TORSIONAL VIBRATION PROBLEM WITH …

1 Troy Feese is a Senior Project Engineerat Engineering Dynamics Incorporated(EDI), in San Antonio, Texas. He has morethan 17 years of experience performingtorsional VIBRATION , lateral critical speed,and stability analyses as well as evaluatingstructures using finite element methods. Healso conducts field studies of rotating andreciprocating equipment. Mr. Feese is alecturer at the EDI seminar and has writtentechnical articles and papers on TORSIONAL VIBRATION , lateral criticalspeeds, and balancing. He contributed toAPI Standard Feese received a BSME from The University of Texas atAustin in 1990 and has an MSME from UTSA. He is a member ofASME, VIBRATION Institute, and a registered Professional Engineerin the State Maxfield is an Engineer at TesoroRefining & Marketing Company, in Martinez,California.

2 Since March 2005, he hasmanaged the prediction/prevention programfor the refinery ( VIBRATION , ultrasonics,thermography,lubrication,and reciprocatingcompressors). Under his leadership, theprediction program has seen an increase inreal finds and respect among the personnelwithin the refinery. The Golden EagleRefinery has roughly 1300 pumps, multiple critical unsparedmachines, 130 turbines, and 34 main reciprocating Maxfield oversees four technicians that acquire and analyzedata Maxfield graduated with a degree (MechanicalEngineering, 2001) from California Polytechnic University draft (ID) fan systems often use louvers or variable inletguide vanes to control the flow of exhaust through the IDfan system can be made more efficient by using a variablefrequency drive (VFD) to control the fan speed, which controls theair flow without adding restrictions to the flow path.

3 To reduceenergy costs, many companies are using more in a motor/ID fan system operating a sufficient marginaway from the TORSIONAL natural frequencies. This type of systeminstability is not a classical TORSIONAL resonance and would bedifficult to predict in the design stage. As a result of the hightorsional VIBRATION , several couplings were damaged and a motorshaft experienced a fatigue failure before the PROBLEM could beclearly identified and different fan systems at the refinery were tested and shownto exhibit similar TORSIONAL behavior, although only one of thesesystems actually failed. Test data showed that the dynamic torquein the couplings was excessive when the original VFD wasoperated at electrical frequencies above the first TORSIONAL naturalfrequency of the system (21 to Hz depending on the fan andcoupling arrangement).

4 Within the normal operating speed range,therewascontinualreversingtorque,w hichisconsideredunacceptablefor centrifugal demonstrate that the VFD was the source of the excitation,the VFD was reconfigured as a soft starter so that it could bebypassed and the fan could then be operated at constant speed usinginlet damper control. When operating across-the-line without theVFD, the dynamic torque was significantly reduced (approximately10 percent of the motor rated torque).After the test results were reviewed, a new VFD was developedand installed by the manufacturer to prevent the PROBLEM fromreoccurring within the normal operating speed range.

5 Finalmeasurements are presented that show significant reduction indynamic torque after the drive modifications were technology has resulted in reduced torque modulationproduced by a variable frequency drive (VFD) driven motor. Pulsewidth modulation (PWM) is generally thought to have smoothoperation compared to older drive are several different types of control for PWM ,whichisreportedtobeacceptablefor applications like fans and pumps. However, if not properlytuned, these VFDs can still excite TORSIONAL natural frequenciesresulting in high TORSIONAL VIBRATION and damaged draft (ID) fans are commonly used in crude units atrefineries.

6 Older control methods utilized dampers that could beopened and increase energy efficiency, aVFD motor canbe used to adjust fan speed instead of throttling flow with , any downtime of the crude unit due to problems with thefans can quickly offset the energy this case, the end user experienced reliability problems withthe ID fan system after installing aVFD for the motor control. Oneunit referred to as 50 Unit consists of a 500 hp induction motordriving an ID fan for the furnace exhaust (Figure 1). Fan speedsrange up to 1200 rpm during normal VIBRATION PROBLEM WITHMOTOR/ID FAN SYSTEM DUE TO PWM VARIABLE FREQUENCY DRIVEbyTroy FeeseSenior Project EngineerEngineering Dynamics IncorporatedSan Antonio, TexasandRyan MaxfieldEngineerTesoro Refining & Marketing CompanyMartinez, CaliforniaFigure 1.

7 Picture of Motor at 50 spacer in the coupling between the motor and ID fan failedseveral times, including one catastrophic failure during original coupling size was verified by the manufacturer andfound to have a sufficient service factor of 3 for ,theenduserdecided to install a much larger coupling with a service factor of was thought that this would prevent further coupling the underlying PROBLEM was not understood, increasingthe coupling size just eliminated the fuse and caused the nextweakest link in the system to fail, the motor shaft. As discussed inthis paper, it was determined through testing that the fatiguefailures were caused by excessive TORSIONAL VIBRATION due to anexcitation source from the OF EQUIPMENTID Fan 50 UnitThe atmospheric furnace F-50 heats the crude oil entering therefinery (between 60,000 and 108,000 barrels per day).

8 An induceddraft fan is needed for the furnace. The fan system consists of aninduction motor driving an ID fan through a nonlubricated couplingas listed in Table 1. The motor speed is controlled by a 1. Equipment for 50 coupling catalog lists allowable torque values for continuousand peak overload. The sizing procedure from the couplingmanufacturer normally recommends a service factor of for amotor driven ID fan. The service factor (SF) was 3 for the originalcoupling. According to the catalog, the coupling was sufficient forthis allowable dynamic torque during continuous operation is notgiven in the coupling catalog, but can be determined from amodifiedGoodman allowable alternating torque variesdepending on the transmitted torque.

9 Goodman plots are constructedwith mean torque along the horizontal axis and alternating torqueplotted along the vertical allowable dynamic torque was 28,000 in-lb zero-to-peak forthe original coupling. At the allowable level, the dynamic torquewould be fully reversing (exceeding the transmitted torque). Forsmooth operation of centrifugal equipment the alternating torqueshould be much lower (normally less than 10 percent of thetransmitted torque).The larger coupling (SF = 15) has a much higher torque ratingthan the original coupling and was oversized for this allowable dynamic torque is 161,000 in-lb zero-to-peak basedon the modified Goodman Fan No.

10 3 ReformerA second ID fan (located at the No. 3 Reformer) was also sister system has the same model motor and VFD as the 50 Unit, but has not experienced any TORSIONAL failures. However, thecoupling and fan are smaller. Based on the fan curves, themaximum fan load is only 370 hp. Therefore, the coupling for theNo. 3 Reformer is smaller and has a lower torque rating. Flexible disc coupling Service factor = 2 Maximum continuous torque = 40,400 in-lb Peak overload torque = 80,800 in-lbBased on the modified Goodman diagram, the allowablealternating torque would be 9000 in-lb zero-to-peak for a maximumload of 370 hp (transmitted torque = 19,500 in-lb at 1195 rpm).