AGMA6011-I03_Specification for High Speed Helical Gear Units

ANSI/AGMA 6011- I03 (Revision of ANSI/AGMA 6011--H98)

ANSI/AGMA 6011- I03

Specification for High Speed Helical Gear Units

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AMERICAN NATIONAL STANDARD

American National Standard

Specification for High Speed Helical Gear Units ANSI/AGMA 6011--I03 [Revision of ANSI/AGMA 6011--H98] Approval of an American National Standard requires verification by ANSI that the requirements for due process, consensus, and other criteria for approval have been met by the standards developer. Consensus is established when, in the judgment of the ANSI Board of Standards Review, substantial agreement has been reached by directly and materially affected interests. Substantial agreement means much more than a simple majority, but not necessarily unanimity. Consensus requires that all views and objections be considered, and that a concerted effort be made toward their resolution.

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The use of American National Standards is completely voluntary; their existence does not in any respect preclude anyone, whether he has approved the standards or not, from manufacturing, marketing, purchasing, or using products, processes, or procedures not conforming to the standards. The American National Standards Institute does not develop standards and will in no circumstances give an interpretation of any American National Standard. Moreover, no person shall have the right or authority to issue an interpretation of an American National Standard in the name of the American National Standards Institute. Requests for interpretation of this standard should be addressed to the American Gear Manufacturers Association. CAUTION NOTICE: AGMA technical publications are subject to constant improvement, revision, or withdrawal as dictated by experience. Any person who refers to any AGMA technical publication should be sure that the publication is the latest available from the Association on the subject matter. [Tables or other self--supporting sections may be referenced. Citations should read: See AGMA AGMA 6011--I03, Specification for High Speed Helical Gear Units, published by the American Gear Manufacturers Association, 500 Montgomery Street, Suite 350, Alexandria, Virginia 22314, http://www.agma.org.]

Approved February 12, 2004

ABSTRACT This standard includes design, lubrication, bearings, testing and rating for single and double helical external tooth, parallel shaft speed reducers or increasers. Units covered include those operating with at least one stage having a pitch line velocity equal to or greater than 35 meters per second or rotational speeds greater than 4500 rpm and other stages having pitch line velocities equal to or greater than 8 meters per second. Published by

American Gear Manufacturers Association 500 Montgomery Street, Suite 350, Alexandria, Virginia 22314 Copyright  2003 by American Gear Manufacturers Association All rights reserved. No part of this publication may be reproduced in any form, in an electronic retrieval system or otherwise, without prior written permission of the publisher.

Printed in the United States of America ISBN: 1--55589--819--X

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AMERICAN NATIONAL STANDARD

ANSI/AGMA 6011--I03

Contents Page

Foreword . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . iv 1 Scope . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1 2 Normative references . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1 3 Symbols, terminology and definitions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1 4 Design considerations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3 5 Rating of gears . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7 6 Lubrication . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9 7 Vibration and sound . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12 8 Functional testing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15 9 Vendor and purchaser data exchange . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17 Bibliography . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 51

Annexes Service factors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A simplified method for verifying scuffing resistance . . . . . . . . . . . . . . . . . . . . . Lateral rotor dynamics . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Systems considerations for high speed gear drives . . . . . . . . . . . . . . . . . . . . . Illustrative example . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Efficiency . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Assembly designations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Purchaser’s data sheet . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

21 24 26 32 41 44 47 48

Figures 1

Amplification factor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14

Tables 1 2 3 4 5 6 7

Symbols used in equations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2 Recommended accuracy grades . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3 Recommended maximum length--to--diameter (L/d) ratios . . . . . . . . . . . . . . . . . 4 Hydrodynamic babbitt bearing design limits . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6 Dynamic factor as a function of accuracy grade . . . . . . . . . . . . . . . . . . . . . . . . . 8 Recommended lubricants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10 Casing vibration levels . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15

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A B C D E F G H

ANSI/AGMA 6011--I03

AMERICAN NATIONAL STANDARD

Foreword [The foreword, footnotes and annexes, if any, in this document are provided for informational purposes only and are not to be construed as a part of ANSI/AGMA Standard 6011--I03, Specification for High Speed Helical Gear Units.] The first high speed gear unit standard, AGMA 421.01, was adopted as a tentative standard in October, 1943. It contained formulas for computing the durability horsepower rating of gearing, allowable shaft stresses, and included a short table of application factors. AGMA 421.01 was revised and adopted as a full status standard in September, 1947 and issued as AGMA 421.02. The High Speed Gear Committee began work on the revision of AGMA 421.02 in 1951, which included: classification of applications not previously listed; changing the application factors from “K” values to equivalent Service Factors; revision of the rating formula to allow for the use of heat treated gearing; and develop a uniform selection method for high speed gear units. This Uniform Selection Method Data Sheet became AGMA 421.03A. AGMA 421.03 was approved as a revision by the AGMA membership in October, 1954. The standard was reprinted as AGMA 421.04 in June, 1957. It included the correction of typographical errors and the addition of a paragraph on pinion proportions and bearing span, which had been approved by the committee for addition to the standard at the October, 1955 meeting. In October, 1959 the Committee undertook revisions to cover developments in the design, manufacture, and operation of high speed units with specific references to high hardness materials and sound level limits. The revisions were incorporated in AGMA 421.05 which was approved by the AGMA membership as of October 22, 1963. The significant changes of 421.06 from 421.05 were: minimum pitch line speed was increased to 5000 feet per minute (25 meters per second); strength and durability ratings were changed; and some service factors were added. AGMA 421.06 was approved by the High Speed Gear Committee as of June 27, 1968, and by the AGMA membership as of November 26, 1968. ANSI/AGMA 6011--G92 was a revision of 421.06 approved by the AGMA membership in October, 1991. The most significant changes were the adaptation of ratings per ANSI/AGMA 2001--B88 and the addition of normal design limits for babbitted bearings. ANSI/AGMA 6011--G92 used “application factor” and not “service factor”. ANSI/AGMA 6011--H98 was a further refinement of ANSI/AGMA 6011--G92. One of the most significant changes was the conversion to an all metric standard. The rating methods were changed to be per ANSI/AGMA 2101--C95 which is the metric version of ANSI/AGMA 2001--C95. To provide uniform rating practices, clearly defined rating factors were included in the standard (ANSI/AGMA 6011--H98). While some equations may slightly change to conform to metric practices, no substantial changes were made to the rating practice for durability and strength rating. In addition, minimum pitch line velocity was raised from 25 m/s to 35 m/s and minimum rotational speed increased to 4000 rpm. AGMA has reverted to the term “service factor” in their standards, which was reflected in ANSI/AGMA 6011--H98. The service factor approach is more descriptive of enclosed gear drive applications and can be defined as the combined effects of overload, reliability, desired life, and other application related factors. The service factor is applied only to the gear tooth rating, rather than to the ratings of all components. Components are designed based on the service power and the guidelines given in this standard. In continued recognition of the effects of scuffing in the rating of the gear sets, additional information on scuffing resistance was added to annex B of ANSI/AGMA 6011--H98. iv Copyright American Gear Manufacturers Association Provided by IHS under license with AGMA No reproduction or networking permitted without license from IHS

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AMERICAN NATIONAL STANDARD

ANSI/AGMA 6011--I03

AGMA 427.01 has been withdrawn. The information found in AGMA 427.01 was included in annex D of ANSI/AGMA 6011--H98. ANSI/AGMA 6011--I03 is a further refinement to ANSI/AGMA 6011--H98. Symbols have been changed where possible to conform with ANSI/AGMA 2101--C95 and ISO standards. The minimum rotational speed has been increased to 4500 rpm. Helix angle limits have changed, and a minimum axial contact ratio limit has been added. The L/D limits have changed, and use of modified leads is now encouraged with the use of predicted rotor deflection and distortion. Bearing load design limits have also changed. For gear tooth accuracy, reference is now made to ANSI/AGMA 2015--1--A01 rather than to ANSI/AGMA 2000--A88. The Zn and Yn life factors now have a maximum rather than a minimum limit when the number of load cycles exceeds 1010. A table of dynamic factor as a function of accuracy grade has been added. References to AGMA oil grades have been removed; now only ISO viscosity grades are listed. To facilitate communications between purchaser and vendor, an annex with data sheets has been added. Realistic evaluation of the various rating factors of ANSI/AGMA 6011--I03 requires specific knowledge and judgment which come from years of accumulated experience in designing, manufacturing and operating high speed gear units. This input has been provided by the AGMA High Speed Gear Committee.

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The first draft of AGMA 6011--I03 was made in May, 2001. It was approved by the AGMA membership in October, 2003. It was approved as an American National Standard on February 12, 2004. Suggestions for improvement of this standard will be welcome. They should be sent to the American Gear Manufacturers Association, 500 Montgomery Street, Suite 350, Alexandria, Virginia 22314.

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ANSI/AGMA 6011--I03

AMERICAN NATIONAL STANDARD

PERSONNEL of the AGMA Helical Enclosed Drives High Speed Unit Committee Chairman: John B. Amendola . . . . . . . . . . . . . . . . . . . . . . . . . MAAG Gear AG

ACTIVE MEMBERS E. Martin . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Lufkin Industries, Inc. J.M. Rinaldo . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Atlas Copco Compressors, Inc. W. Toner . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Siemens Demag Delaval Turbomachinery, Inc. --`,``,`,`,`,``````,`,``,,`,,`-`-`,,`,,`,`,,`---

ASSOCIATE MEMBERS A. Adams . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . K.O. Beckman . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.S. Cohen . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . W. Crosher . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . G.A. DeLange . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . H. Ernst . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . R. Gregory . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . M. Hamilton . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . L. Hennauer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . O.A. LaBath . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . L. Lloyd . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . M.P. Starr . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . F.A. Thoma . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . F.C. Uherek . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . U. Weller . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . D.G. Woodley . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

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Textron Power Transmission Lufkin Industries, Inc. Engranes y Maquinaria Arco, S.A. Flender Corporation Hansen Transmissions HSB Turner Uni--Drive Company Flender Graffenstaden BHS Getriebe GmbH Gear Consulting Services of Cincinnati, LLC Lufkin Industries, Inc. Falk Corporation F.A. Thoma, Inc. Flender Corporation MAAG Gear AG Shell Oil Products U.S.

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AMERICAN NATIONAL STANDARD

ANSI/AGMA 6011--I03

American National Standard --

2 Normative references

Specification for High Speed Helical Gear Units

The following standards contain provisions which, through reference in this text, constitute provisions of this American National Standard. At the time of publication, the editions indicated were valid. All standards are subject to revision, and parties to agreements based on this American National Standard are encouraged to investigate the possibility of applying the most recent editions of the standards indicated below.

This high speed helical gear unit standard is provided as a basis for improved communication regarding: --

establishment of uniform criteria for rating;

--

guidance for design considerations; and,

-- identification of the unique features of high speed gear units. 1.1 Application Operational characteristics such as lubrication, maintenance, vibration limits and testing are discussed. This standard is applicable to enclosed high speed, external toothed, single and double helical gear units of the parallel axis type. Units in this classification are: -- single stage units with pitch line velocities equal to or greater than 35 meters per second or rotational speeds greater than 4500 rpm; -- multi--stage units with at least one stage having a pitch line velocity equal to or greater than 35 meters per second and other stages having pitch line velocities equal to or greater than 8 meters per second. Limits specified are generally accepted design limits. When specific experience exists for gear units of similar requirements above or below these limits, this experience may be applied. Marine propulsion, aircraft, automotive, and epicyclic gearing are not covered by this standard.

ANSI/AGMA 1010--E95, Appearance of Gear Teeth -- Terminology of Wear and Failure ANSI/AGMA 2015--1--A01, Accuracy Classification System -- Tangential Measurements for Cylindrical Gears ANSI/AGMA 2101--C95, Fundamental Rating Factors and Calculation Methods for Involute Spur and Helical Gear Teeth ANSI/AGMA 6000--B96, Specification for Measurement of Linear Vibration on Gear Units ANSI/AGMA 6001--D97, Design and Selection of Components for Enclosed Gear Drives ANSI/AGMA 6025--D98, Sound for Enclosed Helical, Herringbone, and Spiral Bevel Gear Drives ISO 14635--1, Gears – FZG test procedures – Part 1: FZG test method A/8,3/90 for relative scuffing load carrying capacity of oils

3 Symbols, terminology and definitions 3.1 Symbols The symbols used in this standard are shown in table 1. NOTE: The symbols and terms contained in this document may vary from those used in other AGMA standards. Users of this standard should assure themselves that they are using these symbols and terms in the manner indicated herein.

1

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1 Scope

ANSI/AGMA 6011--I03

AMERICAN NATIONAL STANDARD

Table 1 -- Symbols used in equations Symbol A Act AF CSF CRE cp DJ d Fd Ft KB KH KHe KHma KHmc KHpm Ks KSF Kv L Ncm Ncp Nct Nmc nL Pa Payu Pazu PL PS QLUBE SJ SM Umax W Wcpl Wr YN Yθ ZN ZR ZW ∆T σFP σHP

Term Allowable double amplitude of unfiltered vibration Amplitude at Nct Amplification factor Service factor for pitting resistance Critical response envelope Specific heat of lubricant Nominal bearing bore diameter Pinion operating pitch diameter Incremental dynamic load Transmitted tangential load Rim thickness factor Load distribution factor Mesh alignment correction factor Mesh alignment factor Lead correction factor Pinion proportion modifier Size factor Service factor for bending strength Dynamic factor Face width including gap Initial (lesser) speed at 0.707 × peak amplitude (critical) Final (greater) speed at 0.707 × peak amplitude (critical) Rotor first critical, center frequency Maximum continuous rotor speed Number of stress cycles Allowable transmitted power for the gear set Allowable transmitted power for bending strength at unity service factor Allowable transmitted power for pitting resistance at unity service factor Power loss Service power of enclosed drive Lubricant flow Diametral clearance Separation margin Amount of residual rotor unbalance Journal static loading Half weight of coupling and spacer Total weight of rotor Stress cycle factor for bending strength Temperature factor Stress cycle factor for pitting resistance Surface condition factor for pitting resistance Hardness ratio factor for pitting resistance Change in lubricant temperature Allowable bending stress number Allowable contact stress number

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Units mm mm -- --- -rpm kJ/(kg°C) mm mm N N -- --- --- --- --- --- --- --- --- -mm rpm rpm rpm rpm -- -kW kW kW kW kW kg/sec mm rpm g--mm kg kg kg -- --- --- --- --- -_C N/mm2 N/mm2

Reference paragraph 7.5 7.3.3.3 7.3.3.3 5.2 7.3.3.3 8.2.5 Table 4 4.6 5.3.3 5.3.3 5.4 5.3.2 5.3.2 5.3.2 5.3.2 5.3.2 5.3 5.2 5.3.3 4.6 7.3.3.3 7.3.3.3 7.3.3.3 4.1 5.3.1 5.1 5.1 5.1 8.2.5 4.1 8.2.5 Table 4 7.3.3.3 7.4 7.4 7.3.3.2 7.3.3.2 5.4.1 5.3 5.3.1 5.3 5.3 8.2.5 5.5 5.5

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AMERICAN NATIONAL STANDARD

3.2 Nomenclature

The terms used, wherever applicable, conform to the following standards: AGMA 904--C96, Metric Usage ANSI/AGMA 1012--F90, Gear Nomenclature, Definitions of Terms with Symbols ISO 701, International gear notation – Symbols for geometrical data

ANSI/AGMA 6011--I03

All components shall be capable of transmitting the service power. 4.2 High transient torque levels Where unusual torque variations develop peak loads which exceed the application power by a ratio greater than the service factor, CSF or KSF, specified for the application, the magnitude and frequency of such torque variations should be evaluated with regard to the endurance and yield properties of the materials used. See annex D and also ANSI/AGMA 2101--C95, subclause 16.3. 4.3 Torsional and lateral vibrations

This standard should be used in conjunction with appropriate current AGMA standards. External loads must be considered as acting in directions and rotations producing the most unfavorable stresses unless more specific information is available. Allowances must be made for peak loads. 4.1 Service power, PS Service power of an application is defined as the maximum installed continuous power capacity of the prime mover, unless specifically agreed to by the purchaser and vendor. For example, for electric motors, maximum continuous power will be the motor nameplate power rating multiplied by the motor service factor. For gear units between two items of driven equipment, service power of such gears should normally not be less than item (a) or (b) below, whichever is greater. a. 110 percent of the maximum power required by the equipment driven by the gear; b. maximum power of the driver prorated between the driven equipment, based on normal power demands. If maximum torque occurs at a speed other than maximum continuous speed, this torque and its corresponding speed shall be specified by the purchaser. Maximum continuous speed, Nmc, is normally the speed at least equal to 105% of the specified (or nominal) pinion speed for variable speed units and is the rated pinion speed for constant speed units.

When an elastic system is subjected to externally applied, cyclic or harmonic forces, the periodic motion that results is called forced vibration. For the systems in which high speed gears are typically used, two modes of vibration are normally considered. a) Lateral or radial vibration, which considers shaft dynamic motion that is in a direction perpendicular to the shaft centerline; and b) Torsional vibration, which considers the amplitude modulation of torque measured peak to peak referenced to the axis of rotation. In certain cases, axial or longitudinal vibration might also be considered. Because of the wide variation of gear driven systems, clause 7 of this standard outlines areas where proper assessment of the system may be necessary. In addition, appropriate responsibility between the vendor and purchaser must be clearly delineated. 4.4 Tooth proportions and geometry Any practical combination of tooth height, pressure angle and helix angle may be used. However, it is recommended that the gears have a minimum working depth of 1.80 times the normal module, a maximum normal pressure angle of 25 degrees, a helix angle of 5 to 45 degrees, and a minimum axial contact ratio of 1.1 per helix. 4.5 Recommended accuracy grade Table 2 presents recommended ANSI/AGMA 2015--1--A01 accuracy grades as a function of pitch line velocity. Based on experience and application, other accuracy grades may be appropriate.

