revamping a gas compressor drive train … · experience mainlyin turbomachinery, ... this means...

Charles W. Yeiser is a PrincipalEngineer at Rotor Bearing Technology &Software, Inc. (RBTS), a software andconsulting services company, specializingin rotating systems, in Phoenixville,Pennsylvania. His responsibilities includemodeling, analyzing, and measuring theperformance of rotating systems. Mr. Yeiserhas been extensively involved in thedevelopment and implementation of

computational solution algorithms related to rotordynamics, torsionalvibration, and fluid-film/rolling element bearings. Prior to working atRBTS, he was employed by the Franklin Research Center (formerly theFranklin Institute Research Laboratories), as a mechanical engineer.His fields of expertise include torsional vibration, rolling elementbearings, rotordynamics, finite element analysis, and customizedfield test/sensor development (straingauges, proximity probes,accelerometers, data acquisition, etc.).

Mr. Yeiser has a B.A. degree (Physics) from Franklin andMarshall College, and B.S.E. and M.S.E. degrees from theUniversity of Pennsylvania. He has authored numerous technicalreports and papers for private industry and government agencies.

Volker Hütten has been the Head of theMechanical Design Department of SiemensPG I, in Duisburg, Germany, since 2001.During his more than 15 years in thiscompany, he has been responsible for themachinery dynamics of compressors andcompressor trains of order related tasks. Hehas been active in correlating analyticalresults with field data in numerous trou-bleshooting and problem diagnosis situations.

Mr. Hütten received his Diploma degree(1990) from the University of Krefeld.

Akram Ayoub is the EngineeringManager for Voith Turbo, Inc., in York,Pennsylvania. He has 17 years ofexperience mainly in turbomachinery,but also including other industries suchas marine, mining, and steel.

Mr. Ayoub is an active member ofASME and a permanent member of BetaGamma Sigma, the national honorsociety for collegiate schools of business.

Mr. Ayoub has a degree (Mechanical Engineering andIndustrial Engineering) from Texas A&M University, andan MBA degree with honors from the University of Baltimore.Mr. Ayoub has coauthored several papers on connectiontechnologies.

ABSTRACT

The successful application of synchronous motors requiresspecial design consideration to account for their speed-dependentcharacteristics, such as the large oscillating torque levels that occurduring startup. This paper focuses on the torsional design andanalysis process associated with revamping a 7000 hp synchronousmotor-driven compressor train with an 8000 hp synchronous motordriver. Torsional analyses indicated that the replacement motor’sstartup torque would produce unacceptable levels of torsionalvibration. An accurate overtorque protection slip clutch wasintegrated into the low-speed coupling assembly to limit startuppeak torque to about three times the normal operating torques.

The paper details multiple issues associated with revamping thecompressor drive train, including the selection of the replacementdriver and low-speed coupling assembly, torsional vibrationanalysis, and the mechanical operation of the controlled slip clutchmechanism. The paper also presents the acquired torsional fieldmeasurements that were used to confirm the clutch’s operation.

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REVAMPING A GAS COMPRESSOR DRIVE TRAIN FROM 7000 TO 8000 HP WITH ANEW SYNCHRONOUS MOTOR DRIVER AND A CONTROLLED SLIP CLUTCH MECHANISM

byCharles W. Yeiser

Principal Engineer

Rotor Bearing Technology & Software, Inc.

Phoenixville, Pennsylvania

Volker HüttenHead, Mechanical Design Department

Siemens AG

Duisburg, Germany

Akram AyoubEngineering Manager

Voith Turbo Inc.

York, Pennsylvania

andRobert Rheinboldt

Senior Engineer

Air Liquide Process & Construction, Inc.

Houston, Texas

INTRODUCTION

Over the last half-century, industrial gas producers have madehuge capital investments in large horsepower machinery.Synchronous motors remain a popular driver for many of theseapplications. In response to market demands, many older machineshave been rerated to increase their output. However, over time, evenequipment that has provided reliable service for decades needs to bereplaced. Major revamps of machinery are often undertaken for avariety of reasons, again, most of them market-driven. Mostimprovements are made to increase the equipment’s efficiency andreliability and, of course, to reduce maintenance.

Unlike the engineering of new equipment, the revamping of anexisting drive train often poses unique challenges. In most cases,the modified equipment has to coexist with the available plantinfrastructure. Typically, this means that the new equipment mustfit within the existing plant real estate and be able to be integratedinto the available power grid.

This paper focuses on the motor selection, torsional design, andanalysis process associated with revamping two existing 7000 hpsynchronous motor-driven industrial gas compressor trains with 8000hp synchronous motor drivers. Unlike other types of motor drivers, thesuccessful application of synchronous motors requires special designconsideration to account for their speed-dependent characteristics,such as the oscillating torque performance that occurs during startup.Typical of many revamp projects, however, the requisite engineeringstudies were still being completed after the replacement equipment hadbeen ordered. In this application, the final torsional analysisdetermined that the selected motor would create unacceptable levels oftorsional vibration during startup. To remedy this problem, severaldifferent design alternatives were evaluated via torsional vibrationanalysis. The successful solution to the problem involved modifyingthe low-speed coupling assembly to include Voith Turbo’s Safeset®autoreset coupling. This coupling, a variation of a proven torquelimiter coupling design, permits slippage when a preset thresholdtorque is reached, thereby isolating the drive train from excessivelevels of torque. In addition to computer simulations, the performanceof the coupling was verified via torsional field measurements. The twonew motors and controlled slip couplings were installed in the fourthquarter of 2003 and have been providing reliable service. A summaryof the end-user’s operational experiences with the revampedcompressor trains is also presented in this paper.

While the successful application of torque-limiting devices insteel making machinery is not new, its utility in solving torsionalvibration problems in process machinery is just now beingrecognized, particularly in revamp applications. It is the authors’hope that the entire turbomachinery community will benefit fromthis thorough review of the logistical and technical issuesassociated with this revamp project and that the “lessons learned”will be of some value for future applications.

OVERVIEW OF THE REVAMP PROJECT—THE PLANT’S PERSPECTIVE

Background

Two identical three-stage integrally geared gas compressorswere installed in 1975 at an air separation plant in the USA thatsupplied oxygen and nitrogen gas to a pipeline system. Initially,excess production was liquefied and sold to the merchantmarket. Seven thousand horsepower (7000 hp) laminated rotorsynchronous motors were used to drive the compressor trains.These motors were selected due to their relatively high efficiency,lower starting inrush current, together with a severe penalty forpower factor (kVA demand) from the local utility. Over time,pipeline sales increased and the plant was “debottlenecked” toaffect a substantial increase in its saleable product output. Theoutput from the gas compressors was increased, which in turnraised the horsepower requirement of the electric motors.

Fourteen identical compressors were later installed at other

locations. These trains were driven by 7500 hp synchronousmotors with laminated rotors. Increasing the gas flow and motorhorsepower to the shaft also “debottlenecked” these locations.

By the year 2000 the two original 7000 hp motors were nearingend of life due to being run in an overloaded condition. A studywas made of the maintenance history of the motors in thecompany system to determine the best options for the purchase ofnew electric motors. At that time, the company had over 50synchronous motors ranging in size from 1000 to 24,000 hp, withmotor speeds ranging between 327 to 1800 rpm. These evaluatedmotors included both laminated and solid pole rotor designs.

Motor and Low-Speed Drive Coupling Selection Criteria

The criteria for selection of the motors and required modificationsfor this location were driven by the following:

• The new motors would have to start across the line like the originalmotors and therefore motor inrush had to be limited in the design.

• The new motors would be selected for a requirement of minimummotor shop visits during the projected life of 30 to 35 years.

• The new motors would have to be synchronous in order to avoidpaying a high monthly penalty for power factor (kVA demand).

• The new motors would meet the current mechanical andelectrical specifications.

• The efficiency of the new motors must be at least 11/2 percenthigher than the original motor efficiency.

• The new motors would have to be able to replace both the 7000hp and 7500 hp motors with no foundation/sole plate modificationto provide a solution for the other 14 compressor trains.

• The new motors would be 8000 hp to eliminate the possibilityof overloading them under all foreseeable current and futurecompressor operating conditions.

• A torsional vibration analysis of the revamped compressor trainwith the new motor driver and selected low-speed coupling assemblymust be performed to ensure that it will provide 5000 starts.

