Operators of petroleum plantsoften contract compressor orturbine OEMs to supply anentire turbomachinery train. The scopeof supply normally includes the designof the peripherals systems, such as theanti-surge control system for centrifugaland axial compressors. While there areadvantages to having a single partyresponsible for the entire train, it canresult in "finger pointing" and confusionwhen the various equipment OEMs dis-agree on the design of the peripheralsystems.

Engineers at a gas processing plantrecently found themselves in this situa-tion when adding a medium-pressurenatural gas compressor as part of amajor plant expansion. The gas turbinemanufacturer, who was the responsibleOEM, designed an anti-surge protectionscheme for the process compressor, butthe compressor manufacturer felt thedesign was inadequate. Specifically, thecompressor manufacturer felt that inaddition to the cold recycle loop provid-ed in the initial design, a hot recycleloop was also needed to protect thecompressor from surge, especially in theevent of an emergency shutdown (ESD).

Good anti-surge control systemdesign is difficult, and it encompassesthe proper sizing, selection, and loca-tion of all of the following elements:piping that comprises the recycle loops,the recycle valves, volumes of vessels inthe recycle path, check valves, and last but no less important the anti-surgecontroller.

It is very difficult, using steady-stateanalysis, to predict the effectiveness of aproposed anti-surge control system forvarious possible process-upset scenarios,due to the rapid onset of surge (approx-imately 300 milliseconds) and the inter-action of the various components of thesystem. Because both manufacturerswere basing their decisions on steady-state analysis, the plant engineers lackedconfidence in either partys opinion.

Validating anti-surge control systemdesign is an area where advanced, high-

fidelity simulation can be applied suc-cessfully for scenario testing, and in anattempt to determine the true require-ments for machine protection, plantpersonnel commissioned CompressorControls Corporation (CCC) to performa dynamic simulation of the two anti-surge control system proposals.

Using readily available data, the abili-ty of the control systems to preventcompressor surge under the followingscenarios was validated through simula-tion: ESD of the compressor train whileoperating in steady-state at design con-ditions Full closure of a process valve on thecompressor discharge side with the com-pressor operating at rated condition insteady-state Full closure of a process valve on thecompressor suction side with the com-

pressor operating at rated condition insteady state.

The results of these simulation sce-narios assisted the plant engineers inanalysing the competing solutions andchoosing an appropriate design for thecompressor anti-surge system.

Simulation modelA schematic of the dynamic simulationmodel used to describe the various ele-ments of the compression systemincluding both hot and cold recycleloops is presented in Figure 1. The entiresystem was broken down into the com-pressor, valves, and a number of vol-umes for associated pipes, knockoutdrums and the gas-cooler. The parametersof every element were determined fromthe data provided by the equipmentmanufacturers and the gas plant engi-neers. In the few cases where certain

Validating anti-surge control systems

At a gas processing plant, real time simulation was used to analyse the designof a surge protection scheme for a natural gas compressor, after questionsarose over the systems effectiveness in the event of an emergency shutdown

Nikhil Dukle and Krishnan NarayananCompressor Controls Corporation

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Figure 1 Simulation model schematic

data were not available, values wereassumed based on field experience andcommon-sense engineering practice.

In addition, the simulation modelwas connected to real controllers in anin-the-loop setting. This enableddirect comparison of a generic anti-surge controller employing standardPID-type algorithms and a purpose-builtanti-surge controller employingadvanced control methods and predic-tive algorithms.

Model assumptions The following is a list of the major fea-tures of the model and associated sim-plifying assumptions:Natural gas parameters molecular

weight, pressure and temperature in the suction and discharge headersare assumed to be constant and areset up as boundary conditions for themodel.

Three invariant coordinates compres-sion ratio, reduced-speed andreduced-flow are used to approxi-mate process compressor performancemap. Compressor reduced-speed andcompression ratio are used to calcu-late compressor inlet flow. Compres-sion ratio and inlet flow are used tocalculate shaft power. It is alsoassumed that compressor perfor-mance curves provided are valid insteady-state as well as transient condi-tions. The compressor map is definedonly over the speed range providedby the manufacturer (that is, 70% to105%). No attempt has been made toextrapolate the map for speed beyondthis range.

