COMPRESSOR HANDBOOK phần 5 potx

ob-VERY HIGH PRESSURE COMPRESSORS 7.357.3 PACKING AND CYLINDER CONSTRUCTION 7.3.1 Technical Solution for Cylinder Components Two solutions have been used for this special pressure vessel

VERY HIGH PRESSURE COMPRESSORS 7.31

FIGURE 7.26 Cross-section of multi-bored plate.

By the principle of superimposition of effects, the stress conditions generated

by external pressure, internal pressure and axial preload can be considered rately

sepa-The holes are assumed to be of the through type and have a diameter which isconstant, with the geometry of the valve section unchanged in any plane perpen-dicular to the valve body axis Without axial stress, the calculation approach brings

up the problem of an elastic body in a plane stress condition Consequently, theproblem consists of establishing the stress condition due to external and internalpressure in a plate geometrically schematized in Fig 7.26

The plate has three axes of symmetry, 60⬚ apart, which correspond to the ameters through the hole centers In this structure, the greatest stresses are on theinner edges of the holes, particularly on the points lying on the axes connectingtwo adjacent holes and on the axes of symmetry The most interesting points (Fig.7.26) are used to compare different calculations

di-The stress condition of this elastic body could be determined through an exactprocedure, i.e., analytically, by solving the elastic problem, or through approximateprocedures using:

• Existing formulas for comparable geometrical bodies

• The finite element method25,26

• Strain gages on the piece boundary

• Photoelastic models

Evaluation of the stress distribution on the valve body can also be made usingequations for thick-walled cylinders under external and internal pressure.28In thecase of cylinders with a central hole, the formulæ are to establish the stress distri-bution in any point of the radial thickness More complex are the equations forcylinders having eccentric holes,28giving circumferential stress in any point of theexternal and internal boundary A further evaluation of circumferential stress can

be made, (only for the points in Fig 7.26) by utilizing existing studies on stressconcentration factors in plates, whose notches are represented by holes.29 In thiscase, the plate is assumed to be compressed uniformly, as in a solid cylinder, withthe pressure acting on the outside Variations in the circumferential and radialstresses on the required points referring to the center of the valve body beingknown, the circumferential stresses, resulting from the presence of the holes, can

be determined Furthermore, holes of different diameters require further simplifyingassumptions

Strain Gage Method. A model of the plate was equipped with strain gages onexternal and internal surfaces to measure the trend of the circumferential stresses

on the boundaries, with pressure acting inside and outside.24The model was biggerthan the valve, to allow positioning of the strain gages on the internal surface andbecause of seal problems in the passage area of the connecting wires to the straingages The test pressure value was kept under 30 MPa (4350 psi) To minimizeeffects of systematic and accidental errors of the measuring instruments, the value

of the microstrains undergoing measurement was increased, by adopting a light

alloy model instead of steel, having a normal modulus of elasticity E ⫽ 72500MPa (10,512,500 psi) (about 1 / 3 that of the steel used for the valve) To eliminateuncertainties as to the elastic properties of this material, some specimens were takenfrom the piece the model was made from, to obtain the Young’s modulus andPoisson’s ratio for converting the microstrains into stresses

FEM Application. The calculations were made with pressure acting separately

on the external and internal peripheries It was assumed, according to the symmetry

of the system, that there was no rotation in the nodes determining the diameters ofthe half-plate, and that displacement would occur only in the direction parallel tothe circumference The procedure used for calculation involved finite elements withtriangular elements having three nodal points, with the general element having 6degrees of freedom and a linear shape function,24whose trend of stresses is shown

in the graphs in Fig 7.27 in relation to pressure

The trend of circumferential stress with pressure acting on the outside is similar

on hole edges In fact, its lowest values comply with those predicted in points A2.1,A3.1, A2.2 and A3.2 The lowest value (␴c / p e⫽ ⫺2.9) is assumed to be at point

VERY HIGH PRESSURE COMPRESSORS 7.33

FIGURE 7.27 Circumferential and radial stresses on plate edge and symmetry axes.

A.3.2, i.e., the internal boundary point of the hole having the smallest diameterand also related to the straight line joining the centers of two adjacent holes Thehighest value (␴c /p e⫽ ⫺1.9) is at point A4.2, i.e., at the smallest hole, toward theplate center and along a symmetry axis Furthermore, with internal pressure, thecurves of circumferential stresses on the inner edge of the holes show a similartrend, the highest value being point A3.2 The trend of circumferential and radialstresses is alike (Fig 7.27), both in the case of external pressure and that of pressure

in the holes

The sum of circumferential or radial stresses in the case of external pressureand unit internal pressure is constant and equal to ⫺1, i.e

(␴c /p e⫹ ␴c /p ) i ⫽ ⫺1The foregoing can be proved analytically for thick cylinders with centered or ec-centric holes, as formulæ exist for stresses along the thickness and at the boundaryrespectively In any case, if unit pressure exists inside and outside a cylinder, thestress condition is the same at any point of the thickness and the hoop and radialstresses are:

␴c /p⫽ ␴r /p⫽ ⫺1This is the result of two different loading conditions, with external and internalpressure; the above equation can thus be obtained by the superimposition effect.These statements apply to any type of stress (hoop, radial or direct, according tothe reference axes) involving multiconnected domains, regardless of boundary

of greatest concentration (points A3.1 and A3.2), the results practically coincide.The use of conventional equations led to results sufficiently in accordance withone another and generally lower than those obtained through the finite elementmethod This occurs especially at the point of greatest concentration when the thickcylinder formulæ are used At the same points, according to the theory of notches,the results practically coincide with those obtained through the finite elementmethod and experimental measurements.

Knowledge of the effective stress condition, proper choice of materials and taining a high degree of finite elements in the zones of greatest stress concentrationmakes it possible to arrive at the actual safety coefficient and thus ensure reliabilityagainst fatigue failure

ob-VERY HIGH PRESSURE COMPRESSORS 7.35

7.3 PACKING AND CYLINDER CONSTRUCTION

7.3.1 Technical Solution for Cylinder Components

Two solutions have been used for this special pressure vessel:

• A hard metal liner (sintered tungsten carbide with 9 percent cobalt binder),shrink-fit into a steel cylinder, on which a piston equipped with special pistonrings (Fig 7.16) was sliding

• A packing arrangement cup housing the seal rings, with a hard metal plunger(Fig 7.8)