• Elastomeric couplings were not to be used, as they have beenknown to require maintenance about every five years. Furthermore,there were also concerns regarding the shelf life of the replacementelastomeric elements.

Motor Selection and Drive Train Design

After review of the maintenance records and discussion withmotor vendors it was decided that the new motors would have solidpole rotors, as the design met the criteria established for thisrevamp. Solid pole synchronous motors from various vendors hadbeen in use since the late 1970s and had experienced a minimalnumber of problems. Experience had shown that a laminated rotorsynchronous motor would normally require a motor shop visitapproximately every seven years to do rotor repairs. Moreover, thesolid pole rotor’s relatively compact design permitted the vendor tosupply the 8000 hp replacement motor with a base plate that hadthe same footprint as the 7000 hp motors. When needed, an adapterplate could be furnished to also allow the new motors to replace theother 7500 hp motors.

Two new motors were purchased and the compressor vendor wascontracted to do a torsional study and to interface with the motorvendor during design. There was a problem with getting a viablesolution due to the increased motor horsepower and use of the solidpole rotor. The compressor vendor proposed three solutions. Howevereach solution was, in turn, withdrawn after it was determined that itfailed to either resolve potential long-term service/reliabilityconcerns, meet the 5000-start requirement, or both.

During this time period, motors were also being replaced on gascompressors at other air separation plants. One solution to a similar

PROCEEDINGS OF THE THIRTY-FIFTH TURBOMACHINERY SYMPOSIUM • 200634

problem was solved by the use of a controlled slip clutch coupling.The compressor vendor was contacted and it was suggested that asolution be tried by integrating a controlled slip clutch couplinginto the low-speed coupling assembly. A new torsional analysiswas performed and all parties accepted the proposed solution. Twocontrolled slip clutch couplings were ordered and all material wasshipped to the plant awaiting either a motor failure or a scheduledshutdown, to install the new motors and couplings. In addition, arepair kit for the controlled slip coupling with oil, charging pump,and spare parts was ordered and shipped to the plant.

The two new 8000 hp motors and controlled slip couplings wereinstalled in the fourth quarter of 2003. At that time, the compressorvendor sent personnel to the site to acquire field torsional data. Thesensor suite was configured to measure the dynamic torque on thecompressor input shaft (straingauges) and the differential slipacross the coupling (keyphasors). The measurements confirmedthat the coupling was performing as designed; however, the slipoccurred at about 80 percent of its threshold slip setting. As thisdiscrepancy did not jeopardize the torsional performance of thedrive train, no attempt was made to adjust the coupling’s sliptorque threshold. Both couplings have provided reliable servicesince commissioning.

Assessment of Revamp and Recommendations for Similar Trains

The revamp of the two air compressors with new 8000 hp motorsand controlled slip couplings is considered a success as all theplant’s criteria were met and the equipment has been in operationwith few problems. One issue did occur when one of the controlledslip couplings rotated during an “event” and sheared the brass pin.The oil drained out and the motor freewheeled until it was shutdown. The motor, compressor, and coupling system was inspectedand no problems were found. A new brass shear pin was installedand the controlled slip coupling was recharged with oil. The motorwas restarted without incident. The most likely cause was a distur-bance on the utility power. The two gas compressors are beingoperated in excess of 7000 hp.

The use of controlled slip couplings has become an acceptedpractice, as one was recently installed on a revamped 13,000 hpcompressor train in the second quarter of 2003 and another one isscheduled for installation on an 11,000 hp gas compressor train in thesecond quarter of 2005. The technology is proven and is currentlyconsidered as a viable solution for potential torsional vibrationproblems that may arise on compressor revamp projects where solidpole rotor synchronous motors are used to replace turbines, inductionmotors, or laminated rotor synchronous motors. Moreover, this devicehas been recognized for its ability to provide overload protectionduring startups, electrical short circuits, or any other electromechanicalabnormality that may be encountered during service.

GENERAL STARTING CHARACTERISTICSOF SYNCHRONOUS MOTORS

Synchronous motors as compressor drives are normally used forhigher power applications, i.e., 10 MW (13,000 hp) and more. Thiskind of motor offers the opportunity to balance the electricalpower factor of a plant. In other words, electrical idle power can becompensated.

From the mechanical point of view, the torque occurring in theair gap between rotor and stator of the motor must be thoroughlyanalyzed. Transient torques that arise in electric motors can beseparated into two categories: those that occur when the machineis running at operating speed, and those that occur when the train isbeing started. Excitations that are transient in nature, such as thestarting of a synchronous motor, any kind of electrical fault, ortorque ripples from variable speed drive systems (VSDS), mayexcite torsional natural frequencies. Some of these events cannotpossibly be avoided, e.g., excitations caused by starting motors andinteger and noninteger harmonics from VSDS. Hence, in principle,when evaluating the occurring peak torque caused by any kind of

excitation mechanism, the realistic probability of occurrence of theconsidered loading condition has to be taken into account. Thisevaluation requires a good understanding of the electrical gridcondition in the field. Nevertheless, this paper focuses on thetorsional characteristics of synchronous motors during asynchronousstarting. A discussion of this transient behavior and its influence onthe torsional dynamics of drive trains is presented herein.

Why it Occurs

The name synchronous motor is derived from the fact that therotor rotates in synchronism with the rotating magnetic fieldproduced by the three phases of electrical power supply. Hence, thespeed is directly proportional to line frequency and the number ofpole pairs. During acceleration of the train with the motor averagetorque, a strong oscillating torque is developed because of slippagebetween the rotor and stator fields.

The mean and the oscillating torques are superimposed anddescribe the main starting characteristics. The frequency of thetorque pulsation is the frequency at which the stator’s rotating fieldpasses a rotor pole. Since the stator’s magnetic field rotates at syn-chronous speed, the excitation frequency is a function of differencebetween synchronous speed and rotor speed, which is commonlyknown as the slip speed.

The synchronous motor speed and the slip frequency are definedin Equations (1) and (2) as follows:

(Definition of synchronous speed)

(Definition of slip frequency)

where:nsyn = Synchronous speed (rpm)n = Rotor speed during starting (rpm)fgrid = Electrical grid frequency (Hz)NPP = Number of pole pairs (-)

The motor starting characteristic including mean and alternatingtorque are often illustrated in a diagram like Figure 1. Torqueamplitudes are given as a function of motor speed. Curves in thediagram are related to the considered motor for this project. Theassumed starting conditions onsite are as follows:

• SCC = 195 MVA @ 4160 V Bus

Figure 1. Synchronous Motor Torque Versus Speed CharacteristicsGiven by the Manufacturer (SCC = 195 MVA @ 4160 V Bus).


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n f Nsyn grid pp / (1)

�� f f n n nslip grid syn syn � / (2)

n

When the electric power is switched on, the motor generatesfluctuating torque. This kind of transient torque at line frequencydies out rapidly. During speed increase, the frequency of theoscillating torque is double the slip frequency. This means that it isequal to twice the line frequency initially at standstill and it goesdown to zero when the motor is synchronized. Accordingly, as themotor accelerates to its synchronous speed, the torsional excitationfrequency decreases toward zero. Therefore, during such startup,the torsional system is excited at its resonant frequencies belowtwice line frequency (that means: up to 100 Hz for 50 Hz linefrequency, and up to 120 Hz for 60 Hz line frequency).

After approximately 97 percent of rated speed is reached, a directcurrent (DC) field is applied to the rotor, producing a “pull-in” torqueand the rotor “pulls-in-step.” At this point, the rotor is synchronizedwith the rotating field in the stator and is able to produce torque atsynchronous speed. If the field excitation is applied before thesynchronizing speed is reached, an additional large component ofpulsating torque is created at a single slip frequency.

The severity of the train’s torsional response as the electricalexcitation stimulates each natural frequency mainly depends on thedistribution of mass moments of inertia and stiffnesses, the modeshapes, the magnitude and frequency of excitation, location ofexcitation, and the system damping. To determine the occurringtransient torques, a thorough torsional simulation analysis of theentire train must be performed. This analysis, which should be anessential part of the standard design process of the rotatingequipment vendor, must consider the most realistic transient airgap torques for the local electrical grid condition (API Standard546, 1997). The benefit of this procedure is that well-knownpotential torsional vibration problems can be avoided.

