled cylinders arereduced suction gas preheat, better cylinder lubrication, prolonged partslife, and reduced maintenance

Operating experience during the last 30 years has proven that pressor cylinders designed without cooling water jackts (non-cooled) cansuccessfully operate in most natural gas compession applications Some

com-of the perceived advantgaes com-of non-cooled cylinders are simplified der designs that reduce cost and improve efficiency, reduced initial sys-tem costs due to reductions in the cooling water system, improved valveaccessibility, and reduced weight

cylin-Many manufacturers, users, and compressor applications still requirethat compressor cylinders be supplied with liquid-cooled cylinders Fig-ure 11-21 includes schematics of several types of liquid coolant systems

In static systems, the cooling jackets are normally filled with a glycoland water mixture to provide for uniform heat distribution within thecylinder This system may be used where the AT of the gas is less than150°F and discharge gas temperature is less than 190°F

Thermal siphons use the density differences between the hot and thecold coolants to establish flow This system may be used where the AT ofthe gas is less than 150°F and discharge gas temperature is less than 210°F.Forced coolant systems using a mixture of glycol and water are themost common for natural gas compressors Normally, the compressorcylinder cooling system and compressor frame lube oil cooling system iscombined A single pump is used to circulate the coolant through thecylinders and the lube oil heat exchanger and then to an aerial coolerwhere the heat is dissipated

When forced coolant systems are used, care must be taken to providethe coolant at the proper temperature If the cylinder is too cool, liquidscould condense from the suction gas stream Thus, it is desirable to keepthe coolant temperature 10°F higher than that of the suction gas If thecylinder is too hot, gas throughput capacity is lost due to the gas heatingand expanding Therefore, it is desirable to limit the coolant temperature

to less than 30°F above that of the suction gas

Figure 11-21 Cylinder cooling systems (Reprinted with permission from API, Sid.

618, 3rd id., Feb 1986.)

Frame Lubrication System

The frame lubrication system circulates oil to the frame bearings, necting rod bearings, crosshead shoes, and can also supply oil to thepacking and cylinder lubrication system Splash lubrication systems are

con-the least expensive and are used in small air compressors Forced-feedsystems are used for almost all oilfield gas compression applications.Figure 11-22 shows a splash lubrication system where an oil ring ridesloosely and freely on the rotating shaft, dipping into the oil sump as itrotates The ring rotates because of its contact with the shaft, but at aslower speed The oil adheres to the ring until it reaches the top of thejournal when it flows onto the shaft.

Figure 11-22 Splash lubrication system (oil stinger).

lo a forced-feed lubrication system, a pump circulates lubricating oilthrough a cooler and filter to a distribution system that directs the oil to allthe bearings and crosshead shoes Figure 11 -23 is a schematic of a typicalsystem The details of any one system will vary greatly Major componentsand considerations of a forced feed lubrication system are as follows:

* Main oil pump

- Driven from crankshaft

- Should be sized to deliver 110% of the maximum anticipated flow rate

* Auxiliary pump

- Backup for the main oil pump

- Electric motor driven

Figure 11-23 Forced-feed lubrication system.

- Should start automatically when supply pressure falls below a

- Keeps oil temperature below 165°F

- Can use shell-and-tube exchanger with jacket cooling water or cooled exchanger

air Sized for 110% of the maximum anticipated duty

»Overhead day tank

- Sized to handle one month of oil consumption

- Should be equipped with a level indicator,

Packing/cylinder lubrication can be provided from a forced feed pressor lube oil system For very cold installations, immersion heaters andspecial lube oils must be considered If the lube oil temperature gets toocold, the oil becomes too viscous and does not flow and lubricate properly

com-Cylinder/Packing Lubrication System

The flow required to lubricate the packing and cylinders is quite small,and the pressure necessary to inject the lubricant at these locations isquite high Therefore, small plunger pump (force-feed lubricators) sys-

terns are used The force-feed lubricators are usually driven by the pressor crankshaft,

com-The two basic types of cylinder lubrication systems are the point system and the divider-block system The pump-to-point systemprovides each lubrication point with its own lubricator pump Thus, if thecompressor cylinders and packing require six lubrication points, the lubri-cator box would be supplied with six cam driven pumps The divider-block system uses one or more lubricator pumps to supply a divider block,which then distributes the flow to each of the lubrication points The twosystems are sometimes combined such that each stage of compression isprovided with its own pump and a divider block to distribute the flowbetween the cylinders and packing of that particular stage

pump-to-Oil is supplied to this system from the frame lube oil system or from

an overhead tank This oil comes in contact with and thus contaminatesthe gas being compressed Gas/oil compatibility should be checked

