proc imeche part e: dry gas sealing solutions for high

Technical Note

Dry gas sealing solutions for high-pressurecompressor applications

Ken Tacon1, Colin Twiss2, Vugar Mammadov1 andFarhad Aslan-Zada3

Abstract

Although the use of dry gas seals (DGSs) in process centrifugal compressors has become an industry standard since the

1980s end users are still facing operational issues related to premature failures which incur significant maintenance and

operational costs. It has been recognized that the majority of the failures are due to a lack of properly conditioned seal

gas and this has driven more recent development of Gas Conditioning Units (GCUs) to deliver appropriate quantities of

clean, superheated gas to the seal at the required pressure. However, there are a large number of operating units where

retrofit of a GCU represents significant cost and engineering challenge. The aim of this paper is to share some recent

experience of a DGS reliability improvement program that does not involve retrofitting a GCU. A basic introduction to

DGS operating principles and an overview of DGS design and control issues is also included with a focus on tandem seals

in high-pressure natural gas compression applications.

Keywords

Centrifugal compressors, dry gas seals, separations seals, seal gas, buffer gas

Date received: 23 November 2012; accepted: 13 March 2013

Background

Dry gas seals (DGSs) are used in centrifugal compres-sors to prevent loss of containment of the process gasto atmosphere. Large-scale application of DGSsbegan in the late 1970s and today DGSs havebecome the industry standard for turbomachinesused in pipelines, offshore applications, refineries,petrochemical, and gas processing plants.

In late 1980s, the American Petroleum Institute(API) recognized DGSs for the first time (the fifth edi-tion of Standard 617 for centrifugal compressors). Thisrecognition played a key role in the acceptance ofDGSswithin the industry and resulted in accelerated growthof DGS applications in centrifugal compressors.

DGSs have numerous quoted advantages over wetseals. These include lower operating cost, reducedpower consumption, reduced leakage rates, and nooil contamination of the process gas. Generally,their requirements for reliable operation are a tightlycontrolled supply of clean dry seal gas for a maximumshaft speeds around 100m/s and maximum seal cart-ridge temperatures of around 200�C. Regardless ofhow proven and reliable the technology is, each sitelocation is unique and can introduce factors whichadversely impact the advantages offered by DGSsand can result in lower than expected reliability.

This paper explores the basic operating principlesof DGS and presents a general overview of DGSdesign and control issues with a focus on tandemseals in high-pressure (HP) natural gas compressionapplications. Operating experience suggests that sig-nificant operational issues are unavoidable if the ori-ginal seal gas support system is not designed carefully.It is the authors’ opinion that the inclusion of a sealGas Conditioning Unit (GCU) during detailed designphase of a project offers reliable and trouble free oper-ation of tandem DGS in HP centrifugal compressionapplication. While it seems to be a relatively easy taskof specifying Seal Gas Conditioning Skid as part ofstandard scope of supply, the authors indicate thatprecise knowledge of seal working environment isparamount in understanding true reliability values inorder to tackle operational issues.

Proc IMechE Part E:

J Process Mechanical Engineering

2014, Vol. 228(3) 238–252

! IMechE 2013

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DOI: 10.1177/0954408913489546

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1BP Exploration & Production Company Limited, London, UK2Azeri Operations, BP Exploration Caspian Sea Ltd, Baku, Azerbaijan3Department of Mechanical Engineering, Azerbaijan Technical

University, Baku, Azerbaijan

Corresponding author:

Farhad Aslan-Zada, Department of Mechanical Engineering, Azerbaijan

Technical University, H. Javid ave. 25, Baku AZ1073, Azerbaijan.

Email: [email protected]

at PENNSYLVANIA STATE UNIV on September 13, 2016pie.sagepub.comDownloaded from

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Seal design and operating principles

In normal operation a dynamic seal is formed betweenthe stationary primary ring (the ‘‘face’’) and the rotat-ing mating ring (the ‘‘seat’’) float on a thin film of gasgenerated by the grooves recessed onto the surfacemating ring (the ‘‘seat’’). Grooves recessed into theseat create a hydrodynamic effect which generates athin film of gas which separates the two faces andallows their relative rotation without contact friction.Typically the separation between face and seat isapproximately 5 mm. When stationary the sealingfaces are held closed by a combination of spring andhydrostatic forces.1–3

The precise geometry of the sealing faces and thegeometry of the grooves is complex and crucial to theperformance of the DGS. The grooves can bedesigned for either uni-directional or bi-directionaloperation of the DGS as required. Figure 1 showssome typical groove designs.