Figure 2 presents a Campbell diagram for a typical synchronousmotor-driven train. This diagram provides a graphic overview ofthe rotor system’s torsional natural frequencies and its correspon-ding frequencies of potential excitation. In this example, the firstthree torsional natural frequencies (eigenfrequencies) are givenalong with the excitation line corresponding to double slipfrequency. It can be easily seen that motor excitations interferewith the first three natural frequencies, which are all below twicethe line frequency (120 Hz). The motor speeds at which resonancewith the natural frequencies occurs can be calculated as follows:

where:nsyn = Synchronous speed (rpm)ncritical motor speed = Motor speed at torsional resonance (rpm)fgrid = Electrical grid frequency (Hz)fTi = Torsional natural frequency, i =

1, 2, 3, . . . (Hz)

Figure 2. Campbell Diagram Including Synchronous MotorExcitation Line.

The fundamental natural frequency is of primary importance forlarge synchronous motor-driven drive trains. During resonance,this mode of vibration is capable of facilitating large amplificationsof dynamic torque that can result in a catastrophic failure of one,or more, drive components. In general, fundamental torsionalfrequencies for typical motor-driven turbocompressor trains rangebetween 8 Hz and 25 Hz. Consequently, depending upon the gridfrequency, the motor speed at which torsional resonance occurs isusually between 75 percent up to 93 percent its synchronous speed(e.g., 75 percent for 50 Hz and 93 percent for 60 Hz).

Types of Rotor Construction (Solid and Laminated)

The available types of synchronous motors can be separated intotwo designs: the solid pole rotor, and the laminated pole rotor. Interms of their characteristic mean and pulsating torque startupbehavior, these kinds of motor designs are quite similar. Butnevertheless, these designs are different in magnitude of averageand slip torque versus speed. Compared with laminated poledesigns, solid pole rotors produce a significantly higher torsionalexcitation during starting. However, the general construction of thelaminated rotor offers many design alternatives to optimize itsstarting characteristics. These alternatives are not available forsolid pole designs. As mentioned earlier, the customer elected topurchase a solid pole synchronous motor.

Torsional Vibrations and Mechanical Consequences

During startup, the motor has to generate a mean torque toaccelerate the driven system. The actual acceleration rate of thesystem mainly depends on the driver’s starting torque, the massmoments of inertia of the train components, the breakaway torque,and, finally, the load torque (includes friction and windage losses).To determine the occurring loads during startup, a transientsimulation analysis is required. This analysis takes into account thetorsional model of the entire train, including distribution of inertiasand stiffnesses, system damping, and motor characteristics.

Rapid starting with high acceleration minimizes the time periodin which the system is in a state of resonance. Under theseconditions, the resonant response will hopefully not have time tobuild up to steady-state levels. In other words, for slower motorstartups, the system stays at a resonant frequency for a longerperiod of time, allowing amplitudes and stresses to be amplified.Unfortunately, synchronous motors with a higher amount of meantorque normally show a higher magnitude of pulsation torque, too.This fact should always be taken into account. Based on theauthors’ experience in the analysis and testing of manysynchronous motor driven trains, it has been observed that motorstarts with lower levels of average torque, which also tend to havereduced levels of exciting torque, typically produce lower peaktorques during resonance. In practice, by reducing the motor’svoltage during starting, this phenomenon has been effectivelyutilized to reduce the peak torques throughout the train.Alternatively, direct-online (DOL) starting at full voltage, wherethe power supply is assumed to have an infinite short circuitcapacity (SCC), produces the “worst case” motor excitations. Thistheoretical condition produces the maximum torsional responseduring resonance and is often used in torsional startup simulationsto determine the minimum value for the number of “safe starts”that a drive can accrue during its service life.

STRATEGIES TO REDUCEPEAK TORQUES DURING STARTUP

Reduced Voltage Starting

Synchronous motors are often started using several reduced voltagemethods (e.g., autotransformer starting, reactor starting, ...). In mostcases, the electrical grid onsite requires this method of starting. Themaximum allowable starting current is typically specified to preventpower supply overload and/or excessive voltage drop.


n nffcritical motor speed synTi

grid �

§

©¨̈

·

¹¸̧1

2(3)

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With reduced voltage starting, the motor-driven train is startedon a lower voltage and switched to full voltage before, or after,synchronization. This approach to lessening the motor’s impact onthe local power supply has a fortunate consequence in that it oftenreduces the transient torsional loads experienced during startup.Therefore, as stated above, reducing excitation potential of thepulsation torque can lower the peak torques during resonance.Recognizing that the level of the motor torque is proportional to thesquare of the voltage, this method of starting also reduces theaverage torque that is used to accelerate the drive train to itssynchronous speed. Whether, or not, it is helpful to reduce thevoltage as a means to reduce the peak torsional loads should beclarified in a torsional simulation analysis. In theory, there is aminimum starting voltage where the increased torsionalamplifications facilitated by longer dwell times at resonance“balance out” the positive effects of the lower excitation torque.Based on the design of hundreds of synchronous motor-driventrains, it has been the authors’ experience that a lower level ofvoltage normally reduces the peak torque. Of course, the optimalvoltage must also satisfy the “pull out” torque and minimumacceleration requirements for the entire startup sequence.

Soft-Starter (Starting Inverter)

In addition to reduced voltage starting, the starting inverter,often referred to as a soft-starter, or load-commutated-inverter(LCI), is commonly employed to protect the power supply line. Incontrast to the asynchronous starting methods, the motor is runningwithin the synchronism from zero speed up to operating speed bystarting the synchronous motor with an inverter. Afterwards, themotor is switched so that it is drawing power directly off of theelectrical grid. Therefore, by starting the motor in a synchronousmanner, its pulsating torque with double slip frequency is totallyeliminated.

To start the motor, inverters synthesize the time-varyingelectrical voltage from a series of integer and noninteger harmonicwaveforms. Integer harmonics coincide with multiples of themotor supply frequency, whereas the noninteger harmonics aregenerated by a difference between line frequency and supplyfrequency. Owing to imperfections in the synthesized electricalsupply waveforms, these devices induce torque ripples withmagnitudes that are usually in a range of less than 1 percent and upto 20 percent of the motor’s rated torque. The resulting transientresponse should be evaluated via torsional analysis. Nevertheless,compared with DOL starting, torsional simulations havedemonstrated that the maximum torsional response can be greatlyreduced with this method of starting. It is worth noting that theinverter-induced harmonic torques disappear once the motor isswitched onto the local electrical grid.

The use of an LCI provides an elegant, but a rather expensivemeans of reducing the motor torques during startup. As mentionedearlier, the customer did not want to use an LCI.

Minimizing Torsional Stiffness of Steel Coupling

In most cases, the initial drive train design is formulated withoutthe benefit of torsional vibration analysis. Once complete, theproposed torsional system is analyzed. Afterwards, the system mayhave to be tuned if adequate separation margins do not existbetween its torsional natural frequencies and potential excitations.Furthermore, coupling tuning may be necessary to reduce theoccurring transient torques caused by synchronous motor starting.

Adjusting the torsional stiffness of the coupling is a commonmethod of tuning the torsional system. Turbocompressor trains aremostly designed to operate supercritical in terms of thefundamental torsional natural frequency related to the minimumdriver speed. It can be seen in Figure 3 that reducing a train’storsional stiffness tends to decouple the main masses of thetorsional system (as example: driver and compressor), which, inturn, leads to lower dynamic torque levels.

Figure 3. Results of the Torsional Stiffness Variation Map.

For a given distance between shaft ends (DBSE), the minimumstiffness can be achieved by using a solid coupling spacer. Materialswith adequate strength and lower shear modulus than steel can beused to minimize the stiffness. For example, fabricating thecoupling out of titanium can reduce its torsional stiffness, as ratioof the shear modulus values of titanium and steel is 0.534(Gtitanium/Gsteel = 53.4 percent).

However, reducing a coupling’s torsional stiffness by reducingits spacer diameter also tends to lower its transferable torquecapability in terms of static and dynamic torque. Therefore, theabsolute minimum stiffness that can be realized is mostly limitedby the required peak torque capability of the coupling.

As will be shown in a later section, the compressor vendorinitially considered this approach to reduce the train’s startuptorques. Unfortunately, the torsional calculations determined thatthe startup torques would still be excessive.