PIPE SIZING CONSIDERATIONS

Because of the reciprocating action of the piston, care must be cised to size the piping to minimize acoustical pulsations and mechanicalvibrations As a rule of thumb, suction and discharge lines should besized for a maximum actual velocity of 30 ft/sec (1,800 ft/min) to 42ft/sec (2,500 ft/min) Volume 1 contains the necessary formulas for deter-mining pressure drop and velocity in gas piping

exer-Analog or digital simulators can be used to establish the pulsation formance of any compressor piping system in detail API 618 Section3.9.2 provides guidelines for piping pulsation and vibration control based

per-on compressor discharge pressure and horsepower In practice, manyoperators do not "analog" compressors of 1,000 horsepower or lower, butrather rely on extrapolations from proven designs For larger horsepowersizes or where unusual conditions (e.g., unloading and loading cylinders)exist, an analog is recommended

For smaller, high-speed compressors the piping sizing rules of thumbdiscussed above, in conjunction with pulsation bottles sized from Figure

11 -24, should be sufficient for individual field compressors These rules

of thumb can also be used for preliminary sizing of piping and bottles inpreparation for an analog study

To minimize pipe vibrations it is necessary to design pipe runs so thatthe "acoustic length" of the pipe run does not create a standing wave that

Figure 11 -24 Pulsation bottle sizing chart (approximation) (Reprinted wilh

permission from GPSA Engineering Data Book, 10th Ed.)

amplifies the pressure pulsations in the system The acoustic length is thetotal overall length from end point to end point including all elbows,bends, and straight pipe runs Typical pipe runs with respect to acousticlength are considered to be:

* Pipe length from suction pipeline to suction scrubber

* Pipe length from scrubber to suction pulsation dampeners

» Pipe length from discharge pulsation dampeners to cooler

« Pipe length from cooler to scrubber

* Pipe length from discharge scrubber to pipeline

The end of a pipe ran can be classified as either "open" or "closed."Typically, closed ends are where the pipe size is dramatically reduced, as

at orifice plates and at short length flow nozzles A typical open end iswhere the pipe size is dramatically increased

Where the pipe run contains similar ends (closed-closed or open), prohibited pipe lengths are:

open-0.5X,X 1.5A,,2X

where X = acoustic wave length, ft

Where the pipe ran contains dissimilar ends (closed-open or open-closed),prohibited pipe lengths are:

0.25X 0.75X, 1.25X, 1.75X

The wave length may be calculated from:

where A, = acoustic wavelength, ft

k = ratio of specific heats, dimensionless

recip-to ensure that the natural frequency of all pipe spans is higher than the culated pulsation frequency The pulsation frequency is given by:

cal-where fp = cylinder pulsation frequency, cps

n = 1 for single-acting cylinders and n = 2 for double-acting

Rc = speed of compressor, rpm

Refer to Volume 1, Chapters 8 and 9 for the calculations of natural quency of pipe

fre-Foundation Design Considerations

Satisafactory compressor installations many times depend on how wellthe foundation or support structure was designed An inadequate founda-tion design can result in equipment damage due to excessive vibration Themoney saved by cutting corners on foundation design effort may be spentmany times in costs associated with high maintenance and lost production.Due to the basic design of the compressor, its rotating and reciprocatingmasses produce inertia forces and moments tha cannot be completely elim-inated and must be absorbed by the foundation The manufacturer has theability to rninimize the magnitude of these forces and moments by addingcounterweights to the crossheads but cannot totally eliminate them

In addition to the unbalanced forces and moments, the foundationmust absorb the moments produced by the gas torque This is the torquecreated by the gas pressure forces as the compressor goes through a revo-lution The compressor manufacturer must provide the magnitude of theresulting forces and moments and the gas torques,

Typically foundation design engineers have only used the compressorunbalanced forces and moments in their design calculations Recentexperience has found that the moments created by the gas torque canhave a significant impact on foundation design Detailed information andgood design practices for compressor support structures and foundations

may be found in Design of Structures and Foundations for Vibrating

Machines by Suresh Arya, Michael O'Neill, and George Pincus.

