F L U I t) F Lütfu A ru ü R ü"I'A-f l ru {i §: Q u § F ['i¡ E i\i T

,\;q'g'g5q¡li;Fr

'r"i,,,..,4,

HerearethevariablestomeaSurean(lnecessaryinstrumentation..-

E. WEtCOX, Chevror-r Energy Technology Company, Houstc¡n, Texas

Performance testing guidelinefor centrifugal compressors

ffi";,'.

performance can be an excellent indication of internal *e'iij'¿i'[ouling, which, il-allor¡,ed to conrinue, may result i,n,u_nschedrrledourages or reduced thiou'ghpur, In conrrasr, perci:lvád pérFor-mance problems may be a result of a compressor operating farfrom its original design. However, obtaining accurare per[or-mance data in the field can be very challenging.

This a¡ticle wiil explain the relarive importance that differ-ent process variables have in performance calculations, as ,.yeil

as specify the necessary instrumentation to obtain process datawith an acceptable uncertainty. Since original design condi-tions almost neve¡ match actual operaring conditions, how to

. , compare actual field data to design data usingn-ondimeqpionalhead and efficiency will be demonst¡ated. Likewise','the limitson these comparisons wiil be tiúilined fo¡ users. Several examplefiel{ performance evaluations are discussed with some common

' ' pitfalls that can be avoided. Examples of the effects of inaccurateprocess data are also included in the discussion.

Obtaining accurate fieltl data. Accurate centrifugal com-pressor performance measurement depends on the quality o[thefield data. In general, testing should follow the conditions set forthin ASIvIE PTC rc-D97. However, most petrochemical plants don'thave the inst¡umentation specified in this test code. Compressorpiping should be designecl to accommodate flowmeter runs ancl tomeet location requirements for pressure and temperature as weli.Small inaccuracies (in certain areas) can make a large differencebetween the measured versus actual conditions.

The field data required for an accurate performance evalu-ation are:

. Suction and discharge pr€ssures :

. Suction and discharge ternperatures 'I , ,

. Ivf ass flowrate \.' '

. Gas composition

. Rotational speed

. Dliver load. .'

The most important factors in obtaining accurare field dataare:

. Pressure and temperature transmitters shouid be locacedwithin a few feet of the compressor case at least 10 pipe diamete¡sfrom any valve, tee, elbow or other obstruction.

. Transmitters (both pressure and temperature) should be

calibrated prior to the test.

: , 'i: i. ,'r,a , , ',,. . il,,i., ;

. I)ata should only be taken during steady-state condirions.This is usually specified by requiring the discharge remperarLireto remain constant over a 15-minute inte¡val.

' RTDs should be used instead of the¡mocouples, if possible,due to improved accuracy.

. Test points at different flowrates or speeds are incrediblyvaluable. Inaccuracies in flow, pressure or temperature may bedifficult to determine with only one test point. Bur a map olpointr; wili usually reveal the problem.

Flowrate. The flowrate reported by the plant DCS is usuallynot absolutely correct. The meter facor, K, which is used to con-vert the measured differential p¡essure into a flow¡ate, is always a

function of the gas pressure, temperature and molecular weight(Eq. l). The flowrate reported by the meter must be corrected forthe actual process conditions. However, the co¡rection dependson che type offlow units. The correction equations fo¡ standardvolume, mass and actual volume flows are given in Eqs. 2,3 aná4, respectiveiy. It is cornmon for flowmeters to be "compensated"

in the DCS for the actual pressure and temperature. However, itis rare fo¡ the compensation to include the gas moiecular weighr,since this requires an online gas aoalyzer. The flowmeter design

and location should meet the requirements of the guidelinesestablished by ASME PTC-10.

Q= KJtr where . ¡¡: f (p,T,Mw) (l)

ffifu etailed performance analysis of centrifugal compressors

Elj § in the field is essential to evaluare their existing condi-ffion# ,¡o,.,. Current perform-agcq-qf e qompressor crr, ,lso be a

val,rable tool in evaluaiint i,r'íálíUility. Á a.L.._rre in comprytlg.f

e o =e.tr"lr-w. forstd. vorume flow 12)\á -\ P, T^ Z^ Mw^

h, = ,t.trW" Y, f,or nrass flow (3)^ ! Po Tn Z^ Mut,

o.=o[ut"z"**.\á -\ P^To Zo Mtanfor actual volume 0ow (4)

Example-inclrrect floameter. An incorrect flow measure-

ment rvill cause the compressor to appear either low or high inhead because the operating point is marked incorrectly on theperformance map. The best method to determine if a flow mea-surement is incor¡ect is co obtain several data points to compareagainst the entire cu¡ve. As can be seen in Fig. 1, tf.re r-r¡aximurnhead should ¡emain the same. The curve is just shifted to theright or left.