Elastomeric Coupling with Low Stiffness and High Damping

One of the common methods to reduce the torsional vibrationamplitude within a resonance is to increase system damping.Elastomeric couplings are often used in such cases.

This style of coupling typically consists of a steel design combinedwith an integrated elastomeric element. The main tasks of theseelastomeric elements are reducing significantly the torsional stiffnessand, in addition, increasing the damping in the system in contrast to asolid steel coupling. The torsional stiffness of such elastomericelements leads to nonlinear behavior. Stiffness characteristics varywith transmitted power, as well as operating temperature.

To be effective, these dampers need to be located at a positionwith high twisting angle. Due to the fact that the normal stiffnessdistribution leads to vibration nodes and, therefore, to hightorsional displacements within the coupling, the coupling locationstend to be “well suited” for these types of damping devices.

During operation, the elastomeric elements absorb vibrationenergy. As a consequence of the material properties, the rubberelements degrade over time due to thermal loading andenvironmental factors. Compared with steel couplings, the maindisadvantage of this design is that it requires increasedmaintenance to provide reliable operation.

Particularly for higher horsepower applications, it should benoted that elastomeric couplings have a tendency to be extremelyheavy devices. This fact should always be born in mind as it mayhave a negative influence on the lateral rotordynamics of theequipment on either side of the coupling. Finally, in someapplications elastomeric couplings are principally not allowed, percustomer specification.

Integrating Controlled Slip Clutch

Synchronous motor-driven trains are designed to transmit therated power, as well as the dynamic torques that occur during


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starting. In most cases, the dynamic torque requirements are fairlyhigh and, therefore, govern the design of the drive components(selection of shaft sizes, gearing, fits, etc.). Hence, the train’ssteady-state design factors are usually fairly high and can easilyaccommodate a modest increase to its rating.

The state-of-the-art aerodynamic technologies offer a potentialfor upgrading existing compressors of older design within the samecompressor frame size. These upgrades require higher ratedtorques and a larger driver may be necessary. As pointed out, thecomponents are often capable of transferring increased amounts ofstatic (driving) torque, but their capacity to withstand dynamicloads is limited. If the more powerful motor increases the dynamicpeak torques during starting, a controlled slip clutch presents theopportunity to limit the dynamic torques during starting. As only acoupling hub needs to be exchanged, the main parts of the originalcoupling and the gear set can often be reused. This assumes, ofcourse, that these reused components are capable of providingreliable operation at the upgraded power rating.

Furthermore, this kind of safety coupling offers an overloadprotection against transient excitations such as those that might occurduring starting when a motor is switched on to the power grid, as aconsequence of an electrical fault that may occur during continuousoperation (e.g., short circuit, network change over, reconnection afterpower interruption, ...). Especially the load cases like network changeover and reconnection to power line may lead to excessive high airgap torques of up to 20 times of rated torque.

CONTROLLED SLIP CLUTCH

Background

The basic design principle of the “autoreset” controlled slipclutch is to transmit the torque through a friction joint where thetorque capacity is controlled by hydraulic pressure (Appell, 1997).This basic layout of the torque-limiting coupling is shown inFigure 4. This torque limiting principle has been implementedin various critical turbomachinery applications, includingaeroderivative-driven power generation units. Moreover it has beenimplemented on extremely large synchronous motor-drivenapplications, such as a 100 MW (134,000 hp) wind tunnel drive.

Figure 4. Torque Limiter Coupling.

Although the implementation of the controlled slip clutch is newto turbocompressor drives, it has been applied on refiner drives inthe wood and pulp industries since 1987 (Figure 5). In some ofthese applications the startup peak torque is experienced on a dailybasis.

Figure 5. Typical Paper Refiner Drive.

Mechanical Overview and Operating Principles

The controlled slip clutch uses the mean of a controlledhydraulic pressure to achieve a desired slip torque. The simplifiedhub’s torque carrying capability can be calculated using thefollowing equation (ANSI/AGMA 9003-A91, 1991):

where:P = Pressure (psi or MPa)P0 = Pressure at which the friction surface grips (psi or MPa)L = Hydraulic cavity length (inches or mm)m = Coefficient of frictionOD = Friction diameter (inches or mm)

The clutch consists of two welded sleeves to create a sealedclamping disc (Figure 6). The disc is mounted on an intermediateflanged sleeve. The keyed ring is lightly shrunk on the motor shaftalong with the shear ring. For the purpose of this paper, the keyedring and the shear ring are referred to as the driver side while theclamping disc and the flanged sleeve are referred to as the drivenside. The driver and driven sides of the clutch are centered to eachother by means of a high performance ball bearing. An accurateslip torque is achieved between the two sides by controlling thehydraulic pressure inside the clamping disc. During normaloperation, the controlled slip clutch acts as a rigid connectiontransmitting 100 percent of the torque. During startup, the clutch’sdriver and driven sides will slip relative to each other every timethe torque exceeds the clutch’s preset slip torque limit. Whileslipping, the clutch will transmit torque at a level equal to its presetslip torque. In typical applications, the clutch will slip up to 5degrees and it is capable of slipping in both directions.

Figure 6. Cross-Section of Clutch Assembly.


� �T P P LOD

u � u u uS P0 2(4)

Protection Against Transient TorquesDuring Startups, Shutdowns, and Overloads

The controlled slip clutch is designed to limit the peakoscillating torque within the design speed range by sliding ashort distance. The clutch implements a spring loaded triggermechanism that automatically resets every time the machine isturned off. The trigger mechanism extends under the influenceof centrifugal force. The mechanism is fully retracted at zerospeed and fully extended at speeds within its design speedrange. Figure 7 shows the cross-section of the triggermechanism and its position at different speeds. Position “A”shows the trigger position while running below the designspeed range. At this position, the clutch will only slip andrelease due to abnormal overtorque. Position “B” shows thetrigger position within the design speed range. At this position,the clutch will slide a short distance after exceeding thepredetermined peak torque limit. Once the machine is turnedoff, the trigger mechanism retracts (Position “A”) and resetsits position.

Figure 7. Clutch Trigger Mechanism Position During Operation.

The design also incorporates a complete release capability forcatastrophic overtorques produced from abnormal transients,such as motor malsynchronization and compressor surge. Inthese rare occurrences, the controlled slip clutch will release andcompletely disengage the driver side from the driven side (Figure7, Position “C”). Resetting the clutch requires replacing the sheartubes and pressurizing the clutch to the desired pressure. Thetypical time required to reset the clutch is less than 30 minutes.Without this added feature, a catastrophic failure of thecompressor train is likely which could put it out of service for anextended period of time.

Mechanical Requirements, Clutch Selection, and Integration

The controlled slip clutch is selected and designed based onthe slip torque level required for the application. Once the sliptorque is specified (a common slip torque level includes 2PU,2.5PU, 3PU) the clutch size and geometrical dimensions are setbased on the interface constraints. Typically, the clutch isdesigned to fit within the confined space of a flexible coupling’shub. The weight and inertia of the clutch are closely matched tothat of a standard flexible coupling hub.

The controlled slip clutch is installed on the low-speed portionof the drive train, between the electric motor-driver and the load(gearbox, compressor, etc.). In this circumstance, there are twopossible locations for the clutch: the first as a flanged connectionat the electrical motor shaft, and the second as a center spacerbetween the flexible coupling halves. The first location offers thesimplest solution, as the clutch can be designed to replace theflexible coupling’s motor-side hub. In this configuration, theclutch bolts directly to the flexible coupling without the use ofany special adapters. Figure 8 presents a typical installation as aflanged connection.

Figure 8. Typical Clutch Installation. (Photo taken during flexiblecoupling laser alignment.)

Installation and Operation

For a flanged design, the controlled slip clutch is installed by fittingit onto the electric motor shaft. Alternatively, when designed as aspacer-type connection, it is simply bolted in position. Regardless of itsarrangement, no special tooling is required to install the clutch.

The controlled slip clutch is set by connecting a high-pressurehydraulic pump to the oil charge port (Item 1 of Figure 9) andunseating the shear tube to permit pressurization of the oil sleevecavity (Item 2 of Figure 9). As shown in Figure 10, the required oilpressure is selected from the clutch-specific torque versus pressurecalibration diagram. Once pressurized, the shear tube is tightenedto seal the pressure inside the clutch after which the hydraulicpump is removed. The clutch is ready for operation.