All pipes are modelled using a lumped-parameter approach to calculate pres-sures, temperatures and flows alongthe process. Pipes resistances areincluded into resistance of associatedvalves, and pipe volumes are splitequally between the main upstreamand downstream volumes connectedby the pipe.

Pressures and temperatures in processvolumes are calculated using ordinarydifferential equations describing theaccumulation of mass and enthalpyin the volumes.

The volume associated with the gas cool-er in the compressor discharge sectionis split equally between adjacentupstream and downstream volumes.

Process valves are modelled using stan-dard ISA sizing equations, and takeinto account the actual valve size andflow conditions.

Suction and discharge block valves areassumed to have linear trim charac-teristics and a stroke-time of 2sec.

Suction and discharge valve closure testswere performed at constant compres-sor speed.

The OEM anti-surge controller was mod-elled as a conventional PID algorithmwith a 20 millisecond loop-executionrate. Anti-surge controllers providedby machinery OEMs typically employstandard PID algorithms or a minorvariation thereof.

The ESD scenario is simulated withoutany anti-surge controller in the con-trol loop. On ESD, the anti-surgevalves are opened and go full openbased on their stroke times. The delaybetween the ESD signal and the open-ing of anti-surge valves is set to 20milliseconds to simulate the timerequired for a solenoid valve to de-pressurise. Actuator stroke-time forfull opening was assumed as 2.0secfor the hot recycle valve and 2.2secfor the cold recycle valve. The stroketime for the cold recycle is set 0.2seclonger for the purpose of display inan effort to distinguish its trace fromthat of the hot recycle valve. Thisassumption does not have any signif-icant impact on the results.

Shaft deceleration rate on ESD is 4%per second.

ValidationOnce the model was developed, a com-parison was made between the designdata provided by the OEM and the com-pression system model results understeady-state conditions with the com-pressor at the rated operating point. Thevalues of speed, suction pressure, dis-charge pressure, discharge temperature,compressor power and inlet flow werecompared, with a maximum errorbetween design values and model-calcu-lated values of 0.9%.

The small error between the valuesvalidates the model under steady-stateconditions, and provides good initialreconciliation for the analyses of tran-sient process disturbances.

In addition, gas flows through thecold recycle valve and hot recycle valve

were compared. The model-generatedflows exactly matched the design-calcu-lated flows in steady-state conditions,which validated the recycle valve modelwith respect to valve size and trim.

Simulation resultsResults for each test scenario are present-ed as a plot of time traces of pertinentcompressor parameters. A proximity tosurge variable, SS, describes the locationof the compressors operating point ascompared to its surge limit, and is invari-ant to changes in suction conditions. IfSS < 1, the compressor is operating in thesafe zone of the available operatingenvelope; if SS = 1, the operating point isat the surge limit of the compressor; if SS>1, the operating point is in the non-sta-ble zone to the left of the surge limit andthe compressor is in a surge condition.

Scenario 1Compressor ESD

Scenario 1 analysed the ability of thehot and cold recycle-loops, as proposedby the compressor manufacturer, to pro-tect the compressor in the event of anESD. This scenario was further brokendown into three separate sub-scenarios(Cases 1 to 3) to analyse the effective-ness of the hot and cold recycle loopsindividually and together.Case 1: open cold recycle valveAs seen in Figure 2, fully opening thecold recycle valve (ASV6511) immediate-ly on ESD does not prevent surge. In thefigure, the ESD occurs at time t = 0sec;the cold recycle valve begins to open20millisec after ESD and opens fully in2.2sec. The hot recycle valve remainsclosed at all times. Note, however, thatthe operating point reaches the surgelimit less than 1sec after the ESD.

