EMAIL TIPS A Learning Publication from Full Spectrum Diagnostics
Balancing & Certification
Vol. 86 August 2012
EMAIL TIPS Volume 86 August 2012
New White Papers Available
2012 Training Schedules
The current Email Tips library includes a White Paper on the Vibration Analysis Periodic Table and a close-up view of the Table itself.
CORE TRAINING SERIES: • Intro (IVA) – VAI – VAII – VAIII CONCENTRATED TRAINING TRACK: • TWF – Spectrum – Phase – Bearings SPECIALTY & ADVANCED TRAINING • ODS/Modal – Precision Balancing • On-site Mentoring
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This Month’s Features: Severity Criteria for Large Machinery Applications The Inner Workings of Balance Quality Grades for the Vibration Analyst Certifications Requirements for ASNT and ISO – How We Compare
The gearing (below) comes from a shipyard crane application that is definitely not included in your typical manufacturing facility. As with many things in life, it’s the outliers that provide the biggest challenges. This month’s main topic concentrates on the problems categorizing larger scale equipment applications with respect to vibration severity
Specification Tip #1:
structure. The crane framework included a flexible beam network that allows the lighter high-strength
“Allowable” vibration levels are historically dependent on machinery class. There are various ANSI, ISO, and API Standards to help guide the analyst in this regard. Standard classifications are typically a function of machine size, power (hp, kW), speed and mounting conditions. The dominant limitation of this classification approach is in the vast amount of the equipment governed by these specifications. The majority of industrial equipment trends to a smaller size and scale and include rotating speeds that are much higher than the larger, slower speed classed systems. Along these lines, the typical specifications are tailored to concentrate on allowable residual unbalance in the machinery drive and base their criteria on the “overall” vibration (energy) levels. The specifications are not designed to determine machinery “specific” faults; rather they are more suited to defining adverse conditions that may exist that might require further investigation.
Class I - Individual parts of engines and machines, integrally connected with the complete machine in its normal operating condition. (Production electric motors up to 15 kW are typical examples of machines in this category.
All published specifications (ANSI, ISO, API) provide examples, but also include disclaimers. One such disclaimer notes that
Class III - Large Prime Movers and other large machines with rotating masses on rigid and heavy foundations which are relatively stiff in the direction of vibration measurement.
“amplitudes of vibration may be super-ceded by original equipment manufacturers (OEM) guidelines, or from experience with similar machinery, or when circumstances permit”.
Class IV - Large Prime Movers and other large machines with rotating masses on foundations which are relatively soft in the direction of vibration measurement (for example turbo-generator sets, especially those with lightweight substructures).
Such variables give the analyst significant leeway in choosing the most appropriate classification for the given machinery application and to its interpretation.
Class V - Machines and mechanical drive systems with unbalance able inertial effects (due to reciprocating parts), mounted on foundations which are relatively stiff in the direction of vibration measurement.
EXAMPLE FROM THE FIELD
Class VI - Machines and mechanical drive systems with unbalance able inertial effects (due to reciprocating parts), mounted on foundations which are relatively soft in the direction of vibration measurement; machines with rotating slack-coupled masses such as beater shafts in grinding mills; machines, like centrifugal machines, with varying unbalances capable of operating as self contained units without connecting components; vibrating screens, dynamic fatigue-testing machines and vibration exciters used in processing plants.
Recent field analysis on heavy lift crane systems suggested that this “unique or one-of-a-kind” equipment type does not readily fit into newer ISO specifications due to size, speed and operating characteristics. The closest applicable specification comes from a 1974 document; ISO 2372-1974E that provides some extended reference for larger machinery classes (summary for the machinery classes are provided in the image and table to the right). The chart has been converted from the original metric to imperial velocity amplitude units. The heavy-lift crane hoist systems would seem to fit into a Class III or Class IV category depending on the interpretation of the foundation stiffness. Based on operating analysis and natural frequency testing, the FSD interpretation is that the system is supported by a relatively “soft” backup frame
Class II - Medium-sized machines, (typically electric motors with 15 to 75 kW output) without special foundations, rigidly mounted engines or machines (up to 300 kW) on special foundations.