Gas composition. The gas composition is the most impor-tant data required to evaluate a compressor's performance.

tlqHyDRocARBoN pRocESsrNG AUGUsT zooz | :s

TAffiB.E '8. Effects of incorrect gas composition oncorrected flow and calculated horsepower

Correct gas sample, 275'F lncorrect gas sample, 75"F

i:i"{"I{il Fii ülJr/ i{il'jr,'

Actual curve

\'**o.y' Ftowtoohigh

Flow too low

§1tI

ffi f,iTA-T§ tru G fi ú"4 [J ii F,] i1,4 r fui'l-

Like¡n,ise, in a wet ga.s application, an accurate gas compusi-tion ir; also the most difficult piece of tesr data to obcain. r\tleast two gas samples should be taken during the performancetest. lfwo samples are required to vaiidate the gas analysis(for comparison) as well as when one of the samples is iosr orinvalid (i,e.,. qirhout the gas composition, the other processdara are wUrthlei$, fhe sample should be obtained using a

free-flowing arrangement (Ftg.2 shows two exarlpies). Aninsertion probe should bS qq.d, ifpossibie, to obtain the gas

sample instead of a wall tá'p. Samples obtained f¡on.r r.,alltaps u.i11 generally be ieaáe? in the higher molecular r,",eight

components due to the bounda¡r¡ layer effect. The samplebomb should be heated to the process temperature before itis analyzed. This is to p¡event any condensation of liquidsthat could alter the gas composition. Obviously, this is niuchmore important for higher molecular lveight services such as

FCC wet gas compared to reformer or hydrocracker recyclehydro¡1en.

As an example, a sample taken from the discharge of a cokcrwet gas compressor was analyzed at the lab ambient ternperature(approximately 7 5"F) and at 27 5"F (sample temperature, Table 1).

As can be seen, the inco¡¡ect gas composition has a pronouncedeffect on the calculatecl gas horsepower. This eflect is magnifiedbecause the molecular.veight is used to calculate the head as u,elias correct the flo.¡'.

Thc gas composition should be reviewed to determine if it isfeasible. One of the most common methods used to cleterminethe gas sample composition is a gas chromatograph. A gas chro-matogr:aph dete¡mines the components by burning them in thepresence ofa ca¡rier gas. For this reason, a gas chromatographwill not show any water vapor (i.e., water will not burn). Horv-ever, it is very common for process gases to be saturated withwater. For rhis reason,.the mEasured gas composition mu:r b.adjusrcd if ir is ihdeed saturated wirh water. The mole fra. r ionof any constituent in a mixed gas stream is defined b1, Eq. 5.

Since the gas stream temperature .is known, the water vaporpartial pressure is equal ro its saturation pressure at the giventemperature, Afte¡ the water vapor mole fraction is found, thenormalized gas composition must be adjusted so that the toralmole fraction equals 1.0.

. constituent Parlial Pressure

ur..,*u., ,", ^^::::,::,::;:,,"r, ,he gas composi,ionwill rypically become k:ane¡ as liquids are condensed out i¡r theintersection knockouts. If the gas samples clo not re [iect this.there is usualiy a problem. Also, certain components (i.c., F{2 orI-i25) are normally noncondensabie. The mass fractions of ¡hese

comporlents should actually increase as the heavier liqr-iids are

removed fiom the gas srream.

Y&ffiLE §" Effects olf incorrect gas composition on

"***;ffi sffi *f:ffi @a*?**s!

oo ] nucusr 2007 HyDRocARBon pRocESsrNG

7,036

e/l

Effy. mf 1.