Although the first solution was providing fairly good results, it was more affected

by plant conditions, low polymers and catalyst carrier as the lubrication was tained by injecting oil into the gas suction stream The packed plunger solution isless influenced by such factors, considering that the lubricant is injected directlyonto the sealing elements through holes and grooves on the packing cups

ob-The technological development of sintering WC (11 to 13 percent Co) plungers

of large size in one piece, the lower quantity of oil consumed, the excellent formance, and other process considerations21led to preferring packed plungers overliners on the compressors manufactured in the last 25 years

per-The selection of materials for components under pressure is very important.Mechanical properties must always be carefully analyzed and, when extremefatigue conditions exist, aircraft-quality electroslag or vacuum arc remelted steelsshould be utilized To obtain adequate fatigue strength of pressure components, it

is necessary to use autofrettage when operating pressures are very high

Sealing surfaces between cylinder components play an important role in ing good cylinder performance These are normally flat annular surfaces lapped to

achiev-a finish of 0.2 microns CLA* achiev-and pressed together by tie rods so thachiev-at their resultingload provides sufficient contact pressure to achieve seal Since little can be done

to modify the actions the cups are subjected to during operation, care should betaken to prevent the consequences of accidental surface defects by performing localprecompression treatments, such as cold rolling, shot peening, ionitriding etc.Special attention is required for the surface finishing of elements in direct contactwith the fluid subjected to pulsating pressure In order to eliminate superficial faults

as much as possible, which could cause fatigue failure, very high grade finishesare required Tungsten carbide plungers and liners have surfaces with 0.05 micronsCLA; with the additional advantage of reducing to a minimum the coefficient offriction between the moving parts It is difficult to obtain these low roughnessvalues on the gas passages in the cylinder heads and on the surfaces of steelcylinders in general, without the use of special machinery

* CLA ⫽ Center Line Average.

Cast iron and bronze or various combinations of these metals were used in thepast for piston rings Special bronzes are still utilized for packing sealing elements,although plastic elements can be used up to 250 MPa (36250 psi) when the processrequires low heat generation to avoid decomposition in the cylinder Relating tothe plunger material, in the past, nitrided steels were used for plungers in ammoniacompressors up to 100 Mpa (14500 psi) Usually, today piston rods are made ofsteel coated with tungsten carbide (11 to 13% Co) up to pressure of 60 MPa (8700psi).

In polyethylene plants, with more severe pressure conditions and more ous lubrication by white oils, liners or plungers are made of tungsten carbide withcobalt bonding When the cobalt content is increased, the hardness decreases, butthe toughness increases, and this quality is more important for plungers than forliners Today, the steel plunger coated with tungsten carbide can be used up to 140MPa (20300 psi), usually on the first stage of secondary compressors

precari-The sliding surface of plungers and liners should be machined to the maximumdegree of finish obtainable in order to reduce the friction coefficient to a minimum.Values of 0.025 to 0.05 microns (1 to 2 microin.) CLA of roughness are normallyachieved In case of WC coated plungers, the surface roughness is 0.1 micronsCLA (4 microin.) The surfaces of sealing elements do not require the same highquality, since they are softer and on the plunger they are polished during operation,but still need lapped mating surfaces and more accurate geometry to prevent leaksand failures

The life of the sealing elements is influenced by other factors The stroke andrevolutions per minute (RPM) determining the average piston speed influence thelife, since heat generation increases with speed The RPM are limited by compres-sor size and arrangement, dynamic loads on the foundation, operation of the cyl-inder valves, and pressure pulsation in the gas pipes

VERY HIGH PRESSURE COMPRESSORS 7.37

The stroke is selected to have a mean piston speed between 2.7 and 3.3 m / s(530 to 650 ft / min) A long stroke is generally desirable since this exposes a longerpart of the plunger out of the packing, for more effective cooling The life of sealingelements is influenced by the system supplying the oil to the cylinder, the amountand quality of oil, the shape of sealing elements, and the linearity of plungermovement A continuous film of oil must be applied to the sliding surfaces Thetype of oil is selected mainly for process reasons (i.e., the need to keep the productpure), and also its lubricating properties It is current practice to use white oil.The shape of the sealing elements used is similar to those used in conventionalmachines

The piston rings solution, with lubricating oil entrained by the gas, needs onlyfew rings for efficient sealing, but also to enable the one most distant from the gas-oil mixture to be lubricated Each combined piston ring is made of two rings inthe same groove, with a further ring mounted beneath The ring gaps are positionedout of alignment to give a complete seal effect On the top of the rings there is abronze insert, improving the anti-friction properties and the running-in

A packing arrangement is usually composed of 5 elements, for pressures up to

350 MPa (50750 psi) In the past, solutions with 3 to 8 sealing elements were alsoapplied The ring nearest the pressure is a breaker ring of special shape, suitablefor damping the high pressure fluctuations but not designed to provide effectiveseal, as this function is performed by the following ring couples, whose life isconsequently increased

The amount of oil applied must be controlled accurately, since trouble can arisefrom either excessive or insufficient lubrication If excessive oil is injected and theseal rings are providing perfect seal, the oil pressure can rise to a value above that

of normal conditions and the contact pressure between rings and plunger couldcause seizure Of great importance is the linearity of the piston movement, since

it ensures that the sealing elements will not be subjected to irregular operatingconditions and thus forced to assume an incorrect position in their housing, withconsequent overstressing and reduction in life

It is necessary to keep the temperature low by cooling the plunger with oilaround it, outside of the main packing This is important mainly to reduce the risk

of thermal cracks on the plunger surface

7.3.3 Autofrettage of Various Cylinder Components

General Aspects. The use of autofrettage, applied to tubular and vessel-reactors,has been extended to pumps18 and to machines operating particularly in tubular-reactor plants, as it is effective where the probability of fatigue failure is high Thistechnique allows components to be built using materials with lower mechanicalproperties

Autofrettage is performed on cylinder heads with combined axial valves, whenhigh pressures are involved, as gas pulsations are still present and fatigue mustalways be taken into consideration Cylinder chambers and packing cups are ex-cluded, as they can reach adequate prestress levels through shrink-fitting Packing

Autofrettage of injection quills and check valves operating on ethylene ary compressor second stages is also common practice when pressures are veryhigh Cylinder heads with radial valves are shrink-fit and are autofrettaged onlywhen differential pressure between suction and discharge is very high Autofrettagepressure is determined by operating conditions, geometry, presence of prestresses(due to shrink-fitting), and properties of the material Autofrettage pressures forhypercompressor cylinder parts range between 500 MPa and 1300 MPa (72500 to

second-188500 psi).30 Autofrettage of axial holes is performed after shrink-fitting of thecup on the finished piece, only upon completion of machining before final lapping

of the mating surfaces In this case, autofrettage pressure has been applied up to

• The cone solution (Fig 7.29), typical of high pressure tubing, has been applied

up to 1300 MPa (188500 psi)

• Annealed copper gaskets are used up to 1300 MPa (188500 psi) (Fig 7.30)

• Viton O-rings are employed for small-diameter seats, tapered (Fig 7.31) or flat(Fig 7.32), protected against extrusion by the metallic contact between the parts

VERY HIGH PRESSURE COMPRESSORS 7.39

FIGURE 7.30 Metal seal.

FIGURE 7.31 Plastic O-ring with conical seat.