Maximum pressure across sealing faces is achievedat point P3 and drops linearly across sealing dam,thus reducing the pressure differential and leakage(Figure 1). Generally small gap results in low leakage.Usually primary ring is convex shape and mating rigis grooved. Therefore in DGS faces are not flat.

The sealing dam is the area from the inner diameterof the spiral groove to the inside diameter of the faceof the opposing face. Sealing dam generates liftingeffect at the inner diameter, which provides a resist-ance to the gas flow resulting in a pressure increase.All grooved seals operate by using the principles offluid mechanics. At shutdown, hydrostatic forcesalong with the spring load act to close the faces.Seal balance and the design of the grooves preventdamage to the faces at start up and shutdown priorto separation.4–9

There are several designs to achieve the non-con-tacting feature of DGS (Figure 2a and b):

. spiral grooves

. T-slots

. U-grooves

. Stepped grooves

. Wavy faces

Three-dimensional aerodynamic gas grooves(Figure 2c and d) believed to sustain reliable face liftand self-cleaning feature ^through a progressivegroove bottom. Variable groove depth also influencesgas film stability.

Performance comparison chart between bi- anduni-directional seals is presented in Figure 3. At opti-mum seal gap uni-directional seals seem to demon-strate superior film stiffness.

DGS arrangements

There are three common DGS arrangements: single,double, and tandem which are described briefly below.

Single seals. Single seals are used mostly for non-toxicand non-explosive applications for sealing pressuresup to 60 Bar (Figure 4).

Double seals. Double or back to back DGS systemscan be used for toxic and explosive applicationswhen sealing pressures are below 25 Bar (Figure 5).In double seal arrangements it is common to usenitrogen as the sealing gas. Inboard seal leakageflow path is usually to the suction of the compressorvia the inboard process labyrinth. Outboard seal leak-age is vented to the safe area. Double seals believed tohave significantly less seal gas consumption comparedto tandem arrangement. The use of clean and drynitrogen makes their operating environment free ofcontamination; hence, this explains their higher reli-ability compared to tandem seals. One of the criticalfactors in the application of double seals is theirdynamic pressure limitation which results from smallseal gas consumption and inability to provide ade-quate seal face cooling at higher loads.

Tandem seals. Tandem arrangements are used for toxicor explosive applications at sealing pressures up to

Figure 1. Gas film pressure distribution (! John Crane Inc. All rights reserved. Used with permission).

Tacon et al. 239



300 Bar. A typical tandem seal arrangement is shownin Figure 6.

Tandem seals can be furnished with or withoutintermediate labyrinth. Best industry practice is touse an intermediate labyrinth between the primary(inboard) and secondary (outboard) seals with a

supply of nitrogen gas being supplied to the secondaryseal. In applications where product leakage is inadmis-sible this ensures that gas vented from the outboardseal will always be Nitrogen. The primary seal ventwill contain a mixture of process gas and nitrogenand will usually be vented to a closed flare system.

Figure 2. (a,b) Various face groove arrangements (a, b ! John Crane Inc. All rights reserved. Used with Permission) (c,d) Various

face groove arrangements; (c, d, – ! Eagle Burgmann, All rights reserved. Used with Permission) (e) Various face groove arrange-

ments; (e – ! Flowserve Corp. All rights reserved. Used with Permission).

240 Proc IMechE Part E: J Process Mechanical Engineering 228(3)



Tandem seals for HP applications use L-shapedcarrier behind primary ring which allows using poly-mer ring as secondary sealing elements instead ofelastomer O-rings (Figure 6).

Secondary sealing elements play an important rolein the overall seal performance and reliability. Forextreme pressure applications elastomers are replacedby advanced U-cup polymers. Elastomer free seals areused for higher pressure/temperature sealing applica-tions (>20 Bar/<200�C).

Leakage control and monitoring

Usually in multistage compression tandem sealarrangements confirming to API 614 the seal gas issupplied from the compressor casing at the dischargeof the first or second stage impeller. Generally it isaccepted that the seal gas source must be availableat a pressure slightly higher than sealing pressure cov-ering entire operating range from steady state runningto transient conditions such as startup, idle, and whenrotor is at rest position. Maintaining 3.0 BarG differ-ential pressure between seal gas source and requiredsealing pressure can provide sufficient safety margin.This ensures that the seal gas will be at pressure higherthan the compressor suction pressure in order to havepositive flow of the seal gas across inner seal and

Figure 4. Single seal arrangement (! John Crane Inc. All rights reserved. Used with permission).