The reason for surge is the inability ofthe cold recycle valve to quickly andeffectively de-pressurise the large vol-ume in the cold recycle flow loop. Thislarge volume is essentially made up of

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Figure 2 Operating parameter trends on ESD, cold recycle only

the cooler and associated piping. Thetime constant of this volume whenoperating at design conditions is about70sec, which is too long for effectiveand quick de-pressuring. Since the rateof speed reduction (and therefore com-pressor flow reduction) is greater thanthe rate of compressor pressure reduc-tion, the compressor continues to surgeas it coasts down. This is indicated bythe proximity to surge indicator SS,which remains greater than 1 for theduration of coast-down.Case 2: open hot recycle valveAs seen in Figure 3, fully opening thehot recycle valve (FV8016) on ESD pre-vents surge, but the compressor opera-tion reaches its surge limit (SS = 1). Inthe figure, the ESD occurs at time t =0sec; the hot recycle valve begins toopen 20millisec after ESD and opensfully in 2sec. The cold recycle valveremains closed at all times. Within1.2sec after ESD, the compressor operat-ing point reaches its surge limit, but thecompressor is protected, albeit with verylittle margin for error.

Comparing this to Case 1, the hotrecycle loop is obviously much moreeffective in de-pressuring the compres-

sor discharge as the flow through thecompressor reduces when speed beginsto drop. The check valve between thehot and cold recycle loops is very effec-tive in isolating the large cooler volumefrom the hot recycle loop as compressordischarge pressure decreases. In thiscase, the time constant of the volumebetween the compressor discharge andcheck valve upstream of the cooler isabout 2sec; much smaller than the 70secfor Case 1.

Because of the short time periodbetween the disturbance and theapproach-to-surge event, the influenceof hot recycle valve opening on com-pressor suction temperature is ignored.Case 3: opening hot and cold recyclevalvesAs seen in Figure 4, fully opening boththe hot and cold recycle valves simulta-neously on ESD is very effective at pre-venting surge. The compressoroperation never reaches the surge limit,as compared to Case 2 (maximum SS =0.75 as compared to SS = 1). In the fig-ure, the ESD occurs at time t = 0sec; thehot and cold recycle valves both beginto open 20millisec after ESD and openfully in 2.0 and 2.2sec respectively. The

compressor operating point reaches theclosest to its surge limit 1.1 seconds afterESD. This is very similar to Case 2,except that the compressor operationdoes not reach the surge limit.

Scenario 2Discharge valve closure

Scenario 1 tests confirmed the compres-sor manufacturer was correct in regard-ing the absolute necessity of a hotrecycle loop for protection during anESD. Scenario 2 then analysed the abili-ty of the hot and cold recycle loops toprotect the compressor in the event of afull closure of the compressor dischargevalve, which was ramped from a fully-open position to a fully-closed positionover a period of 15sec. Initial testsrevealed that the cold recycle valve, byitself, should be adequate to protectagainst slow disturbances, while boththe hot and cold recycle valves may berequired for protection against fast dis-turbances.

This scenario was initially run using aconventional PID controller, represent-ing the OEM controller, as described inCase 1. A purpose-built anti-surge con-troller with several advanced-controlfeatures was substituted for the conven-tional PID controller in Case 2.Case 1: discharge valve closure (DVC)with OEM controllerAs seen in Figure 5, the discharge valvebegins to close at time t = 33sec, andreaches the fully-closed position in15sec at time t = 48sec. The cold recyclevalve comes into operation 17sec afterthe discharge valve begins to close. Thetime trace of SS, the proximity-to-surgevariable, indicates the oscillations expe-rienced by the compressor. Note thatthere is virtually no safety margin, thatis, SS nearly reaches a value of 1 (surge).The compressor operating point is final-ly stabilised with the cold recycle valveat the 38% open position.

Both hot and cold recycle controllerswere configured with a surge-controlmargin of 5.5% of suction flow.Although the conventional controllerwas able to prevent surge, it had to betuned very aggressively (Kp = 8, Kr = 4).This lends itself to potential instabilityin case of faster disturbances or whenthe compressor speed is changed to con-trol a process parameter at the sametime that the recycle valve is modulated.Both of these occurrences are very com-mon, especially during start-up and par-tial-load operation.