Positive results were obtained on diameters up to 76 mm (3 in.) and up to 900MPa (130500 psi) for the latter solution

• Self-sealing arrangements (Fig 7.33) are used for wider diameters, in order tofollow the bore, subject to considerable strain under high pressures

These seals are made as follows:

• A seamless plastic O-ring with hardness between 75 to 90 Shore A, with goodsurface finish

• Hard plastic (a polyamide resin) and geometrically precise shoulder rings mensions have to be carefully checked, as plastics are subject to alteration withthe passage of time

Di-• Bronze antiextrusion rings with a 45⬚ angle

• Bronze rings to preload the seal assembly and to guide the inner core of thedevice

In autofrettage of radial valve cylinder heads, similar seals are used and internalmandrels are applied to reduce fluid volume Axial valve cylinder heads are auto-frettaged (Fig 7.34) with special seals (Fig 7.33) to achieve seal on the large innerdiameter which can be accomplished by providing a smooth surface finish and

7.40 CHAPTER SEVEN

FIGURE 7.32 Plastic O-ring with flat seat.

FIGURE 7.33 Special seal with O-ring.

using great care in assembling the rig to avoid local damage in the seal zone Aninternal bar reduces fluid volume The seals are preloaded, the assembly is balancedand no additional support is required for the inner core Lateral (suction and dis-charge) holes are plugged by flanges using combined metallic and O-Ring seals(Fig 7.32) Autofrettaged packing cup axial holes (Fig 7.35) use metal seals (Fig.7.30) The test rig for the oil distribution cups uses axially-directed seal (Fig 7.30)and radial seal (Fig 7.31) Autofrettage of injection quills utilizes cone seals (Figs.7.29 and 7.31)

Autofrettage Procedure. In equipments operating at very high hydrostatic sures, the fluid must be able to transmit pressure without undergoing freezing ef-fects, related to fluid properties, operating temperatures and tubing size Pressuremay increase at the pump and, due to solidification problems within the tubing,may be much lower inside the piece to be autofrettaged

pres-Brake oils have been used up to 500 MPa (72500 psi) with some drawbacks(i.e., corrosion on pump seal rings caused poor performance) Prexol 201 over-comes solidification problems and gives adequate intensifier plunger seal life, up

VERY HIGH PRESSURE COMPRESSORS 7.41

FIGURE 7.34 Apparatus for

autofret-tage of axial valve heads.

FIGURE 7.35 Apparatus for autofrettage of packing

cups.

trans-of the cylinder component.

Strain gages on the outer surface of the piece are used when an autofrettageprocedure must be defined for the first time, or in case of complex shapes, andmay detect internal pressure or deviations in mechanical properties For safetyreasons, dimensional checks are performed after autofrettage

The inner diameter of axial valve cylinder heads should be checked, to assessthe amount of metal to be removed, which should be as small as possible in order

to preserve the benefits of prestressing At plastic strain conditions, duration oftests appreciably affect the final results Pressure should rise slowly to allow strain

to take place completely during each loading condition In short tests, yield pointand ultimate tensile stress are increased while strain decreases In the case of steel,

a pressure rise of 10 MPa (1450 psi) per second is on the safe side Generally, thetest requires a pressure rise of 5 minutes minimum Pressure increase is related tothe volume of the fluid in the whole system and its components (tubing and theintensifier) The autofrettage pressure is maintained for 15 minutes (5 minimum)and a slow pressure decrease takes place in about 5 minutes Slow return to finalconditions eliminates errors in dimensional measurements, allows time to checkthe autofrettage effect, and allows the special seals to return to their original po-sitions in their housings after having undergone severe strain, thus reducing dis-assembly problems

Very high pressure systems have potential hazards, although risks are not asgreat as when gases are handled, due to the great energy involved (the fluid pos-sesses compressibility and can be trapped inside the system) If gaskets in thehydraulic system fail, the jettisoned particles could cause injury to people or dam-age objects Fluid leak at high speeds, reduced by the small volumes involved, isanother risk To prevent air from being trapped in the hydraulic circuit during testrig assembly, a vent valve is temporarily opened at the highest point of the circuitand oil is allowed to drip out, prior to tightening To reduce risks from storedenergy, the volume of the system is reduced: the piping is made as short as possibleand suitable inner cores are used in large components like cylinder heads Thecompact system is positioned in a safe area (bunker with fencing around the equip-ment to protect the surroundings) Steel shield between assembly and pump andmetal sheets around the pressure tubes are added protection The operator’s workstation is separate Before disassembling any part of the test rig, the pressure isrelieved from the circuit

Some authors,31,32 advise heat-treating the material at about 250⬚C for an hour

to allow component dimensions (i.e., eliminating flexural stresses without affectingresidual body stresses)33and the material elasticity to be restored (Others recom-mend higher temperatures.) At the same or higher temperatures, decarburizingproblems might arise on the surfaces This is not common practice with polyeth-

VERY HIGH PRESSURE COMPRESSORS 7.43

FIGURE 7.36 Packing assembly.

ylene compressors, as components have proven successful field operation In anycase, this heat treatment cannot be performed when the tempering temperature ofthe material is lower than the heat treatment temperature

Axial valve cylinder heads, requiring accurate inner bore dimensions, must bemachined after autofrettage Remachining is also performed in the seat area quills(oil distribution cup side) and thus the modified prestress level area is quite limited.Appropriate allowances must be considered, and material removal must take intoaccount the reduction in the prestress level

It is generally advisable to perform autofrettage on finished parts The nation of autofrettage and shrink-fitting, especially when high ultimate tensilestrength materials are used, is complex Autofrettage before shrink-fitting is nor-mally carried out on radial valve cylinder heads, allowing use of lower autofrettagepressure, with advantages Lube oil holes of packing cups are autofrettaged at apressure of 1100 MPa (159500 psi) Autofrettage contributes to increasing theavailability of secondary ethylene compressors which operate in plants with tubularreactors or in general when pressures exceed 200 MPa (29400 psi)

combi-7.3.4 Typical Behaviour of Packings

Packings today consist typically of one (or two) split breaker rings and five radialtangential sealing rings (Fig 7.36) The rings are made of special bronze alloys,usually with high lead content, uniformly distributed, so as to guarantee sufficientstrength, low friction coefficient and high thermal conductivity, for a rapid dissi-

7.44 CHAPTER SEVEN

pation of the friction heat through the packing cups The hardness of the ringsvaries from 55 to 80 Brinell (measured with a 10 mm ball and 500 kg load).The plunger on which the sealing elements slide is made of solid tungsten car-bide, with surface finish of 0.05 microns [2 microin.] CLA The synthetic lube oil

of the cylinders has lower lubricating properties than oils used for normal services,since for ethylene polymerization, pollution of the final product must be reduced

to a minimum

Packing performance is greatly influenced by the above parameters and by theefficiency of the breaker rings, whose action is very important, as can be seen byanalysing the operating conditions of a packing The pressure inside the cylindercan be considered as consisting of a constant portion (suction pressure) and afluctuating portion (the difference between discharge and suction) The static pres-sure distribution tends to overload the last ring (frame side), which has to handlealmost the whole load.23This is similar to packings, operating at constant pressure,for example on ammonia synthesis compressors The variable pressure increasesdue to polytropic compression, and then decreases due to the expansion of the gasremaining in the clearance volume, and assumes constant values during dischargeand suction effect