Figure 5. Double seal arrangement (! John Crane Inc. All rights reserved. Used with permission).

Figure 3. Bi-directional vs. Uni-directional groove perform-

ance (! John Crane Inc. All rights reserved. Used with per-

mission). Speed, 13,300 rpm, seal temperature, 200�C; seal

diameter, 162 mm.

Tacon et al. 241



inboard labyrinths-back to the process side. Seal gassupply velocity across inboard labyrinths should bemaintained between 5m/s and 15m/s during normaloperation. For the safe operation of any DGS it isimportant that the supply pressure to be kept higheror at least equal to the pressure downstream of theDGS.

In order to operate efficiently any DGS system willrequire some auxiliary components which should bedesigned to provide sealing faces with clean and drygas at narrow process conditions. DGS basic auxiliarysystem should include following key components:

. Filter to remove particles.

. Coalescer to remove fluids.

. Seal housing lube oil jacket heating to avoid con-densation at cold starts (optional).

. Pressure control valve to adjust and controlrequired pressure.

. Pressure, flow, and temperature measuring devices.

. Booster pump to increase seal gas pressure(optional).

. Trace heating and lagging to ensure adequate heatconservation.

. Seal gas heater to avoid condensation (optional).

Although DGS does not require complicated ancil-lary support but in most cases it requires reliable fil-tration and careful control and monitoring of flowsand pressures from various locations. Primary role ofthe seal control system is to control the environmentof the gas seal, continuously monitor its performanceand initiate alarms and shutdowns should inadmis-sible operating condition been detected.

We will focus more on a tandem gas seal with inter-mediate labyrinth because of their dominance in theindustry involving HP hydrocarbon (HC) gas com-pression. Normally in HP applications leakage fromthe inboard and outboard seals are continuouslymonitored and any malfunction will result in alarm

and immediate shutdown if condition worsens inorder contain safely the process gas within themachine.

In order to provide the inboard gas seals with dryand clean gas the process gas extracted from the com-pressor first stage is re-routed via coalescing element(usually duplex 3 mm range) to the inboard stage inter-face. Flow of the filtered gas is controlled. Generallythere are two accepted methods of controlling thesupply of the seal gas. These are differential pressurecontrol and flow control. In differential pressure con-trol approach seal gas pressure is regulated to prede-fined value (1–1.5 BarG) above the reference sealingpressure via differential pressure control valve.Simplified schematic of differential pressure controlarrangement is presented in Figure 7.10

If flow control philosophy is applied, the supply ofseal gas is regulated by flow through a dedicated ori-fice placed after filter element upstream of each seal(Figure 8). In more automated solutions flow is con-trolled using flow control valve monitoring the flow ofseal gas through an orifice in seal supply line.Automated control is deemed to be preferred optionfrom reliability point of view.

The major designation of the controlled seal gassupply function is to ensure a positive flow of cleancondition seal gas across the inboard seal and processside labyrinth at velocities above 5m/s. One of theimportant aspects is related to the fact that seal gasvelocity will vary with labyrinth clearance and systemshould be capable of maintaining required velocity atall scenarios before compressor will be taken outof service for major maintenance. In order to accountfor a normal operating wear of the inner labyrinthsystem must be designed for higher velocities (i.e.,10–15m/s).

Unfortunately conservative approach in designingseal gas control systems results in end user coveringcosts for recycled flow impacting machine efficiencyand resulting in more expensive seal support system

Figure 6. Tandem seal arrangement (! John Crane Inc. All rights reserved. Used with permission).




components and oversized separation equipment andGCU. In order to overcome these problems majorityof compressor and seal vendors recommend flow con-trol arrangement which is capable to reduce seal gasconsumption and maintain the required velocityacross inner labyrinth.

Advantages of the flow control over differentialpressure control systems were clearly demonstratedbased on the numerical investigations using 25moleweight HC mixture (Figure 9).10 It is evident that atthe sealing pressure below 2900 psi (200 Bar) the seal-ing gas mass flow for differential pressure controlarrangement is much higher than that of the flow con-trol arrangement. With change in process conditionssealing pressure at which both types of control

arrangements demonstrate equivalent seal gas massflows is inversely proportional to the change in themole weight.10 One of other advantages of flow con-trol arrangement is elimination of the requirement forreference pressure measurement which potentiallyrepresents added complication to the controlschematic.2

Inner seals

Integrity of the inner gas seals is usually monitored byleakage indicating instrumentation. Leakage from theinner seals is routed to a safe flare. Normal dynamicleakage from inner seals can be within 60–100 stdL/min depending on seal size and duty. Inboard seal

Figure 8. Flow control arrangement.