_________________________________________ structure to flex under load as not to develop local high stress points, as in more massive inflexible structures of past designs. This frame also includes highly damped natural frequencies near the dominant fundamental gear meshing frequency ranges which implies design flexibility. The most accurate classification for this machinery based on this specification was a Class IV designation.
The Severity Ranges (see the chart) for a Class IV machine are as follows: GOOD FAIR ALERT ALARM
0.000 to 0.156 ips-pk 0.156 to 0.395 ips-pk 0.395 to 1.000 ips-pk Above 1.000 ips-pk
As noted above, the Severity Ranges are typically based on acceptable levels of residual unbalance, not bearing, gear, or other rotor faults. If the dominant vibration in the hoist systems occurred at drive rotor speed and was attributed to residual unbalance, then any overall response exceeding the “Fair” range would be unacceptable. This was NOT the case with the example hoist systems. Residual Unbalance was NOT the dominant response; rather the gear meshing and alignment forces were the concern. This is why Vibration Analysis is so valuable when determining acceptability of machinery. Unlike overall trends, spectrum analysis can determine the frequency content and diagnose specific machinery problem(s)
Figure 1.0 Crane Machinery Deck
Operation of the higher speed drive motor in the “Alert” range would be considered problematic if the source of the vibration response was a Motor Unbalance condition. Considerable damage could result in this state. Operation in the “Alert” range was interpreted by FSD to be acceptable for the hoist systems due to their unique class of machinery, their size and speed, and that the dominant response was determined to be unrelated to shaft balance. The dominant vibration in the hoist systems was found to be gear-related vibration resulting from minor alignment error due to mounting conditions on a relatively flexible base. The “Alert” designation in this case would be to queue the analyst to investigate the potential failure path and attempt to reduce the vibration response accordingly. There is a fine line between designing a flexible system to accommodate stress and allow greater alignment tolerances and a more rigid system that strives to eliminate all relative motion in the system. This line is increasingly blurred when the size and power of the system is increased. As with all severity criteria the preferred approach to creating a criterion is to statistically evaluate a “family” of similar machines. In this case, the unique crane designs create “orphans” requiring some interpretation.
Motor Vibration Gearbox Vibration
Figure 2.0 Vibration Measurements from 6 Hoist Systems
Specification Tip #2: ISO 1940/ANSI 1975 Balancing Specs are intended for Rigid Rotors. Most motor rotors are rigid rotors (not the flexible type that operate at speeds above their first critical speed). A rigid rotor is one that can run from zero up to about 70% of its first critical speed. In that range it does not flex so the phase stays constant and the amplitude varies to the square of the speed (like the centrifugal force). Rotors with long narrow shafts, such paper machine rolls, turbines, generators, etc., run above their first critical speed so they flex or bend when the operating speed is near the first critical or the second critical speed. They cannot be balanced with simple static (force) corrections only because a couple is created also. So static-couple balancing procedures may need to be followed for those. In many cases a balancing grade (G) would not be applied in one correction plane, but be split over two correction planes. Since a portion of the weight of the rotor is supported by each bearing, it’s the balancing tolerance not the Grade that would be split at each bearing. The Balancing Grade and the rotor RPM is applied to the ISO 1940 Balancing Quality of Rotating Rigid Bodies chart to determine the eccentricity or [U/W = lb·in/lb]. Once the eccentricity has been determined it is multiplied by the fraction of rotor weight that each bearing supports. The result is the ISO Balancing tolerance in (oz·in) per bearing. An ISO recommended Grade of G6.3 for fans, pumps, process machinery, standard motor armatures (and rotors) is equivalent to an Imperial unit value of 0.351 in/sec Peak. Most experienced vibration analysts recognize that a vibration velocity amplitude above 0.300 in/sec Peak as being “slightly rough” or “rough”. Acceptance standards for NEW and REBUILT machines recommend that the tolerable vibration level for each bearing be about 1/3 of the vibration alarm (tolerance level) used for that machine. Therefore, Full Spectrum Diagnostics and others recommend that it would be preferable and reasonable to use an ISO Balancing Grade of between G2.4 and G1.0 would be a better ISO choice than G6.3 for modern machinery. Also, note that there is not a direct relationship between the (oz·in) of unbalance and the actual vibration response at a machine bearing in mils of displacement or in/s velocity. The "oz·in" value does