Figure 9. Cross-Section of Clutch’s Charging Port.

Figure 10. Typical Calibration Diagram.


39

Maintenance and Reliability

The controlled slip clutch requires minimum service, especiallyin continuous operating applications, such as a compressor drive.For this type of application, the recommended maintenance intervalof the clutch matches that of the major drive components. Typicalmaintenance requires checking the level of the clutch’s lubricationoil. The clutch’s hydraulic pressure does not need to be checked.The recommended overhaul interval is 5000 starts, which, for thisapplication, is equivalent to 25 years of compressor operation.

The main factors affecting the torque carrying capability of theclutch mechanism include, but are not limited to, the coefficient offriction, speed, manufacturing tolerances, and temperature. Thedesign and selection of the clutch assembly are based on extensivetribological research to quantify the variables that influence thefriction and wear of its internal contact surfaces, the results ofwhich could be the subject of another technical paper. The clutch’smaterial, heat treatment, and surface finish, as well as its frictionsurface lubricant and interface pressure, are carefully selected andcontrolled to achieve consistent static and dynamic coefficients offriction. Moreover, each clutch is statically calibrated to achieveeven tighter slip torque accuracy. When necessary, this calibrationcan also be modified to accommodate higher-speed applications(speeds greater than 3000 rpm). The clutch’s driver and the drivensides are dynamically balanced as separate components andchecked for balance at several relative locations as an assembly.

TORSIONAL VIBRATION ANALYSIS

Among numerous other requirements, to ensure that synchro-nous motor-driven drive trains provide reliable service, acomprehensive torsional vibration analysis must be performed.The analysis must address all aspects of operation, whichincludes system startup, steady-state operation, and transientevents. API Standard 684 (1996) provides an excellent summaryof the requisite procedures and design considerations that mustbe undertaken to successfully accomplish a comprehensivetorsional vibration analysis that will result in a sound drive traindesign. Briefly stated, as a minimum, a torsional vibrationanalysis of a synchronous motor-driven train must include thefollowing elements:

• Undamped torsional natural frequency and mode shape analysis

• Synchronous motor startup torsional vibration analysis

• Steady-state torsional vibration analysis at system resonances

While it is beyond the scope of this paper to provide a completetreatment of the subject torsional vibration analysis, its salientelements will be discussed in the context on presenting thesimulations performed for this air compressor train. There aremany exceptional references that are readily available on thissubject, the API’s Standard 684 (1996) being one of them.

Creating the Torsional Model

Unlike most lateral rotordynamic models, torsional models mustinclude a mathematical idealization that accurately describes thedynamics of the complete drive train. As such, the requisiteengineering data often come from several vendors. For example,the motor, coupling, and compressor vendors provided the data thatwere used to create the models of the subject compressor train. Inmore complicated arrangements of equipment, there can easily befour, or more, vendors who will need to provide torsional data fortheir hardware. Accordingly, each vendor must be made aware ofwhat data are required and how their information will be utilized inthe torsional vibration analysis. Once collected, it is the analyst’sjob to decipher the information and come up with thetorsional mathematical model. This model, often referred to as themass-elastic data, is essentially an assemblage of interconnectedmassless torsional springs (often arranged sequentially) that is

connected to concentrated inertias. The resulting mathematicalmodel is arranged into a system of equations, commonly know asthe equations of motion. Figure 11 presents a simplified torsionalmass-elastic system, illustrating this concept. Depending upon theanalyses performed, these equations can be solved to determine thesystem’s natural frequencies and mode shapes, time-transientresponse, or steady-state response. Anwar and Colsher (1979)provide a detailed presentation of several methods of solution thatare commonly used to facilitate these analyses.

Figure 11. Simple Torsional Mass-Elastic System and Equationsof Motion.

Torsional Model and AnalysisResults of System without Clutch

As mentioned earlier, a common strategy to reduce the transienttorsional vibrations during a synchronous motor startup is to lowerthe system’s fundamental torsional twist natural frequency. In mostarrangements this can be accomplished by utilizing a “torsionallysoft” coupling to connect the motor driver with the drivenequipment. Prior to selecting the slip clutch mechanism, thisapproach was investigated by the compressor vendor. A presentationof the torsional model and results are presented herein.

Figure 12 presents the torsional mass-elastic system for theequipment arrangement with the “torsionally soft,” steel, low-speedcoupling. The arrangement evaluated is a three-branched system,whose low-speed branch operates at the motor’s synchronous speedof 1200 rpm. Referring to this graphic, the torsional elements(springs) are arranged sequentially from left-to-right. Each torsionalelement is bounded by a left and right-hand side station. Theballoons and inverted triangles represent station-based inertia andexternal torque input, respectively.

Figure 12. Torsional Model of System without Clutch.


DYNAMIC ANALYSIS – FLEXIBLE DRIVE TRAINS

MULTI-DEGREE OF FREEDOM SYSTEMS (TORSIONAL TWIST)

EACH DEGREE OF FREEDOM IS REPRESENTED BY ITS SECOND ORDER DIFFERENTIALEQUATION IN THE FORM OF:

WHEN ALL SYSTEM DEGREES OF FREEDOM ARE COMBINED IN A MATRIX FORM:

fkXXdXm ��

>> @@ ^̂ `̀ >> @@ ^̂ `̀ >> @@ ^̂ `̀ ^̂ `̀FXKXDXM ��

ROTOR SYSTEM REPRESENTATION

K1

K2

T1

J2

J3

J1

T2

Where:

J = Inertia

K = Torsional Spring

T = Torque

T3

F

Branch #1 contains the low-speed drive components, namely thesynchronous motor, low-speed coupling, and compressor inputshaft and bull gear. The inertias of the motor rotor’s exciter andcore are located at stations #1 and #2, respectively. The main drivecoupling is represented by torsional element #3 and is bounded bystations #3 and #4. The inertia of the compressor bull gear islocated at station #5. The torsional mass-elastic data for thecompressor pinion assemblies are contained in branches #2 and #3.Each of these branches is connected to branch #1 via a torsionalspring at the centerline of the pinion-to-gear mesh. The torsionalstiffness values of these elements account for the flexibility of thegeared connections. The inertias of the compressor wheels arelocated at the extremities of each pinion branch.

Figure 13 presents the torsional natural frequency results in thecontext of a torsional critical speed map (Campbell diagram). Thediagram provides a graphical presentation of the torsional naturalfrequencies (horizontal lines) along with the excitation frequency ofthe synchronous motor driver as a function of speed to expeditiouslyidentify interferences (intersections or resonance conditions). Thismap presents interferences with the first five modes of vibration(Modes #1 to #5). The critical speed map has been annotated with thecorresponding motor speeds at each interference point. Referring tothis figure, note that there are interferences with the first four torsionaltwist natural frequencies that will enter into resonance with thesynchronous motor excitations during startup. Figure 14 presents thetorsional twist mode of vibration corresponding to the system’sfundamental torsional twist natural frequency. An examination of thismode shape reveals where, if excited, the maximum torsionalresponse is most likely to occur.

Figure 13. Torsional Critical Speed Map (System without Clutch).

Figure 14. Fundamental Torsional Twist Mode of Vibration (Systemwithout Clutch).

For example, if Mode #1 is excited (Figure 14) the maximumdynamic torque, as identified by its node (location where it crosses theshaft centerline), is likely to occur in the compressor’s input shaft. Thenode also provides a qualitative indication of which torsional spring

rates have the greatest influence on the corresponding torsionalnatural frequency. Springs closest to the node will have the greatestinfluence on the frequency. Accordingly, as one can see, by virtue ofits close proximity to the nodal point, in addition to the compressor inputshaft, the torsional stiffness of the low-speed coupling has asignificant influence on this train’s fundamental torsional natural frequency.