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4 Design considerations

AMERICAN NATIONAL STANDARD

Table 2 -- Recommended accuracy grades Pitch line velocity, m/s 35 –100 100 – 160 Over 160

ANSI/AGMA 2015--1--A01 accuracy grade A5 A4 A3

4.6 Pinion proportions Table 3 presents maximum length--to--diameter (L/d) ratios for material hardening methods in current use. The L/d values shown in table 3 apply to helical gears when designed to transmit the service power. Generally, higher L/d ratios are permitted when analytical load distribution methods are employed that yield load distribution values, KH, that are less than the value calculated per 5.3.2 at the maximum L/d ratio per table 3. A detailed analytical method should include, but not be limited to, bending and torsional deflection and thermal distortion. Table 3 – Recommended maximum length--to-diameter (L/d) ratios Hardening method Through hardened Case hardened

Maximum L/d ratio Double Single helical helical 2.2 1.6 2.0 1.6

NOTE: L = face width including gap, mm; d = pinion operating pitch diameter, mm

No matter what the L/d ratio is, if the combination of tooth and rotor deflection and distortion exceeds 25 mm for through hardened gears, or 18 mm for case hardened gears, then an analytically determined lead modification should be applied in order to reduce the total mismatch to a magnitude below these values. Determination of the combined tooth and rotor deflection shall be based on the service power. The modification is intended to provide a uniform load distribution across the entire face width. Working flanks of the pinion or gear wheel should be modified when necessary to compensate for torsional and bending deflections and thermal distortion. Gears with pitch line velocities in excess of 100 m/s are particularly susceptible to thermal distortion. Consideration should be given to the relationship of lead modifications to gear tooth accuracy. When a higher L/d ratio than tabulated in table 3 is proposed, the gear vendor shall submit justification

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in the proposal for using the higher L/d ratio. Purchasers should be notified when L/d ratios exceed those in table 3. When operating conditions other than gear rated power are specified by the purchaser, such as the normal transmitted power, the gear vendor shall consider in the analysis the length of time and load range at which the gear unit will operate at each condition so that the correct lead modification can be determined. When modified leads are to be furnished, purchaser and vendor shall agree on the tooth contact patterns obtained in the checking stand, housing or test stand. 4.7 Rotor construction Several configurations may be applied in the construction of rotors. The most commonly used are listed below: a) Integral shaft and gear element. This configuration is commonly used for pinions, smaller gears, or rotating elements operating above a pitch line velocity of 150 meters per second. The pinion or gear, integral with its shaft, is machined from a single blank; b) Solid blank shrunk on a shaft. The shrink fit may be used either with or without a mechanical torque transmitting device (such as key or spline). When no torque transmitting device is used, the shrink fit must provide ample capacity to transmit torque when considering centrifugal and thermal effects. When a torque transmitting device is used, the shrink fit must provide ample location support when considering centrifugal and thermal effects; c) Fabricated gear. A forged rim is welded directly to the fabricated substructure producing a one--piece welded gear. The shaft may be a part of the weldment. Fabricated gears should be analyzed to consider centrifugal and thermal stresses and fatigue life. Maximum pitch line velocity for welded gear construction is 130 meters per second; d) Forged rim shrunk onto a substructure. The substructure may be forged, cast, or fabricated. The shaft may be a part of the substructure. Shrunk rims shall consider stresses and torque transmitting capacity due to fit, centrifugal, and thermal effects (refer to item b). The normal design limit for this type of construction is 60 meters per second. Combinations of the above are often used on multistage units.

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ANSI/AGMA 6011--I03

AMERICAN NATIONAL STANDARD

4.8 Gear housing The gear housing should be designed to provide a sufficiently rigid enclosed structure for the rotating elements that enables them to transmit the loads imposed by the system and protects them from the environment in which they will operate. The vendor’s design of the housing must provide for proper alignment of the gearing when operating under the user’s specified thermal conditions, and the torsional, radial and thrust loadings applied to its shaft extensions. In addition, it should be designed to facilitate proper lubricant drainage from the gear mesh and bearings. The user’s design of the supporting structure must maintain proper and stable alignment of the gearing. Alignment must consider all specified torsional, radial and thrust loadings, and thermal conditions present during operation. 4.8.1 Special housing considerations Certain applications may be subjected to operating conditions requiring special consideration. Some of these operating conditions are: -- temperature variations in the vicinity of the gear unit; -- relative thermal growth between mating system components; -- environmental elements that will attack the unit housing, rotating components, bearings or lubricant; --

inadequate support for the housing;

-- high pitch line velocities which may affect lubricant distribution, create excessive temperature rise, or cause other adverse conditions.

4.8.2 Shaft seals Where shafts pass through the housing, the housings shall be equipped with seals and deflectors that shall effectively retain lubricant in the housing and prevent entry of foreign material into the housing. Easily replaceable labyrinth--type end seals and deflectors are preferred. The seals and deflectors shall be made of nonsparking materials. Lip--type seals have a very finite life and can generate enough heat at higher speeds to be a fire hazard. Surface velocity should be kept within the seal manufacturer’s conservative recommendation. 4.9 Bearings Proper design of bearings is critical to the operation of high speed enclosed drive units. The bearing design shall consider normal service power. Radial bearings are normally of the hydrodynamic sleeve or pad type. Thrust bearings are usually flat land, tapered land, or thrust pad type. Rolling element bearings are occasionally used when speeds are at the very low end of the high speed range. Bearing design shall consider start--up and unloaded conditions, as well as normal service power. 4.9.1 Hydrodynamic bearings Hydrodynamic bearings shall be lined with suitable bearing material. Tin and lead based babbitts (white metal) are among the most widely used bearing materials. Tin alloy is usually preferred over lead alloys because of its higher corrosion resistance, easier bonding, and better high temperature characteristics. Hydrodynamic bearings shall have a rigid steel or other suitable metallic backing, and be properly installed and secured in the housing against axial and rotational movement. Bearings are generally supplied split for ease of assembly. Selection of the particular design (sleeve, pad type or land bearing) shall be based on evaluation of surface velocity, surface loading, hydrodynamic film thickness, calculated bearing temperature, lubricant viscosity, lubricant flow rate, and bearing stability. Heat is generated at running speeds as a result of lubricant shear. Temperature is regulated by controlling the lubricant flow through the bearing and external cooling of the lubricant. The anticipated peak babbitt temperature as related to bearing lubricant discharge temperatures should be kept within a range that is compatible with the bearing material and lubricant characteristics. See table 4 for design limits.

5

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Stresses and deflections at high speeds often dictate limits for a specific type of construction. High pitchline velocity, especially when combined with high loads, may require special material specifications and/or testing. Construction features such as holes in the gear body should be analyzed for their influence on the stress. The influence of real or virtual inclusions and/or cracks may need to be considered using the methods of fracture mechanics, with testing of the material to ensure that there are no inclusions greater than the assumed maximum. Overall, a careful analysis of actual operating stresses and deflection should be made to ensure reliable operation.

ANSI/AGMA 6011--I03

ANSI/AGMA 6011--I03

AMERICAN NATIONAL STANDARD

Table 4 -- Hydrodynamic babbitt bearing design limits1)

Type of bearing Radial bearing -- Fixed geometry -- Tilting pad Thrust bearings -- Tapered land -- Flat face -- Tilt pad

Projected unit load,3) N/mm2

Minimum lubricant film thickness, mm

Bearing metal temperature,2) 3) °C

Maximum velocity,3) m/s

3.8 4.2

0.020 0.020

115 115

100 125

2.5 0.5 3.5

0.020 N/A 0.015

115 115 115

125 50 125

NOTE: Table limits will generally not occur all together; one parameter alone may dictate the design. 1) Limits are for babbitt on steel backing. When other materials are used, established limits for these materials are permissible. Bearing clearances should be chosen to yield proper temperature, high stiffness and stability, as well as to ensure adequate clearance to cope with thermal gradients, whether steady, static, or transient. The average ratio of diametral clearance (SJ) to the nominal bore size (DJ), SJ/DJ, for radial bearings is approximately 0.002 mm/mm. 2) Bearing temperature measurements are taken in the backing material within 3 mm of the backing material/babbitt interface at the hottest operational zone of the bearing circumference. 3) Higher values are acceptable if supported either with special engineering or testing and field experience.

4.9.2 Radial bearing stability Hydrodynamic radial bearings shall be designed such that damaging self generated instabilities (e.g., half frequency whirl) do not occur at any anticipated operational load or speed. Hydrodynamic instability occurs when a journal does not return to its established equilibrium position after being momentarily displaced. Displacement introduces an instability in which the journal whirls around the bearing axis at less than one--half journal speed. Known as “half frequency whirl”, this instability may occur in lightly loaded high speed bearings. --`,``,`,`,`,``````,`,``,,`,,`-`-`,,`,,`,`,,`---

4.9.3 Thrust bearings Thrust bearings shall be furnished with all gear units unless otherwise specified. Thrust bearings are generally provided on the low speed shaft for all double helical gears and on single helical gears fitted with thrust collars (see 4.9.4). Thrust bearings are generally provided on each shaft for all single helical gears not fitted with thrust type collars. If the axial position of any of the shafts depends on items outside the gear unit, the purchaser and vendor shall agree to the arrangement relative to the thrust bearings. When gear units are supplied without thrust bearings, some type of end float limitation shall be provided at shaft couplings to maintain positive axial positioning of the gear rotors and connected rotors. Provisions to prevent contact of the rotating ele-

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ments with the gear casing shall be provided unless otherwise specifically agreed to by the purchaser. The design of a hydrodynamic bearing to sustain thrust is as complicated as the design of a radial hydrodynamic bearing. Complete analysis requires consideration of heat generation, lubricant flow, bearing material, load capacity, speed and stiffness. Thrust bearing load capacity should consider the possibility of torque lock--up loads from couplings. When other external thrust forces are anticipated, the vendor must be notified of their magnitudes. 4.9.4 Thrust collars Thrust collars (also known as rider rings) may be used to counteract the axial gear thrust developed by single helical gear sets. Thrust collars arranged near each end of the teeth on a single helical pinion and having bearing surface contact diameters greater than that of the pinion outside diameter may be used to carry the gear mesh thrust forces. Typically the thrust collars have a conical shape where they contact a similarly shaped surface on the mating gear rim located below the root diameter of the gear. Other designs also exist and may be used. Single helical gear sets using thrust collars may be positioned in the housing in a similar fashion to that of double helical gear elements. 4.9.5 Rolling element bearings Selection of rolling element bearings shall be based upon the application requirements and the bearing

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AMERICAN NATIONAL STANDARD

ANSI/AGMA 6011--I03

PS

is service power, kW.

It is recognized that all prime movers have overload capacity, which should be specified.

4.10 Threaded fasteners

5.2 Service factor, CSF and KSF

Refer to ANSI/AGMA 6001--D97, Design and Selection of Components for Enclosed Gear Drives, clause 8.

The service factor includes the combined effects of overload, reliability, life, and other application related influences. The AGMA service factor used in this standard depends on experience acquired in each specific application.

4.11 Shafting The pinion and gear shafts may normally be designed for the maximum bending and maximum torsional shear stresses at service power (see 4.1) by the appropriate methods and allowable values from ANSI/AGMA 6001--D97, clause 4, or other equivalent standards. In some instances, this may result in an oversized or undersized shaft. Therefore, an in--depth study using other available analysis methods may be required.

In determining the service factor, consideration should be given to the fact that systems develop a peak torque, whether from the prime mover, driven machinery, or transitional system vibrations, that is greater than the nominal torque. When an acceptable service factor is not known from experience, the values shown in annex A should be used as minimum allowable values. 5.3 Pitting resistance power rating The pitting resistance of gear teeth is considered to be a Hertzian contact fatigue phenomenon. Initial pitting and destructive pitting are illustrated and discussed in ANSI/AGMA 1010--E95.

5 Rating of gears 5.1 Rating criteria The pitting resistance power rating and bending strength power rating for each gear mesh in the unit must be calculated. The lowest value obtained shall be used as the allowable transmitted power of the gear set. The allowable transmitted power for the gear set, Pa, is determined: P a = the lesser of

P ayu P azu and K SF C SF

(1)

where Pazu is allowable transmitted power for pitting resistance at unity service factor (CSF = 1.0);

The purpose of the pitting resistance formula is to determine a load rating at which destructive pitting of the teeth does not occur during their design life. Ratings for pitting resistance are based on the formulas developed by Hertz for contact pressure between two curved surfaces, modified for the effect of load sharing between adjacent teeth. The pitting resistance power rating for gearing within the scope of this standard shall be determined by the rating methods and procedures of ANSI/AGMA 2101--C95, clause 10, when using service factors, with the following values: ZW

is hardness ratio factor, ZW = 1.0;

Yθ

is temperature factor, Yθ = 1.0;

Ks

is size factor, Ks = 1.0;

ZR

is surface condition factor, ZR = 1.0;

KSF is service factor for bending strength; recommended values are shown in annex A.

ZN

is stress cycle factor (see 5.3.1);

KH

is load distribution factor (see 5.3.2);

The service power shall be less than, or equal to, the allowable transmitted gearset power rating:

Kv

is dynamic factor (see 5.3.3).

Payu is allowable transmitted power for bending strength at unity service factor (KSF = 1.0); CSF is service factor for pitting resistance; recommended values are shown in annex A;

PS ≤ Pa where:

(2)

5.3.1 Stress cycle factor, ZN Stress cycle factor, ZN, is calculated by the lower curve of figure 17 of ANSI/AGMA 2101--C95, and

7

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manufacturer’s recommendations and rating methods. For normal applications, an L10 life of 50 000 hours minimum is required.

ANSI/AGMA 6011--I03

AMERICAN NATIONAL STANDARD

should be based on 40 000 hours of service at rated operating speed. Z N = 2.466 n −0.056 L

(3)

where nL

is number of stress cycles.

When the number of stress cycles exceeds 1010 (i.e., speed above 4167 rpm for 40 000 hours), ZN should be less than or equal to 0.68. If less than 40 000 hours is used for rating, it must be with the specific approval of the purchaser and must be so stated along with the rating. 5.3.2 Load distribution factor, KH KH is the load distribution factor. Values are to be per ANSI/AGMA 2101--C95. The following values shall be used with the empirical method: KHma is mesh alignment factor. Use values from curve 3, precision enclosed gear units, of figure 7 and table 2 of ANSI/AGMA 2101--C95; KHmc is lead correction factor, KHmc= 0.8; KHpm is pinion proportion factor, KHpm= 1.0; KHe is mesh alignment correction factor, KHe = 0.8. The calculated value of KH shall not be less than 1.1. NOTE: The above empirical rating method assumes properly matched leads whether unmodified or modified, teeth central to the bearing span and tooth contact checked at assembly with contact adjustments as required. If these conditions are not met, or for wide face gears, it may be desirable to use an analytical approach to determine load distribution factor. AGMA 927--A01 provides one such approach.

5.3.3 Dynamic factor, Kv --`,``,`,`,`,``````,`,``,,`,,`-`-`,,`,,`,`,,`---

Dynamic factors account for internally generated gear tooth dynamic loads, which are caused by gear tooth meshing action at a non--uniform relative angular velocity. The dynamic factor is the ratio of transmitted tangential tooth load to the total tooth load, which includes the dynamic effects. F + Ft Kv = d Ft where:

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(4)

Fd

is incremental dynamic tooth load due to the dynamic response of the gear pair to transmission error excitation, N;

Ft

is transmitted tangential load, N.

Dynamic forces on gear teeth result from gear transmission error, which is defined as the departure from uniform relative angular motion of a pair of meshing gears. The transmission error is caused by: --

inherent variations in gear accuracy as manufactured;

--

gear tooth deflections which are dependent on the variable mesh stiffness and the transmitted load.

The dynamic response to transmission error excitation is influenced by: --

masses of the gears and connected rotors;

--

shaft and coupling stiffnesses;

--

damping characteristics of the rotor and bearing system.

The AGMA accuracy grades per ANSI/AGMA 2015--1--A01, specifically tooth element tolerances for pitch and profile, and the pitch line velocity may be used as parameters to guide the selection of dynamic factors. Within the 1.09 to 1.15 dynamic factor range, the trend is for Kv to vary in nearly a direct relationship with AGMA accuracy grades from A5 to A2 as shown in table 5. Table 5 -- Dynamic factor as a function of accuracy grade ANSI/AGMA 2015--1--A01 accuracy grade A5 A4 A3 A2

Dynamic factor, Kv 1.15 1.13 1.11 1.09

The dynamic factor, Kv, does not account for dynamic tooth loads which may occur due to torsional or lateral natural frequencies. System designs should avoid having such natural frequencies close to an excitation frequency associated with an operating speed, since the resulting gear tooth dynamic loads may be very high. Refer to ANSI/AGMA 2101--C95 for additional considerations influencing dynamic factors.

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ANSI/AGMA 6011--I03

5.4 Bending strength power rating The bending strength of gear teeth is a measure of the resistance to fatigue cracking at the tooth root fillet. The intent of the AGMA strength rating formula is to determine the load which can be transmitted for the design life of the gear drive without causing root fillet cracking or failure. The gear rim thickness must be sufficient for the calculated rim thickness factor to be 1.0. Occasionally, manufacturing tool marks, wear, surface fatigue, or plastic flow may limit bending strength due to stress concentration around large, sharp cornered pits or wear steps on the tooth surface. The bending strength power rating for gearing within the scope of this standard shall be determined by the rating methods and procedures of ANSI/AGMA 2101--C95, clause 10, when using service factors, with the following values: Yθ

is temperature factor, Yθ = 1.0;

Ks

is size factor, Ks = 1.0;

KB

is rim thickness factor, KB = 1.0;

YN

is stress cycle factor (see 5.4.1);

Kv

is dynamic factor (see 5.3.3);

KH

is load distribution factor (see 5.3.2).