Figure 7. Differential pressure control arrangement.

Tacon et al. 243



static leakage rates are usually two to three times lessthan dynamic.

Primary vent backpressure is tightly controlled by‘‘Inner Leakage Backpressure Control Valve’’ with adefined set point and alarmed for low and high values.If pressure exceeds predefined limit unit will betripped. Condition of the inner seal is maintainedvia pressure indicator transmitters with differentvoting architecture. Usually 2oo2 voting sequence ismore common in the industry. In case of catastrophicfailure of the inner seals bursting disk ruptures andrelieves the excess of process gas to low pressure (LP)flare. Leakage flow from the inner seals is monitoredby flow indicator transmitters and high leakage alarmsignal will be annunciated to inform the operator ifpressure exceeds a defined limit.

Outer seals

Outer seals operate on nitrogen buffer gas. Supply ofdry and clean buffer gas is controlled by pressure con-trol valve and restriction orifice. Flow of the buffergas is monitored via flow indicator transmitters withlow- and high-flow alarm. Condition of the outer sealis monitored via pressure indicator transmitters withhigh alarm (compressor outer seal leakage pressurelimit high) and trip set at pre-defined pressure (com-pressor outer seal leakage pressure limit high–high).In case of trip initiated due to high HP limit from theouter seal leakage compressor blow down is normally

initiated. Outer seal leakage pressure indicator trans-mitters are usually built-in with 2oo2 voting sequence.Bursting disks are utilized to protect against moment-ary pressure increase in the buffer gas supply line tothe outer seals.

Back in 2007 Siemens Turbomachinery has circu-lated a safety bulletin concerning centrifugal compres-sors equipped with Tandem DGSs. In applicationswhere Outboard Seal is not monitored, there was asignificant risk of undetected gas release in the eventof loss of both seals (inboard seal and Outboard seal)due to compressor shutdown function becoming inac-tive. Two years later API developed a new documentspecifically for compressor DGSs (API 692) whichaddressed the problem of outboard seal monitoring.

Bearing housing buffer gas supply

In order to separate DGS Cartridge from the bearinghousing and prevent from oil penetrating into gas sealsBearing Separation Gas should be continuously sup-plied at specified pressure and flow. It also preventsfrom build-up of an explosive gas mixture in case ofa seal malfunction or failure. Buffer gas supply to sep-aration seal is controlled via ‘‘Bearing Housing BufferGas Pressure Control Valve’’. It is quite common to setbearing housing buffer gas Pressure Control Valve(PCV) in the range of 50–150mBar. Two pressure indi-cator transmitters monitor the pressure of the buffergas supply and annunciate a low alarm if ‘‘Bearing

Figure 9. Gas flow across inner labyrinth seal (25 mole weight gas).10




Housing Nitrogen Pressure Limit Low’’ at supply pres-sure 420mBar with trip at 415mBar. If compressortripped on ‘‘Bearing Housing Nitrogen Pressure LimitLow-Low’’ compressor blow down is normallyinitiated. Pressure indicator transmitters monitoringbearing housing buffer gas supply pressure are con-nected by 2oo2 voting sequence.

For tandem seals used in HP applications againstHC process gas preferable buffer gas is Nitrogenfiltered via 3 mm filters. The buffer gas flow acrossthe intermediate labyrinth at normal operating condi-tions should usually be maintained close to 5m/s.Injection of the buffer gas at pressure of approxi-mately 2 Bar higher than the maximum flare backpressure serves to prevent process gas leakage to theoutside and also keeps the outer seal in a dry andclean environment.

Separation seals are required to carry two funda-mental functions – preventing migration of lube oilfrom bearing cavity into compressor seals and pro-tecting lube oil system from entry of hazardous pro-cess gas (Figure 10). These seals can be of various

design, among which non-contacting clearance bushdesign (JC-93FR) or Burgmann CSR type receivesmore attention and wide spread.

With compressor moving from static to dynamiccondition the clearance gap between sleeve andcarbon busing reduces (Figure 11). This leads to thereduction of the separation gas consumption.Consumption in hot condition for JC-93FR sealscan therefore be 20 times less compared to coldcondition.