Typically, as a consequence of its potential to facilitate largeamplifications of the dynamic torques/angular motions, thefundamental torsional twist mode of vibration is commonly regardedas the “mode of concern” in this particular type of drive trainconfiguration. The higher frequency torsional modes of vibrationinvolve some degree of torsional twist in the pinions and, despitebeing resonant with the synchronous motor torques during motorstartup, are fairly “well isolated” from these excitations by thesizable inertia of the compressor’s bull gear. Furthermore, thesemodes of vibrations are also fairly “well damped.” Estimates of themodal damping for this arrangement of equipment vary between 2 to5 percent of the critical value. For the purposes of this study, theauthors have used a value of 31/2 percent for this configuration ofequipment. This value is consistent with those presented by Chen, etal. (1983), as well as by Anwar and Colsher (1979). It should benoted, however, that the selection of the appropriate modal dampingvalue(s) is often viewed as an “experience based” endeavor, andtherefore should be consistent with the engineer’s understanding ofthe train’s design/operation and the end-user’s requirements.Fortunately, however, in contrast to steady-state responsecalculations, the time-transient response at, or near, resonance ismuch less sensitive to small-to-moderate variations of modaldamping (Smalley, 1983).

Torsional response analyses were performed to calculate thedynamic torques along the drive train, resulting from the input(synchronous motor) and load (compressor) torques, using anonsynchronous time transient analysis. This type of analysisconsiders the synchronous motor’s speed-dependent mean andpulsating torque characteristics as well as the compressorcountertorque (load). Thus, the analysis captures the speed-dependent loading behavior of the motor and compressor toaccurately simulate the dynamic torsional response at variouslocations along the drive train. Newmark’s linear response solutionalgorithm was utilized to simulate the time-transient response of thecompressor drive train (Anwar and Colsher, 1979). The full voltage(infinite bus) startup characteristics for the synchronous motor, aswell as the compressor load torques, are presented in Figure 15. Asmentioned earlier, it should be recognized that the synchronousmotor starting characteristics are typically application-specific.Accordingly, with certain synchronous motor designs, the magnitudeof the oscillating component torque can reach as high as 120 percentof the motor’s rated torque, while others may remain below 70percent. As such, it is imperative that the time-transient simulationsutilize the vendor-supplied startup data. Moreover, the full-load(infinite bus) motor starting characteristics should be used, as thiswill produce the most severe time-transient response.

Figure 15. Synchronous Motor Torque Versus SpeedCharacteristics (Infinite Bus).


41

Figures 16 and 17 present the simulated torsional response of thedrive train. The upper plot on Figure 16 presents the simulated motorspeed versus time for the compressor drive train. Referring to thisgraph, the motor’s synchronous speed is achieved in approximately13.5 seconds. The remaining plots on Figures 16 and 17 present thesimulated dynamic torques for the motor drive extension, low-speedcoupling and compressor input shaft, respectively. Note that both ofthese figures clearly illustrate the presence of large dynamic torquesat approximately 11 seconds into the startup sequence.

Figure 16. Synchronous Motor Startup Response (System withoutClutch).

Figure 17. Synchronous Motor Startup Response (System withoutClutch).

Referring again to the upper graph on Figure 16, 11 seconds alsocorresponds to approximately 90 percent of the motor’ssynchronous speed and is consistent with the resonance with thedrive train’s fundamental torsional twist mode of vibration that isshown on the critical speed map of Figure 13. Furthermore,referring to the motor startup performance shown in Figure 15, theresonance at 90 percent of synchronous speed is also coincidentwith the slip speed at which the motor’s oscillating torquemagnitude reaches its maximum value.

In addition to numerous other torsional performance criteria,the API specifies that, throughout the drive train, the maximumdynamic torque must not exceed five times the motor ratedtorque for a full voltage start (API Standard 546, 1997). Theoriginal 7000 hp rating was used to establish this torque limit.As can be seen from the time-transient simulation results, thislimit is clearly exceeded during resonance. Subsequent torsionalfatigue calculations also determined that the simulated responsewould not meet the end-user’s requirement of 5000 starts.Recognizing that the motor’s starting characteristics, as well asthe train’s layout, cannot be changed, the only ways to reducethe startup torques would be to dramatically reduce thetrain’s fundamental torsional twist natural frequency, and/orincorporate a torque-limiting clutch mechanism at the drive endof the motor. To further reduce the train’s fundamental torsionalnatural frequency would require utilizing a torsionally-soft,elastomeric coupling. This type of retrofit was not viewedfavorably by the end-user as this style of coupling requiresperiodic maintenance. Furthermore, owing to its increasedweight and inertial characteristics, the coupling could also havea negative influence on the lateral rotordynamics of the motorand compressor components.

Torsional Model and Analysis Resultsof System with Slip Clutch Mechanism

Similar to the previous drive train configuration, a mass-elasticmodel of the modified system with the slip clutch mechanismwas prepared to assess its torsional dynamics. Figure 18presents the torsional mass-elastic system for this equipmentarrangement. For the purposes of performing naturalfrequency and response calculations, owing to its relativelyhigh torsional stiffness and low inertia, the mass-elasticcharacteristics of the slip clutch mechanism have beencombined with those of the low-speed coupling. These data arerepresented by element #3 of branch #1 and its adjoiningleft- and right-hand side stations.

Figure 18. Torsional Model of System with Slip Clutch.

Figure 19 presents the torsional natural frequency results in thecontext of a critical speed map. Referring to this figure, similar tothe previous model, note that there are interferences with the firstfour torsional twist natural frequencies that will enter intoresonance with the synchronous motor excitations during startup.Figure 20 presents the train’s fundamental torsional twist mode ofvibration. Albeit with slightly different frequencies, theremaining mode shapes are quite similar to those of the previousmodel.


Figure 19. Torsional Critical Speed Map (System with Clutch).

Figure 20. Fundamental Torsional Twist Mode of Vibration (Systemwith Clutch).

The performance of the clutch component, by virtue of its abilityto permit slippage, constitutes a nonlinear system and, thereforerequires a slight modification to the Newmark numericalintegration procedure to simulate its behavior. An overview of themodified numerical integration solution procedure is presentedherein. For this application, the slip clutch was set to slip when theinstantaneous torque exceeded three times the motor’s rated torque(145 kNm [107,000 ft-lb]). Below this value, the coupling behaveslinearly, and the aforementioned linear time-transient solutionalgorithm is employed (Anwar and Colsher, 1979). The slip torquethreshold is monitored at every solution timestep. Once the sliptorque threshold is reached, the Newmark damping and stiffnesssystem matrices are augmented to reflect a torsional system that isno longer elastically connected through the clutch assembly. Inthis state, continuity across the clutch element is maintained bycomputing an external frictional torque that is applied to itsleft- and right-hand side stations. The value of this torque isdetermined by taking the product of the ratio of the clutch’skinematic and static coefficients of friction and its slip torquethreshold. For this simulation, the clutch’s frictional torque was setto 113.93 kNm (84,000 ft-lb). As long as the clutch continues toslip, the uncoupled equations of motion are integrated and thefrictional torque is applied to the ends of the clutch element. Whileslipping, the elastic connection across the clutch is reintroducedinto the torsional equations of motion the instant the differentialangular velocity across the clutch changes sign. At this point, the

linear numerical integration solution routine continues with theoriginal equations of motion until the next time the slip torque limitis exceeded.

Figures 21 and 22 present the simulated torsional response of themodified drive train with the slip clutch assembly under fullvoltage (infinite bus) startup conditions. The upper plot on Figure21 presents the simulated motor speed versus time for thecompressor drive train. The motor’s synchronous speed is achievedin approximately 13.7 seconds. The remaining plots on Figures 21and 22 present the simulated dynamic torques for the motor driveextension, low-speed coupling and compressor input shaft,respectively. Note that both of these figures clearly illustrate thepresence of large dynamic torques at approximately 11 secondsinto the startup sequence; however, the maximum torsionalamplitude is controlled by the clutch’s slip torque threshold.Accordingly, the maximum dynamic torque values do not exceedthe five times the motor rated torque limit (API Standard 546,1997). Moreover, subsequent torsional fatigue calculations alsodetermined the simulated response would meet the end-user’srequirement of 5000 starts.

Figure 21. Synchronous Motor Startup Response A (System withClutch).

Figure 22. Synchronous Motor Startup Response B (System withClutch).


43

Finally, as shown in Figure 23, the performance of the slip clutchmechanism was also quantified by evaluating its cumulative slipduring the startup. The clutch begins to slip when its maximum sliptorque setting is reached at approximately 10.4 seconds. From thispoint forward, the “stepped” appearance of the simulated slipindicates that the clutch disengages (vertical slope) and reengages(horizontal slope) approximately seven, or eight, times as the trainpasses through the resonance. Based on these data, the cumulativeslip for the entire startup was estimated to be 4.2 degrees. The slipclutch vendor determined this performance to be well within theacceptable limits for their device.