If the likelihood of oscillations is tobe reduced, the controller will eitherhave to be de-tuned, which will result incompressor surge, or the surge controlmargin will have to be increased, whichreduces the efficient operating envelopeof the compressor and will result in

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Figure 3 Operating parameter trends on ESD, hot recycle only

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Figure 4 Operating parameter trends on ESD, hot and cold recycle

excessive wasteful recycling during start-up and partial-load operation.

Another detrimental aspect of con-ventional PID algorithms is that con-stant tuning parameters are employedfor disturbances in both directions.Thus, there is always a possibility thatthe algorithm will be unable to provideappropriate control response for a new,untested large disturbance. This prob-lem reveals itself later during the suctionvalve closure tests.Case 2: DVC with purpose-built anti-surgeAs seen in Figure 6, the compressor dis-charge valve (SDV80051) begins to closeat time t = 19sec. The valve reaches thefully-closed position in 15sec at time t =34sec. The cold recycle valve comes intooperation at time t = 34sec, approxi-mately 15sec after the discharge valvebegins to close, 2sec faster than in Case1. The cold-recycle anti-surge controllereffectively turns the operating pointback toward the stable zone before itreaches the surge limit. SS reaches a max-

imum value of 0.94, and the compressordoes not experience any surge cycles.

This test clearly shows the effective-ness of the advanced, surge-preventionalgorithms of the purpose-built con-troller when compared to a convention-al PID controller.

In order to understand why the pur-pose-built anti-surge controller is moreeffective than the OEM controller, it isnecessary to look at the action of thecontrol algorithms employed. The OEM(PID) controller acts on the errorbetween the compressors operatingpoint and the surge control line, and noaction is taken until the operating pointmoves to the left of the control line. Thepurpose-built controller uses a combina-tion of closed-loop, open-loop, andanticipatory control responses, eachwith different tuning and controlpoints. These algorithms allow the con-troller to begin moving the recycle valveeven before the operating point crossesthe surge control line.

The response of the previously men-

tioned algorithms can be seen by look-ing at the detailed movements of therecycle valves. The anticipatory controlalgorithms begin to open the cold recy-cle valve at approximately t = 35sec,based on the rate-of-change of the com-pressors operating point measured as afunction of SS. Note that the openingbegins when SS = 0.8, even though thecompressor operation is to the right ofthe surge control line (SS = 0.9 on thesurge control line and SS = 1.0 on thesurge limit line).

With larger and/or faster distur-bances, it may be necessary to manipu-late both the hot and cold recycle valvessimultaneously. In such cases, the hotrecycle valve will open for a very shorttime, arresting the movement of theoperating point towards the surge limit.Thus, the control strategy employed isto use the cold recycle loop exclusivelyto protect against small disturbancesand the hot recycle loop to protectagainst large disturbances. Once the dis-turbance is sufficiently arrested, the hotrecycle valve will close to prevent over-heating the gas at the compressor suc-tion, while the recycle flow required forsafe operation is slowly transferred tothe cold recycle loop.

The two controllers (the hot-recycleanti-surge controller and the cold-recy-cle anti-surge controller) collaborate toprovide effective and efficient anti-surgeprotection.

Scenario 3Suction valve closure

Scenario 3 analysed the effectiveness ofthe cold recycle loop in protecting thecompressor in the event of a full closureof the compressor suction valve, whichwas ramped from a fully-open positionto a fully-closed position over a periodof 25sec. This is analogous to a loss-of-load condition.

Scenario 3 was run using both a con-ventional PID controller model(described in Case 1) and a purpose-builtanti-surge controller (described in Case2). The tuning constants for the twocontrollers are identical to those usedfor Scenario 2.Case 1: Suction valve closure (SVC)(SDV80002) with OEM controllerAs seen in Figure 7, the suction valve(SDV80002) begins to close at time t =29sec. The valve reaches the fully-closedposition in 25sec at time t = 54sec. Thecold recycle valve comes into operationat time t = 51sec, approximately 22secafter the suction valve begins to close.