Due consideration should be given to additional testing, such as ultrasonic or magnetic particle inspection of high speed gear rotors which are subject to high fatigue cycles or high stress, or both, during operation. For details on tooth failure, refer to ANSI/AGMA 1010--E95. 5.6 Reverse loading

5.7 Scuffing resistance

Stress cycle factor, YN, is calculated by the lower curve of figure 18 of ANSI/AGMA 2101--C95, and should be based on 40 000 hours of service at rated operating speed. (5)

where nL

Three grades of material have been established. Grade 1 is normal commercial quality steel and shall not be used for gears rated by this standard. Grade 2 is high quality steel meeting SAE/AMS 2301 cleanliness requirements. Grade 3 is premium quality steel meeting SAE/AMS 2300 cleanliness requirements. Both Grade 2 and Grade 3 are heat treated under carefully controlled conditions. The choice of material, hardness and grade is left to the gear designer; however, values of σHP and σFP shall be for grade 2 materials.

For idler gears and other gears where the teeth are completely reverse loaded on every cycle, use 70 percent of the allowable bending stress number, σFP, in ANSI/AGMA 2101--C95.

5.4.1 Stress cycle factor, YN

Y N = 1.6831 n −0.0323 L

specifies the treatment of momentary overload conditions.

is number of stress cycles.

Scuffing failure (sometimes incorrectly referred to as scoring) has been known for many years and is a concern for high speed gear units. When high speed gears are subject to highly loaded conditions and high sliding velocities, the lubricant film may not adequately separate the surfaces. This localized damage to the tooth surface is referred to as “scuffing”. Scuffing will exhibit itself as a dull matte or rough finish usually at the extreme end regions of the contact path or near the points of a single pair of teeth contact resulting in severe adhesive wear.

5.5 Allowable stress numbers, σHP and σFP

Scuffing is not a fatigue phenomenon and may occur instantaneously. The risk of scuffing damage varies with the material of the gear, lubricant being used, viscosity of the lubricant, surface roughness of the tooth flanks, sliding velocity of the mating gear set under load, and geometry of the gear teeth. Changes in any or all of these factors can reduce scuffing risk.

Allowable stress numbers, which are dependent upon material and processing, are given in ANSI/ AGMA 2101--C95, clause 16. That clause also

Further information is provided in annex B. Annex B is not a requirement of this standard. However, it is recommended that either annex B or some other

When the number of stress cycles exceeds 1010, YN should be less than or equal to 0.80. If other than 40 000 hours is used for rating, it must be with the specific approval of the purchaser and must be so stated along with the rating.

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9

ANSI/AGMA 6011--I03

method be used to check for the probability of scuffing failure. See AGMA 925--A03 for further information.

AMERICAN NATIONAL STANDARD

lubricant is requested, the vendor shall provide calculations and an experience list to support a request for an alternate lubricant selection. 6.2.1 Lubricant viscosity

6.1 Design parameters High speed gear units shall be designed with a pressurized lubrication supply system to provide lubrication and cooling to the gears and bearings. A normal lubricant inlet pressure of 1 to 2 bar is an industry accepted value. Special applications may require other lubricant pressures. If a gear element extends below the lubricant level in the gear casing, it is said to be dipping in the lubricant. Dipping at high speed can result in high power losses, rapid overheating, possible fire hazard, and should be avoided. The following minimum parameters should be considered to ensure that proper lubrication is provided for the gear unit: --

type of lubricant;

--

lubricant viscosity;

--

film thickness;

--

surface roughness;

--

inlet lubricant pressure;

--

inlet lubricant temperature;

--

filtration;

--

drainage;

--

retention or settling time;

--

lubricant flow rate;

--

cooling requirements.

6.2 Choice of lubricant Certain lubricant additives, such as those in extreme pressure (EP) lubricants, may be removed by fine filtration. Changes to the level of filtration should only be done in consultation with both the gear unit and lubricant manufacturers. Extreme pressure lubricants are not normally used in high speed units. To avoid dependency on extreme pressure additives, unless otherwise specified, the gear unit shall be designed for use with a lubricant that fails ISO 14635--1 load stage 6. The lubricant used shall pass ISO 14635--1 load stage 5. When an alternate

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Selection of an appropriate lubricant viscosity is a compromise of factors. In addition, lubrication systems are oftentimes integrated with other drive train equipment whose viscosity requirements are different from the gear unit. This complicates the selection of the lubricant. Load carrying capacity of the lubricant film increases with the viscosity of the lubricant. Therefore, a higher viscosity is preferred at the gear mesh. Development of an adequate elastohydrodynamic lubricant film thickness and reduction in tooth roughness are of primary importance to the life of the gearset. However, in high speed gear units, particularly those with high bearing loads and high journal velocities, heat created in the bearings is considerable. Here, the viscosity must be low enough to permit adequate cooling of the bearings. Lubricant viscosity recommendations are specified as ISO viscosity grades. Recommendations for high speed applications are listed in table 6. For turbine driven speed increasers where the lubrication system supplies both the bearings and the gear mesh, an ISO VG32 is usually provided for the gear drive. A lubricant with a viscosity index (VI) of 90 or better should be employed. Special considerations may require the use of lubricants not listed in table 6. The gear vendor should always be consulted when selecting or changing viscosity grades. Table 6 -- Recommended lubricants ISO viscosity grade (VG) 32 46 68 100

Viscosity range mm2/s (cSt) at 40°°C 28.8 to 35.2 41.4 to 50.6 61.2 to 74.8 90.0 to 100.0

Minimum viscosity index (VI) 90 90 90 90

NOTE: When operating at low ambient temperatures, the lubricant selected should have a pour point at least 6°C below the lowest expected ambient temperature.

6.2.2 Synthetic lubricants Synthetic lubricants may be advantageous in some applications, especially where extremes of temperature are involved. There are many types of synthetic

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6 Lubrication

AMERICAN NATIONAL STANDARD

ANSI/AGMA 6011--I03

lubricants, and some have distinct disadvantages. The gear vendor should be consulted before using any synthetic lubricant. 6.3 Lubrication considerations 6.3.1 Ambient temperature Ambient temperature is defined as the temperature of the air in the immediate vicinity of the gear unit. The normal ambient temperature range for high speed gear unit operation is from --10°C to 55°C. The vendor should be informed what the ambient temperature will be, or if a large radiant heat source is located near the gear unit. Furthermore, if low ambient temperature causes the sump temperature to drop below 20°C at start--up, the vendor should be advised. Special procedures or equipment, such as heaters, may be required to ensure adequate lubrication. 6.3.2 Environment If a gear unit is to be operated in an extremely humid, salt water, chemical, or dust laden atmosphere, the vendor must be advised. Special care must be taken to prevent lubricant contamination. 6.3.3 Temperature control The lubricant temperature control system must be designed to maintain a lubricant inlet temperature within design limits at any expected ambient temperature or operating condition. Design inlet temperature may vary, but 50°C is a generally accepted value. Lubricant temperature rise through the gear unit should be limited to 30°C. Special operating conditions, such as high pitch line velocity, high inlet lubricant temperature, and high ambient temperature may result in higher operating temperatures. 6.3.4 Gear element cooling and lubrication The size and location of the spray nozzles is critical to the cooling and proper lubrication of the gear mesh. Spray nozzles may be positioned to supply lubricant at either the in--mesh, out--mesh, or both sides of the gear mesh (or at other points) at the discretion of the vendor. 6.3.5 Lubricant sump The lubricant reservoir may be in the bottom of the gear case (wet sump) or in a separate tank (dry sump). In either case, the reservoir and/or gear case

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should be sized, vented, and baffled to adequately deaerate the lubricant and control foaming. In dry sump applications, the external drainage system must be adequately sized, sloped and vented to avoid residual lubricant buildup in the gear case. Drain velocities may vary, but 0.3 meters per second in a half full opening is a generally accepted maximum value. 6.3.6 Filtration A good filtering system for the lubricant is very important. The design filtration level may vary, but filtration to a 25 micron or finer nominal particle size is a generally accepted value. Filtration finer than 25 microns is recommended when light turbine lubricants are used, particularly for higher operating temperatures. ISO 4406 may be used as a more complete specification of the oil cleanliness required. An ISO 4406:1999 cleanliness level of 17/15/12 is recommended if there is no other recommendation from the gear unit manufacturer. To remove the finer particles, systems may be installed downstream of the filters. It has been found that removing very fine particles can greatly extend lubricant life. It is good practice to locate the filter as near as possible to the gear unit lubricant inlet. Further, it is recommended to provide a duplex filter to facilitate cleaning of the filter when the unit can not be conveniently shut down for filter change. Any kind of bypass of the filter is prohibited. A mechanism to indicate the cleanliness of the filter is recommended. Systems that take a portion of the filtered lubricant and further clean it are acceptable. 6.3.7 Drain lines Location of the drain line must be in accordance with the vendor’s recommendations. Drain lines should be sized so they are no more than half full. The lines should slope down at a minimum of 20 millimeters per meter and have a minimum number of bends and elbows. 6.4 Lubricant maintenance The lubricant must be filtered and tested, or changed periodically, to assure that adequate lubricant properties are maintained. Prior to initial start--up of the gear unit, the lubrication system should be thoroughly cleaned and flushed. It is recommended that the initial charge of lubricant be changed or tested after 500 hours of operation.

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11

ANSI/AGMA 6011--I03

6.4.1 Change interval Unless the vendor recommends different intervals, under normal operating conditions subsequent change or test intervals should be 2500 operating hours, or 6 months, whichever occurs first. Extended change periods may be established through periodic testing of lubricants. With periodic lubricant testing and conditioning, it is not uncommon to operate lubrication systems without lubricant changes for the life of the gear drive. 6.4.2 Water contamination Where operating conditions result in water collecting in the lubrication system, the lubricant should be processed, or changed as required, to keep water content below the lubricant manufacturer’s recommendation. Failure to control moisture may result in damage to the gear unit. Some lubricants are hygroscopic (absorb water) and may need special consideration to eliminate or control the water content and total acid number.

7 Vibration and sound 7.1 Vibration analysis When the frequency of a periodic forcing phenomenon (exciting frequency) applied to a rotor--bearing support system coincides with a natural frequency of that system, the system may be in a state of resonance. A shaft rotational speed at which the rotor--bearing support system is in a state of resonance with any exciting frequency associated with that speed, is called a “critical speed”. Vibration of any component of the gear unit can result in additional dynamic loads being superimposed on the normal operating loads. Vibration of sufficient amplitude may result in impact loading of the gear teeth, interference in the gear mesh, or damage to close clearance parts of the gear unit. Where torque variations exceed 20 percent of the rated torque at the service power, the magnitude and frequency of such torque variations should be evaluated with regard to the endurance properties of the materials used. The types of vibration which are generally of concern for gear units are the torsional, lateral and axial modes of the rotating elements, since these can have a direct influence on the tooth load. Of these,

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the two that are normally reviewed analytically during design are the lateral critical speeds of the gear unit rotating shafts and the torsional critical frequencies of all connected rotating elements. 7.2 Torsional vibration analysis Any torsional vibration analysis must consider the complete system including prime mover, gear unit, driven equipment and couplings. Dynamic loads imposed on a gear unit from torsional vibrations are the result of the dynamic behavior of the entire system and not the gear unit alone. Thus, a coupled system has to be analyzed in its entirety. A common method used is to separate the system into a series of discrete spring connected masses. When applied to a multi--mass system, this method is known as using lumped parameters. These parameters are developed into a model in order to analyze the system as a whole and solve its torsional mechanical vibrations. It is important to note that this result is only as good as its model. In fact, the process of lumping parameters could be the largest source of errors. The result of the torsional system analysis is not within the control of the vendor, since the gear unit itself is only one of several elements in a coupled train. The gear unit vendor is seldom the system designer and in normal cases the gear unit vendor is responsible only for providing mass elastic data. The system designer, not the gear vendor, is responsible for the torsional vibration analysis. 7.3 Lateral vibration analysis The rating equations used in this standard assume smooth operation of the rotors. To insure smooth operation, these rotors should be analyzed for lateral critical speeds. It is imperative that slow roll, start--up, and shutdown of rotating equipment not cause any damage as the rotating elements pass through their critical speeds. See annex C. 7.3.1 Undamped lateral critical speed map An undamped lateral critical speed analysis is sufficient in some cases to determine rotor suitability. If this method is chosen as the sole criterion for determining the suitability of a rotor, it should be based upon significant experience in designing high speed gear drives utilizing this method. It includes a lateral critical speed map, showing the undamped critical speeds versus support stiffness or percentage of torque load. This graphic display shows all

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The undamped lateral critical speed map for gear rotors is used to determine potential locations of the critical speeds by locating the intersection of the principal bearing stiffness values with the undamped critical speeds. If no intersections are indicated, with experience this can be used to determine rotor suitability. Note that these undamped speeds can be significantly different from the critical speeds determined from a rotor response to unbalance analysis. The differences are due to the cross coupled stiffness and damping effects from the bearings. 7.3.2 Analytical methods Coupling moments and shear force transfer effects between rotors with properly designed and installed couplings will be minimal. As a result, each coupled element can generally be analyzed independently. The mathematics of this analysis are complex and beyond the scope of this standard (see C.6.2). Commercial computer software is available and analysts should assure themselves that the method they use gives accurate results for the type of rotors being analyzed. Most high speed rotors are supported in hydrodynamic journal bearings; therefore, of equal importance is the method used to analyze the support (bearing) stiffness and damping. The analyses should include the following effects on the critical speeds: -- bearing--lubricant film stiffness and damping for the range of bearing dimensions and tolerances, load, and speed; -- bearing structure and gear casing support structure stiffness; -- coupling weight to be supported by each gear unit shaft (the weight of the coupling hub plus 1/2 the weight of the coupling spacers). The coupling weight shall be applied at the proper center of gravity relative to the shaft end. The weight and center of gravity will be specified by the purchaser of the coupling; -- potential unbalance of the gear rotor and coupling.

7.3.3 Lateral critical speeds Lateral critical speeds correspond to resonant frequencies of the rotor--bearing support system. The basic identification of critical speeds is made from the natural frequencies of the system and of the forcing phenomena. If the frequency of any harmonic component of a periodic forcing phenomenon is equal to or approximates the natural frequency of any mode of rotor vibration, a condition of resonance may exist. If resonance exists at a finite rotational speed, the speed at which the peak response occurs is called a critical speed. The speed or frequency at which these occur varies with the degree of transmitted load, primarily as a result of the change in stiffness of the bearing lubricant film. Critical speeds are normally determined using a rotor response analysis and are deemed to be acceptable if: (a) the separation margin is greater than 20 percent; or (b) the vibration levels are within the specified limit and the amplification factor is less than 2.5 (see 7.3.3.3). In some cases a simple undamped lateral critical speed analysis may be sufficient to properly analyze the rotor. 7.3.3.1 Forcing phenomena A forcing phenomenon or exciting frequency may be less than, equal to, or greater than the synchronous frequency of the rotor. Potential forcing frequencies may include, but are not limited to, the following: --

unbalance in the rotor system;

--

coupling misalignment frequencies;

--

loose rotor--system component frequencies;

--

internal rub frequencies;

--

lubricant film frequencies;

--

asynchronous whirl frequencies;

-- gear--meshing and side--band frequencies, as well as other frequencies produced by inaccuracies in the generation of the gear tooth. 7.3.3.2 Rotor response analysis The rotor response to unbalance analysis is used to predict the damped vibration responses of the rotor to potential unbalance combinations (i.e., critical speeds). The critical speeds of a gear rotor determined from the rotor response analysis should be verified by shop and field test data. The rotor response analysis should consider the following parametric variations in order to assure

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applicable loading conditions and no--load test conditions (approximately 10 percent of the rated torque) at the maximum continuous speed.

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ANSI/AGMA 6011--I03

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that the vibrations will be acceptable for all expected conditions: 1. Unbalance, g--mm --

6350 W r ; midspan unbalance N mc

--

overhung mass unbalance

--

out--of--phase unbalance

63 500 W cpl ; N mc

63 500 W cpl N mc

at cou-

The response of a critical speed is considered to be critically damped if the amplification factor is less than 2.5 (see figure 1). The shape of the curve in figure 1 is for illustration only and does not necessarily represent any actual rotor response plot. In most cases the amplitude does not decrease to Ncp (0.707 of peak); therefore calculate Ncp from the “flip” of Ncm, or use another method such as the amplification factor in the “Handbook of Rotordynamics” by F.F. Ehrich, page 4.28.

3175 W r at the furthermost mass staN mc tion on the gear tooth portion of the gear. pling, and

Nmc is maximum continuous speed of rotor, rpm; Wr

is total weight of the rotor, kg;

Wcpl is half weight of the coupling and spacer, kg. 2. Gear loading

Vibration amplitude

where

Operating speed

--

unloaded, or minimum load, or both;

--

50 percent load;

--

75 percent load;

Key:

--

100 percent load.

Nmc

SM CRE

Act 0.707 Peak

Nmc

Ncm Nct Ncp

Shaft speed, rpm is maximum continuous rotor speed, rpm;

--

minimum clearance and maximum preload;

Ncp --Ncm is peak width at the half power point; N ct AF is amplification factor= ; N cp − N cm

--

maximum clearance and minimum preload.

SM

is separation margin;

CRE

is critical response envelope;

Act

is amplitude at Nct.

3. Bearing clearances

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4. Speed range from zero to 130 percent of maximum rotor speed. 7.3.3.3 Amplification factor

Figure 1 -- Amplification factor

The amplification factor, AF, is defined as the critical speed divided by the band width of the response frequencies at the half power point. N ct AF = N cp − N cm

(6)

where Nct

is rotor first critical, center frequency, rpm;

Ncm is initial (lesser) speed at 0.707 × peak amplitude (critical), rpm; Ncp is final (greater) speed at 0.707 × peak amplitude (critical), rpm.

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7.3.4 Stability analysis Damped eigenvalues (damped natural frequencies) may occur below 120% maximum rotor speed due to a variation in load, bearing properties, etc. These damped eigenvalues are the frequencies at which the rotor will vibrate if there is sufficient energy or insufficient damping in the system. Therefore, a damped stability analysis is performed to ensure that these damped eigenvalues have a large enough logarithmic decrement (log dec) to insure stability. The stability analysis calculates the damped eigenvalues and their associated logarithmic decrement.