Breaker rings oppose a rapid pressure increase in the cylinder, limiting gas age and reducing the propagation of the pressure wave towards the seal rings Theirmost important function, however, is to delay the ‘‘backflow’’ from the packingrings towards the cylinder chamber, when the plunger begins its back stroke Ifthis action is inadequate, the pressure upstream of the first sealing element willsuddenly drop to the suction value, due to the steep slope of the expansion curve.The resultant of the forces acting on the first sealing ring is suddenly inverted,causing rapid expansion of gas under the radial and especially under the tangentialring, which exerts a stronger sealing action, with the following problems:

leak-• Breakage of the dowel pin between radial and tangential ring

• Breakage of the lips of the tangential cut rings

• Damage to the garter springs of the sealing element

When, after a certain period of operation, the first sealing pair no longer forms its function, the problems occur in the second pair and the process of pro-gressive damage continues through the various rings of the packing To analyzethe operating conditions, behaviour and performance of packings, measurementswere taken at the lube oil injection quills and in the compression chamber (Fig.7.37) of a first and second stage cylinder on a compressor having a capacity of53,000 kg / hr (1945 lb / min), operating in a plant with a vessel reactor Packingshad a three piece pressure breaker ring, with small circumferential clearance andfive grooves of radial tangential seals (with axial clearance of about 0.15 mm [.006in.]) The distribution of the pressures along the packing in relation to the crankangle (Fig 7.37) and during the suction and discharge strokes (Fig 7.38) is quitesimilar on first and second stage.23

per-In general, the first three sealing elements are affected by the pressure fluctuation

of the cylinder, while the last two are subjected to an almost steady pressure (Fig

VERY HIGH PRESSURE COMPRESSORS 7.45

FIGURE 7.37 Operating pressures on a 1st and on

a 2nd stage cylinder.

FIGURE 7.38 Pressure distribution

on the 2nd stage packing during

dis-charge and suction stroke.

is evidenced in both stages during compression and expansion The breaker ring,

in fact, withstands about 80% of the pressure fluctuation, (100 MPa [14500 psi] inthe second and 70 MPa (10150 psi) in the first stage) The efficiency is higher inthe first stage, due to the greater variability of the specific volume of the gas (5%

in the second and 16% in the first stage) This may be partly explained by thedifference between the polytropic coefficient in first and second stage It should berecalled that when the physical conditions of a gas are close to those of a liquid,the task of the breaker ring is more difficult and its effect is lower

The fluctuating part of the pressure affects the first three seal rings, with thesecond and third withstanding a differential pressure of 50% and 30% as compared

to the first sealing couple (Figs 7.37 and 7.38) The steady part of the pressure ismainly supported by the last two sealing elements

Some packings were dismantled and analyzed after 10,000 to 20,000 hours ofoperation The wear of each radial and tangential element was compared (Fig 7.39)

VERY HIGH PRESSURE COMPRESSORS 7.47

and there was a similar wear pattern curve for radial and tangential rings For first

or second stage, the trend for higher wear is on the first and last elements Thewear rate is higher in tangential rings as compared with radial ones On first stage,the first pair did not wear completely, as the dowel broke, due to ‘‘backflow’’ andthen the pressure loaded the second pair, causing accentuated wear Wear on thetangential ring higher than the amount allowed by the butt gap is frequently ob-served due to non-uniform wear on the rings, resulting from the high pressures andforces acting on them On the second stage, the ‘‘backflow’’ caused breakage ofthe springs (of the coil type) of the first pair and later breakage of the dowel pin

of the second pairs The work of withstanding the variable pressure was then carriedout by the third sealing element

The maximum wear on the frame side elements of both stages is due to theconstant pressure to which they are subjected, considering that lubricating condi-tions are not optimal Wear on the radial ring of the last pair of the first stagepacking is an exception, encountered in other compressors, which can be explained

as follows: The radial rings, subject to steady pressure, tend to remain in theirposition without effecting an appreciable sealing action towards the plunger, butsimply creating a barrier to the pressure at the cuts of the tangential rings In thezone subject to variable pressure, the first radial rings are forced to exert a sealingaction on the gas that tends to re-enter the cylinder during the suction phase Thesealing effect is not complete, since the radial cuts allow the gas passage

A general wear pattern can be derived connected with the pressure distributionalong the packing (Fig 7.39) The steady portion of the pressure causes a type ofwear with maximum values reached on the frame side sealing ring The fluctuatingportion of the pressure causes wear with an opposite trend, with the highest values

on the first sealing pair The resultant wear will be a curve with its maximumvalues at the extremities of the packing Generally, the theoretical maximum value

is either towards the first ring (pressure side) or towards the last (frame side)depending on the predominance of the fluctuating or the steady portion The prac-tical wear pattern is different as the ‘‘backflow’’ can make some sealing elementsinefficient The performance of the sealing elements is strongly influenced by op-erating conditions, lubrication and alignment The normal plunger runout is within0.075 mm (.003 in.), as easily measurable by proximity probes, with alarm 0.15

mm (.006 in.) and trip 0.2 mm (.008 in.) Long life of packing rings has beenreported up to 65,000 operating hours, with 180 MPa (26100 psi) final pressure

7.4 BIBLIOGRAPHY

1 Crossland, B., K E Bett, and Sir Hugh Ford: Review of Some of the Major Engineering Developments in the High-Pressure Proc Polyethyene Process, 1933–1983, Institute of

Mechanical Engineering, 1986, Vol 200, Ne A4.

2 Andrenelli, A., ‘‘Reciprocating Compressors for Polyethylene Production at Pressures

Higher Than 3000 Atmospheres,’’ Quaderni Pignone 13.

4 Traversari, A., and F Bernardini, ‘‘Aspects of Research on Secondary Compressors for

Low Density Polyethylene Plants,’’ Quaderni Pignone 25, June 1978, pp 123–124.

5 Vinciguerra, C., U.S Patent 3, 581.583 to Nuovo Pignone S.p.A., January 15, 1969.

6 Andrenelli, A., ‘‘Special Features in Reciprocating Compressors for Polyethylene

Pro-duction,’’ Proceedings of the Industrial Reciprocating and Rotary Compressors: Design and Operational Problems, Institution of Mechanical Engineers, Vol 184, Part 3R, Oc-

tober 13–16, 1970, pp 106–113.

7 Traversari, A., P Beni P., Approaches to Design of a Safe Secondary Compressor for High Pressure Polyethylene Plants, High Pressure Symposium: Safety in High Pressure

Polyethylene Plants, Tulsa, Oklahoma, March 12–13, 1974.

8 Giacomelli, E., and M Agostini, Safety, Operation and Maintenance of LDPE Secondary Compressors, ASME PUP Division Conference, New Orleans, Louisiana, 1994.

9 Manning, W R D., ‘‘Ultra-high-pressure Vessel Design, Pt 1,’’ Chem Proc Eng.,

March, 1967.

