A SOLUTION TO YEARS OF HIGH VIBRATION PROBLEMS IN THREE REINJECTION COMPRESSOR TRAINS RUNNING AT 33 MPa DISCHARGE PRESSURE Jong Kim, PhD/Presenter
Authors Jong Kim, PhD Senior Principal Engineer, Waukesha Bearings Corporation Bachelor of Science, Mechanical Engineering, 1985 Busan National University Master of Science, Mechanical Engineering, 1987 KAIST (Korea Advanced Institute of Science and Technology) Doctorate of Philosophy, Mechanical Engineering, 1991 KAIST(Korea Advanced Institute of Science and Technology) Marcio Felipe dos Santos Senior Maintenance Engineer, Major South American Oil Company Bachelor of Science, Mechanical Engineering, 1979 Universidade Federal of Rio de Janeiro Barry J. Blair Chief Engineer, Waukesha Bearings Corporation Bachelor of Science, Mechanical Engineering, 1990 University of Virginia Master of Science, Mechanical Engineering, 1990 University of Virginia Master of Product, Design and Development, 2015 Northwestern University
Abstract Over a 13 year span, a major South American oil company s maintenance department fought high vibrations in three gas reinjection compressor trains. To reduce the chances of machine trips, technicians field balanced the compressors every year and replaced worn point contact pivot tilt pad journal (TPJ) bearings and O ring squeeze film dampers (SFDs) with new ones yearly. The downtime from implementing these preventative measures and from actual trips in the trains resulted in a loss of capacity of 1% a year and additional flaring of the gas. After a thorough analysis of the compressors and inspection of damaged components, it was determined that the reoccurring problems would be solved by installing optimized Flexure Pivot tilt pad journal bearings with Integral Squeeze Film Damper (ISFD) technology into the compressors. In 2013 the reinjection compressors were placed back into service with Flexure Pivot TPJ bearings with ISFD technology. Since then the reinjection compressors have exhibited lower vibration levels that do not grow over time, have had ZERO trips, and have not required field balancing for continuous operation. Overall efficiency has increased by approximately 1%.
Contents Background Problem Statement Modeling Upgraded Bearing and Damper Design Optimization Summary Lessons Learned
Unit Background Information 3 gas reinjection compressor trains operating since 2000 Each compressor train has two casings First casing (LP) experienced vibration issues Second casing (HP) did not have vibration issues Discharge pressure: 33 MPa Rated speed: 11,456 rpm OEM bearing information 114.3 mm bore x 50.8 mm long TPJ 5 pad, load on pad 60% offset Vibration Issue
Problem Statement Rotor vibration levels increased over time Downtime due to high vibrations resulted in about 1% loss in production time and additional flaring gas Field balanced every year Installation of new OEM bearings every year 76.2 Vibration with worn pivots Vibration pk pk (microns) 100 80 60 40 Vibration trend of LP compressor with OEM bearings over 5 month span 20 0 0 24 Jul 2012 Compressor DE X Compressor DE Y Compressor NDE X Compressor NDE Y 18 Nov 2012 Vibration pk pk (microns) 61.0 45.7 30.5 15.2 Vibration with new OEM bearings of the original design Compressor DE X Compressor DE Y 0 01 Nov 2011 19 Nov 2011 6 Dec 2011 24 Dec 2011 10 Jan 2012 Compressor NDE X Compressor NDE Y
Root Cause Severe pivot wear in TPJ bearings Bearing clearance increased by 63+ microns O ring damper performance changed Damper film eccentricity ratio change (bottoming out) Damper Axial Pressure Profile Wear mark on pad Pivot wear Wear mark on housing Bearing Shell O-ring Top view Fretting damage on bearing OD Bottom view End Seals Tilt Pad Oil Inlet Casing O ring groove