Another concept (CSR-Burgmann) is based ongenerating aerodynamically a radial lifting force bypockets worked into inner face. This results in main-taining of a concentric gap of a few micrometers.In standstill mode the seal and the shaft are in con-tact with each other and have a good static sealingeffect.

In both arrangements inside diameter design of acarbon ring incorporates recessed pockets which aimto create a sealing dam at LP side and forces carbonring to float on the gas cushion reducing likelihood ofunwanted sleeve contact.

Figure 10. Non-contacting clearance bush separation seal (! John Crane Inc. All rights reserved. Used with permission).

Figure 11. Comparative sketch of separation seal clearance for cold and hot conditions (! John Crane Inc. All rights reserved. Used

with permission).

Tacon et al. 245



Field operating problems and reliabilityissues. Learning from experience

It is widely accepted that DGS reliability is closelylinked to the quality of the gas being sealed. DGSs oper-ate reliably for long periods when sealing gas is dry andclean. It has been reported earlier that in the presence ofclean and dry seal gas, DGS demonstrate a negligiblewear even after 10 years of operation.11 However, if thegas is dirty and/or wet the DGS performance is put atrisk. There are number of precision parts in any DGSarrangement requiring clean and dry gas to maintainnon-contacting clearances. Hence, the role of the sealsupport system is to ensure that DGS operates withclean dry gas under all operating conditions includingsteady state operation, start-up, shut down, and settleout. This requires two conditions to be met:

1. The seal support system conditions the seal gas sothat the gas is clean and dry.

2. There is a continuous positive flow from the sealsupport system onto the sealing interface andacross the process side labyrinth at all times.

Should either of these conditions not be met at anytime then there is a risk of DGS contamination andfailure. Usually during settle out and recycle the dif-ferential pressure may be not enough to overcome theresistance of the coalescer and piping and thisincreases the likelihood of reverse flow when the pro-cess gas enters the seal chamber. Hence, there are twoprimary means of the contaminants to find their wayto the sealing interface. Contamination can be depos-ited on the sealing interface during transient condi-tions from the dirty process gas as a result of thereversed flow from process side to the seal chamberand due to isenthalpic expansion of the filtered sealgas across inboard seal faces.

In case of isentalpic expansion potential condensa-tion of liquid HC across sealing interface and tem-perature reduction due to Joule–Thompson effect

may cause two major problems – sticky seal andfrozen seal. Both conditions have increased likelihoodat start-up. During compressor operation the liquidends flash across the faces and re-deposit at theinboard seal vent line to flare (Figure 12). This con-dition may continue unnoticed till compressor isstopped for operational reasons. Problems appearduring re-start. Usually seat and face stick or freezetogether and get damaged. In case of sticky seal dis-turbance of the fluid film gap by introduction of liquidcan result in upsetting fine design balances and asconsequence a hard contact of the rotating and sta-tionary sealing faces leading to damage due to exces-sive thermo-mechanical stresses.

Probability of dirty process gas migrating acrossinner labyrinth seal when compressor differentialhead is not sufficient to overcome the piping andfilter resistance is very likely scenario when there isno means of forced circulation. This is more relevantfor configurations with variable speed drive as it takestime before sufficient discharge pressure is built inorder to maintain a positive circulation of conditionedseal gas.

The problem of poor quality buffer gas supply hasbeen known for a long time and nowadays seal ven-dors offer customized buffer gas conditioning skidswhich are designed to clean and heat the sealing gasin order to move the liquid drop out point away fromthe operating envelop. These systems very often inte-grate a pressure booster for maintaining positive flowof clean gas across the seal cavity and back to com-pressor suction and a gas heater. Air operated seal gaspressure booster is normally operated during pressur-ization sequence and at settle-out conditions. Thishelps to keep inboard seal faces clean as the flow ofthe clean seal gas is always from seal chamber to theprocess side. The maximum pressure boost is equal tothe drive piston area divided by the boost piston areamultiplied by the pressure feeding the drive cylinder.Extensive field experience shows that the absence ofthe booster pump does not fully address transientconditions as there is no enough differential pressureavailable to maintain the positive flow across theinboard seal. Seal gas pressure booster is an ultimatesolution to the problem of keeping dirty process gasaway from entering seal chamber at various operatingconditions.