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The rotor should have minimum log dec of +0.1 at any of the damped eigenvalues to be considered stable. 7.3.5 Mode shape

7.4 Balance

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All gear rotating elements shall be multiplane dynamically balanced after assembly of the rotor. Rotors with single keys for couplings shall be balanced with their keyway fitted with a fully crowned half--key so that the shaft keyway is filled for its entire length. The balancing machine shall be suitably calibrated, with documentation of the calibration available. The rotating elements should be balanced to the level of the following equation: (7)

where Umax is amount of residual rotor unbalance, g--mm; W

is journal static loading, kg;

Nmc is maximum continuous speed, rpm. 7.5 Shaft vibration During the shop test of the assembled gear unit operating at its maximum continuous speed or at any other speed within the specified range of operating speeds, the double amplitude of vibration for each shaft in any plane measured on the shaft adjacent and relative to each radial bearing shall not exceed the following value or 50 mm, whichever is less:

is allowable double amplitude of unfiltered vibration, micrometers (mm) true peak to peak.

7.5.1 Electrical and mechanical runout When provisions for shaft non--contact eddy current vibration probes are supplied on the gear unit, electrical and mechanical runout shall be determined by rolling the rotor in V--blocks at the journal bearing centerline, or on centers true to the bearing journals, while measuring runout with a non--contacting vibration probe and a dial indicator. This measurement will be taken at the centerline of the probe location and one probe tip diameter to either side and the results included with the test report. 7.5.2 Electrical/mechanical runout compensation If the vendor can demonstrate that electrical/mechanical runout is present, the measured runout may be vectorially subtracted from the vibration signal measured during the factory test. However, in no case shall the amount subtracted exceed the smallest of: --

measured runout;

-- 25 percent of the test level determined from 7.5; --

6.4 micrometers.

7.6 Casing vibration During shop no--load test of the assembled gear drive operating at its maximum continuous speed or at any other speed within the specified range of operating speeds, casing vibration as measured on the bearing housing shall not exceed the values shown in table 7. 7.7 Vibration measurement Vibration measurements and instrumentation shall be in accordance with ANSI/AGMA 6000--B96 unless otherwise agreed upon by the purchaser and vendor. 7.8 Sound measurement Sound level measurement and limits shall be in accordance with ANSI/AGMA 6025--D98 unless otherwise agreed upon by the purchaser and vendor.

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(8)

where A

Each finite resonant frequency has an associated mode shape. Knowing the mode shape that the rotor will assume when responding to a critical speed is important in understanding the consequences of bearing placement and residual unbalance. In most high speed gear unit rotors, the mode shape of the first critical speed is mostly conical with a node point between the bearings, vibration at the bearings approximately 180° out of phase, and the point of highest vibration at the drive (coupling) end of the shaft. A slight bending shape of the rotor is common. The amplitude at the bearing locations is usually high enough to allow the damping inherent in hydrodynamic journal bearings to limit maximum vibration amplitudes. However, the location of highest amplitude at the coupling makes most gear units sensitive to unbalance at this location and extra care in coupling balance is recommended.

U max = 6350 W N mc

A = 2800 N mc

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Table 7 -- Casing vibration levels Frequency range Unfiltered (peak) Filtered component

Velocity 10 Hz -2.5 kHz 4 mm/sec 2.5 mm/sec

Acceleration 2.5 kHz -10 kHz 4 g’s

NOTES: 1) The above vibration levels are for horizontal offset gear units only. The allowable vibration levels for vertical offset gears are twice those shown in the table. 2) Filtered component means any vibration peak within the frequency range.

8 Functional testing 8.1 General Each unit conforming to this standard should be functionally tested at full speed. Additional tests may also be done at other speeds. Functional testing provides a means of evaluating operational characteristics of the unit. The procedures may be the vendor’s standard or one agreed upon by the vendor and purchaser.

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Functional testing presents an opportunity to evaluate the operational integrity of the design and manufacture of gear drives. Functional test procedures provide a means of evaluating the entire gear system for noise, vibration, lubrication, gear tooth contact, bearing operating temperatures, bearing stability, lubricant sealing, mechanical efficiency, instrument calibration and other unit features, and provide data that parallels the expected on--line operational characteristics. 8.2 Procedures Functional testing may also include procedures ranging from partial speed and no load spin testing to full speed and full power testing. Following testing, the unit may be disassembled for bearing and gear tooth contact inspection. 8.2.1 No load testing The unit under test is normally driven in the same rotational direction and with the same input shaft as in the design application. The output shaft will have no load applied to it. Test speeds may range from partial speed to over speed. The test duration should be no less than one hour after temperature stabilization.

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8.2.2 Full speed and partial load testing The unit under test is normally driven in the same rotational direction and with the same input shaft as in the design application. The output shaft will be connected to a loading device which applies a resisting torque less than the design full load torque. Test duration should be no less than one hour after temperature stabilization. 8.2.3 Full speed and full power testing Full speed and full power testing can be carried out in the same manner as described in 8.2.2 for units with lower operating powers. Full power testing of units with higher power ratings may require back--to--back locked torque testing. In this procedure two identical ratio units are shaft coupled together, input to input and output to output. Full operational torque is applied by disengaging one of the shaft couplings, rotating the shafts relative to one another until the proper torque is achieved, then re--engaging the shaft coupling. The unit shafts are then rotated at full speed. Full power testing duration is usually not less than four hours after temperature stabilization. When performing back--to--back locked torque testing the following risks should be considered: -- Bearings with full load applied at the static condition will start with full load and no hydrodynamic lubricant film until “some” rotational speed is reached; -- Gear and pinion teeth with full load applied at the static condition will start with full load and no lubricant film to separate the teeth until “some” rotational speed is reached. Scuffing may occur; special procedures such as coating of the gear teeth with an EP lubricant may be required (this problem may be avoided if the method of torque application allows start up at low torque); -- Bearings of one unit will be loaded in a direction opposite normal operation; -- Slave unit bearing loads are in the opposite direction, stub shafts used to complete the torque path may have to be removed, and if the gear elements of the slave unit are not flipped end for end, they will be loaded on the flanks that are not normally loaded. Therefore the slave unit, and often also the tested unit, will have to be modified after the test; -- For purposes of this test the slave unit may require a lead and profile modification suitable for

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ANSI/AGMA 6011--I03

loading in the testing mode. When the leads are modified specifically for test, then after back--to-back testing the slave gears may require final modification suitable for the contract application. The vendor and purchaser shall agree on the extent of this work. At the conclusion of back--to-back tests, the slave unit will require a test of its own, since the back--to--back configuration cannot be duplicated for that purpose. The vendor and purchaser shall agree on the test to be performed. 8.2.4 Special testing In the case of very high rotational speeds or multiple input/output shafts, conventional testing may become impractical. In such cases, special test procedures specific to the application should be developed between the vendor and purchaser. 8.2.5 Power loss testing --`,``,`,`,`,``````,`,``,,`,,`-`-`,,`,,`,`,,`---

When testing for power loss in a high speed gear unit, one method is to measure the heat removed by the lubricant flowing through the gear unit. Lubricant flow rate and lubricant inlet and outlet temperatures are measured. Power loss is then calculated using: P L = Q LUBE c p ∆T

(9)

where PL

is power loss, kW;

QLUBE is lubricant flow, kg/sec; ∆T cp

is change in lubricant temperature from inlet to outlet, °C; is specific heat of the lubricant, kJ/(kg°C).

Aeration of the lubricant may result in the indicated flow rate being higher than the actual mass flow, so the indicated flow may need to be adjusted to a lower value. Accuracy of the power loss calculation may be improved if all other heat transfer to or from the gear unit is properly accounted for. Other methods of measuring power loss may be used, such as the difference in the power in and out as measured with torque meters, if agreed to by the purchaser and vendor.

9 Vendor and purchaser data exchange 9.1 Rationale for data requirements In order to promote consistency and reduce errors, recommended information to be furnished to the vendor and data provided by the vendor is specified in this section. A detail of the schedule for transmission of drawings, curves and data should be agreed to at the time of the proposal or order. The purchaser should promptly review the vendor’s data when he receives them; however, this review does not constitute permission to deviate from any requirements in the order unless specifically agreed upon in writing. After the data has final approval, the vendor should furnish certified copies in the quantity specified. A complete list of all vendor data should be included with the first issue of major drawings. This list contains titles, drawing numbers, and a schedule for transmission of all data the vendor will furnish. Inquiry documents should be revised to reflect any subsequent changes. These changes will result in the purchaser’s issue of completed, corrected data sheets as part of the order specifications. 9.2 Document identification Transmittal (cover) letter title blocks or title pages should contain the following information, when available: --

purchaser/user’s corporate name;

--

job/project;

--

equipment item number;

--

inquiry or purchase order number;

-- any other identification specified in the inquiry or purchase order; -- vendor’s identifying proposal number, shop order number, serial number, or other reference required to completely identify return correspondence. 9.3 Data provided by purchaser To allow the gear unit to be properly selected or designed, the vendor must have adequate information from the purchaser. The following is a guide to data that should be sent along with a request for quotation: -- a data sheet is provided in annex H. All of the data on the left hand side of that form should be included in the request for proposal;

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--

scope of supply;

--

information on the couplings that will be used;

--

testing requirements;

-- measurement units to be used in drawings and other communications (SI or U.S. customary); -- list of specifications;

applicable

standards

and

-- copies of any applicable purchaser specifications; -- any other special requirements, such as painting, shipping, storage or environmental protection requirements. 9.4 Proposal data The following is a guide to proposal data that should be furnished by the vendor: -- general arrangement or outline drawing for each gear unit showing overall dimensions;

-- conditions and period of the vendor’s warranty. 9.5 Items needing resolution The following items normally should be resolved after purchase commitment. This may be done at a coordination meeting, preferably at the vendor’s plant or by other suitable means of communication. -- purchase order, vendor’s internal order details and sub--vendor items; --

any required data sheets;

-- applicable specifications, standards, clarifications and previously agreed upon exceptions; -- that the system and all its components are in accordance with specified standards; -- schedules for transmittal of data, production and testing; -- quality assurance program, procedures and acceptance criteria; --

inspection, expediting and testing;

-- purchaser’s data sheets, with completed vendor’s information entered thereon and literature to fully describe details of the offering (a suggested data sheet is provided in annex H);

-- schematics and bills of material (B/Ms) of auxiliary systems;

-- if applicable, a list of requested exceptions to the specifications;

--

-- schedule for shipment of the equipment, in weeks after receipt of the order, and all approved drawings; -- list of recommended start--up spares, including any items that the vendor’s experience indicates are likely to be required; -- complete tabulation of the utility requirements, including the required flow rate of lubricant and supply pressure, heat load to be removed by the lubricant, and nameplate power rating (approximate data shall be defined and identified as such); -- description of tests procedures, as required;

and

inspection

-- when requested, the vendor should furnish a list of the procedures for any special, or optional tests, that have been specified by the purchaser or proposed by the vendor; -- any start--up, shut--down, or operating restrictions required to protect the integrity of the equipment;

-- physical orientation of equipment, shaft rotation, piping and auxiliary systems; final coupling selection.

9.6 Contract data The following lists contract data normally supplied by the vendor: a. Certified dimensional outline drawing and parts list, including the following: -- size, rating and location of all purchaser’s connections; -- approximate overall and handling weights; -- overall dimensions; -- dimensioned shaft end(s) for coupling mounting(s); -- height of shaft centerline; -- dimensions of baseplates or soleplates (if furnished), complete with the diameter, number and location of bolt holes and thickness of the metal through which bolts must pass; -- shaft position diagram, including recommended limits during operation, with all changes in shaft end position and support growths from an ambient reference or 15°C noted; --`,``,`,`,`,``````,`,``,,`,,`-`-`,,`,,`,`,,`---

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-- journal bearing clearances and tolerances; -- axial rotor float or thrust bearing clearance, as applicable; -- number of teeth on each gear. b. When a lubricant system is supplied, a schematic, certified dimensional outline drawing, and parts list including the following: --`,``,`,`,`,``````,`,``,,`,,`-`-`,,`,,`,`,,`---

-- control, alarm and trip settings (pressures and recommended temperatures); -- utility requirements, including electrical, water and air; -- pipe and valve sizes; -- instrumentation, control schemes;

safety

f. When mechanical running test is supplied, test reports, including the following (see clause 8): -- vibration; -- lubricant flow and inlet and outlet temperatures; -- bearing temperatures. g. Nameplates and rotation arrows shall be of Series 300 stainless steel or of nickel--copper alloy (Monel or its equivalent) attached by pins of similar material and located for easy visibility. As a minimum, the following data should be clearly stamped on the nameplate: -- vendor’s name;

devices

and

-- size, rating and location of all purchaser’s connections; -- instruction and operation manuals; -- maximum, minimum and normal liquid levels in the reservoir; -- quantity of lubricant required to fill reservoir to the normal level. c. Electrical and instrumentation schematics and bills of materials, including the following:

-- size and type of gear unit; -- gear ratio; -- serial number; -- service power, Ps; -- rated input speed, in revolutions per minute; -- rated output speed, in revolutions per minute; -- gear service factor, as defined in this standard; -- purchaser’s item number;

-- vibration warning and shutdown limits;

-- number of gear teeth;

-- bearing temperature warning and shutdown limits;

-- number of pinion teeth;

-- lubricant temperature warning and shutdown limits. d. Lateral critical speed analysis, which may include any or all of the following: -- method used; -- graphic display of bearing and support stiffness and their effects on critical speeds (undamped lateral critical speed map); -- graphic display of the rotor response to unbalance, including damping (rotor response analysis); -- journal bearing stiffness and damping coefficients; -- damped stability analysis, including identified eigenvalues and associated logarithmic decrement. e. Torsional data for the gear unit and any shaft couplings supplied by the vendor, sufficient for a third party to do a system torsional analysis.

-- date of manufacture: month and year unit was successfully tested. h. Statement of any special protection required for start--up, operation, and periods of idleness under the site conditions specified on the data sheets. The list shall clearly identify the protection to be furnished by the purchaser, as well as that included in the vendor’s scope of supply. 9.7 Installation manual When specified by the purchaser, an installation manual shall be supplied. Any special information required for proper installation design that is not on the drawings shall be compiled in this manual. This manual shall be forwarded at a time that is mutually agreed upon in the order. The manual shall contain information such as special alignment and grouting procedures, utility specifications (including quantities), and all other necessary installation design data, including drawings and data specified in 9.6. The manual shall also include sketches that show the location of the center of gravity and rigging

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ANSI/AGMA 6011--I03

provisions, to permit removal of the top half of the casing, rotors, and subassemblies that have a mass (weight) greater than 140 kilograms. 9.8 Operation, maintenance and technical manuals

be forwarded at a time that is mutually agreed upon in the order. This manual shall contain a section that provides special instructions for operation at specified extreme environmental conditions, such as temperatures. 9.9 Recommended spares When the vendor submits a complete list of spare parts, the list should include spare parts for all equipment and accessories supplied. The vendor should forward the list to the purchaser promptly after receipt of the reviewed drawings and in time to permit order and delivery of the parts before field start--up. 9.10 Special tools A list of special tools required for maintenance shall be furnished.

--`,``,`,`,`,``````,`,``,,`,,`-`-`,,`,,`,`,,`---

The vendor shall provide sufficient written instructions and a cross--referenced list of all drawings to enable the purchaser to correctly operate and maintain all the equipment ordered. This information should be compiled in a manual or manuals with a cover sheet containing all reference--identifying data specified in 9.2, an index sheet containing section titles, and a complete list of referenced and enclosed drawings by title and drawing number. The manual shall be prepared for the specified installation; a generic manual is not acceptable. This manual shall

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ANSI/AGMA 6011--I03

Annex A (informative) Service factors [The foreword, footnotes and annexes, if any, are provided for informational purposes only and should not be construed as a part of ANSI/AGMA 6011--I03, Specification for High Speed Helical Gear Units.]

A.1 Purpose This annex provides detailed instructions for the determination and use of service factors for enclosed high speed helical gear units as described in ANSI/AGMA 6011--I03. A.2 Determination of service factors The determination of service factor is based on the equipment characteristic overload of the gear unit as a result of operation, desired reliability of the gear unit during its design life, and length of time that is considered the design life. It relies heavily on experience acquired in each specific application. A broad explanation of the factors involved are: -- The causes of service overloads are broken into three broad categories: those produced by the prime mover, those produced by the driven equipment, and those resulting from system considerations unique to the equipment train; -- The reliability of a geared system depends on many factors both internal to the gear unit itself and external to the unit. Increases in service factor to influence reliability normally take into consideration external sources of failure such as abuse and unexpected operating conditions; -- The desired life of most high speed enclosed drives is usually longer than other types of enclosed drives. At high operating speeds this can translate into a very large number of stress cycles on the components. A.2.1 Prime mover characteristics Some different types of prime movers are: electric or hydraulic motors, steam or gas turbines, and single or multiple cylinder internal combustion engines. Each of these prime movers is designed to produce some nominal power, but each will produce this power with some variation over time. The variation of power output with time may be lower or higher depending on the prime mover and also the way the prime mover is applied in a particular machinery train, but any variation over nominal power is an overload and must be considered.