Rotordynamic Model Shaft Radius, mm 400 200 0-200 20 25 30 35 40 45 50 55 1 5 10 15 60 65 68 69 70 Division wall -400 0 400 800 1200 1600 2000 24 Axial Location, mm
Baseline Mode Shapes with OEM Bearings Small log dec for baseline model (no SFD and no aero cross coupling stiffness from seals and impellers) Shaft Radius, mm 400 200 0-200 Cmin (minimum clearance) Cmin (minimum clearance) f=4400.1 cpm d=.2065 logd N=11456 rpm -400 0 400 800 1200 1600 2000 Axial Location, mm
Stability with OEM Bearings Level I stability predicts that the rotor is unstable without SFD 15000 3 63 8 121.932 2.13 07 / 19.64 0.78 12142 Shaft Radius, mm 400 200 0-200 Cmin (minimum clearance) f=4427.8 cpm d=-.4406 logd N=11456 rpm -400 0 400 800 1200 1600 2000 2 Axial Location, mm
Division Wall Seal Contribution 0.6 0.4 0.2 Existing Brg (Cmax) Existing Brg (Cmin) Estimated aero cross coupled stiffness Log Dec 0.0-0.2 da QA With DW (sw ratio=0.5) Insignificant improvement to stability with division wall hole pattern seal For stability, SFD required -0.4-0.6 Without Division Wall (DW) seal -0.8 0.0E+00 1.0E+07 2.0E+07 3.0E+07 Applied Cross Coupled Stiffness, Q N/m
O ring Squeeze Film Damper (SFD) Damper radial clearance (c) = 0.110 mm Damper radius (R) = 95.25 mm Effective damper length (L) = 37.85 mm Stability is very sensitive to damper eccentricity ratio ( ) Added O ring stiffness Damper stiffness and damping coefficients (without O ring stiffness): K=1.94E+07 N/m, C=1.88E+05 Ns/m at =0.25 K=6.06E+07 N/m, C=2.62E+05 Ns/m at =0.50 K=1.70E+09 N/m, C=2.06E+06 Ns/m at =0.90 Log dec of 1st mode 1.25 1.00 0.75 0.50 0.25 0.00-0.25-0.50 2 1 With Applied Cross Coupled K=2.13E+07 N/m OEM bearing with SFD OEM bearing without SFD -0.75 0.0 0.2 0.4 0.6 0.8 1.0 O-ring SFD eccentricity ratio 1 / Unstable
Pivot Wear Effects on Synchronous Vibration Pivot wear increased operating bearing clearance and reduced preload, resulting in increased synchronous vibrations A bearing with an SFD can make the rotor less sensitive to pivot wear than a bearing without an SFD Response, mm p-p 0.08 0.07 0.06 0.05 0.04 0.03 0.02 0.01 Pivot wear effect (without SFD) No pivot wear 0.012 mm pivot wear 0.025 mm pivot wear 0.038 mm pivot wear 0.051 mm pivot wear MinCOS RatedCOS MaxCOS Response, mm p-p 0.08 0.07 0.06 0.05 0.04 0.03 0.02 0.01 Pivot wear effect (with SFD) SFD:K=4.375E+07 N/m, C=2.625E+05 N s/m No pivot wear 0.013 mm pivot wear 0.025 mm pivot wear 0.038 mm pivot wear 0.051 mm pivot wear MinCOS RatedCOS MaxCOS 0 0 2000 4000 6000 8000 10000 12000 14000 Rotor Speed, rpm 0 0 2000 4000 6000 8000 10000 12000 14000 Rotor Speed, rpm
Bearing Upgrades (Journal Bearing) 1. Flexure Pivot Tilt Pad Journal Bearing No pivot wear Integral pivot Maintains bearing clearance High pivot stiffness No pivot stiffness effect on bearing dynamic coefficients Conventional Tilt Pad Journal Bearing (Rocker Back) Tight control of clearance and preload Electrical Discharge Machining (EDM) No pad flutter Flexure Pivot Tilt Pad Journal Bearing
Bearing Upgrades (Damper) 2. Integral Squeeze Film Damper (ISFD) Accurate stiffness control by eliminating O ring support No change in stiffness and damping over time Designed to counter static load Optimized damping Damper Film S spring Less cavitation 0.11 mm Original Damper Clearance Conventional SFD O ring 0.35 mm ISFD Clearance ISFD