10 Morrison, J L M., B Crossland, and J S C Parry, ‘‘Fatigue Strength of Cylinders

with Cross Bores,’’ J Mech Eng Sci 1959 1 (N 3).

11 Parry, J S C., ‘‘Fatigue of Thick Cylinders: Further Practical Information,’’ Proc Inst Mech Engrs 1965–66 180 (Pt 1), 387.

12 Chaaban, A., K Leung, and D J Burns, ‘‘Residual Stress in Autofrettaged Thick-Walled

High Pressure Vessels,’’ PVP, Vol 110, 1986, pp 56–60.

13 Kendall, D P., ‘‘The Influence of the Bauschinger Effect on Re-Yielding of

Autofret-taged Thick-Walled Cylinders,’’ ASME Special Publication, P.V.P., Vol 125, July, 1987,

pp 17–21.

14 Yang, S., E Badr, J R Sorem, Jr., and S M Tipton, ‘‘Advantages of Sequential Cross

Bore Autofrettage of Triplex Pump Fluid End Cross Bores,’’ P.V.P., Vol 263, High

Pressure—Codes, Analysis and Applications, ASME, 1993.

15 Manning, W R D., Design of Cylinders by Autofrettage, Engineering (April 28, May 5

and May 19, 1950).

16 Chaaban, A., and N Barake´, ‘‘Elasto-Plastic Analysis of High Pressure Vessels with

Radial Cross Bores,’’ P.V.P., Vol 263, High Pressure—Codes, Analysis and

20 Rees, D W A., ‘‘Autofrettage Theory and Fatigue Life of Open-Ended Cylinders,’’

Journal of Strain Analysis, Vol 25, pp 109–121, 1990.

21 Parry, J S C., ‘‘Fatigue of Thick Cylinders: Further Practical Information,’’ Proc Inst Mech Eng., 1965–66, 180 (Part I).

VERY HIGH PRESSURE COMPRESSORS 7.49

22 Kendall, D P., and E H Perez., ‘‘Comparison of Stress Intensity Factor Solutions for

Thick-Walled Pressure Vessels,’’ P.V.P.—Vol 263, High Pressure—Codes, Analysis and

Applications, ASME, 1993.

23 Traversari, A., and E Giacomelli, ‘‘Some Investigation on the Behaviour of High sure Packing Used in Secondary Compressors for Low Density Polyethylene Produc-

Pres-tion,’’ Proceedings of the 2nd Int Conf on H.P Engineering, University of Sussex,

Brighton, England, July 8–10, 1975, pp 57–58.

24 Giacomelli, E., ‘‘Finite Element Method on Polyethylene Compressor Valves Design,’’ Quaderni Pignone 26, January 1979, pp 19–25.

25 Zienkiewicz, O C., ‘‘Axi-Symmetric Stress Analysis,’’ The Finite Element Method in Engineering Science, (London, Eng.: McGraw Hill, 1971), pp 73–89.

26 Tottenham H., and C Brebbia, Finite Element Techniques in Structural Mechanics

(Southampton, Eng.: Millbrook).

27 Muschelisvili, Some Basic Problems of the Mathematical Theory of Elasticity, Moscow,

30 Giacomelli E., P Pinzauti, and S Corsi, Autofrettage of Hypercompressor Components

up to 1.3 GPa: Some Practical Aspects, ASME PUP Division Conference, Orlando,

Florida, 1982.

31 Vetter, C., and H Fritsch, ‘‘Zur Berechnung und Gestaltung von Bauteilen mit

Bean-spruchung durch schwellende Innendruck,’’ Chemie Ingr Tech., 1958, 40 (n 24).

32 Morrison, J L M., B Crossland, and J S C Parry, ‘‘Strength of Thick Cylinders

Subjected to Repeated Internal Pressure,’’ Proc Inst Mech Eng., 1960, 174 (no 2).

33 Giacomelli, E., and P F Napolitani, ‘‘Ricerca Sperimentale sul Comportamento degli Accoppiamenti Forzati Albero-mozzo,’’ Thesis, Dept Mech Eng., University of Pisa,

Italy, 1969.

34 Cosimi, L., ‘‘Il Compressore a Pistone a Secco con Tenuta a Labirinti,’’ Il calore, 1961—

N 3.

35 Faupel, J H., and F E Fisher, Engineering Design, (New York, N.Y., Chichester,

Bris-bane, Toronto: John Wiley and Sons, 1981).

36 Whiteley, K S., Ullmann’s Encyclopedia of Industrial Chemistry, Vol A21, Section

1.5.1, Polyofins, 1992.

to the importance of natural gas as an alternative to crude oil based fuels.

Natural gas is a mixture of gases in which the primary constituent is methane,typically at 85.0 to 95.0 mole percent

As a transportation fuel, stored natural gas must be compressed for an increase

in energy density The compressor is used to boost the pressure of natural gas and

is the primary equipment of the compressed natural gas (CNG) refueling station

8.2.1 Suction And Discharge Pressures

Discharge pressures of 3600 to 5000 psig preclude the use of the multi-stage ciprocating piston compressor Suction pressures are site specific and dependent onthe operating pressures of the local gas utility distribution pipeline Suction pres-sures can range from inches water column to 1000 psig Most often pressure reg-ulation and metering is supplied by the gas utility providing stable suction pressures

re-to the compressor To minimize energy consumption, CNG compressor

is due to the lower specific heat ratio property (k⫽ C p /C v) of natural gas relative

to air For this reason, natural gas compressors of similar technology can operate

at higher pressure ratios than air compressors The gas discharge temperature of acompressor stage is one of the limiting factors determining maximum stage pressureratio The maximum discharge temperatures allowable are a function of the ac-ceptable operating temperatures of the sealing materials used, including pistonrings, rod rings, o-rings and gaskets Avoiding high discharge temperatures alsodecreases compression horsepower To maintain satisfactory discharge temperatures

a suitable number of compression stages must be selected Table 8.1 provides aguide to the number of compressor stages required for a given suction pressure.There is some overlap of suction pressure ranges Some compressors, such as thosewith oil lubricated and cast iron piston rings can operate at higher pressure ratiosthan compressors using nonlubricated and special material piston rings such as thefilled TeflonsTM

CNG COMPRESSORS 8.3

FIGURE 8.2 Compressor brake horsepower vs suction pressure.