A.2.2 Driven equipment characteristics Driven equipment can generally be divided into rotary and reciprocating types of machines. Rotary machines generally have smoother power requirements than reciprocating machines, but each type is unique and the equipment characteristics of each must be known to be properly evaluated. A.2.3 System conditions The gear unit is a part of a system, and this system can have dynamic (vibratory) response to time, varying (dynamic) power transmission that may overload the gear unit. This is most commonly found as torsional vibration in the rotating shafts, but can be any vibratory response to dynamic exciting forces. Generally, overloads are assumed to be transmitted with no amplification through the gear. However, when there is a resonant response to a dynamic power overload, a much higher load may occur at the gear unit. Thus, the dynamic overloads that are caused by prime movers and driven machines may be amplified in such a way as to greatly increase their magnitude at the gear unit, and primarily at the gear tooth mesh. The normal rating of gear units and the normal service factors used assume that these responses (resonances) do not appreciably affect the gear unit load. Therefore, careful system analysis is recommended to ensure that no unexpected overloads due to resonances are present. A.2.4 Reliability and life requirements There is a reliability factor in the power rating equations, but it deals only with the statistical nature of material testing and probability of failure for materials at a given stress level. In a gear unit there are many separate components that may fail, many modes of failure, and many factors that can contribute to those modes of failure. For this reason, quantifying factors associated with reliability and life to account for these external issues can be extremely difficult. A.3 Service factor table Service factors have served the industry well when they have been identified by knowledgeable and

21

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experienced gear design engineers. The service factors shown in table A.1 have been used with success in the past. These values may be used as general guidelines, but they do not eliminate the responsibility of defining any unusual system requirements that would alter the listed values.

netic fields have dissipated can produce very high torques;

A.3.1 General selection guidelines

-- Generators have extremely high loads when they are out of phase with the main system, and across--the--line electrical shorts can produce very high torque loads. For this reason torque limiting devices or higher service factors are advisable;

-- Synchronous electric motors can produce very high torsional forcing functions during start-up. This can cause very high transient torsional torques on the gear unit;

There is no way to list all the possible considerations that may affect selection of service factors, but the following are some guidelines. -- Induction electric motors can produce high torques on start--up. Therefore, on an application with many starts, higher service factors may be warranted;

-- Brakes or other decelerating devices can produce loads on the gear unit larger than the transmitted power. The list could be much longer, but the intent here is to give a general idea of items to consider when selecting service factors.

-- Electric motors that have electric power interrupted and then re--applied before induced mag-

Table A.1 -- Service factors, CSF and KSF Service factor, with prime mover Application Blowers Centrifugal Lobe Compressors Centrifugal process gas, except air conditioning air conditioning service air or pipe line service Rotary axial flow -- all types liquid piston (Nash) lobe -- radial flow Reciprocating 3 or more cylinders 2 cylinders Dynamometer -- test stand Fans Centrifugal Forced draft Induced draft Industrial and mine (large with frequent starts) Generators and exciters Base load or continuous Peak duty cycle

Synchronous motors

Induction motors

Gas or steam turbine1)

Internal combustion engine (multi--cylinder)

1.7 2.0

1.4 1.7

1.6 1.7

1.7 2.0

1.6 1.6 1.7

1.4 1.2 1.4

1.6 1.4 1.6

1.6 1.6 1.7

1.7 2.0 2.0

1.7 1.7 1.7

1.7 1.7 1.7

1.7 2.0 2.0

2.0 2.3 1.3

2.0 2.0 1.1

2.0 2.0 1.1

2.0 2.3 1.3

1.7 1.7 2.2 2.2

1.4 1.4 1.7 1.7

1.6 1.6 2.0 2.0

1.7 1.7 2.2 2.2

1.4 1.7

1.3 1.4

1.3 1.4

1.4 1.7

(continued)

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ANSI/AGMA 6011--I03

Table A.1 (concluded) Service factor, with prime mover Application

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Paper industry Jordan or refiner Paper machine -- line shaft Pumps Centrifugal (all service except as listed below) Centrifugal boiler feed descaling (with surge tank) hot oil pipe line water works Reciprocating 3 or more cylinders 2 cylinders Rotary axial flow -- all types gear type liquid piston lobe sliding vane Sugar industry Cane knives Crushers Mills

Synchronous motors

Induction motors

Gas or steam turbine1)

Internal combustion engine (multi--cylinder)

-------

-------

1.5 1.3

-------

1.7

1.3

1.5

1.7

---------2.0 2.0

1.7 2.0 1.7 1.5 1.5

2.0 2.0 2.0 1.7 1.7

---------2.0 2.0

2.0 2.0

2.0 2.0

1.7 2.0

2.0 2.0

1.8 1.8 2.0 2.0 1.8

1.5 1.5 1.7 1.7 1.5

1.5 1.5 1.7 1.7 1.5

1.8 1.8 2.0 2.0 1.8

1.8 2.0 2.3

----------

1.5 1.7 1.7

1.8 2.0 2.3

NOTES: 1) Gas turbines seldom operate at full design power while steam turbines often operate at or above rated power. Appropriate design considerations should be made to assure adequate torque capacity.

23

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Annex B (informative) A simplified method for verifying scuffing resistance [The foreword, footnotes and annexes, if any, are provided for informational purposes only and should not be construed as a part of ANSI/AGMA 6011--I03, Specification for High Speed Helical Gear Units.]

B.1 Purpose

v′

is pitch line velocity, m/s;

This annex provides information concerning the scuffing (scoring) of high speed gear units.

ν40

is viscosity of lubricant at 40° C, mm2/s (cSt);

Cw

= 1.10 (conservative value);

Cw

= 1.15 (nominal value);

Cw

= 1.20 (maximum value).

ANSI/AGMA 6011--I03 is concerned with two failure modes in gear teeth. They are surface pitting and root bending fatigue failure of the tooth material for a given number of stress cycles. There is another known failure type: scuffing (sometimes referred to as scoring). The calculation of the scuffing load capacity is a very complex problem. While this type failure has been known for many years and mathematical methods have been devised to assess relative risk (see AGMA 925--A03), a simplified scuffing criterion is suggested that is suitable for general high speed design work. From the values of tooth loading, pitch line velocity, and viscosity of the lubricant, a condensed load function, F (load), is formed, which, to assure scuffing resistance, must be less than (or equal to) the geometric function, F (geometric). The geometric function is based on gear characteristics such as number of teeth of the pinion and gear, center distance and gearset ratio. As long as the value of the load function, F (load), does not exceed that of the geometric function, F (geometric), there is adequate safety against scuffing.

NOTE: Cw values are suggested values. Vendor’s own experience may change these values with supporting data. Value of Cw = 1.20 should only be used if total helix deviation meets ANSI/AGMA 2015--1--A01 accuracy grade A3.

Table B.1 -- Lubricant viscosities ISO viscosity grade VG VG -- 22 VG -- 32 VG -- 46 VG -- 68

NOTE: For high speed gearset applications, lubricant viscosity means light turbine oil with little or no additives based on a viscosity range of: 32 ≤ ν40 ≤ 68. The standard FZG oil test, ISO 14635--1, gives approximations for the lubricant with respect to scuffing tendency.

Geometric function, F (geometric):

Therefore: (B.1)

F (load) ≤ F (geometric)

 

0.25 46 F (load) = w′ [ v′ ] ν 40 Cw

0.22

(B.2)

where w′

is specific tooth load on the pitch circle, N/ mm;

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F(geometric) =

50 + z 1 + z2(a ) 0.5 A

[C u] (B.3)

where:

Load function, F (load):

 

Nominal viscosity at 40°C, mm2/sec (cSt) 22 32 46 68

z1

is number of teeth of the pinion;

z2

is number of teeth of the gear;

a

is center distance, mm;

A

is taken from table B.2;

Cu

is taken from table B.2.

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B.2 Scuffing considerations

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ANSI/AGMA 6011--I03

Table B.2 -- Values A and Cu for calculating F (geometric) αp

A

Cu @ 1 ≤ u < 3

Cu @ 3 ≤ u ≤ 10

15

350

95 + 28.6 (3 -- u)

130 -- 10 [112.5 -- (13 -- u)2]0.5

17.5

300

90 + 30 (3 -- u)

120 -- 10 [90 -- (12 -- u)2]0.5

20

300

100 + 33.3 (3 -- u)

130 -- 10 [109 -- (13 -- u)2]0.5

22.5

250

95 + 28.5 (3 -- u)

130 -- 10 [112.5 -- (13 -- u)2]0.5

25

250

105 + 31.4 (3 -- u)

140 -- 10 [133.5 -- (14 -- u)2]0.5

NOTE αp is pressure angle, degrees; u is gear ratio (z2/z1).

B.3 Field of application The above scuffing criterion is applicable to: a. High speed gears with a modified addendum (rack shift or x factor) resulting in reasonably balanced sliding and rolling conditions between the tooth flanks at the tip of the pinion and mating gear; b. Gear tooth accuracy grade, per ANSI/AGMA 2015--1--A01, shall be equal to or better than: A5 for single pitch deviation, fpt A5 for total cumulative pitch deviation, Fp A4 for total profile deviation, Fα A4 for total helix deviation, Fβ c. Surface roughness of tooth flanks after grinding, Ra ≤ 0.5 µm (20 rms); d. Basic rack profile with: pressure angle, αP = 20 deg addendum, hap = 1 module. The working flanks of the pinion or gear shall be provided with profile modifications to obtain a trapezoidal tooth load distribution along the path of contact.

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The working flanks of the pinion or gear shall be provided with longitudinal modification to compensate for bending and torsional deflections and thermal deformations of the gear rotors in order to obtain a uniform tooth load distribution over the entire rated face width. The lubricant used shall pass ISO 14635--1 load stage 5. B.4 Scuffing design criteria As stated, there are no firm criteria for designing to prevent scuffing at this time. However, it is hoped that the use of methods such as those in this annex and those in AGMA 925--A03 can lead to a set of design criteria. There are other methods for predicting scuffing and there is no intent to deny the validity of any method at this time. B.5 Conclusion Predicting scuffing is very important in high speed gearing. It is hoped that industry consensus can be reached on scuffing prediction. To achieve this consensus, industry must utilize available methods and gain experience.

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25

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Annex C (informative) Lateral rotor dynamics [The foreword, footnotes and annexes, if any, are provided for informational purposes only and should not be construed as a part of ANSI/AGMA 6011--I03, Specification for High Speed Helical Gear Units.]

C.1 Purpose In the dynamic analysis of a high speed gear box, it is necessary to verify that the drive is inherently stable, and that any actual harmful critical speeds are sufficiently removed from any operating speed or load range of the equipment. This annex provides information on rotor dynamics for high speed gear drives.

bearing load is mainly a result of the rotor weight and is therefore constant.

C.2 Modes

Figure C.1 -- Typical modes of rigid rotor lateral vibration Typical turbomachinery equipment can pass through what is called flexural type critical speeds within their operating speed range. Here the rotor will actually deflect to create mode shapes similar to those shown in figure C.2, in addition to any vibration resulting from shaft displacement in its bearings. C.3 Bearings In gear rotor dynamics, bearing oil film stiffness varies with speed as well as torque load applied to the drive. This is quite different from turbomachinery driven through a high speed flexible coupling where

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Figure C.2 -- Typical modes of flexural lateral vibration High speed gear drives use fluid film or sleeve type bearings. They frequently are manufactured with non--cylindrical bores. Gear drive bearings generally have a large length to diameter ratio to gain the bearing area required to support the torque load as well as rotor weight loading and still be able to maintain high efficiencies. This type of bearing design lends itself to asymmetrical oil film stiffness rates in the X and Y directions. High stiffness values occur in the direction of the applied load. Relatively large cross coupled stiffness and damping coefficients are common. Bearing cross coupling spring and damping, in simple terms, means that, in addition to a resulting resisting force being generated in the direction of displacement or velocity, another force is created 90 degrees from the direction of motion. This phenomenon has a more pronounced effect in gear drives than in turbo equipment, which frequently uses tilting pad type bearings. For an accurate analysis of a gear drive, a complete eight element matrix of spring and damping rates should be obtained (see figure C.3). Stiffness terms: Kxx is force in X resulting from a displacement in the X direction, in Newtons per millimeter; Kxy is force in X resulting from a displacement in the Y direction, in Newtons per millimeter; Kyy is force in Y resulting from a displacement in the Y direction, in Newtons per millimeter; Kyx is force in Y resulting from a displacement in the X direction, in Newtons per millimeter.

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High speed gear drives are frequently coupled to turbomachinery. Although the gear drive operates at turbomachinery speeds, its dynamic behavior is significantly different from compressors or turbines. Gear shafting is generally of the rigid rotor design. This means that throughout the operating speed range of the machine, most vibration that occurs is caused by shaft displacements in the bearing system oil films rather than deflections of the rotor (see figure C.1).

AMERICAN NATIONAL STANDARD

ANSI/AGMA 6011--I03

Bearing film

Journal

Dyy Kyy

Y

Dxy Kxy

Kyx Dyx

vibrations. Lightly loaded fluid film bearings can get into sub--synchronous vibration problems, particularly in the qualification testing process, which is generally a no load test. Oil whirl and oil whip are the names for this type of problem. This vibration is usually at a frequency of around 0.4 times rotational speed. If not properly detected in the analysis of the drive, undesirable or even destructive vibrations may be exhibited in testing or lightly loaded field running.

Kxx Dxx

Qe Te

X Figure C.3 -- Cross coupled bearing schematic representation

QLG TG Bearing groove

Damping terms: Dxx is force in X resulting from a velocity in the X direction, in Newtons per millimeter;

Q1 T1

Dxy is force in X resulting from a velocity in the Y direction, in Newtons per millimeter;

Obtaining these coefficients is the first step to an accurate gear drive rotor dynamics analysis. Sophisticated bearing analysis techniques are available to determine these coefficients. A typical method will solve the Reynolds and energy equations over a grid network of the bearing area for the particular geometry in question by finite difference techniques. The results from each grid point are numerically combined to produce the performance characteristics of the complete bearing. A detailed heat balance of the bearing system under its operating conditions must be performed to ensure that the actual oil film viscosities are being utilized. This is normally accomplished in an iterative type technique, where an assumed temperature is chosen for performance calculation and then is compared with the final calculated temperatures resulting from the heat balance. If the two do not agree, a new assumed temperature is chosen and the process continues in the program until convergence occurs (see figure C.4). C.4 Stability A stability analysis is required to ensure that the drive will not exhibit self sustaining non--synchronous

Qv T3

Qi

Px

Bearing films

TG

External source

Qe Te Q2 Q1 T2

Figure C.4 -- Heat balance model C.5 Critical speed A critical speed is defined as the speed at which the peak response amplitude actually will occur when the rotor bearing system is in resonance with a periodic forcing frequency. There are many possible forcing frequencies in a gear drive system but the one most likely to excite the system is the harmonic force generated at rotor rotational speed due to mass imbalance. Gears generally are designed to have their actual critical speeds above 120 percent of their maximum operating speed. Undamped and damped natural frequencies may be calculated below running speed. Damping may completely suppress the response of these modes or significantly shift the frequency at which these modes will actually experience peak response or critical speed by the above definition. Damping tends to lower calculated natural frequencies. For simple systems they are related by: Wd = 1 − ξ 2 Wo

(C.1)

where ξ

is the damping ratio;

Wd

is the damped natural frequency;

27

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T2

H2 P 2

Dyy is force in Y resulting from a velocity in the Y direction, in Newtons per millimeter; Dyx is force in Y resulting from a velocity in the X direction, in Newtons per millimeter.

Q2

Drains

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Bearing shell

ANSI/AGMA 6011--I03

AMERICAN NATIONAL STANDARD

is the undamped natural frequency. ξ= D (C.2) Dc Wo

where D Dc

C.6.1 Undamped critical speed analysis is the actual damping; is the critical damping.

Damping, however, tends to raise the frequency at which the actual response amplitude or critical speed due to imbalance occurs. For simple systems they are related by: Wr 1 = Wo 1 − 2 ξ 2

(C.3)

where Wr

critical speed analysis, the damped critical speed stability type analysis, and the damped unbalance response analysis.

is the actual response frequency.

The damped, undamped, and response frequencies will agree only when the damping ratio is small. Large discrepancies will be seen at damping ratios larger than 0.3. Another way of expressing damping ratio is by a logarithmic decrement which defines how quickly a vibration will decay with time. Log decrement

S=



4 π2 ξ2 1 − ξ2

(C.4)

C.6 Analysis types There are three main tools used in natural frequency and critical speed analysis, each having its own strengths and weaknesses. They are the undamped

The undamped critical speed analysis is an excellent simple tool for preliminary evaluation of a rotor bearing system. It allows the analyst to identify approximately the magnitude of oil film stiffness required to obtain the desired regime of operation of the system (i.e., rigid or flexible rotor design). Approximate mode shapes are obtained. Effectiveness of bearing damping can be seen. If motion of the rotor occurs at the bearing, damping will be very effective. If the motion occurs other than at the bearing, damping will be ineffective. While the undamped critical speed map is a useful tool in estimating performance, it is lacking in several major areas. First, it does not consider the cross coupled effects in the oil film; and second, it does not consider the direct or cross coupled damping terms. In gear drives, which generally have large damping values as well as large cross coupled terms, the result can tend to yield critical speed predictions less than what an actual machine may exhibit. Lastly, no indication of stability characteristics is obtained. The map should display the effect of load variations. Stiffness values for the range of applied load are generally plotted on the map (see figure C.5).

105

Mode 3

Critical speed, cpm

Mode 2

Mode 1

104 1 x pinion

KXX -- 50% LD

8000 cpm

KXX -- 75% LD

1 x bull gear

4000 cpm

KXX--100%LD KYY -- 50% LD KYY -- 75% LD KYY--100%LD

103

105

106 107 Bearing support stiffness, N/mm

108

Figure C.5 -- Undamped critical speed map

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C.6.2 Damped critical speed analysis The damping approach is similar to the undamped map except that it is evaluated using full bearing spring and damping characteristics, including cross coupling terms. Damping in gear bearings is significant and the first two mode shapes generally show significant movement in the bearings, thereby utilizing the available damping (see figure C.6). This tends to give a result closer to the real world when evaluated, considering that frequencies with damping ratios greater than 0.2--0.3 will not be responsive where indicated. It gives results which agree very closely with the damped response analysis for the flexural mode of vibration which is generally the real critical speed where response will occur. This is because of little movement at the bearings and corresponding small damping in the system for this mode.