Optimized stiffness and damping are 4.375 E+07 N/m (250,000 lb/in) and 2.625E+05 N s/m (1500 lb s/in) Additionally, the ISFD is designed to center the Flexure Pivot TPJ under gravity load by countering the static deflection Optimization of ISFD
Redesigned Bearing Optimized Original Design Optimized Design Conventional TPJ with SFD Flexure Pivot TPJ with ISFD Technology 5 pad, Load On Pad 4 pad, Load Between Pad Shaft diameter 114.300 +0/ 0.013 mm Shaft diameter 114.300 +0/ 0.013 mm Bearing bore 114.414 +0.025/ 0 mm Bearing bore 114.427 +0.025/ 0 mm Clearance range 0.124/0.156 mm Clearance range 0.124/0.156 mm Preload range 0.293/0.501 Preload range 0.230/0.273 L/D 0.444 L/D 0.500 Pad arc 60 Pad arc 72 Pivot Offset 60% Pivot offset 55%
1 st Mode Shapes: Original and Upgraded Bearings Shaft Radius, mm Applied destabilizing cross coupling stiffness of 2.13E+07 N/m to mid span Without SFD, unstable Both O ring SFD and ISFD can make the rotor stable No change in stiffness and damping of ISFD over time 400 200 0-200 Upgraded bearing (Flexure Pivot TPJ + ISFD) f=3594.5 cpm d=1.0362 logd N=11456 rpm Shaft Radius, mm Shaft Radius, mm 400 200 0-200 -400 400 200 0-200 OEM bearing w/out O ring SFD 0 400 800 1200 1600 2000 2 Axial Location, mm OEM bearing w/ O ring SFD f=4427.8 cpm d=-.4406 logd N=11456 rpm f=3624.7 cpm d=.3504 logd N=11456 rpm -400 0 400 800 1200 1600 2000 Axial Location, mm -400 0 400 800 1200 1600 2000 Axial Location, mm
No Subsynchronous Vibration (SSV) with Upgrade With OEM bearing (Acceptance test) Small SSV with O ring SFD No SSV with ISFD Upgrade bearing
Vibration Improvement (Comp A & B) Compressor A vibration dropped to less than half and maintained that level over time Compressor B vibration also down to below 50 µm from 90 µm (original) and kept the same level over time Vibration pk pk (microns) Vibration pk pk (microns) 100 80 60 40 20 100 80 60 40 20 C-1252 27,60 01B VIBRAÇÕES e DESLOCAMENTO AXIAL DO COMPRESSOR 25,10 D12 19,20 15,50 0 10050,60 100 100 85,40 58,60 90,70 0 0 0 0 24 Jul 2012 Compressor A with OEM bearing 0 2 Jun 2013 Compressor B with OEM bearing 12 Dec 2013 75,50 72,50 25,60 12,90 14 Mar 2013 21,90 22,40 14,10 25,30 Compressor A with Upgrade (4 month span) 8 Aug 2013 Compressor DE X Compressor DE Y Compressor NDE X Compressor NDE Y 15 Oct 2014 Compressor B with Upgrade (10 month span)
Vibration Improvement (Comp C) Compressor C was also upgraded with a Flexure Pivot TPJ with ISFD technology Again, the vibration level decreased to below 30 µm with the upgrade and maintained that same level due to no change in bearing clearance and SFD performance over time Vibration pk pk (microns) Vibration pk pk (microns) 100 80 60 40 20 0 8 Mar 2005 100 80 60 40 20 0 3 Jun 2013 Compressor C (vibration trend over 10 years) Compressor DE X Compressor DE Y Compressor NDE X Compressor NDE Y 30 Dec 2014 17 Apr 2014 14 Oct 2014 Compressor C with OEM bearing Compressor C with Upgrade
Summary Three reinjection compressor trains suffered from excessive vibration over many years Original configuration: point contact pivot tilt pad journal bearings with an O ring SFD The root cause was excessive pivot wear and degradation of the O ring SFD Pivot Wear mark on housing Fretting Damage Bearing bore increased Stiffness and damping changed over time
Summary The compressors were retrofitted with optimized Flexure Pivot tilt pad journal bearings with ISFD technology Operating exceptionally well Since 2013 Low vibration levels 50% drop pk pk compared to OEM bearings Do not grow over time No field balancing required so far (2 years) No trips (continuous production) No expensive bearing replacements Overall efficiency increased by 1%