Figure 23. Synchronous Motor Startup Response C (System withClutch).

FIELD TESTING AND DATA ANALYSIS

Purpose

Torsional vibration problems may be caused by various sources.Most of them are well understood and can be determined within atorsional simulation analysis.

In some cases, gear vibrations and/or noise may give the firstindication of a torsional vibration problem due to the fact thattorsional and lateral movements are coupled via the gear mesh.Under these conditions clattering of the gear teeth may be heard asthe gear teeth oscillate within their backlash. A fatigue failure oftrain components during operation can be seen as worst indicationof a torsional vibration problem.

To avoid any kind of risk related to the train modifications it wasthe intention of the compressor manufacturer to measure theoccurring torque in the field. In principle the used controlled slipclutch was well understood. But nevertheless, the function of thetorque limiting effect should be demonstrated by field testingincluding torsional vibration measurement. This test should alsoverify that the modified train will be safe over the whole life cycle.Static and dynamic friction and slip torques as well are complexmechanisms that are difficult to predict accurately. Consequently,it was essential to check whether the real slip torque correlates withthe predicted figure to enhance the confidence in this design.

Sensors and Data Acquisition

The core element of the measurement method was the couplingequipped with a straingauge. A full bridge circuit was applied tothe coupling spacer (Figures 24 and 25). The straingauge waswired to a sensor signal amplifier. The output of the sensor signalamplifier was transmitted via a telemetry system to a nonrotatingevaluation unit. The torque equivalent output of the evaluation unit

as well as the later in particular described keyphasor signals weresampled by a digital data acquisition recorder and additionallystored on a tape recorder (Figure 26).

Figure 24. Schematic of Torque Measurement Configuration.

Figure 25. Torque Measurement Configuration.

Figure 26. Data Acquisition System.


The straingauge is able to measure the torsional deflection withinthe spacer material. Based on the torsional flexibility this deflectionis proportional to the torque. To achieve an unambiguous signal thestraingauge was installed far away from a stress intensified zone ascan be seen in Figure 24. The system is prepared to detect thetorque transmitted from the motor via the coupling into thecompressor on an instantaneous basis. In other words, the systemis capable of observing the quasistatic and the dynamic torqueduring starting and afterwards during continuous operation. This isthe most reliable method to obtain accurate response frequenciesand stresses during a rapid start of a motor.

In addition to torque, the drive train also had instrumentation tomeasure the slip angle of the clutch during starting. To get animpression of this angle two keyphasor signals were installed; oneon the motor shaft and the other on the coupling spacer (refer toFigures 24 and 25 speed measurement motor/spacer). Based on anadequate data acquisition frequency the differential anglebetween motor shaft and coupling spacer could be measured in asufficiently precise manner.

Presentation of Test Plan,Acquired Data, and General Observations

In November 2003 the dynamic torque measurements were carriedout onsite. Four starts were recorded by the installed equipment.During start number 1 and 2 the train did not achieve synchronousspeed due to general plant reasons. However, start number 3 and 4were successful and the motor brought the compressor train tosynchronous speed. The observed responses for both successful startsare similar in shape and level of peak torques.

To give an impression of the characteristics, the measured resultsof the fourth run are shown in Figure 27. Figure 27demonstrates the synchronous motor start, which in principlerepresents a shape of the transient torque as expected. To provide abetter evaluation of the behavior of the controlled slip clutch, Figure28 zooms into the resonance condition where the motor air gaptorque with double slip frequency excites the train’s fundamentaltorsional twist mode of vibration. Referring to this figure, note thatthere are eight main torque peaks. Seven of them are on the samepeak torque level. Additionally, the shapes of these peaks show someindication of nonlinear behavior, similar to a slip-stick mechanism.The detected peak torque values are summarized in Table 1. The clutchwas designed to limit the startup peak torque at 145 kNm (107,000ft-lb). During field testing the clutch slipped seven times at a torque ofabout 120 kNm (88,500 ft-lb) and limited the peak torque. Theevaluation of the slip torque level shows a deviation of 62 percent(Table 1). That means in fact, that the range is between 117 kNm(86,300 ft-lb) up to 121 kNm (89,200 ft-lb). Considering the complexityof the mechanical interactions at the clutch’s slip interface, therelatively small variation in the measured slip torque of 62 percent isquite impressive, and moreover, verifies the accuracy of the device.

Figure 27. Measurement Results of Starting the Synchronous Motor.

Figure 28. Measurement Results Motor (Zoomed in the ResonantCondition).

Table 1. Determination of Peak Torques During ResonanceCondition.

The torque level in reverse direction were measured in a rangefrom 248 kNm (235,400 ft-lb) up to -70 kNm (251,600 ft-lb). Noindication of slipping in the opposite direction was observed due toits relatively low level. By means of the two keyphasor signals adifference angle between motor and coupling of 3.3 degrees wasmeasured.

Based on these field test results it is proven that this slip torqueprotection works during startup. The slip torque was in fact 120 kNm(88,500 ft-lb), whereas the clutch was adjusted to limit the torque atabout 145 kNm (107,000 ft-lb). The deviation between predicted andmeasured torque is about 20 percent. About 50 percent of this deviationcan be attributed to differences in the diametrical tolerances of theclutch’s calibration mandrel and that of the electric motor’s driveextension shaft. If required, the clutch could have been recalibrated in


45

Peak

NumberPeak Torque

Deviation from

average Peak

Torque

1121.17 kNm

(89,370 ft-lb)+1,4%

2121.57 kNm

(89,665 ft-lb)+1,7%

3117.09 kNm

(86,361 ft-lb)-2,0%

4117.07 kNm

(86,346 ft-lb)-2,0%

5121.04 kNm

(89,274 ft-lb)+2,2%

6119.07 kNm

(87,821 ft-lb)-0,3%

7119.09 kNm

(87,836 ft-lb)-0,3%

Average 119.44 kNm (88,094 ft-lb)

the field. However, this procedure was not determined to be necessary,as the clutch’s design slip torque was not exceeded and its measuredangular slip was found to be acceptable.

To check the plausibility of the strain measurement, the torquesat steady-state (continuous) operation were compared to theelectrical power input. The measured results correlate well with theelectric and related thermodynamic power. The deviation is in arange of about 62 percent. Based on this comparison it can bestated that the field test was a reliable diagnosis.

Comparison with Computer Simulation Results

To facilitate a direct comparison of the measured torque valueswith the simulation results, the nonlinear time-transient responsestartup simulations were rerun with a reduced slip torque settingof 120 kNm (88,500 ft-lb). Similar to the previous analyses, thesimulation utilized the same infinite bus motor startingcharacteristics, modal damping parameter, load torques, and slipcoupling coefficients of friction. Figure 29 presents time-waveformdata for the measurement (top) and simulation (bottom) data. Aninitial inspection of these data shows that there is relatively goodagreement between the two sets of results. The measurement dataindicate that the clutch begins to slip around 10.75 seconds, whilethe simulation predicts initial slippage about 0.25 seconds earlier(10.50 seconds). During resonance, the maximum and minimumvalues of the predicted torsional oscillations are almost identical tothe measured values.

Figure 29. Measurement Versus Simulation Results (System withSlip Clutch).

The only notable difference in the two sets of data is that thesimulation predicts about twice as many slip-stick events duringresonance. Accordingly, as shown in Figure 30, the predicted slipof 6.9 degrees is about twice as the measured value (3.3 degrees).Nevertheless, the simulation results are considered to be quiteacceptable. It is likely that further refinements of the simulationslip-stick algorithm, and/or the simulation input data wouldprovide even better predictions. Possible refinements include:

• Implementation of a damping model at the simulated slip-stickinterface.

• Reduction in the time-transient response integration timestep.

• Refinement of the motor and load startup data.

• Refinement of the static and kinetic coefficients of friction forthe slip clutch.

• Refinement of the torsional mass-elastic model.

Figure 30. Simulation Clutch Slip During Startup.

SUMMARY AND CONCLUSIONS

Review of Revamp Project and Lessons Learned—The Plant’s Perspective

The main objective in revamping the compressor trains wasto extend the life of the plant while being able to operate atsubstantially higher than design production. The second objectivewas to reduce the amount of electrical power consumed per unitvolume of product produced. The third objective was to reduce theamount of projected maintenance for the compressor train. Thefourth objective was to minimize the amount of modificationrequired to install the new motors. The fifth objective was to designthe revamped compressor train for 5000 starts. The final objectivewas to minimize the risk of the revamp.