ξ= 0.01

Figure C.6 -- Bearing damping The degree of damping or likelihood of response is shown via logarithmic decrement or damping ratio values. See figure C.7. This stability type analysis can also identify sub--synchronous vibration potential such as half

ξ = 0.011

ξ = 0.011

ξ = 0.012

NF3 Bend

8 7

ξ = 0.6

6 5

NF1 Bounce

ξ = 0.5

4 ξ = 0.47

ξ = 0 .45

3

NF1 Rock

ξ = 0.50

ξ = 0.46 (Max cont speed) Mating shaft

2 ξ= 0.4

ξ= 0.43 1 2000

3000

4000

(Max cont speed) Shaft studied

Natural frequency, cpm X 10 3

10 9

5000 6000 Shaft rotating speed (rpm)

7000

8000

9000

Figure C.7 -- Damped critical speed map--natural frequency versus rotational speed load

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C.6.3 Damped response analysis The damped response analysis is generally considered to be the most useful of the tools for evaluating rotor synchronous vibration. It gives excellent correlation with actual machines. By definition, a critical speed is the speed which corresponds to the 110

frequency of the peak in vibration response to excitation. The damped response analysis includes all the effects from both damping and cross coupling. It will not indicate stability problems. In the analysis, it is generally best to specify unbalance forces several times larger than the actual rotor balance specification allows. Unbalanced force stations must be selected to excite the particular mode of vibration in question. The unbalance should be applied at several places along the rotor in successive runs to ensure that each mode will be excited. Coupling end, midspan, and blind end locations should be run as a minimum. Coupling end unbalance will usually excite the most common mode seen (see figures C.8, C.9 and C.10).

Max AMP 99 mm at 20 800 rpm

Dia amplitude ( m m)

90 70 50 30 10

0

9000

18 000

27 000

Speed (rpm) Figure C.8 -- Unbalance modeled at coupling 110

MAX AMP 86 mm at 20 800 rpm

Dia amplitude ( m m)

90 70 50 30

10

0

9000

18 000

Speed (rpm)

27 000

Figure C.9 -- Midspan

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frequency whirl, which can occur with unloaded gears. Here a growth factor is calculated for each mode. If the factor has a negative value, the system is inherently stable. If the value is positive, the system may be unstable. This analysis should also be performed over the load range if applicable. The damped natural frequency analysis yields more information but can be difficult to interpret if one is not familiar with evaluating the effect of the damping ratio.

AMERICAN NATIONAL STANDARD 110

ANSI/AGMA 6011--I03 MAX AMP 98 mm at 20 800 rpm

Dia amplitude ( m m)

90 70 50 30 10

0

9000

18 000

27 000

Speed (rpm)

In high speed gear drives with large L/D bearings, it is generally accepted that NF1 (bounce mode) and NF2 (rock mode) are heavily damped and unresponsive. When heavily damped (damping ratio greater than 0.3), these bearing modes may fall within the 20% band width around the rotating speed--natural frequency line. The acceptability may be proven either by response analysis or by the damping ratio of actual damping/critical damping. A term called the amplification factor determines when a response peak is to be treated as a real critical speed or if the frequency tends to be critically damped. Amplification factors less than 2.5 are considered to be critically damped. It is not the normal case to be able to evaluate the accuracy of a critical speed calculation for a gear

Vibration level

RRE Act 0.707 PEAK

Nmc Ncm Nct Ncp

drive. This is because the criticals are usually designed to be at operating speeds higher than the rest of the drive may be able to withstand. Bearing temperature or centrifugal stress considerations usually limit the maximum operating speed. The only thing that can usually be verified is that the actual critical is above design speeds, but not the actual critical speed frequency. This is determined by not measuring any peak in response over the speed range of the machine. Evaluating the undamped and damped natural frequencies as well as the damped response analysis is the most complete way to determine if a gear drive rotor will have dynamics problems. If only one tool can be available, the most reliable overall results will be obtained with the damped response analysis.

= rotor first critical center frequency, cycles per minute = initial (lesser) speed at 0.707 x peak amplitude Ncm (critical) = final (greater) speed at 0.707 x peak amplitude Ncp (critical) Ncp -- Ncm = peak width at the “half power” point AF = amplification factor N ct = N cp − N cm RRE = resonance response envelope = amplitude at Nct Act Nct

Figure C.11 -- Amplification factor

31

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Figure C.10 -- Blind end

ANSI/AGMA 6011--I03

AMERICAN NATIONAL STANDARD

Annex D (informative) Systems considerations for high speed gear drives [The foreword, footnotes and annexes, if any, are provided for informational purposes only and should not be construed as a part of ANSI/AGMA 6011--I03, Specification for High Speed Helical Gear Units.]

D.1 Purpose

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The need for high mechanical reliability in geared drives can best be satisfied by a “systems approach” to the entire train of machinery including foundations, lubrication, vibration, the forces and moments associated with piping, couplings, etc. The purpose of this annex is to point out common problems that may occur in geared systems, an explanation of these problems, and the possible effects. It is not the intent of this annex to present detailed methods of analyzing or solving the problem, nor will there be any attempt to set design criteria or limits. D.2 Responsibility A gear unit is susceptible to a variety of problems when it becomes a part of a rotating machinery system, the severity of which generally increases with speed. Even though these problems are generally beyond the vendor’s control, they adversely affect system reliability and/or performance and may cause damage to the gear unit. The party having contractual responsibility for system performance should investigate and resolve these problems in the design stage and thereby avoid the conflicts that may develop between the component manufacturers and users. It is recommended that the party having contractual responsibility for the system analysis involving a critical service gear drive be clearly identified in the specifications, contract or purchase order. Because of the substantial cost involved in a system analysis, and in some cases the system performance, it should be emphasized that all parties supplying components to the system have a responsibility to furnish correct and accurate data so that the analysis will be meaningful.

coupling, gear, driven equipment, or any other component. The increasing demands for system “mechanical reliability” can best be satisfied by a coordinated technical exchange between designer, equipment supplier, erecting engineers, and user. The various system analyses, in at least preliminary form, should precede detailed equipment purchase specifications. This sequence will permit the design to be based on more nearly correct load and operating conditions. This coordinated effort can be properly called “system engineering” and is normally performed by the design agent or his technical representative. Gear vendors may not have the expertise nor the detailed information to adequately analyze system overload. This function must be performed by specialists under the responsibility of the systems engineer. There is no set format for communicating this data. The required information is the magnitude of overload and a description of the operational conditions under which it occurs, such as when, how long, and nature. Gear units and couplings can be adversely affected by one or more system generated problems. Failures that result from these system induced causes can be categorized under three main headings: -- those resulting from overstressing component parts, which are grouped under “overload”; -- alignment related, such as distorted foundations or poor alignment with connected machinery; -- those resulting primarily from a lubrication related failure.

D.3 Introduction

D.4 Overloads

It is not uncommon to find daily process system operating costs many times the cost of the gear unit. This downtime cost makes it desirable to avoid failure of any part in the system ---- be it prime mover,

For the purpose of this discussion overload will be defined as:

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“That load which is in excess of the nominal design point load.”

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Overload can be of momentary duration, periodic, quasi--steady state, or vibratory in nature. Depending on its magnitude and the number of stress cycles accumulated at overload, it can be a fatigue or a yield stress consideration. Overload on a gear drive can result from internal or external causes. Internal cause of overload ---- such as faulty manufacturing (faults of manufacture) are usually found by routine inspections before the gear drive is put into service. External sources of overload result from the operational characteristics of the system into which the gear drive is placed, and are more complex and difficult to identify. The gear vendor has little if any control over the external influences that produce overload. The system engineer who has overall responsibility for performance should include, along with output, unit cost, efficiency, etc., the investigation of overloads as they relate to potential failure, downtime, and system reliability.

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The following material is intended to assist the system analyst by highlighting subjects for his consideration, and to establish better communication between system people and the vendor. D.4.1 Estimated maximum continuous power Operational overload characteristics of various driven equipment vary with the type of machine and should be considered on an individual basis. Pump or compressor designers, for example, can predict the power requirements at the design point with fairly good accuracy. However, continuous power (service power) is a combination of: -- changes in specific gravity or density of the media being pumped; --

carry out;

--

overspeed;

-- variations in pressure ratio across a compressor due to abnormal operating conditions. Changes in specific gravity of the fluid medium handled by a pump, or change in density of the gas handled by a compressor, affect the power transmitted in direct proportion. On boiler feed pumps, for example, this occurrence can be encountered during startup, upon malfunction of pre--heating

ANSI/AGMA 6011--I03

equipment, or during boiler cool--down following a failure. In the case of air handling centrifugal compressors, design power is usually based on the normal maximum ambient temperature. Consideration should be given to cold weather operation since the density of air varies with absolute temperature. Compressors handling other gases are usually encountered in process systems under greater control where temperature variations are less. However, other variables may become serious. In refinery practice, for example, the composition of the gas can vary widely, and in other process work the inlet pressure may not be a fixed value. Carry out is an expression used by the pump and compressor industries to indicate performance on a head curve beyond the so--called design point. Figure D.1 illustrates a typical compressor percentage performance curve. It will be noted at 100% speed as the head drops off and flow is increased, power increases to a level as high as 115% load. Carry out is an everyday reality. It comes about through such things as improper estimation of system performance during design stages, altered system requirements of existing processes, gradual deterioration of processes, systems employing multiple units where shutdown or failure of one increases the requirements on the remaining units, or through leaks or failures. Figure D.2 illustrates a similar percentage performance curve for centrifugal pumps. Overspeed is just what the name implies, and is obviously limited to applications with variable speed prime movers. Because the power absorption of the driven machine varies approximately with the third power of speed, overspeed is a large contributor to overload. Referring again to figure D.1, the performance curve indicates that at 110% speed and 100% flow, power is increased to 125%. Carry out at this speed can increase the power still further, to levels approaching 140% of service power. Normal practice for a turbine driven centrifugal pump is to set the overspeed trips at 115% design speed. Governor settings are generally established to permit continuous operation between 105% and 110% design speed. It should be borne in mind that operators can and do reset governors to avail themselves of maximum output of the system, regardless of the original settings.

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140

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Head at 110% speed

% Head and % power

130

120

Head at 100% speed

110

100

90

Power at 110% speed Power at 100% speed 60

70

80

90 % Flow

100

110

120

130

140

130

140

Figure D.1 -- Typical centrifugal compressor performance curve Head at 110% speed 140

Head at 100% speed

120

110

Power at 110% speed --`,``,`,`,`,``````,`,``,,`,,`-`-`,,`,,`,`,,`---

% Head and % power

130

100

90

Power at 100% speed

60

70

80

90 % Flow

100

110

120

Figure D.2 -- Typical centrifugal pump performance curve

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D.4.2 Vibratory overloads An essential phase in the design of a critical service system of rotating machinery is the analysis of the dynamic (vibratory) response of a system to excitation forces. The dynamic response of a system results in additional loads imposed upon the system and relative motion between adjacent elements in the system. The vibratory loads are superimposed upon the mean running load in the system and, depending upon the dynamic behavior of system, could lead to failure of the system components. In a gear unit these failures could occur as tooth breakage or pitting of the gear elements, shaft breakage or bearing failure. Due to the backlash between the geared elements of a gear unit, tooth separation will occur when the vibratory torques in the shafts exceed the average torque, resulting in tooth separation and subsequent impacts. Gear tooth loads due to these impacts can be several times the vibratory torque in the gear shafts.

--`,``,`,`,`,``````,`,``,,`,,`-`-`,,`,,`,`,,`---

A vibratory torque which is synchronized to the rotation of a gear element can form a cyclic wear pattern on the gear. This wear, which varies around the circumference on the gear element, results in tooth spacing errors of the gear causing noise or even can become a self--generating excitation which reinforces the original excitation. Vibratory motion of gear unit components can take up clearances causing interference problems between gearing elements, or between shafting and bearings or seals. D.4.2.1 Vibration analysis Any vibration analysis must consider the complete system including prime mover, gear unit, driven equipment, couplings and foundations. Dynamic loads imposed upon a gear unit are the result of the dynamic behavior of the total system and not that of the gear unit alone. The individual components of the system are usually supplied by different manufacturers. Therefore, responsibility for performing the vibration analysis must rest with the designer of the total system or his designated agent. The vibration analysis must determine all significant system natural frequencies and evaluate the system response to all potential excitation sources. If the analysis indicates a resonant or near resonant

ANSI/AGMA 6011--I03

condition, the recommended solution is to shift natural frequencies by changing stiffness or mass instead of relying on system damping to limit vibratory amplitudes. Normally, a linear vibration analysis is adequate. However, under certain conditions nonlinear responses can occur and the possibility of their existence should be recognized. It is also advantageous to perform a preliminary vibration analysis early enough in the design procedure to allow for any changes which might be required for detuning purposes. D.4.2.2 Torsional vibration The vibratory load caused by a steady state torsional vibration of a system is due to the interaction of a periodic excitation, and a natural frequency of the system. The magnitude of the dynamic load caused by this type of vibration is dependent on three factors: magnitude of the excitation, amount of damping in the system, and proximity of the excitation frequency to resonance. Typical sources for steady state excitation are: --

internal combustion engines;

--

reciprocating pumps and compressors;

--

pump or compressor impellers.

A torsional vibration in a system can also be caused by a transient excitation which is often called a shock or impact loading. Transient conditions occur due to sudden changes in load or speed, or the accelerating or decelerating through system natural frequencies, including the A.C. component of synchronous motors during startup. This type of disturbance will produce oscillations at all the natural frequencies of the system. These oscillations will decay and eventually disappear due to damping. The peak dynamic loads occur during or directly after the disturbance and their magnitudes are not substantially reduced by the damping in the system. Effects of the transient class of vibration can be most severe in the case of gear teeth due to their ability to separate, thus producing impact loadings on the teeth. D.4.2.3 Lateral vibration Dynamic loads at a gear mesh can be caused by a lateral vibration of a gear element in response to an excitation source. The lateral vibration of a rotor system should consider all flexibilities and restraints which will influence the vibratory response of the rotor. In the case of a rotor system comprised of a

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gear element and shaft, this should include the influence of bearings, foundations, couplings, connecting adjacent rotors and the mating gear element. The most common sources of lateral excitation in a rotor system are unbalance and misalignment. Therefore, care should be given to minimize these factors in the design, manufacture and installation of a rotating system. The lateral response of the system should be evaluated based on the design tolerances for system unbalance and misalignment. Consideration must be given to operation in the proximity of lateral natural frequencies because large vibratory loads may result with relatively low excitation. Fluid film bearings are generally used to support rotors in critical service systems. These bearings possess stiffness and damping properties which vary with speed and load. These non--linear properties should be considered when calculating the lateral natural frequencies of the system. Under certain conditions of operation, these bearings can cause instabilities in the rotor motion which will impart dynamic loads on the gear mesh. D.4.2.4 Axial vibration Dynamic loads on a gear mesh are sometimes caused by what appears to be an axial vibration. This axial motion is most often the response of the gear element to unbalanced thrust forces. Common sources for these forces are malfunctioning or misaligned couplings, electric armatures mounted off their magnetic center, face runout of thrust collars or compressor wheels, and assembly errors. D.4.2.5 Vibration measurements and design considerations The results of any theoretical vibration analysis are only as accurate as the mathematical model which is developed to perform the calculations. The correctness of the model of the system is dependent on the accuracy to which the inertia, stiffness, damping and excitation can be ascertained. Since there is always the possibility of the actual system responding differently than the theoretical evaluation, consideration should be given to physically measuring the vibratory loads in the system at the time of initial startup. Obtaining test data related to operational loading on a system has the following advantages:

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--`,``,`,`,`,``````,`,``,,`,,`-`-`,,`,,`,`,,`---

-- establishes confidence that the rotating system will perform satisfactorily or indicate areas where corrective actions are required prior to a system failure; -- provide a basis for evaluation of systems that may be designed or manufactured in the future; -- pinpoint system excitations or non--linear responses which were not considered in any theoretical evaluation. In the design stages it is advantageous to provide design features in the system which would facilitate testing, such as ground surfaces and proper access points for pickups or strain gages. Also in the system design, if it is feasible, consideration should be given to field modifications that could be made with a minimum of operational downtime if damaging vibratory loads were encountered. An example of this would be providing both access to couplings and additional space for coupling changes for detuning purposes. D.5 Alignment D.5.1 Drive train alignment A gear unit by the nature of its operation is always connected to at least two other pieces of equipment. The successful operation of the gear unit is largely dependent on the alignment of these components. There are three distinct types of misalignment which must be considered between connecting component shafting. -- Parallel offset misalignment ---- when two shafts are not coaxial, but their axes are parallel; -- Angular misalignment ---- when two shafts are not coaxial, and their axes are not parallel; -- Axial misalignment ---- when the ends of the two shafts are not positioned to provide the required shaft separation under operating conditions. Misalignment during operation not only causes vibration, but superimposes bending stress on the shear stress due to transmitted torque. These stresses cannot be readily calculated but they warrant discussion so the designer can take precautions to minimize their effect. Perfect alignment is almost impossible to obtain; therefore, flexible couplings are used to minimize the effects of the inherent misalignment. However, “flexible” couplings, whether of the gear tooth, spring elements, flexing disc, or elastomeric type, produce forces and moments on their support-

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ing shafts when operating misaligned. The analytical determination of the magnitude of these forces and moments is not fully understood. It can be generalized that: -- the sense and direction are such that they try to bring the supporting shafts in line; -- significant bending moments may be imposed on supporting shafts; -- the magnitude of the forces and moments increases with larger angularity across the coupling; -- notwithstanding catalog claims for angular capacity, flexible couplings should not be looked upon as universal joints; they should be given the best possible alignment.