TABLE 8.1 Compressor Stages vs Suction Pressure

Suction pressure (psig) No of stages Discharge pressure (psig)

green-in the pipgreen-ing system or compressor caused by static seal failures Controlled leakage

is expected leakage from compressor rod packings and seals As industrial sions standards tighten, consideration must be given to CNG compressor manufac-turer’s gas leakage rate data

Pressurized crankcases are most often used on trunk piston type compressors.Trunk pistons have a linear guide and piston as one integral part There is no rodsealing between the piston and linear guide Without a linear traveling piston rodand seals (see Atmospheric Crankcase, below), piston ring leakage flows into thecrankcase To hold leakage gas at suction pressure, the crankcase must be designed

as a pressure vessel with heavy rounded walls and internal or external structuralribs (see Fig 8.3) Some pressurized crankcase compressors use a cantilevered shaft

to eliminate one shaft seal Other components including oil lubrication systems,static seals on inspection plates and cover seals must withstand the elevated pres-sures

Atmospheric Crankcase. Compressors with atmospheric crankcases commonlyuse double acting cylinders and crossheads Crossheads allow the use of a pistonrod which moves linearly and compresses both to the head and crank end (see Fig.8.4) The piston rod is readily sealed using a series of rod packings Rod packingsare assembled in a packing case with gas leakage vented and piped for discharge

CNG COMPRESSORS 8.5

FIGURE 8.4 Atmospheric crankcase compressor (courtesy of Gemini Engine Company).

to atmosphere New rod seal leakage can be very low and commonly less than0.1% of total cylinder mass flow rates Most compressors of this type vent the gas

at source rather than allowing the gas to leak into the crankcase The crankcaseoperates at atmospheric pressure, eliminating the need for special shaft seals, gas-kets, and elevated pressure lubrication systems

The atmospheric crankcase is most suitable for large compressors where designfor pressure containment is difficult Atmospheric crankcase type compressors us-ing rod sealing also allow compressors to be designed for high gas suction pressuresbeyond what is practical for pressurized crankcase type compressors In addition,maintenance procedures are less onerous, allowing crankcase inspections withoutdepressurization

Blow Down Gas Recovery. Similar to air compressors, the natural gas compressormust be depressurized for start up This necessitates that on shutdown, gas en-trapped in the compressor and piping system must be vented Unlike an air com-pressor which can be vented to atmosphere, the natural gas compressor must beprovided with a blow down gas receiver tank This tank must be adequately sized

8.2.3 Lubrication

Compressor lubrication has become an issue of debate within the CNG industry.Lubricated compressors require lubrication of piston rings, rod packings and valves.Nonlubricated or oil free compressors use special materials for these components,eliminating the need for additional oil injection Proponents of nonlubricated com-pressors claim that they achieve the highest discharge gas quality Proponents oflubricated compressors maintain that with well engineered lubrication and filtrationsystems, similar discharge gas quality is attainable A lubrication oil carry overlimit maximum of 0.5 lb / mmscf at compressor discharge has become a commonindustry standard This standard can be met using nonlubricated compressors orlubricated compressors with filtration

In deciding lubricated versus nonlubricated, other factors to consider are outlined

in Table 8.2

8.2.4 Piston Ring and Seal Performance

The extreme gas pressures exerted in the final stages of a natural gas compressorpresent some unique design problems Piston ring wear rates increase dramaticallywith increasing stage pressures High pressure differentials across piston rings con-tribute to ring extrusion between the piston and cylinder clearances Lowering clear-ances reduces ring extrusion, but increases the possibility of piston contact withthe cylinder wall as the piston wear bands deteriorate Extreme pressures alsocontribute to high operating PV (the product of surface pressure and velocity)values of piston rings The result can be high piston and cylinder wear rates High

PV also generates high ring surface contact temperatures These temperatures can

be higher than measured gas discharge temperatures resulting in piston ring materialcreep and extrusion Another less understood factor in piston ring wear is theapparent loss of oil viscosity at high operating pressures

High pressure static sealing using compliant and porous o-ring materials canresult in seal failure upon rapid decompression O-ring materials including Buna-

N and VitonTM are porous and allow high pressure natural gas to permeate thematerial If the o-ring is operating at high pressure for some extended time periodand then the compressor shuts down and rapidly decompresses, the gas entrained

in the o-ring will rapidly expand Failure of the seal is caused by rapid expansion

of entrained gas causing bubbles and lacerations in the o-ring material as the gas

CNG COMPRESSORS 8.7 TABLE 8.2 Lubricated vs Nonlubricated Compressors

Advantages —increased piston ring life

—allows use of metallic rings

—air cooling or noncooling systems sufficient

—higher pressure ratios and discharge temperatures allowable

—fewer stages necessary in some cases

—higher operating speed

—reduced capital cost

—longer overhaul intervals

—low to nil oil contamination of discharge gas

—reduced lubrication requirements

—less filtration needed

—reduced routine maintenance

—reduced noise levels with liquid cooled units

Disadvantages —oil contamination of discharge

—increased vehicle emmissions

—higher compressor oil consumption

—increased maintenance requirements on lubrication systems

—increased noise levels with air cooled compressors

—increased cooling requirements

—lower maximum discharge temperatures

—reduced piston ring and rod packing life

—lower pressure ratios and discharge temperatures required

—more stages may be necessary

—increased capital cost

—lower operational speeds

—reduced valve life

—shorter overhaul intervals

escapes High operating temperatures, and pressures along with the presence ofcompressor lubricating oils, seems to exacerbate the problem The higher durometerseal materials provide some improvement likely due to their lower porosity Ifsuitable, the use metallic gaskets should be considered

8.2.5 Other Compressor Design Considerations

Adequate gas aftercooling is key to the satisfactory performance of high demandCNG refueling stations The cooler the natural gas is when it enters the storagevessels or refueling vehicle, the more dense the fuel, enhancing storage capacity.Final gas discharge temperatures should be a maximum of 20 to 30⬚F above am-bient air temperatures Further reductions in temperature will greatly increase af-tercooler cost

8.8 CHAPTER EIGHT

Compressors used in time filling applications (see Section 8.4.2) operate with alarge variation in discharge pressure Some compressors require operation abovesome discharge pressure minimum This minimum pressure is critical to the properoperation of piston rings Without sufficient back pressure the rings will not seatadequately against the cylinder walls and in some lubricated compressors, oil con-sumption will increase as oil is drawn into the compressor cylinder To address theproblem, a back pressure regulator can be installed at compressor discharge andset to the manufacturer’s minimum required pressure

Compressor package noise levels are often critical to the acceptance of an stallation by approval authorities Noise level requirements are often site-specificand set by city and municipal bylaws Reduced noise levels also enhance fuelmarketing efforts A commonly specified noise level is 75 dba measured at 10 feetfrom the perimeter of the compressor skid Most compressor packagers can meetthis noise level with an enclosed and sound attenuated compressor skid package

in-8.2.6 Compressor Electrical Systems

Natural gas, being flammable, requires that all electrical equipment and wiringwithin a code specified distance from natural gas compressors and gas containingequipment be explosion-proof The explosion-proof classification most used in theCNG industry is class 1, division 1 or 2, group D in accordance with National FireProtection Association (NFPA) standard 70, the National Electrical Code Codessuch as NFPA 52 define the boundaries of explosion-proof areas

Explosion-proof electrical enclosures and junction boxes are designed to stand internal explosions without flame propagation out of the enclosure They aremetallic and of heavy wall construction making them expensive and cumbersome

with-to access The compressor skid of Fig 8.5 shows an explosion-proof disconnectpanel standing to the right of the instrumentation panel Some alternatives to theuse of explosion-proof equipment include:

• Conventional contact closure instrumentation with intrinsically safe circuitry

• Standard electrical enclosures with air purge systems and failure shutdown locks

inter-• Impenetrable gas tight electrical rooms within the hazardous area with gas tection

de-• Locating electrical equipment remote and outside the hazardous area

8.3 CNG STATION EQUIPMENT

In addition to the compressor, other items of equipment are required to completethe CNG refueling system The descriptions of equipment that follow are repre-sentative only The CNG industry has few system design standards Most instal-lations are uniquely designed to meet a performance specification

The Three Bank Priority Panel uses a single multi-port valve or multiple valves

to control the flow of discharge gas from the compressor to a series of storagepressure vessels The pressure vessels are divided into three banks with each bankhaving one or more pressure vessels joined by a manifold The three storage banksare designated the high, medium and low banks, with the high and low bankshaving the highest and lowest filling priority respectively The compressor operates

to maintain maximum pressure in the storage banks through the priority panel byfilling the high bank, medium bank and low bank in turn

If all storage vessel banks are depleted in pressure, most priority panels shift allcompressor discharge flow to feed the refueling vehicles directly At this stage,unless the compressor has a very large flow capacity, refueling flow rates are slowand equal to the rate of compressor flow

at a low or ‘‘empty’’ pressure (e.g 200 to 500 psig) As the difference in headpressure between the low storage bank and NGV fuel tank reduces, the flow ratedecreases At a minimum flow rate, as measured by the dispenser meter, sequencingvalve S2 opens The gas flow rate increases with the increase in head pressure.Again, as the flow rate drops with the head pressure, a minimum flow rate signalsthe opening of sequencing valve S3 When the NGV fuel tank reaches full pressure,the sequencing valve closes, terminating the fill In this way, multiple vehicles can

be filled consecutively or simultaneously, depending on the number of dispensersprovided

Refueling flow rates are dependent on how restrictive the piping system is tween the pressure vessels and the NGV fuel tank Commonly the most restrictivepiping is on board the NGV Since there is no industry standard for on board NGVpiping systems, predicting refueling performance is difficult

be-CNG COMPRESSORS 8.11

FIGURE 8.7 CNG dispenser (courtesy of Fueling Technologies

Inc.).

Typical refueling flow rates for passenger cars and light trucks average 300 to

500 scfm with maximum flow rates reaching as high as 900 scfm Refueling flowrates for large commercial, industrial and public transportation vehicles can average

1500 to 2500 scfm with maximum flow rates reaching 5000 scfm

8.3.3 Emergency Shutdown Systems

The CNG station Emergency Shutdown (ESD) System uses fail safe closed valves,strategically located to shut off gas flow in an upset scenario The ESD valves arelocated in piping systems near pressure vessels, compressor suction and dischargelines, and inside CNG dispensers Most ESD valves use either pneumatic, spring

8.12 CHAPTER EIGHT

return operators or spring return electric solenoid actuators The means of initiatingvalve closure include:

• Loss of electric power

• Manually depressing ESD push buttons located in various station locations

• Seismic activity detection devices

• Proximity or limit switch devices sensing motion of impacted equipment

• Vibration sensing equipment

• Gas, fire and heat detectors

• Loss of pneumatic control pressures from plastic air line ruptures caused by fire

8.3.4 Pressure Vessels Storage

Fast fill type CNG stations store compressed gas in one or more pressure vessels.The size, design pressure and configuration of the pressure vessels determine howmuch stored gas can be used for refueling before replenishment by the compressor

is required The CNG industry commonly uses pressure vessels configured in threebanks Each bank has at least one pressure vessel Higher capacity installations usemultiple pressure vessels joined by a manifold and perform as one larger volumebank The CNG dispenser uses a set of valves which open in sequence duringrefueling Each sequencing valve allows the flow of gas to the vehicle from onebank By opening the valves in sequence, a greater percentage of gas can be with-drawn from the pressure vessel storage than if all the pressure vessels were joined

by a manifold as a single bank

Pressure vessel storage full pressures range from 3600 to 5000 psig Industrystandard NGV fill pressures are 2400, 3000, and 3600 psig at a standard gas tem-perature of 70⬚F To complete NGV refueling at least one bank of a pressure vesselstorage must have a higher pressure than the NGV fuel tank final fill pressure

Storage Utilization. Only a fraction of the total stored weight of gas from fullstorage pressure vessels can be dispensed At some point, there will be insufficientpressure in the storage vessels and the NGV fuel tank fill cannot be completed Atthis point, the amount of gas remaining in the pressure vessels is substantial andmay be 50 to 75% of the full amount Storage utilization is defined as

Total gas weight dispensed / Storage full gas weight⫻100%

Storage utilization increases with the number of storage banks used The increase

in storage utilization becomes marginally less with each additional bank The dustry has found the use of three banks to be the best compromise in terms ofcost, complexity and performance

in-In addition, storage utilization is highly dependent on storage full pressure andfinal vehicle fill pressure (see Table 8.3) Considerable gains in storage utilizationoccur when maximum storage fill pressures are increased This allows the use of

CNG COMPRESSORS 8.13 TABLE 8.3 Storage Utilization (SU)*

no compressor replenishment during fill; three equally sized storage bank volumes for the three bank

SU values; gas critical temperature of 366 ⬚F.; gas critical pressure of 669.84 psia; 70⬚F gas temperature;

53 psig NGV fuel tank start of fill pressure.

* SU values generated with the assistance of Ralph O Dowling, P.E., Christie Park Industries, using the Institute of Gas Technology ‘‘Cascade’’ computer program.

smaller storage vessel volumes and vessel cost savings The benefits of higherstorage fill pressures are offset, however, by an increase in compression energycosts and higher compression equipment and maintenance costs

Other variables affecting storage utilization include

• NGV fuel tank empty pressures

• Dispenser sequencing valve switching set points

• Initial fuel tank gas temperatures

• Heat transfer rates from storage tanks, piping and NGV fuel tanks

• Successive vehicle filling versus periodic filling

• Relative volumes of each storage bank

• Gas critical pressure and critical temperature

8.3.5 Gas Dehydration

New standards for compressor discharge gas quality are making the use of gasdehydration equipment mandatory in many CNG installations The Society Of Au-tomotive Engineers standard SAE J-1616 specifies maximum allowable compressordischarge gas water content This standard ensures that NGV fuel tanks will besafeguarded from corrosion and will operate safely for the life of the vehicle.Gas dehydration equipment is most commonly installed on the suction or finaldischarge line of the compressor or compressor system Gas dehydration equipmentinstalled on compressor suction lines can incur significant piping line losses Aminimum pressure loss specification should be included with other process para-meters when sizing and selecting equipment

8.14 CHAPTER EIGHT

FIGURE 8.8 3 Bank fast fill system.