not take into account the ACTUAL RESPONSE of the machine at a bearing due to external forces such as bearing faults, bearing fits or other outside conditions. Such conditions may include hydro or aero-dynamic back pressure, structural stiffness, inertia, etc. Conversely, the passively measured housing vibration amplitudes measured in “mils” or “in/sec” is the ACTUAL RESPONSE. EXAMPLE FROM THE FIELD A trial weight is attached to a 1500 lb rotor system with an operating speed of 1800 RPM. The 5 oz. weight is mounted at a radius of 8 inches producing an unbalance effect, U= 40 oz·in. Independent of the speed of that rotor, the unbalance is always 40 oz·in. at all speeds. ISO Specification 1940 provides an Equation that can be used to calculate the Permissible Unbalance. Uper = [6.01] · [G] · [Rotor Wt @ one bearing] [RPM] So for this example assume a G 2.5 class. The permissible vibration would be: Uper = [6.01] · [2.5] ·  = 6.26 oz·in.  Now using the ISO Chart:
The ISO chart poin nts to an ec ccentricity of 0.00045 inches (lowest limitt for precise e balance for rotors in light strructures) at 1800 1 RPM so s Uper = 0.00045 x 750 7 lb. x 16 oz. o = 5.4 oz··in. Note: The T Balancin ng Grade G is the Maxim mum Limit in the ra ange for tha at grade: (e = .00 011 in. at 18 800 RPM, so o Uper= 0.00 011 x 750 x 16 = 13.2 oz.in n. for rotorrs in large massive housing gs, the uppe er limit of G2.5) It is ob bvious that that the equ uation is mo ore liberal than th he ISO Ch hart (giving g an avera age value betwee en the low lim mit and high limit of tolerrance that would be b determine ed on the IS SO chart). We W will use the cha art info. When we w check the moveme ent of that rotor r at a bearing g at a particu ular speed, we w are meassuring the actual response r of that machin ne in in/sec of o velocity or in mils m of displlacement. Iff we went through t a balancing procedu ure to correct the unba alance by determining what size s correctio on weight and a where to attacch it on the ro otor, we cou uld measure the effect of the trial t weight alone. On a simple polar graph, let's sa ay we measu ured the effe ect of the trrial weight alone to t be 0.6 in n/s. This gives us the ability to calculatte the Rotorr Sensitivity y. SENSIT TIVITY = 40 oz.in.divide ed by 0.6 in/ss = 66.67 oz.in./ in n/s “o or” 1 oz--in = 0.015 5 in/s. Therefo ore, for our lower limit Balancing Tolerance of 5.4 oz-iin, the vibrattion velocity needs to be e 0.015 5 x 5.4 = 0.0 081 in/s or le ess. If the ve elocity readss above 0.08 8 in/s in this case, the amountt of unbalace e is excessiv ve at that be earing. So the ISO 1940/A ANSI S2.19-1975 Balancce Quality Gradess for Rigid Ro otors Chart offers o the fo ollowing: G 6.3 G 2.5 G 1.0
6.3 mm/s RMS R = 0.351 in/s Peakk 2.5 mm/s RMS R = 0.139 in/s Peakk 1.0 mm/s RMS R = 0.056 in/s Peakk
0.081 in n/s falls betw ween G 2.5 and a G 1.0. At this point it see ems reason nable that th he reader sign-up p for one of Lou Pagliaro’s P P Precision Balanc cing coursess.
Spec cification Tip #3: Certificati C ion ments with Vibratiion Analysiss Certification requirem respecct to th he Ameriican Soc ciety for Nonde estructive Testing an nd the International Standards Orga anization are spelled d out in govern ning specificcations thatt include AS SNT SNTTC-1A A and ISO 18436-1 and a 18436-2 2. These specifiications co over the topics and d training duratio ons that a candidate vibration v analyst must comple ete for a given certtification levvel. The specifiications also o outline ma any requirem ments that must be b included in i the certificcation exam and exam processs. Details include the number of questions, passin ng grades, and statisttical evalua ations that insure that the que estions crea ate a fair and d unbiased test. wing an exttensive inte ernal review w the Full Follow Spectrrum Diagno ostics Vibra ation Analyssis course conten nt, topical outlines, o and d training hours h were compa ared to the re equirementss of the ASN NT and ISO specifiications. FS SD believes that their certification c processs covers both the e ASNT and ISO require ements per their Written n Practice and a Quality Assura ance Progrram. A co omparison of course conten nt for the Vibration Analysis A I course is provided on the ne ext page. Additional A co omparisons of courses and tra aining hours for the FSD classes to the ISO and ASNT req quirements are also presen nted.