Molecular weight

Flow MMscfd 32, s27.4

Shaft horsepower, hp 5,77s

t

l-Lffii!:) Ffl"t'1ji/ lÁl§ü ftr3U",ú{,f II\f {r Hq}{.JIF}lVlE[ri"I

Example-inclrrect gas cornposition. ]f the measured gas com-position is lower than the actual gas composition in the compres-sor, it will have the most pronounced effect on borh the correcredflowrate and calculated polytropic head. An incorrect low molec-ular weight will cause the corrected flowrate (if it is measuredin standard cubic feet) to be higher (Eq. 2). Likewise, the lowmolecular weight will cause the calculared polytropic head tobe higher than it actually is for the given compression ratio. Amolecula¡ weight that is coo low w¡ll al^so.,pause the calculatedpolytropic .ffñi..,.y to be higher,

tl(¿ú4áf "o, as pronounced.

Additionall¡ the lower molecular weight will lower rhe calcu-lated mass flow, which will in tu¡n lowerthe calculated horsepower. Note, these arethe effects produced by an incorrect gas

composition being used as the input fora field test. These are not the results of a

compressor that is operaring in a gas thatis actually lower in molecula¡ weight thanits design. If the measured gas compositionis too high, the results are just the inverseof the above (i.e., the flow, head and effi-ciency are all lower). A summary is shownin Table 2.

Calculating and evaluatingperformance parameters. Oncerhe field performance data are obtained,tlre correct performance parameters mustbe accurateiy calcuiated and evaluatedro ensufe thar rhe field data are realis-tic. The most critical step in calculatingperformance parameters is determiningthe inlet and outlet density, enthalpyand entropy. For hydrocarbon gas mix-tures, performance programs that useequations of state, such as Lee-Kesler,Iledlich-Kwong or Peng-Robinson, wiilprovide much better results than approxi-mations from Mollie¡ diagrams or idealgas relationships. Process simulators canbe used as well. Once the gas propertiesare calculated, the cor¡ect parametersmust be selected to adequately evaluatethe compressor performance. Evaluatingthe results ofthese calculations should be

made to determine the validity of bothrhe calculatioeq -and,the field data (see

Schultz) fo, "

t[li"'j:airhe equations forthese parameters).

. Shaft horsepower-lf a torque meteris avaiiable, the shaft horsepower may be

calculated directly. However, since this is

not commonly the case, the horsepowercan be calculated by the heat balancemethod, as given in ASME PTC-10 andcompared to the calcuiated driver horse-power. This comparison between driverand driven power is the mosr imporrantindicator of accuracy of a performance test(i.e., if the two values are not close, some-

rhing is wrong with the test data).

Applying the first law of thermodynamics to the control vol-ume around the compressor is shown in Fig. 3.

(*, - *r,)(h, - h,)* e* ,rtiP = 2.545 + hP uecu

Shp = hp*, + h?ueca, h?wcn = hput+ bpuz

(6)

(7)

'I'he radiant heat loss, Q¡, is normally negligible, but it can be

approximated by dividing the compressor case into axial secrionsand approximating the heat t¡ansfer from each section.. Likei,r,ise,

: .'\': '. ,

t/EHYDROCARBON PROCESSTNG AUGUSTZO0T | 61

*: I_ffi g [.] í: LüW /\N [} iq ü"8 "¿\r ñ r,d ffi E I {-§ i p ¡\{ H [\{ T"

Wo seal losses

Wseal losses

Loss inpressure dueto higher f7

lncrease involume flowdue to higher11 and leakage

a

en lrn atri-f',rrc¡' ''1r il:': o':ilT 5 ¡i :'¡-!i''the seai leakage on the inlet is normally less than 170, but it canbe easily calculated at the orifice in the ventrofF'the sealipoti.Note, the internal seal losses (i.e., balance piston and impéllerlabyrinth seals) do not affecr the calculated shaft horsepower.However, they do affect the calculated head and efficienc¡ whichin turn affects the discharge pressure and temperature, Theeffects ofseal losses wiil be discussed later.

. Polytropic head--This value is limired to approximately10,000 -12,000 ft-lbf/lbm per impeller for most closed impel-lers. The maximum is set by the impeller tip speed which is lim-ited by the gas sonic velocity as well as the impeller yield stress.

Impellers designed to operate in highly corrosive processes(H2S, CO2) often require a maximum yield suess of 90 kpsiand Rockwell C of less than22,which limirs them ro approxi-mately y,0OO ft-tUgtbm. For example, if a performance rest isdone on a compressor and the calculared head per impeller is20,000 fclbf/lbm, chen either the rneasured compression ratiois roo high or the measured molecula¡ weight is too low.