--`,``,`,`,`,``````,`,``,,`,,`-`-`,,`,,`,`,,`---

The designer, in order to obtain a greater mechanical reliability of a coupled shafting system must make a comprehensive assessment of the operating alignment. This is a system study and must include all elements of the system including bedplates and/or foundations. An accurate evaluation of thermal growth for all components from a valid and common reference line is required. Journal displacement within bearings, though generally smaller in magnitude, should be considered, particularly as it affects cold or static alignment checks. After determining the probable magnitude of alignment change from static and cold to dynamic and hot (including any periodic cyclic changes that may occur), select a coupling arrangement that provides enough length or span between flexible elements to keep angularity low, in the region of 5 minutes or lower. A hot alignment check is recommended at the time the unit is put in service. This should be performed when all temperatures have stabilized, and the system is transmitting rated power at rated speed. D.5.2 Foundations Another kind of alignment problem commonly encountered in geared systems is the misalignment of pinion and gear axes due to foundation or bedplate twistings or deflections. It should be recognized that gear units require foundations with sufficient rigidity to maintain alignment under operating loads. Reinforced concrete foundations with grouted--in soleplate are generally preferable to fabricated steel bedplates in terms of foundation stiffness, mass and damping characteristics. A concrete foundation of adequate section, on good soil or on sufficient piling,

ANSI/AGMA 6011--I03

is the best insurance to avoid unequal settling or twisting from other causes. Fabricated steel bedplates make convenient shipping and handling frames, but are generally designed for strength, not rigidity. They are frequently designed without consideration for the various piping and/or oil sump thermal expansion. Out--of-door installations on steel bedplates are particularly subject to cyclic bowing caused by the daily “rise and fall” of the sun. When steel bedplates are used, the designer should endeavor to achieve two things: -- arrange oil sumps, piping, and weather protection to minimize unsymmetrical thermal expansion; -- thoroughly investigate elastic deformation of the bedplate due to piping forces and moments; then design the bedplate to eliminate twisting at the gear supports. D.5.3 System piping The forces and moments imposed on pumps, compressors and turbines by their inlet and discharge piping are major factors in deflecting this equipment and causing operating misalignment. All efforts should be made to minimize piping effects. Lubricant supply and drain piping for the gear unit should be given similar consideration. D.5.4 Installation instructions The system designer should assemble and integrate complete and comprehensive installation instructions covering, as a minimum, such things as: -- soleplate, bedplate, machinery position and leveling details; --

foundation bolting and grouting details;

-- cold alignment data ---- including method of measuring, relative position, and sequence of alignment; -- keying, pinning and torquing details as required; --

pipe support and flange makeup details;

-- all other relevant details that would otherwise be left to the judgment of the job site mechanic. D.6 Additional lubrication considerations The continued successful operation and long life of a gear unit is dependent on the constant supply of a lubricating oil of proper quantity, quality, and condition. The lubrication system has five functions to perform:

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--

reduce friction;

--

transfer heat;

--

minimize wear;

--

transfer wear particles;

tics should be carefully selected and, if selected, frequently changed to avoid accumulative separation of the additives during operation. When exposed to high operating temperatures in excess of 90° C, rapid degradation will occur.

--

reduce rusting and corrosion.

D.6.2.2 Viscosity and viscosity index

Failure of the lubrication system to adequately perform any one or more of these functions may result in premature failure of the gear drive. D.6.1 Type of lubricant Two basic types of oils are used to lubricate gear drives: --

petroleum base;

--

synthetic.

There can be a wide variation in the lubricating qualities of oils within each of these general types. Oils are compounded to meet specific requirements for various applications such as gear oils, bearing oils, internal combustion oils, worm gear oils, etc. Therefore, it is important that an oil be selected meeting the recommendations supplied with the gear unit. Synthetic oils should never be substituted for petroleum base oils without the gear vendor’s approval, since these oils not only have different lubricating qualities, but also may not be compatible with materials used in the gear unit. D.6.2 Lubricant selection The correct type and viscosity of oil must be supplied in accordance with the vendor’s recommendations. The friction, wear, film strength and corrosion protection characteristics of different types of oils can vary widely. Deviation from the recommended oil for the gear drive can result in premature wear, failure, or both. D.6.2.1 Lubricant quality Lubricating oils for high speed gear units should be high quality, refined, paraffin base petroleum oils. They must not be corrosive and must be free from grit or abrasives. As they are oftentimes subject to large flow rates and high operating temperatures, they must have good antifoaming properties. Oils of a straight mineral type should be used. High quality rust and oxidation resistance is desirable. Oils with additives which enhance these characteris-

Oils refined into lubricants are generally derived from two types of crude oil, either paraffin base or naptha base. Paraffin based oils are preferred because they have better natural extreme pressure characteristics and better resistance to “thinning down” at higher operating temperatures. Naptha based oils, on the other hand, require special additives in order to possess this benefit. The oil’s resistance to “thinning” is measured by the viscosity index. The higher the index value the better the resistance to “thinning”. Oils without additives of the paraffin base type usually have VI values of ninety (90) or above, whereas naptha base oils will exhibit lower values, oftentimes between twenty (20) and thirty (30). D.6.3 Oil film Gear elements and the supporting bearing system require a continuous supply of properly selected and conditioned oil for survival. An oil film of adequate thickness must be established between the rolling and sliding component surfaces to avoid damaging wear and scuffing and to provide component cooling. Hydrodynamic and elastohydrodynamic lubrication theories are commonly used today in analyzing film thickness in bearings and gear teeth. The oil viscosity has the greatest effect on the film thickness. Consequently, failure to use an oil that has both the proper viscosity and viscosity index can result in failure to produce an adequate film thickness for the gear teeth and bearings. Improper oil film thickness may cause several operational problems. Lack of oil film or inadequate oil film thickness may cause metallurgical drawing due to frictional heat of hardened surfaces, destructive wear, scuffing or pitting of the gear teeth, and frictional melting, plastic flow or failure of the babbitted bearing surfaces. Increased oil viscosity increases frictional power losses and therefore increases the temperature rise and may produce heat energy beyond the control of the cooling system. --`,``,`,`,`,``````,`,``,,`,,`-`-`,,`,,`,`,,`---

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D.6.4 Lubricant supply The oil supply must meet the requirements set forth in the gear vendor’s recommendations. D.6.4.1 Quantity The proper flow rate of oil must be supplied to the gear drive to ensure adequate oil film formation on the rotor elements, and in cases where babbitted bearings are employed, in the bearing journals, to prevent metal to metal contact of the respective elements. In addition, sufficient flow must be maintained to assure adequate cooling. Too small a quantity may cause inadequate distribution resulting in potential overheating, whereas too large a quantity may result in excessive churning of the oil which may also result in overheating. D.6.4.2 Pressurized lubrication systems When lubrication systems are self contained, the system should be designed with a flow capacity of a minimum of 10% greater than that initially required to allow for pump wear, slight bearing wear with normal service, or change in oil viscosity due to temperature variations and change of viscosity with use. Where pressurized oil is furnished from a central supply, operating, alarm and shutdown pressures must be in accordance with the gear unit vendor’s specifications. Pressures lower than that recommended may result in reduced flow and overheating. Pressures too high may cause excessive churning and possible gearbox flooding, increasing power loss and also resulting in overheating. Oil pressure to the gear drive should be measured either in the oil passages of the gear unit or at a point as near to the entry of the unit as possible, thus avoiding the inclusion of pressure losses in the piping between the point of measurement and the actual gear supply. D.6.4.3 Lubricant temperature The gear supplier will normally specify the minimum allowable oil temperature for startup. If temperatures lower than this are expected, provisions must be made to heat and, if possible, circulate the oil prior to startup. The gear drive must not be operated for extended periods at this minimum startup temperature.

Oil inlet temperature must be in accordance with the vendor’s specifications. A low supply temperature may result in a change in viscosity causing higher than expected temperature rise in the gear unit and improper oil distribution to the spray jets and bearings. When the oil supply temperature is higher than specified, the oil will be subject to rapid oxidation reducing the life of the oil, and reducing the operating viscosity resulting in an inadequate oil film. This condition can result in overheating, excessive wear and even failure. D.6.4.4 Pressurized system components The system components must be selected and installed to avoid problems. The following are some suggestions to avoid problems: -- Aeration. Care must be taken to avoid excessive aeration of the oil. Aeration may result in pump cavitation and decrease the volume of oil to come in contact with the elements of the gear drive; -- Oil reservoir. The reservoir must be large enough to allow time for the air to separate from the oil. Return lines to the oil reservoir should return below the oil level. This also includes relief valve bypass lines and any other return lines. These lines should be located as far away from the pump suction line as possible. Baffles properly located in the reservoir will ensure the aerated return oil does not find its way to the suction line until air has had time to escape from the oil; -- Drain lines. The location of the drain from the gear drive is critical, and the vendor’s recommendations should be followed. Drain lines should be sized so they run no more than half full of oil. The line should slope down at a minimum of (20 mm/m, 2%) and have a minimum number of bends and elbows. It is desirable to have a vent located in the drain line near the exit from the gear drive to insure proper drainage; -- Vents. Vents must be carefully located and of ample size to avoid pressure buildup and allow ready escape of air from the system without the loss of oil. Vents must be high enough to avoid entry of contaminants from the environment into the oil. Oftentimes it is desirable to place the vent in the drain line near the exit from the gear drive to ensure proper drainage. The oil is filtered prior to returning to the gear drive as well. In this manner direct contamination of the gear drive from the atmosphere outside is avoided;

39

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--`,``,`,`,`,``````,`,``,,`,,`-`-`,,`,,`,`,,`---

The lubrication system design must successfully achieve a balance of the viscosity and the oil film thickness considerations.

ANSI/AGMA 6011--I03

ANSI/AGMA 6011--I03

-- Suction lines. These lines should be generously sized to minimize pressure loss. Suction pressure (net positive suction head) must not be less than that recommended by the pump manufacturer. The total suction loss must include the loss in the piping, valves and fittings, in addition to the distance of the lift. If a check valve is used in the suction line of positive displacement pumps, a pressure limiting device should be installed to protect against the effects of reverse rotation of the pump; -- Flushing. Before oil is circulated through the gear drive, a bridge section containing a removable screen is fitted between the supply point and the drain. The system must be flushed until there is no significant accumulation of dirt on the screen. During flushing the piping should be hammer rapped to dislodge foreign particles. After flushing is completed, the supply and drain lines are connected to the gear drive. D.6.4.5 Lubricant condition

also important the oil be supplied and maintained in the proper condition. Dust, dirt, grit and other particles in the oil supply should be eliminated. These foreign matters act as an abrasive in the bearings and gear teeth, causing abrasive wear. The pressurized oil must be supplied through a filter as specified by the gear unit vendor. These filter systems should be serviced regularly to avoid circulation of contaminants with the oil and to avoid excessive pressure drops through the filters which may reduce the quantity of oil supplied to the gear drive. The oil must be maintained in its correct chemical condition to properly perform. Foreign matter, dirt and moisture can change the chemical properties of the oil. Additives used in many oils are depleted with use and require replacement. Since many factors influence the useful life of the oil, its condition should be analyzed on a regular basis to ensure its properties are within specification.

--`,``,`,`,`,``````,`,``,,`,,`-`-`,,`,,`,`,,`---

Having provided the proper type and grade of oil, it is

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ANSI/AGMA 6011--I03

Annex E (informative) Illustrative example [The foreword, footnotes and annexes, if any, are provided for informational purposes only and should not be construed as a part of ANSI/AGMA 6011--I03, Specification for High Speed Helical Gear Units.]

E.1 Purpose

E.2.3 Rating parameters

This annex provides examples based on the assumption that the gear set power rating is the minimum component rating. In practice all component ratings must be calculated to determine the lowest rated component.

The pitting resistance power rating and bending strength power rating at unity service factor are calculated per ANSI/AGMA 2101--C95 equations. With the factors that have a value of one (1.0) deleted, the equations are:

E.2 Example #1 E.2.1 Operational parameters The gearset to be rated transmits power from an induction motor rated at 2500 kilowatts and 1480 RPM to a centrifugal compressor operating at 5000 RPM. Annex A indicates that a service factor of 1.4 is appropriate for this service. E.2.2 Gearset parameters The through hardened double helical gearset to be rated has the following parameters: Number of teeth, pinion Number of teeth, gear Gear speed Module, normal Pressure angle, normal Helix angle Center distance Outside diameter, pinion Outside diameter, gear Normal circular tooth thickness at reference diameters, diameters pinion (182.76 mm) and gear (617.24 mm) Face width Overall face (gap included) Hardness pinion Hardness gear Pinion speed Material grade Gear quality level Cutter whole depth Cutter 1/2 pitch addendum Cutter tip radius

53 179 1480 rpm 3 mm 20° 29° 32’ 30” 400 mm 188.75 mm 623.24 mm

ω 1 d w1

1.91 × 10 7

where: ω1

= 4998.5 rpm

b

= 255 mm

ZI

= 0.22656 (see AGMA 908--B89)

dw1

= 182.76 mm

ZN

= 0.67313 (pinion) = 0.720 (gear)

ZE

= 190 [N/mm2]0.5

YJ pinion

= 0.56923 (see AGMA 908--B89)

YJ gear

= 0.58766 (see AGMA 908--B89)

mt

= 3.4483 mm (3/cos 29°32’30”)

YN

= 0.79531 (pinion) = 0.82720 (gear)

Kv

= 1.13

KH

= 1.2648 (see ANSI/AGMA 2101--C95)

4.63 mm 255 mm 300 mm 350 HB 300 HB 5000 rpm 2 A4 7.0 mm 3.8 mm 1.28 mm



2

b m t Y J σ FPY N K vK H 1 (see ANSI/AGMA 2101--C95, Eq. 28) P ayu =

CSF= KSF = 1.4 (see annex A) σHP

= 1079 N/mm2 (pinion @ 350 HB) = 958 N/mm2 (gear @ 300 HB) (see ANSI/AGMA 2101--C95, figure 8 Grade 2)

σFP

= 359 N/mm2 (pinion) (see ANSI/AGMA 2101--C95, figure 9 Grade 2)

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

ZI d w1σ HPZ N P azu = 7 ZE 1.91 × 10 K vK H (see ANSI/AGMA 2101--C95, Eq. 27) ω1 b

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= 324 N/mm2 (gear @ 300 HB) (see ANSI/AGMA 2101--C95, figure 9, Grade 2)

σFP

P azu = ×

4998.5 (255) (0.22656) 1.91 × 10 7 (1.13) (1.2648)

) (0.67313) 182.76 (1079  190

2

= 5163 kW (pinion) --`,``,`,`,`,``````,`,``,,`,,`-`-`,,`,,`,`,,`---

P azu = ×

4998.5 (255) (0.22656) 1.91 × 10 7 (1.13) (1.2648)

958) (0.72) 182.76 (190 

2

595.1 mm 0.1 0.0 260 mm 80 mm 58 HRC 2 A4 2.4 mm 14 mm 8 mm 0.25 mm

E.3.3 Rating parameters

= 4657 kW (gear) P ayu =

Outside diameter, gear Profile shift coefficient (x1), pinion Profile shift coefficient (x2), gear Face width Gap Hardness pinion and gear Material grade Quality level Cutter tip radius Cutter depth Cutter 1/2 pitch addendum Cutter protuberance

4998.5 (182.76) 255 (3.4483) (0.56923) 1.13 (1.2648) 1.91 × 10 7

359 (0.79531) = 4782 kW (pinion) 1 4998.5 (182.76) 255 (3.4483) (0.58766) P ayu = 1.13 (1.2648) 1.91 × 10 7 ×

324 (0.8272) × = 4635 kW (gear) 1 Pa is the lesser of 5163 4657 4782 4635 1.4 1.4 1.4 1.4 or 3311 kW E.2.4 Rating conclusions Pa is equal to the lesser of Pazu or Payu divided by the service factor, or Pa = 4635 1.4 = 3311 kW. This is greater than the service power of 2500 kW.

The pitting resistance power rating and bending strength power rating at unity service factor are calculated per ANSI/AGMA 2101--C95 equations. With the factors that have a value of one (1.0) deleted, the equations are:



ZI d w1σ HPZ N P azu = 7 ZE 1.91 × 10 K vK H (see ANSI/AGMA 2101--C95, Eq. 27) ω1 b



2

b m t Y J σ FPY N K vK H 1 (see ANSI/AGMA 2101--C95, Eq. 28) P ayu =

ω 1 d w1

1.91 × 10 7

where: ω1

= 8215.4 rpm

b

= 260 mm

E.3 Example #2

ZI

= 0.1730 (see AGMA 908--B89)

E.3.1 Operational parameters

dw1

= 256.91 mm

The gearset to be rated transmits power from a gas turbine rated at 15 MW and 8215 RPM to an electric generator operating at 3600 RPM on a base load cycle. The service factor is 1.3.

ZN

= 0.6547 (pinion) = 0.6856 (gear)

ZE

= 190 [N/mm2]0.5

YJ pinion

= 0.4722 (see AGMA 908--B89)

YJ gear

= 0.4861 (see AGMA 908--B89)

mt

= 6.5551 mm (6/cos 23.75)

YN

= 0.7826 (pinion) = 0.8038 (gear)

Kv

= 1.13

KH

= 1.2369 (see ANSI/AGMA 2101--C95)

E.3.2 Gearset parameters The carburized and case hardened double helical gearset to be rated has the following parameters: Number of teeth, pinion Number of teeth, gear Gear speed Module, normal Pressure angle, normal Helix angle Center distance Outside diameter, pinion

42 Copyright American Gear Manufacturers Association Provided by IHS under license with AGMA No reproduction or networking permitted without license from IHS

39 89 3600 rpm 6 mm 20° 23°45’ 421.6 mm 268.8 mm

CSF= KSF = 1.3 (see annex A) σHP

= 1550 N/mm2 (see ANSI/AGMA 2101--C95, table 3 Grade 2)

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ANSI/AGMA 6011--I03

= 450 N/mm2 (pinion and gear) (see ANSI/AGMA 2101--C95, table 4 Grade 2) 8215.4 (260) 0.1730 P azu = 1.91 × 10 7 1.13 (1.2369)

σFP



256.91 (1550) (0.6547) × 190



2

Pa is the lesser of 26 060 28 580 22 410 23 690 1.3 1.3 1.3 1.3 or 17 240 kW. E.3.4 Rating conclusions The allowable transmitted power, Pa = 17 240 kW, is greater than the service power of 15 MW.