FSD Course Content Evaluation for VA-1
ASNT & ISO Course Comparison for FSD Training
REMAINING: 2012 CORE VIBRATION TRAINING SCHEDULE IMPLEMENTING A SUCCESSFUL PdM PROGRAM
REMAINING: 2012 SPECIALTY VIBRATION TRAINING SCHEDULE MODAL & ODS ANALYSIS 2 IN-PLANT COURSES AVAILABLE
IN-PLANT COURSES AVAILABLE Tuition: $ 945.00 / 2-day Format Proficiency Test: Included
For In-Plant Training: ModalGuy@aol.com Tuition: $ 1,595.00 / 3-day Format One Month ME’scope Software Included
INTRODUCTION TO VIBRATION ANALYSIS
CONC TIME WAVEFORM ANLAYSIS
IN-PLANT COURSES AVAILABLE
Tuition: $ 945.00 / 2-day Format Proficiency Test: Included
Tuition: $ 945.00 / 2-day Format Proficiency Test: Included
CONCENTRATED SPECTRUM ANALYSIS
VIBRATION ANALYSIS I SEP 11-14 DEC 11-14 DEC 18-21
NOV 27-28 Chillicothe, OH 2012-TWF-02
OCT 02-03 St. Paul, MN
Chicago, IL 2012-VA1-04 Charlotte, NC 2012-VA1-05 New Orleans, LA 2012-VA1-06
Tuition: $ 1,295.00 / 3-day Format Certification Exam: $ 200.00
Tuition: $ 945.00 / 2-day Format Proficiency Test: Included
CONCENTRATED PHASE ANALYSIS SEP 18-19 Davenport, IA 2012-CPA-01 Tuition: $ 945.00 / 2-day Format Hands-on Exercises: Included
VIBRATION ANLAYSIS II SEP 25-28 Cedar Rapids, IA 2012-VA2-06
CONCENTRATED RE BEARING & GEAR ANLAYSIS
Tuition: $ 1,395.00 / 3-day Format Certification Exam: $ 200.00
OCT 16-17 Cedar Rapids, IA 2012-BGA-01 Tuition: $ 945.00 / 2-day Format Proficiency Test: Included
VIBRATION ANALYSIS IIIa OCT 23-26 Leesport, PA
Tuition: $ 1,595.00 / 4-day Format Certification Exam: $ 200.00
PRACTICAL VIBREATION ANALYSIS IIIb AUG 21-24 St. Paul, MN 2012-PVA-01 Tuition: $ 1,595.00 / 4-day Format Certification Exam: $ 200.00
VIBRATION ANALYSIS FOR MOTOR SHOPS AUG 28-30 Charlotte, NC 2012-VAM-01 DEC 04-06 St. Paul, MN 2012-VAM-02 Tuition: $ 1,295.00 / 3-day Format Hands-on Exercises: Included
!!! NEW !!!
AUG 07-08 Brookfield, WI 2012-BAL-03 NOV 13-14 Wheeling, WV 2012-BAL-04 Tuition: $ 945.00 / 2-day Format Hands-on Exercises: Included
THE VIBRATION FAULT GUIDE The Vibration Fault Guide is a 110-page indispensable asset for every vibration analyst, as well as a helpful tool to bridge the gap between the analyst and his management staff. Some 45 Rotating Machinery Faults are Identified and defined. Conveniently designed at 3” x 6”, it easily fits in your pocket for everyday use!
GUÍA DE FALLAS DE VIBRACIÓN Este manual consta de 110 páginas y fue compilado por Full Spectrum Diagnostics como referencia rápida para la industria de Mantenimiento Predictivo y Monitoreo de Condición. La guía incluye ejemplos de espectros, formas de onda, definiciones de fallos y reglasde análisis de fase para aproximadamente unos 45 problemas que se pueden presentar en maquinaria rotativa. También incluimos varios estándares de especificaciones de vibración, guías para definición de bandas de alarmas, fórmulas y definiciones de Procesamiento de Señales. Es suficientemente pequeño como para llevarlo en la bolsa de su camisa ( 3.5” X 6.0”), pero su contenido es tan grande que podría ser considerado como una guía de referencia esencial para la industria del Monitoreo de Condición.