Another important factor which has to be takeninto account is seal gas composition. Seal gas contain-ing C6 and heavier fractions and water vapor repre-sents high risk to seal faces as these heavy HCfractions and water tend to condense as gas flowsthrough various components of the seal gassystem.12,13 Filters, valves, and other restrictionscause seal gas pressure drop. Seal faces create mostfavorable conditions for Joule–Thompson effect. Inorder to avoid this happening sufficient superheatshould be provided for moving heavier ends andwater vapor condensation away from the actual

Figure 12. Inboard seal vent line.




operating envelop. API 614 specifies minimum 20Fsafety margin above seal gas dew point in order toavoid condensation. In real application 40F andabove would give more assurance that condensationprocess on seal faces is potentially inhibited.

As next step in the evolution of the DGS supportsystems and various requests from the compressoroperators majority of DGS vendors developed GCUpanels intended to deliver clean superheated gas atpressures and flows required for a reliable long-termoperation of compressor seals at various operatingscenarios. Unlike conventional DGS panels featuringcoalescer elements GCU modules incorporategas/liquid knock-out drum vessels, seal gas heater,coalescer elements, and pressure booster compressor(Figure 13). It has been proposed that the applicationof these panels eliminate the risk of liquid drop outfrom the seal gas stream and thus ensures trouble freeoperation at all operating conditions including exten-sive settle out. Due to the established fact that themajority of seal failures are due to seal face contam-ination initial estimates suggested that application ofGCU panels could extend mean time between failuresto five years.

An offshore oil and gas production platform in theCaspian Sea was experiencing repetitive seal failureswith both LP and HP (high pressure) gas injectioncompressor seals. There are four gas injection trainson the platform and each train is configured with LPand HP cases driven by 27MW Gas Turbine. Each

train is capable of delivering 250MMSCFD of naturalgas increasing its pressure from 60 BarG to 386 BarG.LP compressor was equipped with John CraneT28XP/XP Tandem seals with dynamic pressurelimit of 150 BarG and HP compressor was equippedwith the same seal with higher dynamic rating of 240BarG. Seals were equipped with interstage labyrinths.Simplified LP seal cross-sectional schematic is repre-sented in Figure 14. Between 2006 and 2009 therewere 12 failures. Initial analysis suggested that over75% of all failures occurred during winter months.Original DGS support system arrangement had noheater and pressure booster pump. Seal gas sourcefor both LP and HP case seals was drawn from therelevant case first impeller discharge and was suppliedto primary seal interface via 3 mm coalescer andorifice.

Size of the orifice plate was configured in order tomaintain certain mass flow and velocity across sealfaces. Leakage was routed to primary vent withback pressure control and constant monitoring withintegrated alarm and shutdown functions. In order tomonitor primary seal condition primary vent wasequipped with flow meter with consequent alarmand safe shutdown function. Spring loaded checkvalves at the connections to the flare header wereinstalled in order to avoid any overpressure eventby back flow of flare gas into the primary vent cham-ber. In case of primary seal catastrophic failure pri-mary vent line was equipped with rupture discs to

Figure 13. Dry gas seal gas conditioning unit (! John Crane Inc. All rights reserved. Used with permission).

Tacon et al. 247



relieve the excess pressure to LP flare system.Secondary seal sealing gas was sourced from platformnitrogen package under constant pressure of 5 BarG.Primary seal vent line was also equipped with PressureSafety Valve (PSV) upstream of the back pressurecontrol valve set at 6 BarG. Secondary seal buffergas flows to secondary sealing interface and entersprimary vent via interstage labyrinth. Therefore,buffer gas is supplied under pressure slightlyhigher than primary vent back pressure. Secondaryseal supply line is also equipped with rupture discsin order to avoid build-up of excessive pressure incase of primary seal catastrophic failure. Leakagefrom secondary seal together with large portion ofnitrogen supplied to oil separation seals flows tosafe area.

Seal reliability improvement solution

Analysis of the damaged seal components was sug-gesting a common failure mechanism for vast

majority of the failed seals. There was a strong evi-dence of inboard faces being stuck together at start-up. High-load transmission during start-up was obvi-ous from the damaged drive notches on the seats.Sticky seal faces get overheated fairly quickly duringstart up as no face separation or insufficient separ-ation takes places. Presence of excessive amount ofliquid phase on the primary sealing surface leads toexcessive thermo-mechanical stresses and facedestruction (Figure 15). It was proposed that key fac-tors contributing to seal faces sticking together werecondensation of heavy HC s and ice build-up due toJoule–Thompson effect.