. Polytropic efficiency-The maximum value depends onthe impeller specific speed, but is limited to approximately75-78o/o for oider 2D impellers with vaneless diffuse¡s and80-85o/o for newer 3D impellers with vaned diffuse¡s.. _ ., i, f :\'

The most important p.rfor-".r.. data'lirmr-,J"t'Jst'ii' ,t. "

driver and driven power comparison. If the difference betweenthe driver and driven power is low, and the calculated headand efficiency are low, then the data are probably good andthe compressor has a performance problem. If the head and/or efficiency is off, but the porver does not agree, this typicallyindicates bad test data.

Losses. Losses are generally grouped into three distinct areas:

mechanical, seal and aerodynamic. Mechanical losses includepower dissipated through bearings, oil or gas seals, shaft-drivenlube oil pumps and gearboxes. Seal losses are the decrease in theamount ofenergy available to conver! into pressure head dueto internal recirculation inside the compressor. Aerodynamic

T'AE§-E 3. Mechanical losses

losses include effects such as friction and pressure losses in theimpeller:s and diffusers. For the purpose of this article , aerody-namic losses are included in rhe impeller efficiencies and notdiscussed.

Mechanical losses. i\4echanical losses a¡e mostly a functionofsize and speed. Larget: bearings and seals at higher speeds dis-sipate rnore power. Mechanical losses are simply added ro thecalculated gas horseporver. These losses can be approximatedfrom:

' IVleasuring the florvrate and lube oil temperature increaseand using Eq. 8. This rnay be difficult due to many differenriube oil return lines and pressure controllers that spill back tothe reservoir.

. OF,M-supplied curves for diffe¡ent bearings and seals.

. Bearing rotordynamic computer models that calculatehorsepower losses as well.

. Tables based on compresso¡ gas horseporver (T'able 3).Approximate gearbox efficiencies are given in Table 4.

thc^LT.,L^ _ t. _o,t (g),'f MÉ:CH 33,000

Seal losses. Balance piston or division wall leakage is theonly seal loss evaiuated in rhis discussion since rhey are usua.llymuch larger than impeller labyrinth seal leakage. Seal losses are

much mo¡e difficult to estimate than mechanical losses because

they are not just added to the calculated gas horsepower. Leak-age through the balance piston seal to the compressor suctionincreases the volume flow through the impellers as well as

increases the inlet temperarure, both of which decrease thecompressor discharge pi.rrr,.. (Fig.4). Balance piston leakagecauses the measured head and efficiency to decrease, but dc¡es

not increase the calculaced gas horsepower. This does not meanthat increased balance piston seal leakage does not affect com-pressor power. In the operating rvorld, compressors have to 1>ut

up a re<¡uired discharge pressure to pump forward. As the bal-ance piston seal leakage increases, either the compressor speed

or suction pressure must be increased to maintain the desireddischarge pressure and flow. This causes the power consumedby the compressor to increase.

A feel for the amount of balance piston leakage can be estab-lished by measuring the differential.pressure across the balance

piston line. Most OElvls design for this to be less than 2 or 3

psid. Anything above this usually means that a balance pistonseal is leaking excessively. A differential p¡essure gauge is almostalways required to make th.is measu¡,:ment due to the smalldifferential.

Seal leakage rate can be estimated by using a leakage equation,Eq. 9 applies for adjacent'teeth)in the l:Lbyrinth. This is used toiteratively solve for the labyrinth cavity ¡,¡s5su¡es and seal leakagerate. If the flow is chbkéd, Eq. 10 is used for the last labyrinth.

YABLÉ 4" Gearbox efficiencies

P2

Gas power requirement , , Mechanical losses,% Gear type Etficiency, %

0 -3,000 3.0 Helical 97-99

3,000 - 6,000 2.5 Herringbone 6-99

6,000- 1 0,000 95-982.0 Straight bevel

62 I Auousr zooT HyDRocARBoN pRocEssrNc

1.5

,r lq

Spiral bevel 96-98I0,000 +

;SPECIALRE

----7

<-

,i:LLJiD FrOkV An§m müTATil\tü EQU¡pf\fitEt{T

;'r;,1','" '.

':..!

polytropic head and efficiency calculations. Adding in the modelfo¡ the balance piston seal leakage allows us to see why the com-pressol efficiency is not as designed.