8.4 CNG STATION SYSTEM DESIGNS

Most CNG refueling station systems are custom designed to meet specific cost andperformance criteria

8.4.1 Three Line Fast Fill System

The three line fast fill system is commonly used for retail sale of CNG It is amongthe most costly, but provides the maximum performance, with CNG dispensersproviding fuel at flow rates comparable to liquid fuel dispensers Fuel flows directlyfrom the pressure vessel storage and is independent of compressor capacity as long

as head pressure remains at the pressure vessels

With the three bank storage, storage utilization is high This enables the system

to best meet random surges in demand High storage utilization also allows thecompressor to operate for extended periods of time, with a minimum of stops andstarts

8.4.2 Time Fill System

The time fill system requires the least amount of equipment and is the least pensive, and no pressure vessels are required The compressor discharges directly

ex-CNG COMPRESSORS 8.15

FIGURE 8.9 Time fill system.

to the refueling vehicles Refueling rates are dependent on the number of vehiclesrefueling at once The system finds application with private vehicle fleets wherefast refueling times are not important and vehicles are parked for an extendedperiod of time during an off shift

Upon connection of the refueling nozzle to the NGV, a pressure loss in thefeeder line will be sensed by a pressure switch at the compressor The compressorwill start to compress gas through a time fill control panel This panel has a pressureregulator and temperature instrumentation with shut off valve to stop the fill whenthe fill pressure is reached

8.4.3 Single Line Fast Fill System

The single line system is similar to the three line fast fill system, but uses only asingle bank storage The dispenser is without sequencing valves and the prioritypanel has only one priority

The use of more than one dispensing hose is not recommended If two vehicleswith different fuel tank pressures are connected to the same supply line, the refu-eling of one of the vehicles will halt until pressures equalize

The system is less costly than the three bank system, but storage utilization islow Surges in demand cannot be met as effectively Low storage utilization results

in frequent compressor start and stop operation

8.4.4 Diverter System

The diverter system is similar to the single line fast fill system, but provides twohose refueling capability A diverter valve directs all compressor flow to the firsthose authorized Upon authorization of the second hose, refueling will begin di-rectly from the single bank storage without robbing compressor discharge flow

8.16 CHAPTER EIGHT

FIGURE 8.10 Single line fast fill system.

FIGURE 8.11 Diverter system.

from the first hose authorized As soon as the fill is complete at the first hose, thediverter valve switches all compressor flow to the second hose

This system finds application in the refueling of fleet vehicles one after another.The toggling of the diverter valve spaces vehicle filling so that at least one vehicle

is always connected to a hose While one vehicle is moved into position and nected to the fueling hose, the second vehicle is nearing fill completion In thisway, refueling is nonstop and the compressor operates continuously with a mini-mum of starts and stops

con-CNG COMPRESSORS 8.17

TABLE 8.4 Daily Refueling Frequency Distribution

8.5 EQUIPMENT SELECTION AND SYSTEM PERFORMANCE

The following example will demonstrate how to select a compressor and pressurevessel storage for a three line fast fill station For other refueling system types asimilar approach can be adopted

Example:

A proposed retail CNG refueling station site will have 25 psig regulated andmetered gas available by the local gas utility Maximum compressor discharge gaspressure will be 3600 psig The station will be of the three line fast fill type, withmaximum NGV fill pressures of 3000 psig The proposed CNG refueling instal-lation is forecast to have a filling frequency distribution as per Table 8.4 Theaverage fuel consumption of each vehicle is also determined to be 4.6 gallonsgasoline equivalent per day

37,500 scf / 8 hrs / 60 min / hr⫽ 78 scfmFrom Fig 8.2, a 25 psig suction pressure and 3600 psig discharge pressurerequires 0.48 BHP / scfm At 78 scfm, the required compressor horsepower is 37.4BHP A 40 BHP, 4 stage compressor is selected, providing a flow capacity of 83scfm

8.18 CHAPTER EIGHT

8.5.2 Pressure Vessel Storage Sizing

The filling frequency distribution is critical to the selection and sizing of the sure vessel storage The hours from 7am to 8am is the busiest with 20 fill ups Gaswithdrawn from the storage pressure vessels during this hour is calculated

pres-20 NGVs⫻ 4.6 gal equiv.⫻ 108.7 scf / gal equiv.⫽10,000 scf

In this time, the compressor has also operated to refill the storage To maximizecompressor running duration, control systems do not start up the compressor untilthere is a substantial drop in pressure of the storage To account for this, it isassumed that the compressor operates for 45 minutes of the hour Storage replen-ishment flow is calculated

45 min.⫻ 83 scfm ⫽3735 scf

In addition, it is assumed that the two vehicles between 6am and 7am did notinitiate compressor operation This accounts for an additional 1000 scf depletedfrom storage

The total amount of gas that must be provided from the storage pressure vessels

is calculated

10,000 scf⫺ 3735 scf⫹ 1000 scf⫽ 7265 scfThe size of the pressure vessel storage can now be estimated From Table 8.3,the storage utilization factor is 36% The size of the pressure vessel storage iscalculated

7265 scf / 36% ⫽20,181 scf natural gas @ 3600 psigThree pressure vessels each having an internal volume of 22.8 cu ft are selected.Each vessel provides a storage capacity of 7,082 scf at 3600 psig working pressure,

or a three vessel total of 21,246 scf

8.5.3 Other Equipment

A priority panel is selected and sized for the compressor discharge capacity of 78scfm Pressure losses through the panel are kept below 50 psid so that the storagepressure vessels are filled to a maximum pressure without compressor shutdown

A two hose metered CNG dispenser is specified so that two vehicles can be filledsimultaneously during peak demand times

8.6 CODES AND STANDARDS

CNG refueling equipment and installations must comply with numerous codes andstandards In North America, work is currently underway to further develop codes

CNG COMPRESSORS 8.19 TABLE 8.5 Industry Codes and Standards t

AMERICAN NATIONAL STANDARDS INSTITUTE (ANSI)

Refueling Connection Devices

Gas Vehicle (NGV) Fuel Containers

ANSI / ASME B31.3 Chemical Plant and Petroleum Refinery Piping ANSI / ASME B31.8 Gas Transmission and Distribution Piping

System

AMERICAN SOCIETY OF MECHANICAL ENGINEERS (ASME)

ASME Section VIII, Division 1 and 2 Boiler And Pressure Vessel Code

Qualifications Standard for Welding and Brazing Procedures, Welders, Brazers, and Welding and Brazing Operators

AMERICAN PETROLEUM INSTITUTE (API)

API Specification 11P Specification for Packaged Reciprocating

Compressors for Oil and Gas Production Services

API Standard 618 Reciprocating Compressors for General Refinery

Service API Standard 661 Air-cooled Heat Exchanger for General Refinery

Services

NATIONAL FIRE PROTECTION ASSOCIATION (NFPA)

Stationary Combustion Engines and Gas Turbines

Systems

Equipment

SOCIETY OF AUTOMOTIVE ENGINEERS (SAE)

Natural Gas Vehicle Fuel

t Industry codes and standard compiled with the assistance of Ray Benish, CNG Systems Inc.