When we look closely to operating conditions it isobvious that there is excessive pressure drop acrossseal faces at standstill periods. HP compressorinboard stage sealing gas expands from 224 BarG to4 BarG. During isenthalpic expansion condensationof heavy ends is most likely to occur. Based onsampled process gas composition phase diagramswere built using ASPEN HYSIS simulation model.

Figure 14. LP compressor dry gas seal (! John Crane Inc. All rights reserved. Used with permission).

Figure 15. Catastrophic seal failure-damaged seal faces.




The model demonstrates that likelihood of condensa-tion exists within the operating envelop (Figure 16).Proprietary CSTEDY software was used to constructa diagram of seal face temperature distribution duringsteady state running conditions as input parameters(Figure 17). It is evident that there is negligible tem-perature rise across seal faces during operation.

As it was indicated earlier in this particulararrangement sealing source for primary seal wasextracted after the first wheel and was filtered to3 mm through coalescing elements. Seal cartridge wasalso equipped with heating jacket. Main source ofheating was sealing oil pre-heated by electricalheater and circulated via circulation pumps before

start-up. Jacket oil heating was designed to add suffi-cient heat to the seal faces before start-up, so no con-densation of primary sealing gas can take place.Sufficient superheat would have helped to inhibit thecondensation as gas expands throughout various sec-tions of the system. However, this was not consideredduring design stage and using oil jacket heating wasconsidered to be sufficient for reliable operation.

Numbers of modifications were carried out inorder to improve the efficiency of the original sealcartridge heating oil system. Seal heating oil supplyand return lines were heat traced and seal oil tank wasinsulated. These arrangements were considered to besufficient to cure the problem. Field experience, how-ever, demonstrated that such arrangement is not suf-ficient to provide reliable seal operation at cold winterseason and multiple seal failures were causing signifi-cant operational and maintenance problems.

Recommendations on using seal gas electric heaterand pressure amplifier as stand-alone retrofits or usingthoroughly engineered GCU to provide clean and drygas at required pressure and temperature could beconsidered as common solution to the well-knownproblem of seal face contamination. When not con-sidered originally their installation may require exten-sive engineering effort. Conventional DGS panels areusually configured with filter/coalescer element to sep-arate the liquid from the seal gas which is not suffi-cient during cold ambient conditions. In thisparticular case few recommendations were taken for-ward which offered a step change in addressing issuesarising from seal gas quality problem. This helped toavoid drop out of liquids on the primary sealing inter-face prior to start-up.

Figure 16. HP compressor inner seal gas phase diagram.

Figure 17. Inner face temperature distribution (! John

Crane Inc. All rights reserved. Used with permission).

Tacon et al. 249



First it was suggested to keep the seal oil heatingand circulation running continuously after unit shut-down. This would help to conserve the heat and keepseal faces warm which could help significantly ininhibiting HC drop out. Second compressor startsequence was modified and some steps resulting inkeeping the case under settle out were removed.New modified sequence considered early torque trans-mission to compressor rotor and elimination of grad-ual pressurization of LP and HP cases with rotor atrest. Forced recycle for both LP and HP stages startedfrom 15 BarG instead of 60 BarG. It was presumedthat heat generated during recycle will keep the sealfaces warm and move the liquid drop out away. Sealcartridge temperature was monitored during start upvia temperature probes and it was evident that notemperature reduction was observed compared to

original starting sequence (Figure 18a and b). It canbe seen from the Figure 19(a) and (b) that as soon asthe LP forced recycle starts seal cartridge tempera-tures increase sufficiently regardless of slight increasein seal gas flow. This helps to inhibit the HC drop outon the seal faces. Another key factor contributing toimproved seal working environment is relatively LPdrop between inner seal interface and primary vent atstart-up compared to the original set up. Since whencompressor is under operation the temperature of thesupplied seal gas is high enough to ensure that there isno liquid drop out may take place it was the authorsopinion that modifications to the startup sequenceshould significantly improve the reliability of theseals.

Special attention had to be paid to the potentialcooling gas starvation at recycle, but this was

Figure 18. Modified start up pressurization sequence process data strip chart.




monitored via temperature element placed at suctionside seal housing close to inboard stage. All four gasinjection trains where proposed software changeswere embedded were accurately monitored.Proposed modifications were successful solution tothe problem as no seal failures occurred since thesemodifications were implemented in February 2009. Inorder to have more supporting evidence on the impactfrom the modifications one the longest running sealswere removed at the planned maintenance outage andseal parts were subjected to visual inspection. It isobvious from the Figure 19 that operating environ-ment for these seals had improvement significantly asthere is no physical contamination present on the sealspost modification compared to seals beforemodification.