Comparing measured field performance to shoptest or predicted performance. Since field conditionsnever exactly match the original design, certain nondimen-sional parameteÍs must be calculated so that the field perfor-mance can be compared to the OEM shop test or predictedperformance data.''§7hile these nondimensional parameterswill enable "apple-to-apple" comparisons for different condi-tions, they have very real limitations based on the aerody-namic characteristics of the impellers. These nondimensionalparameters includei

Polyrropic head coefficient, p7

ric = üoütHP.2. - P.2

,-l ,

NT (e)

(10)

Hllp = .--L -

lnDN)'

Poiytropic efficiepc¡ r¡p

H\t, =;J;

frt-bl

Flow coefficient, @

O: Q. or gND' ¡/

(12)

(1 3)

(14)

Continuecl

. 0.51¡trP*rH

,lznrOnce the leakage ¡ate is clerermined, the measured field data

must be adjusted to find the actual conditions. In the case of a

straight-through compressor with a balance piston seal, rhe hotbalance piston seal leakage is mixed with the gas at the inletusing a mass and enthalpy balance to determine the actual inlett.-[.r.tur. to the firsJimpeller (Fig. 5). This will change the

/t"\

= hi+ hoo

nl1 * flt¡p

?'/tlHyDRocARBoN pRocESSrNc AUGUsT zooz ] o:

,.LL.t

lnlet flange flow,m¡ P¡ T1, h1

lmpeller flow,IIll 't lfl6p¡ Ptu Tt', ht'

Ma

FLI,'ID

::t:i::. .'

Impeller mach number, ,41

The first step in any comparison is to obtain a set of non-dimensional curves for rhe shop test or predicted data. Ifthey were not provided by the OEM, they can be obtainedby iteratively calcuiating the p¡ and r¡2values from the givenvalues of discharge pressure and shaft horsepower. This canbe accomplished by guessing a discharge temperarure for thegiven discharge pressure until the correct shaft horsepowerhas been reached. Once the correcr discharge remperaru¡eis known, the polytropic head coefficient and efficiency canbe calculated to give a set of nondimensional curves (Fig.6). These curves will predict the compressor performancefor the given mach numbe¡. They can be used ro compareagainst the existing operating condirions if rhe field machnumbe¡ is close enough to the mach number for the curves.The ASME PTC-10 test code defines the maximum shift inmach number ratio for a certifiecl shop performance resr (Fig,7). These limitations are good to apply in the field as well. Ifthe mach number shift is too large, the comparison may beinaccurate. Ifthis is the case, a new set ofperformance curves,rhat match the actual iniet conditions, should be obtainedFrom the OEM.

o+ | aucusr 2oo7 HyDRocARBoN pRocESSTNG

fiL0w AtuD RüTAT"IN6 §QL'NF'MEzu'[',

=a 0.1

§ o.o

.0,1

-0.2

Continued

(1 5)

Allowable shift in mach numberASME PTC 1O-1997

blq

y F;g"rrr. 1,0

The polyrropic head coefficient and efficiency ar the exist-ing operating conditions can be plotted on the nondimen-sional graphs to determine if the compressor is operating onirs curve. Also, the nondirlensional curves can be used tocalculate the field discharge conditions (pressure, tempera-ture, horsepower, etc.) based o¡r the field inlet conditions andthe performance curves (Fig. 8). Even though seal losses donot affect the calculated horsepower (i.e., they are includedin the lower measured efficiency), they do greatly affect thepredicted horsepower. The seal losses increase the calculatedvalue of é, which moves the predicced operating point fartherro the right on the performance curves, which always increasesthe horsepower.

Exa m pl e perf ormance eva I uat¡oñ. Reformer hydrogenrecycle compressor configuration:

. Six impellers, straight-through, barrel-type

. Vaneless diffusers

. Interlocking aluminum labyrinth balance piston seal

. Motor-driven through a speed-increasing gearbox

. Contact-type oil face seals.