Lessons Learned Increase in synchronous vibrations may be an indication of bearing clearance increasing from pivot wear and/or change in O ring damper performance Pivot wear may accelerate over time from increasing imbalance due to deposits on impellers Without eliminating pivot wear, just replacing the worn bearing with new build of the same design is NOT a long term solution Proper bearing and damper selection and optimization can reduce or eliminate the likelihood of increasing vibrations and pivot wear Flexure Pivot technology is a proven design to eliminate pivot wear ISFD technology maintains performance over time
Feedback and Questions Case Study: A Solution to Years of High Vibration Problems in Three Reinjection Compressor Trains Running at 33 MPa Discharge Pressure
Appendix: Damper Design Comparison Anti Rotation Pin Oil Damper Flow Oil Oil Oil Damper Flow Cs Shaft Cg S spring Cs Shaft Cg NO Circumferential Flow Bearing Squeeze Film (Outer Oil Film) bearing whirls or orbits (not spins) in a precessional motion due to synchronous (unbalance) or nonsynchronous excitation, squeezing the oil and thus generating an oil film pressure, and subsequently a damping force. Flow can be axial too, depending on sealing. Conventional SFD Damper Flow In/out of Orifices and Axial Gaps Integral Squeeze Film Damper
Authors Jong Kim, PhD, (+1 262.506.3055 jkim@waukbearing.com)is a Senior Principal Engineer at Waukesha Bearings Corporation, headquartered in Pewaukee, Wisconsin (USA) and a Senior Consulting Engineer at Bearings Plus, Houston, Texas (USA), which is a business unit of Waukesha Bearings. Dr. Kim has overall responsibilities for rotordynamic analysis, bearing upgrades and bearing/seal technologies including Flexure Pivot bearings, ISFD technology and brush seals. Dr. Kim joined Waukesha Bearings in 2011. Prior to joining Waukesha Bearings and since 2001, he worked for KMC and Bearings Plus. Dr. Kim received his Bachelor of Science (Mechanical Engineering, 1985) from Busan National University and both his Master of Science (Mechanical Engineering, 1987) and his PhD (Mechanical Engineering, 1991) from KAIST (Korea Advanced Institute of Science and Technology). He has authored several papers and has been granted multiple patents on bearing technologies. Marcio Felipe dos Santos (+55 092.3616.6614 tombodafumaca@gmail.com) is a Senior Maintenance Engineer at a major oil company in South America, in the Amazon. Since 2000, Mr. Dos Santos has provided maintenance in LPG plants and on reinjection compressors. Prior to joining his current company, he maintained drilling rigs. He has his Mechanical Engineering Degree from Universidade Federal of Rio de Janeiro (1979). Throughout his career Mr. Dos Santos has had maintenance duties. Barry J. Blair (+1 262.506.3043 bblair@waukbearing.com) is the Chief Engineer at Waukesha Bearings Corporation, headquartered in Pewaukee, Wisconsin (USA). Mr. Blair has overall responsibilities for research & development activities at Waukesha, including new products and overseeing the refinement of bearing design tools and methods. Mr. Blair joined Waukesha Bearings in 1993 and has served in increasingly responsible engineering and technology roles. Mr. Blair received both his Bachelor of Science (Mechanical Engineering, 1990) and Master of Science (Mechanical Engineering, 1990) from the University of Virginia, completing requirements of both degrees concurrently. He has authored and coauthored several papers on the development of both hydrodynamic and active magnetic bearing technologies.