The time trace of SS, the proximity-to-surge variable, indicates the com-pressors approach to its surge limit.The compressor enters the unstablesurge region at time t = 53sec (SS >1)and oscillates back and forth while the

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PID controller modulates the cold recy-cle valve in an effort to move the com-pressor back into the stable operatingzone.

The tuning constants were the sameas those used in Scenario 2, Case 1.While these settings provided an ade-quate response for Scenario 2, theyresult in very aggressive valve move-ment for this scenario and cause unsta-ble and unsafe operation. In fact, thecompressor operation is initiallypushed into the surge region becauseof the valve oscillation seen at time t =50 to 53sec. The only alternative avail-able to the operator is to manuallyposition the cold recycle valve suchthat surging stops.

If the algorithm is de-tuned to suitthis scenario, then it becomes ineffec-tive for Scenario 2. This is the dilemmapresented by a constant-gain PID algo-rithm. While it can be effectively tunedfor one scenario, it may prove to betotally ineffective for all others.Case 2: SVC (SDV80002) with purpose-built anti-surgeAs seen in Figure 8, the compressor suc-tion valve (SDV80002) begins to close attime t = 23sec. The valve reaches the

fully closed position in 25sec at time t =48sec. The cold recycle valve comes intooperation at time t = 43sec, approxi-mately 20sec after the suction valvebegins to close; the hot recycle valvedoes not open at all. The maximumvalue of SS is 0.93, indicating that thecompressor operating point remainedwell within the stable operating zone.

The advanced algorithms of the pur-pose-built controller are able to preventcompressor surge effectively, withoutany control-loop oscillations. In fact,the cold recycle valve begins to open through a combination of anticipatoryand open-loop algorithms before theoperating point even reaches the surgecontrol line.

ConclusionIn the situation described in this article,where the end user was faced with try-ing to reconcile the differing recom-mendations of the two manufacturers,high-fidelity simulation was useful inevaluating the effectiveness of eachcompeting solution. Clearly, the com-pressor manufacturers recommendationto include a hot recycle loop was appro-priate, a result that is not evident with-

out simulation. Simulation was alsoindispensable in identifying a problemthe end-user was unaware of that is,the inability of the OEM controller toadequately protect the compressor fromsurge in the case of all potential distur-bances. Simulation also helped in iden-tifying the many benefits of apurpose-built anti-surge controller:

The anticipatory algorithms of thepurpose-built controller were able toprotect the compressor in the case ofsuction valve and discharge valve clo-sures. In fact, purpose-built controllerscan tailor their control response to thesize and speed of the disturbance andcan effectively protect against any dis-turbance. The OEM approach of open-ing the recycle valves only when theoperating point reaches the surge con-trol line was neither sufficient nor effec-tive in preventing surge over a widerange of disturbances.

In the case of disturbances larger thanthose tested, coordination between hotand cold recycle loops, as provided bythe purposed-built anti-surge con-trollers, will be more effective both insurge prevention and in minimisingheating of the compressor suction dueto hot recycle.

The use of simulation is obviously notalways required or appropriate, but incases where design questions exist, orthe effectiveness of different controllersis being evaluated, high-fidelity simula-tion can be an effective and useful tool.

AcknowledgementsThe authors wish to thank Gregory Lyulkoand Al Sheldon, of Compressor ControlsCorporation, for their assistance in thepreparation of this article.

Nikhil M Dukle is manager of simulationand sales support with Compressor Controls Corporation, Des Moines, Iowa,USA. He designs, tests, and commissionsintegrated control systems and turbomachinery simulation systems forinternal and external use, and providessales support. He has a BS in mechanicalengineering from the University of Poona,India, and an MS in mechanical engineering from the University of Houston, Texas. E-mail: ndukle@cccglobal.com Krishnan Narayanan is director of application technology development atCompressor Controls Corporation. Hedevelops and implements new algorithmsfor turbo-machinery control systems. Hehas a BS in technology in mechanical engineering from the Indian Institute ofTechnology, Madras, an MS in mechanicalengineering from the University of Minnesota, Minneapolis, and an MBAfrom Iowa State University in Ames, Iowa.

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Figure 7 Operating parameter trends for suction valve closure, OEM controller

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