- Original design:.6,370hp

'7,940 rpm. Inlet llow = 14,600 cfm. Molecular weight = 5.18

66 I AUGUST 2007 HvDRocARBoN pRocESS|NG

Fn'u Í l_i h Lü\flf Á tr\¡ ü RüTAT § N ffi [I{J E iP[Vd ü i§T

'Tr = 100'F

'Pr = 187 Psia' Pz = 275 Psia.* Balance piston labyrinth dimensional data:Stationary labyrinth: nine teeth, 0.375-in. equal spacing,

0, 1 875 -i n.'r.al|, 14.59 5 -i¡t inte¡nal diameter.Rotating labyrinth: eight teeth, 0.375-in. equal spacing,

0.I875-in. :.¿ll, 14.937-in outside diameter.Calculated seal leakage rate = 120 lbm/min.

- -'- I\4easured performance data:Pr = 150 PsiaTr = 80'FPz = 248 PsiaT2 = 191'FBalance¡hamber pressure = 156 psiaFlow = 190 MMscfdSpeed = 7,949 rpmMolecular weight = 7.1

Driver data:Amperage = 978Voitage = 4,0A0Speed = 1,784P¡= o'gzE = 0.9571) Calculate the horsepower supplied to the compretii))r,rr,

Polytropic head coeff., p.p and effieincy, rlp

7h

l7\* \{

\

l:iff:-* | \1.4 1.6 1.8 2.A

opolytropic head coefficient an{ efficiency.

I

If the ge arbox eFficiency is assumed rc be 97o/o, the calculatedshaft horsepower delivered to the compressor is:

hp,o-p = 7,988 x 0.97 = 7,750 hp

The OEM supplied a performance curve showing predicteddischarge pressure and shaft horsepower (Fig. 9). These curvesare valid for only the design conditions listed previously (i.e.,

you cannot plot the measured discharge pressure and shafthorsepower on these curves). Likervise, they are only predictedcurves (i.e., the compressor was not shop performance tested).

ít {-}1";rlr 1l i §\i (;; H i.? tli il [r' íiln ii: i\J T'

'§"íAÉffitl",l$i 5. Reformer hydrogen recycle compressor

Predicted Predictedperformance performance

(w/o seal losses) (w/seal losses)

248

191 l/tr 182

Head, lbf-ft/lbm 64,504 68,898 66,411

Efficiency, % 65 79 76

Shp 7,641 6,129 7,030

These curves must be converted into a nondimensional form so

that they can be used ro predict the existing field performance(Fig. 10).

The discharge condirions are repredicted with and withouuseal losses to compare against the measured field results (Tables

5 and 6). The horsepower fo¡ the existing field performance datamust be close to the horsepower supplied by the driver for theresults to be considered valid. As you can see, the con-rpressor effi-ciency is considerably iowe¡ than what is predicted by the OEMperfornrance curve. Adding the seal losses (approximately 4-5o/oof the r:otal flow) to the predicted curves brings the predictedand actual conditions closer together. However, the rneasu¡eclefficiency is just too low to be a balance piston problem alone.

Likewise, the thrust bearing temperarure and axial position were

relative:ly low. Based on the past fouling history in this com-pressor as well as the high balance piston differentiaI pressure,

the loss in efficiencywas rhought to be a result of fouling. Thecomprc:ssor was still moeting the desired discharge pre.ssure, burthe lou, efficiency was causing excessive horsepower consump-tion, which was limiting unit charge rate. Because the motor was

oversized and the loss irr efficiency had been gradual, operationswas unaware that the pr:oblem was in the comPressor and not the

motor (i.e., they thought the motor was dirty).The compressor wa$ pulled because it couldn't be washed in

place due to a lack of adequate case drains. Alarge amount ofammonia chloride buildup was found in the compressor station-

.\qry.cpryp.p,rlfll§ (the diffuser channels had approximately 40o/o' UtUiUrlC, Figs. 11 and 12). Note: The synchronous vit¡ration

amplitudes were relatively low (< 1 mil) because the fouling was

mostly on the stationary components; the only fouling on the

rotor was on the inside diamete¡ of the impelier eyes.

Afte¡ the compressor was ¡einstalled, the measured Field per-

formance was within 3o/o of the predicted.

NOMENCLATURE

Q = sPecific heat, constant Pressure, Btuilbm, 'FC = specific hear, consrant volume, Btu/lbm, "F

D = imPeller diamerer, in.á = enthalpy, [i¡u/lbm

,? = head, lbf:ftllbmhP = hotstPo*e'

,á = sPecific heat raúo, crlc,l( = meter lactorm = mess flow, lbm/sM = mach number, non-dim.