Conclusions

Ultimate HC recovery is highly dependent on efficientand reliable top side gas injection facilities. Thisimposes strict requirements on operational reliabilityand process safety of centrifugal compressors particu-larly in offshore environment. Reliability of any HPcentrifugal compressor evolves around reliable andtrouble free operation of its shaft sealing and sealsupport systems. Common DGS problems associatedwith supply of clean sealing gas and avoiding liquidcondensation on seal faces has been recognized longtime ago.

Seal and compressor designers have put significantattention on cleanliness control and careful design ofseal support systems during initial project phase.Based on the extensive field experience it is widelyaccepted in the industry these days that reliableDGS system will have GCU integrated to the sealsupport console to provide a clean and superheatedgas at volumes and pressures required during steadystate and transients.

Alternatively compressor start-up and operatingconditions should be carefully investigated for

optimizing pressurization sequence and reducing like-lihood of dirty process entry into the seal cavity andseal gas cooling due to expansion. Field experiencehas shown that this is achievable if conditions affect-ing seal working environment is carefully analyzedand understood.

Funding

This research received no specific grant from any

funding agency in the public, commercial, or not-for-profitsectors.

References

1. Aimone RJ, Forsthoffer WE and Salzmann RM. Dry gas

seal systems, best practices for design and selection,which can help prevent failures. Turbomach Int 2007;48(Jan/Feb): 20–21.

2. Stahley JS. Design, operation, and maintenanceconsiderations for improved dry gas seal reliability incentrifugal compressors. In: Proceedings of the 30th tur-

bomachinery symposium, Houston, Texas, USA, 2001,pp.203–207. USA: Turbomachinery Laboratory, TexasA&M University.

3. Stahley JS. Mechanical upgrades to improve centrifugal

compressor operation and reliability. In: Proceedingsof the 32d turbomachinery symposium, Houston, Texas,USA, 2003, pp.145–152; 161–166. USA:

Turbomachinery Laboratory, Texas A&M University.4. Stahley JS. Dry gas seals handbook. Tulsa, Oklahoma:

PennWell, 2005.

5. Zuk J and Reinkel HE. Numerical solutions for the flowand pressure fields in an idealized spiral groove pumpingseal. In: Proceedings of the 4th international conference onfluid sealing, Philadelphia, PA, 5–9 May 1969, pp.290–

301. UK: British Hydromechanics Research Association.6. Shah P. Dry gas compressor seals. In: Proceedings of

the 17th turbomachinery symposium, Houston, Texas,

USA, 1988, pp.133–139. USA: TurbomachineryLaboratory, Texas A&M University.

7. Carter DR. Application of dry gas seals on a high pres-

sure hydrogen recycle compressor. In: Proceedings of the17th turbomachinery symposium, Houston, Texas, USA,1988, pp. 3–7. USA: Turbomachinery Laboratory, Texas

A&M University.

Figure 19. Seal cartridges before (a) and after (b) compressor start-up modifications.

Tacon et al. 251



8. Cheng HS, Castelli V and Chow CY. Performance char-acteristics of spiral groove and shrouded rayleigh stepprofiles for high speed non-contacting gas seals. ASME

J Lubr Technol 1969; 91: 60–68.9. Miller BA and Green I. Semi-analytical dynamic ana-

lysis of spiral-grooved mechanical gas face seals. ASME

J Tribol 2003; 125: 404–413.10. Stahley JS. Dry gas seal design standards for centrifugal

compressor applications. In: Proceedings of the 31st tur-

bomachinery symposium, Houston, Texas, USA, 2002,pp. 145–152. USA: Turbomachinery Laboratory, TexasA&M University.

11. Power Engineering. Poor buffer gas fluid main cause of

compressor seal failure. Power Eng 2004; 108(11): 10–14.

12. McCraw JF, Bakalchuk V and Hosanna R. Design andmitigation techniques for applications with possibleliquid contamination of the sealing gas dry gas seal

system. In: Proceedings of the 37th turbomachinery sym-posium, Houston, Texas, USA, 8–11 September 2008,pp.37–43. USA: Turbomachinery Laboratory, Texas

A&M University.13. Bridon R and Lebigre O. Seal gas contamination. In:

Proceedings of the 40th turbomachinery symposium,

Houston, Texas, USA, 12–15 September 2011,pp.146–152. USA: Turbomachinery Laboratory, TexasA&M University.




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