THE VIBRATION TECHNIQUES GUIDE The Vibration Analysis Techniques Guide is a 108-page pocket sized information treasure trove. Information on dozens of analysis techniques, specifications and data presentation formats are included. If you liked the Vibration Fault Guide, your next educational step should be the Vibration Analysis Techniques Guide get yours now at:
Order now at: http://www.fullspec.net/store.html Or by Phone @ (763) 577-9959
BULK DISCOUNTS & BUNDLE PACKAGES AVAILABLE!
THE VIBRATION ANALYSIS WALL CHART The Vibration Analysis Wall Chart is a 46” x 36” Full-Color Laminated Reference for your drab office wall. The overall Alarm charts in the center of the chart are surrounded by groupings of over 40 dominant rotating machinery faults. The fault groupings include frequency content and dominant directional response that effectively allow the analyst to “narrow-down” the potential sources and zero-in on current vibration problem. When combined with the Vibration Fault Guide, the Vibration Analysis Wall Chart completes the diagnostic analysis loop. Weather you require one chart or a need to wall-paper your office, we can help!
Order now at: Http://www.fullspec.net/store.html Or by Phone @ (763) 577-9959
BULK DISCOUNTS & BUNDLE PACKAGES AVAILABLE!
THE VIBRATION ANALYSIS PERIODIC TABLE This Full Color Laminated 13 x 17 inch card-stock table provides a “quick-look” method of distinguishing one machinery fault from another and suggests Diagnostic Tests or formula that may be used to build a case and make the call! This Chart is a “Logical” analysis tool that classifies vibration problems by Frequency Content and Directional Response. The potential vibration sources are instantly reduced based on the analyst’s current measurement data. The Table “forces” the user to think logically and classify faults accordingly. Individual Faults are Foot Note Referenced to the Vibration Fault Guide for a more detailed review. If you think the Vibration Fault Guide is valuable, your next educational step should be the Vibration Analysis Periodic Table! Get yours now at: www.fullspec.net, or by phone at 763-577-9959
BULK DISCOUNTS & BUNDLE PACKAGES AVAILABLE!
Full Spectrum Diagnostics’ Lead Instructor & Seminar Author is Dan Ambre. Dan is a graduate of The University of Iowa with a Bachelor’s degree in Mechanical Engineering, and has completed additional graduate level course work in Engineering Dynamics from The University of Illinois at Chicago, and Florida Atlantic University. Dan is a Certified Vibration Analysis Level III Instructor (Category IV) with over 16 years of Vibration Training and Certification Experience. Dan’s 25+ years of vibration experience in the Aviation & Aerospace Industries comes from positions at Sundstrand Aviation Corporation and Pratt & Whitney (United Technologies Corporation). This fieldwork includes Vibration & Acoustic testing, Rotor Dynamics analysis of high speed Rotor Systems, Experimental Modal, and Finite Element Analysis. His consulting experience base comes from positions at Technical Associates and Full Spectrum Diagnostics, which he founded in June of 2000. He is a Registered Professional Engineer in the States of North Carolina and Minnesota.
In 2004, after years of close association, Louis G. Pagliaro joined Full Spectrum Diagnostics. Lou fulfills multiple roles as our Seminar Sales Coordinator, Senior Instructor and course content co-author. Lou is a certified Level III Vibration Analyst (since 1996) and was recently re-certified by American Society of Nondestructive Testing as an ASNT PdM Level III in Vibration Analysis (Category IV). He has over 30 years of varied industrial experience, including Maintenance Management, Vibration Analysis, Vibration Training, Certification, and course development in areas of Noise Control, Precision Maintenance, Precision Alignment, Preventive Maintenance, and Maintenance Skills Enhancement. His worldwide teaching credentials include instruction of TAC, Update International, CSI, Entek and SKF customers in eight countries. Louis is a graduate of Niagara College in Welland, Canada. Lou can be contacted by phone at (704) 577-3953, or via Email at Lpagl@aol.com.
Dan can be contacted by phone at (612) 875-9959, or via Email at ModalGuy@aol.com