MMscfd = million s¡andard cubic feet per dayMw = moleweight, lbmilb-mole

l/= speed, rpml/S = specific spt:ed, non-dim.

P - pressure, psia

Q = volum. [low, cfm

blq

F t" tJ it l1-) F t üifl{ i}, I'{ ffi

Paranreter Measuredfieldp e rfo rnra nce

258P2

fn DVf i ¡

flDJ tkator

+7 I

4,000 x 978x0.957 x0.92= 7,988 hP

431

Select 1 60 at www.HydrocarbonProcessing.com/RS

)F r

§l

Ffl-Lilfl} F[-üW At\X§] ffi i"il SrT t f\i G ffi f; {"¡ f, P $\fi E h§T

't{'r+ñ3ü=ffi 'Í}" Example test and repredicted field per"formance data

Title Hydrogen recycle compressor

Test data Corrected test data Predicted data

150.0 248.0 150.0 252.1

:', Bartos, J. C., "Predicting Centrifugal CompressorPerformance from lmpeller Dimensional Dara,"Power Machinery and Compression Confere ncc,

_/ I990, Housron, Texas.' Boyce, Iv[. P., "Principals ofOperarion and Perfor.

mance Estimarion of Centrifugal Compressors,"Proceedings of the Twenty-Second TrtrbomacbinerlSlmposium, Turbomachinery Laborarory, TexasA6rM University, College Srarion, Texas, 1993, pp.

/ 161_175.

' iCla¡k, W. A., "Measu¡e Porver by rhe Simplest Merhodand Use fo¡ Onstream Computer Monitoring," Pro-ceedings of the knth Tt¿rbamacbinerlt Syntpostum,Turbomachinerl, Laborarory, Texas A&M Univer'

/ siry, College Station, Texas, I981, pp. l5l-153.Davis, H. M., "Compresso¡ Performance, Field Dara

Acquisition and Evaluarion," Proceedings oftbe Sec-

ond Tu' bomac h inerl SIm?o tium, TurbomachineryLaboratory, Texas A&M University, College Sra-

, rion, Texas, 1971. pp.15 22.n Gresh, M. T., Compressor Perfotmance-Select:on,

Operation, and Testing ofAxial and Centrifugal Cont-pretsort, Buterwo¡h-Heinemann, Boston, 1 991.

''Heftmann, E., ed., Cbromatography-A LaboratoryHnndbo o k of Chrom atograp bi c an d E lec trop h o t e t i c

Method¡, Van Nosrrand Reinhold Co., tr-e w York,1.975.

"lUiti, O, J., "Centrifugal Comprcssor Rerares andRedesigns," Powe, fufachiner! and Compression Con-

.- ference, Houston, Texas, 1995

" Hun,, J. \f., "Field Testing of Ce nLrifugal Comprcs-sors," Proceedingt ofthe Tentb Turbonachinery S1m-posium, Turbomachinerl Laboratory, Texas A&M

.7 University, College Station, Texas, 1981.'" Japikse, D., "Efficiency: The Silenr Parrner of

Machine Performance," Proceedings of the EleuenthTur boma c h in ery STmp otium, Turbomachinery Labo -

ratory, Texas A&M Universiry, College Sration,Texas, 1982, pp. 105-111.

,, Kerlin, T, 1i7. and E. M. Katz, "Temperature Mea-suremcnt in the 1990s," INTECL|, April 1990, pp.40 - 43.

,: Lapina, R. P., Estiruating Centrifugal Compressor Per-

formance, GulfPublishing, Houston, Texas, 1982.

: Lee, B. I. and M. G. Kelser, "A Generalized Ther-

Suctionconditions:

Dischargeconditions:

lnlet 0utlet lnlet Discharge

Press, psia 150 248.0

Temp, "F 80 191 ,0 84.9 191 ,0 84.6 182.1

Flow 190

0

0

MMscfd (only one flow is required)

lcfm

lbmimin

Flowmeter

design data

1 Location (1 - suction, 2 - discharge)

0 Mole weight0 Temp., "F

0 Press., psia

Speed 7,940 rpm 1,940

lmpeller Number Diametec in.