For complex offshore structures or where foundations may be critical,finite-element analysis computer programs with dynamic simulationcapability can be used to evaluate foundation natural frequency and theforced vibration response

Industry Standard Specifications

As previously discussed in this chapter and in Chapter 10, ing compressors are generally classified as either low-speed (integral)compressors or high-speed (separable) compressors API has provided astandard and specification for each type of compressor to help the userand the facility engineer provide reliable compressor installations.API Standard 618 "Reciprocating Compressors for Petroleum, Chemi-cal, and Gas Industry Services" covers moderate- to low-speed compres-sors in critical services Integral compressors and low-speed, long strokebalanced-opposed compressors with speeds from 200 to 600 rpm gener-ally fall into this type of construction The use of this standard with high-speed packaged separable compressors generally results in pages ofexceptions by the compressor packager,

reciprocat-API Specification I I P "Specification for Packaged ReciprocatingCompressors for Oil and Gas Production Services" covers packagedhigh-speed separable compressors with speeds from 600 to 1,200 rpm.The majority of reciprocating compressors sold in today's market fallinto this category

The user and facilitiy engineer must determine the critical nature of eachinstallation and determine the type of construction desired He or she mustconsider such things as intended service, compressor location, the conse-quences of downtime, and frequency of up-set or abnormal conditions

When specifying compressor packages to API IIP, it may be necessary

to specify certain sections of API 618 to ensure satisfactory installations,

An example of this would be the supply of multiple compressors to belocated in pipeline booster stations In this case, an analog or digital pul-sation and vibration study per API 618 Section 3.9 would be advisable toimprove reliability and to minimize system problems and potential dam-age caused by gas pulsations and interaction between the individual com-pressor packages

Fugitive Emissions Control

One of the growing environmental concerns for both new and existingreciprocating compressor installations is fugitive emissions Fugitive emis-sions are the leakage of volatile organic compounds (VOCs) into the atmos-phere The local environmental regulations should be checked at the begin-ning of the compression project to avoid delays and field modifications.The major source of fugitive emissions from a gas compressor cylin-der is the piston rod packing Other sources of fugitive emissions arearound the cylinder valve covers, unloader covers, unloader actuatorpacking, and clearance pocket gasket and actuator packing

Fugitive emissions can be reduced by supplying improved O-ring sealdesigns along with piston rod packing cases and actuator stem seal designs that utilize an inert buffer gas purge The purge gas and VOCscan then be collected and sent to either a flare or vapor recovery system.The compressor manufacturer must advise the maximum allowable back pressure on the compressor components A typical compressor cylinderinert buffer gas arrangement is shown in Figure 11-25

EXAMPLE PROBLEM

Grveo:

Late in the field life it is desirable to compress the 100 MMscfd for theexample field downstream of the separator from 800 psig at 100°F to1,000 psig An engine-driven separable compressor is available from sur-plus The engine is rated for 1,600 hp at 900 rpm Horsepower is propor-tional to speed The compressor frame has six 7-in bore by 6.0-in strokedouble-acting cylinders with a minimum clearance of 17.92%, a rod loadlimit of 25,000 Ib, and rod diameter of 1.75 in Assume k = 1.26, Zs =0.88, and Z = 0.85

Figure 11-25 Typical compressor cylinder inert buffer gas arrangement (Courtesy

ofDresser-Rana Company.)

Compute discharge temperature, volumetric efficiency, required ance, rod load, and required horsepower for the given conditions Alsocalculate the lowest suction pressure at which this unit can compress

clear-100 MMscfd

Spjujioji:

I Calculate the gas discharge temperature