The first three objectives were achieved by using a higherhorsepower solid pole rotor synchronous motor whose design wasoptimized for this application. The fourth objective was achievedby having the motor frame designed to fit the existing sole platesand duplicating all-important dimensions of the original motor.The fifth and final objective was achieved by using the controlledslip clutch coupling. Accordingly, the revamp met the objectivesand is considered a success.

The revamp of the two 7000 hp gas compressors was part of alarger program of replacement or upgrading of critical componentsthat are reaching end of life. The replacement of the motors wentvery smoothly during the turnaround. Installation of the controlledslip clutch coupling consisted of fitting it onto the motor shaft,positioning it for correct axial location, and then charging theclutch with oil up to the design pressure. Once charged, the clutchcoupling is fixed onto the motor shaft where it serves as themotor-side, low-speed coupling hub. The performance of the slipclutch mechanism was verified via torsional field measurements.To date, both clutches have continued to provide reliable service.

Some of the lessons learned from this revamp are:

• The use of the slip clutch type coupling is a viable solution tolimiting the transient torsional torque during synchronous motorstartup. This was verified by test during startup after the new solidpole rotor synchronous motors and slip clutch type couplings wereinstalled. Risk of mechanical failure is reduced to an acceptable level.

• The use of a solid pole rotor synchronous motor has manyoperational advantages, but there is a penalty to pay compared tousing induction or laminated rotor synchronous motors. Thepenalty is the more sophisticated replacement coupling sets that arerequired. This is considered a reasonable engineering tradeoff.


• A complete set of motor data has to be sent to the group doingthe torsional analysis. The motor specification package wasupgraded during this project as a result of “lessons learned.”

• The final design of the motor and coupling set may take three orfour iterations.

• The motor vendor, compressor vendor, and the group doing thetorsional analysis should agree to a final solution before the motoris released for manufacture.

• The total time from start of project to having the componentsready to install takes about 18 to 24 months.

• The time to change out a motor varies between two and threeweeks for a motor installed on sole plates and three to four weeksfor a motor whose frame is grouted into the foundation. Thisincludes removal of the old equipment, cleanup, installation,alignment, and startup. Most of the time there will beelectrical/instrument work such as power cable replacement,switchgear relaying upgrades, and compressor instrumentationupgrades going on while the motor is being installed.

From an operator’s perspective, in addition to improvedperformance, the proof of a successful revamp project is not havingany “reminders” that the work was done (e.g., special startingrequirements, additional maintenance items, etc.). The revampedcompressor trains have been in continuous operation since thefourth quarter of 2003 at the increased production rates. The workwas considered a success and other compressor trains have beenupgraded since then.

With the exception of the one event where a brass safety pin wassheared during a utility disturbance, there have been no additionalincidents and/or special requirements placed on the drive trains’operation by virtue of installing the controlled slip clutches. It isstrongly suggested that a repair kit be purchased and stored at the sitewhere controlled slip clutch coupling is in service. A repair kitgenerally consists of a charging pump with pressure gauge; a torquewrench for installation of the sealing plug; and oil, spare gaskets, andbrass shear pins. In addition, it is recommended that the pressuresetting for the controlled slip clutch coupling be marked on thecoupling guard so that there is no question as to what pressure thedevice needs to be set at in the event of having to recharge the unit.

General Remarks and Additional Observations

As with any new technology, the biggest hurdle for the controlledslip clutch is industry acceptance. Although the controlled slipclutch is a well-proven product in the pulp and paper industries, theturbomachinery community, being somewhat conservative, tends tobe very cautious when it comes to implementing new technologies.Considering the investment in equipment, as well as thefinancial and logistical consequences of an unplanned outage, thischaracteristic is completely understandable.

It is the intention of this paper to provide the turbomachinerycommunity with a comprehensive accounting of the torsionalvibration problems, as well as the trials and solutions thereof,encountered during a typical revamp project. It is the authors’understanding that all parties involved were satisfied with theoutcome of this project. Accordingly, when appropriate, the con-trolled slip clutch should be considered as a viable means to solvetorsional vibration problems for new and used equipment. Sincethe implementation of the controlled slip clutch on this project, thetechnology has been successfully applied on different compressortrains at various locations.

Additional observations, recommendations, and conclusions areas follows:

• State-of-the-art aerodynamic designs and analyses offer a viablepotential for upgrading existing compressors of older design withinthe same compressor frame size.

• By carefully applying the appropriate technologies, it is possibleto maintain the equipments’ original design safety factors whenupgrading its performance.

• Nonlinear torsional vibration analyses can be used to accuratelypredict the behavior of slip clutch mechanisms. It is imperativethat the torsional analyst be provided with a complete set ofengineering data from all of the equipment vendors to obtain acomprehensive understanding to the drive train’s dynamics.

• The use of slip clutch mechanism should also be considered innew applications, especially if the dynamic load levels aredecisive in determining the dimensional requirements of the traincomponents. In these situations, by reducing dynamic load levels,the clutch may allow designers to optimize components based onsteady-state operating conditions, rather than transient events.

• Compared to other alternatives (LCI, transformers, elastomericcouplings, etc.), slip clutch mechanisms can offer a relatively lowcost solution for difficult synchronous motor-induced torsionalvibration problems.

• The implementation of a sound engineering approach, alongwith a complete understanding of the drive train equipment and fullcooperation among the involved parties, are vital components toobtain a successful drive train design.

REFERENCES

ANSI/AGMA 9003-A91, 1991 (R1999), “Flexible Couplings—Keyless Fits,” American National Standards Institute,Washington, D.C./American Gear Manufacturers Association,Alexandria, Virginia.

Anwar, I. and Colsher, R., 1979, “Computerized Time TransientTorsional Analysis of Power Trains,” ASME Paper 79-DET-74,New York, New York.

API Standard 546, 1997, “Brushless Synchronous Machines—500 kVA and Larger,” American Petroleum Institute,Washington, D.C.

API Standard 684, 1996, “Tutorial on the API Standard ParagraphsCovering Rotor Dynamics and Balancing: An Introduction toLateral Critical and Train Torsional Analysis and RotorBalancing,” First Edition, American Petroleum Institute,Washington, D.C.

Appell, B., 1997, “New Method for the Protection of Gas TurbinesAgainst Overload and Overspeed,” Voith Turbo Publicationcr580e, Crailsheim, Germany.

Chen, H. M., McLaughlin, D. W., and Malanoski, S. B., 1983, “AGeneralized and Simplified Transient Torque Analysis forSynchronous Motor Drive Trains,” Proceedings of the TwelfthTurbomachinery Symposium, Turbomachinery Laboratory,Texas A&M University, College Station, Texas, pp. 115–119.

Smalley, A. J., 1983, “Torsional System Damping,” Proceedings ofthe Seventh Annual Meeting of the Vibration Institute, TheVibration Institute, Houston, Texas.

BIBLIOGRAPHY

Brakelmann, S., 2003, “Freeport Measurement Report,” SiemensPG I, Duisburg, Germany.

Corbo, M. A. and Malanoski, S. B., 1996, “Practical DesignAgainst Torsional Vibration,” Proceedings of the Twenty-FifthTurbomachinery Symposium, Turbomachinery Laboratory,Texas A&M University, College Station, Texas, pp. 189–222.

IEEE Standard 1255-2000, 2000, “Guide for Evaluation of TorquePulsation During Starting of Synchronous Motors,” Instituteof Electrical and Electronics Engineers, Inc., New York,New York.


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Wachel, J. C. and Szenasi, F. R., 1993, “Analysis of TorsionalVibrations in Rotating Machinery,” Proceedings of the Twenty-Second Turbomachinery Symposium, Turbomachinery Laboratory,Texas A&M University, College Station, Texas, pp. 127–151.

ACKNOWLEDGEMENT

The authors would like to thank the management of Air Liquide,Voith, and Siemens PG I for authorizing the publication of thisproject. They would also like to thank Siemens’ Stanley Stevensonfor his helpful feedback and encouragement.


revamping a gas compressor drive train … · experience mainlyin turbomachinery, ... this means...

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