23

Calculated data

Mole weight a1

Properties lnlet0utletlnletPredicted Predicted

outlet inlet outlet

Compressibility

Enthalpy, Btu/lbm

Specif ic volume, ft3/lbrn

Specific heat, Btu/lbm,'Fcplcv

Entropy, Btu/lbm, "F

1.0026 1.0059

561 .3 673,0

5.44 3.83

1 .130 1 .1121.33 1.32

3.77 3.82

1 .0026

5 56.1

5.39

1.1291 ,33

3.76

I.0059683.42 0/

1.1151,32

3.84

1,0026

561 .6

5.44

1 .131

1.33

3,77

1.0059

683.4

3.94

1.l751.32

3.84

Flow (corrected) Flow (corrected) Flow (predicted)

MMscfd 190.0

lcfm 13,456.9 Seal losses, lbm/min 1 14.0 120.1

lbm/min 2,496.0 lmpeller flow, lbm/min 2,6'10.0 2,616.1

Q/N 1.695 1.789 1.792lmpeller q/n

PerformancePerformance Predicted(corrected) pefformance modynamic Co¡¡elation Based on Three-

Paramete¡ Corresponding Stares," AIChE Journal,, Yol. 2t,No. 3, May t975, pp. 5t0-527.1 M"rrh.*s, T., "Fie ld Performance Tesring to Improve

Compressor Reliability," Proceedings of the TentbTur bomach inerl 51 mpot iun, Turbomachinery Labo -

rarory, Texas A&M Universiry, College Srarion,

¿. Texas, 198I, pp. 165-167.Nathoo, N. S. and \i/. G. Gottenbe rg, "Measuring the

Thermodyaamic Pe¡formance of Mulristage Com-pressors Ope rating on Mixed Hydrocarbon Gase s,"

Proceedingt of the Tenth 71u'bomacb inery STmpo siutn,

Turbomachinery Laboratory, Texas A&M Univer-

,... siry, College Sration, Texas, 1981, pp. 15-22.". Pais, R. B. and G, J.Jacobik,' Field Perfo¡mance fesr-

ing of Ce nrrifugal Compressors," Proceeding of úeTh ird Titr b omachinerl Sympo tium,'furbomachineryLaborarory, Texas AE¿M University, College Sra-

_ tion, Texas, 1974, pp.84-90.llschulrz, J. M. "The Polyrropic Anaiysis of Cenrrifugal

Compressors,",/a urnal of Engineeringfor Pouer, Yol.1, pp. 69-83.

viASME PTC I0-1997 Performance Test Code on

Conpressors and Exhau¡ters, American Society of7 Mechanical Enginee rs, New Yo¡k.

",)Van Laningham, F. L., "Guidelines tbr Pe¡formanceTesting Centrifugal Compressors," ?roceedings of the

Te n th Tro'bo m a ch ine ry Symp ot ium, 1'u rbomach ineryLaboratory, Texas A&M University, College Sta-

rion, Texas, 1981, pp. 169-171.' - "Bullerin TS-l02," Minco Product Catalog, HP

Polytropic exponent 1.61 1.56 1,48

Polytropic head 64,509 64,195 66,417

Polytropic head coeff. 0.545 0.541 0.561

Polytropic wl 1.000 1.000 1.000

Polytropic effy. 0.65 0.68 0.76

Gas horseoower 7,497 7,497 6,892

Shaft horsepower 7,647 7,647 7,030

Mach no. 0.357 0,35 9 0.35 5

Volume ratio 1,361 1.379 1,421

Q¡ = radiative heat rransfer, BtuR = gas constantf = temPe¡ature, 'F/ = specific volume, ft3llbm

Z = compressibilirySubscripts:

I = inlet condirions2 = ourlet condirions

áP = balance pisronD = design condirions

MECH = mechanicalM= mechanicalP = polytropic states = isentropic state.t = corresponds to seal conditions

SymbolsA = differentialI = efficiency, non-dim.<D = sonic velocity ratio, non-dirn.

U. = head coefficient, non-dim.U.o.r = labyrinth seal entrance coefficient,

non-dirn.

- BIBLIOGRAPHY' Aicher, \V., "Tesr of Process Turbocompressors Wirh-

our CFC Gases," Proceedings of the Tuentl-SecondTur b oma ch in cr1 Symp o s ium, Tur b om ac h ine 11 L d bo -

ratory, Texas A&M University, College Station,Texas, 1993, pp. 17)-184.

q/q HYDROCARBO¡I PROCE5STNG AUGUST 2007 | 69