= 26 060 kW (pinion) 8215.4 (260) 0.1730 P azu = 1.91 × 10 7 1.13 (1.2369)



256.91 (1550) (0.6856) × 190



2

= 28 580 kW (gear) 8215.4 (256.91) 260 (6.5551) (0.4722) P ayu = 1.13 (1.2369) 1.91 × 10 7 450 (0.7826) = 22 410 kW (pinion) 1 8215.4 (256.91) 260 (6.5551) (04861) P ayu = 1.13 (1.2369) 1.91 × 10 7 ×

×

450 (0.8038) = 23 690 kW 1

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(gear)

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43

ANSI/AGMA 6011--I03

AMERICAN NATIONAL STANDARD

Annex F (informative) Efficiency [The foreword, footnotes and annexes, if any, are provided for informational purposes only and should not be construed as a part of ANSI/AGMA 6011--I03, Specification for High Speed Helical Gear Units.]

F.1 Gear unit efficiency

F.2 Calculation methods

Most contracts for high speed helical gear units require some guarantee of minimum operational efficiency. When high power is transmitted, a very small increment of efficiency can represent substantial economic gain or loss over the life of the gear unit. To realize optimum gear unit efficiency, a detailed study of the several sources of power loss is required.

F.2.1 Mesh losses

Sources of power loss for high speed helical gear units include: mesh, internal windage, radial and thrust bearing friction, and shaft driven accessory power requirements. F.1.1 Mesh losses Mesh losses result from oil shearing and frictional losses which are dependent on the specific sliding velocity and friction coefficient. Most gear meshes under this standard will operate in the EHD lubrication regime. F.1.2 Internal windage losses Because of the sensitivity of gear and unit specific relationships (such as housing--to--rotor clearances, pitch line velocity, gear blank proportions and design, oil viscosity, method of mesh lubrication and cooling, horizontal or vertical offset, and internal baffling), this component of gear box losses is very difficult to accurately estimate without experimental data from a specific gear unit.

Hydrodynamic journal bearing losses are generated through oil shearing. Bearing losses may be calculated by a modified Petroff equation or by complex computer modeling methods. F.1.4 Accessory losses

z z +z z  1

2

1 2

(F.1)

where αn

is normal pressure angle of basic rack;

z1

is number of teeth in the pinion;

z2

is number of teeth in gear;

P

is transmitted power, kw.

F.2.2 Windage losses Windage and churning loss can be evaluated by the following equation: PW =

d′ 2n 2b cos 3 β′m n 1.42 × 10 −11 A

(F.2)

where PW

is windage power loss per gear, kW;

d′

is operating pitch diameter of gear, mm;

n

is gear speed, rpm;

b

is total face width, mm;

β′

is operating helix angle;

mn

is normal module, mm;

A

is arrangement constant (use 1000 to 4000, based on arrangement).

F.2.3 Bearing losses

d 3 L j 1.723 × 10 −17 P Bh = m n 2b b c

(F.3)

The thrust bearing power loss in kW, PBt, is: P Bt = m

The power consumed by shaft driven accessories can be computed by classic pressure -- displacement methods in the case of fuel or lube oil pumps. Accessories other than pumps require appropriate evaluation.

44

P M = (22 − 0.8 α n) 0.01 P

Hydrodynamic sleeve bearing loss in kW, PBh, can be estimated by the following equation:

F.1.3 Bearing losses

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Mesh power loss (PM), for 171/2° or 20° NPA of basic rack, can be estimated as below:

n 2b

r4o − r4i1.723 × 10−17

(F.4)

t

where m

is oil viscosity, mPa·s;

nb

is bearing speed, rpm;

db

is bearing bore, mm;

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ANSI/AGMA 6011--I03

t

is oil film thickness, mm;

ri

is inside radius of thrust bearing, mm;

ro

is outside radius of thrust bearing, mm;

L

is bearing length, mm;

c

is diametral clearance, mm;

j

is bearing power loss coefficient (see figure F.3).

F.2.5 Accessory losses Oil pump losses may be evaluated based on oil flow for lubrication and operating pressure: Pp =

10 −6

d m nb S= b 2 c w 60 where

(F.6)

where

The Sommerfield Number used in figure F.3 is calculated by the following: 2

Qp 60 000 e

(F.5)

Q

is pump displacement (l/min);

p

is pump operating pressure (kPa);

e

is pump efficiency (85% estimated). 60

w is load per unit area, kPa. 110

1

50 100

Absolute viscosity, mPa s

90 S

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Absolute viscosity, mPa s

80 S

70 60 50

3

2

20

20

10

3 4 5

1 0 40

10 0 40

2

30

40 30

1 2 3

40

50

60 70 80 Temperature, °C ISO Grade 46 ISO Grade 68 ISO Grade 100

90

100

Figure F.1 -- Viscosity of petroleum oil

1 2 3 4 5

50

60 70 80 Temperature, °C

100

Dow Corning XF--258 (Silicone) GE Versalube F--30 (Silicone) MIL--L0286B (Cellutherm 2505A) Mil--7808D Mil--L--25336 (Sinclair L--743) Figure F.2 -- Viscosity of synthetic oil

45

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90

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ANSI/AGMA 6011--I03

AMERICAN NATIONAL STANDARD

(A) j = Power Loss Coefficient

L/db = 0.25 0.375 0.5 0.75 1.0 1.25 1.5

10

(B)

5

2

0.2 0.5 0.1 Sommerfield Number, S Plot of j for elliptical bearings 1

0.05

0.02

0.01

j = Power Loss Coefficient L/db = 0.25 L/db = 0.5

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0.75 1.0 1.25 1.5

10

5

2

0.2 1 0.5 0.1 Sommerfield Number, S Plot of j for cylindrical bearings

0.05

(C)

0.02

0.01

L/db = 0.5

j = Power Loss Coefficient

0.75 1.0 1.25 1.5

10

5

2

1 0.5 0.2 0.1 Sommerfield Number, S Plot of j for four--groove bearings

0.05

0.02

0.01

Figure F.3 -- Bearing power loss coefficient, j

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AMERICAN NATIONAL STANDARD

ANSI/AGMA 6011--I03

Annex G (informative) Assembly designations [The foreword, footnotes and annexes, if any, are provided for informational purposes only and should not be construed as a part of ANSI/AGMA 6011--I03, Specification for High Speed Helical Gear Units.]

L--R

L--L

R--R

R--L

L--LR

LR--L

LR--R

R--LR

Plan views

LR--LR

Plan views

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NOTES: 1. Code: L = Left; R = Right 2. Arrows indicate line of sight to determine direction of shaft extensions. 3. Letters preceding the hyphen refer to number and direction of high speed shaft extensions. 4. Letters following the hyphen refer to number and direction of low speed shaft extensions.

Figure G.1 -- Parallel shaft spur, helical and herringbone gear drives, single or multiple stage

47

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ANSI/AGMA 6011--I03

AMERICAN NATIONAL STANDARD

Annex H (informative) Purchaser’s data sheet [The foreword, footnotes and annexes, if any, are provided for informational purposes only and should not be construed as a part of ANSI/AGMA 6011--I03, Specification for High Speed Helical Gear Units.]

H.1 Purpose Data sheets in SI and U.S. customary units are provided to facilitate communication between pur-

chaser and vendor. The purchaser should fill in the left side of the data sheet.

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ANSI/AGMA 6011--I03

DATA SHEET: ANSI/AGMA 6011--I03 HIGH SPEED GEAR UNITS SI UNITS PURCHASER:

JOB NO.

ITEM NO.

END USER SITE PROJECT NAME REVISION NO.

INFORMATION TO BE COMPLETED BY PURCHASER

INFORMATION TO BE COMPLETED BY VENDOR

1

APPLICABLE TO:

2

REQUISITION NO.

MODEL NO.

3

SERVICE

QUOTE NO.

4

DRIVER TYPE

5

DRIVEN EQUIPMENT

6

NO. REQUIRED

PURCHASE

MANUFACTURER

BASIC GEAR UNIT DATA FULL LOAD POWER LOSS

RATED SPEED, RPM:

kW (MIN)

%

PITCH LINE VELOCITY

m/sec

ANTICIPATED SPL

dBA @

m .

WR2 REFERRED TO LS SHAFT

11

INPUT

SPECIFIED

NOMINAL

BREAKAWAY TORQUE

12

OUTPUT

SPECIFIED

NOMINAL

NET MASS (WT) OF GEAR UNIT

. 2

Nms N--m @ LS Shaft

kg

13

MAX CONTINUOUS SPEED

RPM

MAX. MAINTENANCE MASS (WT) (IDENTIFY)

kg

14

TRIP SPEED

RPM

TOTAL SHIPPING MASS (WT)

kg

15

EXTERNAL LOADS

16

OTHER OPERATING CONDITIONS

TOTAL SHIPPING DIMENSIONS

17

OIL VISCOSITY: CONFIGURATION REQUIREMENTS

18 19

SHAFT ASSEMBLY DESIGNATION

20

HS SHAFT ROT FAC’G CPL’G

CW

CCW

21

LS SHAFT ROT FAC’G CPL’G

CW

CCW

22

HS SHAFT END:

23 24

LS SHAFT END:

25

(LR or RL)

CYLN.

DOUBLE STAGE

SINGLE HELICAL

DOUBLE HELICAL

ELECTRICAL AREA

CLASS

30

MAX ALLOW SPL

31

UNUSUAL CONDITIONS

BAROMETER

2--KEYS

GRP

DIV

dBA @ DUST

LOCATION

FUMES

TYPE

43

NO. AT EACH RADIAL BEARING

44

NO. AT EACH THRUST BEARING

PINION

GEAR RATIO

TOTAL NO.

NO. REQUIRED

NO. REQUIRED

GEAR

HELIX ANGLE

DEGREES

FINISH

NORMAL PRESSURE ANGLE

Ra DEGREES

mm PINION L/d

ADD MOD COEF:

GEAR

MIN HARDNESS:

GEAR

MANUFACTURING METHODS

BACKLASH

PINION PINION

/ GEAR

TEETH GENERATING

/

TEETH FINISHING

/

GEAR HUB TO SHAFT

mm

PINION

/

TEETH HARDENING NO. REQUIRED

mm

AGMA GEOMETRY FACTOR ”J”: PINION

m

GEAR CENTER DIST

NET FACE WIDTH

INTEGRAL

SHRUNK--ON

RIM ATTACHMENT

BEARING METAL TEMPURATURE SENSORS

42

GEAR TOOTH GEOMETRY

NORMAL DIAMETRAL PITCH

KEYPHASORS

LOCATION

°C kPa abs

AXIAL POSITION PROBES

40

TYPE OF GEAR

NUMBER OF TEETH

/

NO. AT EACH RADIAL BEARING

ACCELEROMETER

45

2--KEYS

INTEGRAL FLANGE

INSTRUMENTATION

39

41

TAPER

RADIAL VIBRATION PROBES

LOCATION

°C

CONSTRUCTION FEATURES

INSTALLATION DATA

38

MIN. STARTUP OIL TEMPERATURE

3

1-- KEY

29

37

m /hr

CYLN.

m

35

kPa

UNIT OIL FLOW (TOTAL)

INCREASER

ELEVATION

36

1-- KEY

cSt @ 100° C

REDUCER

28

34

TAPER

cSt @ 40° C

UNIT OIL PRESSURE

SINGLE STAGE

AMBIENT TEMPERATURE (MIN / MAX):

33

m

INTEGRAL FLANGE

27

32

X

HYDR’LC TAPER

HYDR’LC TAPER

26

X

LUBRICATION REQUIREMENTS

TOTAL NO.

ADDITIONAL REQUIREMENTS

46

BEARINGS PINION

GEAR

RADIAL TYPE UNIT LOADING, kPa JOURNAL VELOCITY, m/s

47

THRUST TYPE

48

UNIT LOADING, kPa

49

MEAN DIA VELOCITY, m/s

50

49

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GEAR SERVICE FACTOR

10

OTHER

MECHANICAL EFFICIENCY

GEAR SERVICE POWER

9

kW

∆T LUBE

MEASURED BY:

SPECIFIED RATING REQUIREMENTS

7 8

PROPOSAL

DATE

BY

ANSI/AGMA 6011--I03

AMERICAN NATIONAL STANDARD

1 2 3 4

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DATA SHEET: ANSI/AGMA 6011--I03 HIGH SPEED GEAR UNITS US CUSTOMARY UNITS PURCHASER:

ITEM NO.

END USER SITE PROJECT NAME REVISION NO.

INFORMATION TO BE COMPLETED BY PURCHASER APPLICABLE TO:

PROPOSAL

MANUFACTURER MODEL NO.

SERVICE

QUOTE NO.

DRIVER TYPE

BASIC GEAR UNIT DATA

DRIVEN EQUIPMENT

6

NO. REQUIRED

FULL LOAD POWER LOSS

GEAR SERVICE FACTOR

10

RATED SPEED, RPM:

OTHER

MECHANICAL EFFICIENCY

GEAR SERVICE POWER

9

HP

∆T LUBE

MEASURED BY:

SPECIFIED RATING REQUIREMENTS

7

DATE

BY

INFORMATION TO BE COMPLETED BY VENDOR

PURCHASE

REQUISITION NO.

5

8

JOB NO.

HP (MIN)

%

PITCH LINE VELOCITY

ft/min

ANTICIPATED SPL

dBA @

ft lb ft2

WR2 REFERRED TO LS SHAFT

ft lb @ LS Shaft

11

INPUT

SPECIFIED

NOMINAL

BREAKAWAY TORQUE

12

OUTPUT

SPECIFIED

NOMINAL

NET MASS (WT) OF GEAR UNIT

lb

13

MAX CONTINUOUS SPEED

RPM

MAX. MAINTENANCE MASS (WT) (IDENTIFY)

lb

14

TRIP SPEED

RPM

TOTAL SHIPPING MASS (WT)

lb

15

EXTERNAL LOADS

16

OTHER OPERATING CONDITIONS

TOTAL SHIPPING DIMENSIONS

17

OIL VISCOSITY: CONFIGURATION REQUIREMENTS

18 19

SHAFT ASSEMBLY DESIGNATION

20

HS SHAFT ROT FAC’G CPL’G

CW

CCW

21

LS SHAFT ROT FAC’G CPL’G

CW

CCW

22

HS SHAFT END:

23 24

LS SHAFT END:

25

(LR or RL)

CYLN.

DOUBLE STAGE

SINGLE HELICAL

DOUBLE HELICAL

ELECTRICAL AREA

CLASS

30

MAX ALLOW SPL

31

UNUSUAL CONDITIONS

2--KEYS

BAROMETER GRP

FUMES

LOCATION

NO. REQUIRED

NO. REQUIRED

NO. REQUIRED

TYPE

43

NO. AT EACH RADIAL BEARING

44

NO. AT EACH THRUST BEARING

45

ADDITIONAL REQUIREMENTS

DEGREES

FINISH

NORMAL PRESSURE ANGLE

GEAR

MIN HARDNESS:

GEAR

MANUFACTURING METHODS

Ra DEGREES

in

ADD MOD COEF:

PINION L/d BACKLASH

mil

PINION PINION PINION

/ GEAR

TEETH GENERATING

/

TEETH FINISHING

/ /

GEAR HUB TO SHAFT

INTEGRAL

SHRUNK--ON

RIM ATTACHMENT TOTAL NO.

46

in

GEAR

HELIX ANGLE

TEETH HARDENING

BEARING METAL TEMPURATURE SENSORS

42

PINION ft

NET FACE WIDTH

TOTAL NO.

GEAR CENTER DIST

NORMAL DIAMETRAL PITCH

KEYPHASORS

LOCATION

PINION

GEAR RATIO AGMA GEOMETRY FACTOR ”J”:

DIV

dBA @ DUST

°F ” Hg

AXIAL POSITION PROBES

40

GEAR TOOTH GEOMETRY NUMBER OF TEETH

/

NO. AT EACH RADIAL BEARING

ACCELEROMETER

TYPE OF GEAR

2--KEYS

INTEGRAL FLANGE

INSTRUMENTATION

39

41

TAPER

RADIAL VIBRATION PROBES

LOCATION

°F

1-- KEY

29

38

MIN. STARTUP OIL TEMPERATURE

CYLN.

ft

36

GPM

CONSTRUCTION FEATURES

INSTALLATION DATA

34

psi

UNIT OIL FLOW (TOTAL)

INCREASER

ELEVATION

37

1-- KEY

cSt @ 210° F

REDUCER

28

35

TAPER

SSU @ 100° F

UNIT OIL PRESSURE

SINGLE STAGE

AMBIENT TEMPERATURE (MIN / MAX):

33

ft

INTEGRAL FLANGE

27

32

X

HYDR’LC TAPER

HYDR’LC TAPER

26

X

LUBRICATION REQUIREMENTS

BEARINGS PINION

GEAR

RADIAL TYPE UNIT LOADING, psi JOURNAL VELOCITY, ft/sec

47

THRUST TYPE

48

UNIT LOADING, psi

49

MEAN DIA VELOCITY, ft/sec

50

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ANSI/AGMA 6011--I03

Bibliography The following documents are either referenced in the text of ANSI/AGMA 6011--I03, Specification for High Speed Helical Gear Units, or indicated for additional information.

AGMA 908--B89, Geometry Factors for Determining the Pitting Resistance and Bending Strength of Spur, Helical and Herringbone Gear Teeth AGMA 925--A03, Effect of Lubrication on Gear Surface Distress AGMA 927--A01, Load Distribution Factors -- Analytical Methods for Cylindrical Gears Ehrich, Fredric F., Handbook of Rotordynamics, McGraw--Hill, Inc., 1992

ISO 4406:1999 (SAE J1165), Hydraulic fluid power -- Fluids -- Method for coding the level of contamination by solid particles SAE/AMS 2300, Steel Cleanliness, Premium Aircraft--Quality Magnetic Particle Inspection Procedure SAE/AMS 2301, Steel Cleanliness, Aircraft Quality Magnetic Particle Inspection Procedure

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51

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PUBLISHED BY AMERICAN GEAR MANUFACTURERS ASSOCIATION 1500 KING STREET, ALEXANDRIA, VIRGINIA 22314

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