The NALCO Water Handbook Second Edition

The NALCO
Water
Handbook
Second Edition

Pepared by a staff of experts in the field and sponsored by the largest company in the world specializing in chemicals for water treatment, this is "the" comprehensive source book on water supply and treatment. It offers practical advice on how to improve water quality, optimize water usage and treatment processes, and avoid mistakes when dealing with vendors.
Here, in one convenient reference, is everything you need to know about the best use of water in any situation and the best way to condition it. The book fully covers these important topics:
Chemistry of water
Water sources
Water contaminants
Water treatment
Water disposal
Industrial use of water
Water used by municipalities
This new Second Edition has been revised and updated to reflect major advances in technology, in all aspects of the field, from analysis to unit operations. There is new coverage of coal and coke gasification...electric furnace steel and direct reduction steel production...ultrapure water as used in pharmaceutical plants, the electronics industry, and utility plants...acid rain...treating contaminated groundwater and toxic leachates...and chemical feed systems. New material is presented in the rapidly expanding areas of membrane separation and ion exchange. There's also a new chapter on energy which will be of particular interest to water chemists, especially those in utility plants.

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Erosion Control, Sediment Control and Stormwater Management on Construction Sites and Urban Areas

Erosion Control, Sediment Control
and Stormwater Management
on Construction Sites
and Urban Areas

The Handbook recognizes that erosion and runoff are influenced by the combination of climate, topography, soils, vegetative cover and the extent of land-disturbing activities. Because topography, soils, environmental conditions, and to a lesser extent local climate vary widely over the state, the application of the procedures and criteria in the Handbook should be tailored to local site-specific conditions and user objectives.
Erosion at construction sites and sediment-laden and turbid stormwater runoff impact individuals, our society and the environment. Damages occur on-site and off-site if land, water and related resources are degraded. Similar impacts may occur as a result of erosion in urban areas on non-construction sites.

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Environmental Effects on Engineered Materials edited by Russel H. Jones

Environmental Effects
on
Engineered Materials
edited by
Russel H. Jones

This invaluable reference provides a comprehensive overview of corrosion and environmental effects on metals, intermetallics, glossy metals, ceramics and composites of metals, and ceramics and polymer materials. It surveys numerous options for various applications involving environments and guidance in materials selection and substitution. Exploring a wide range of environments, including aqueous and high-temperature surroundings, Environmental Effects on Engineered Materials examines specific material-environmental interactions; corrosion rates and material limitations; preventive measurements against corrosion; utilization of older materials in recent applications; the use of new materials for existing equipment; and more.

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Contaminated Sediments: Characterization, Evaluation, Mitigation~Restoration, and Management Strategy Performance STP 1442

Contaminated Sediments:
Characterization, Evaluation,
Mitigation~Restoration, and
Management Strategy Performance
STP 1442



The papers gathered in this publication cover the primary goal of the symposium and reflect research
activities in many parts of the world. Keynote papers, selected for this volume, reflect recent work carried out on large coastal investigations (e.g., in the Los Angeles area), and on natural and artificial
capping of contaminated sediments. Other papers in this volume have been assembled into
three groups: (!) sediment characterization, (2) mitigation and restoration methods, and (3) monitoring and performance. Each of these sections begins with the corresponding keynote paper.

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SLURRY EROSION: USES, APPLICATIONS, AND TEST METHODS STP 946

SLURRY EROSION:
USES, APPLICATIONS,
AND TEST METHODS
STP 946


This publication, Slurry Erosion: Uses, Applications, and Test Methods, contains papers presented at the international symposium of the same name held in Denver, Colorado on 26-27 June 1984. The symposium was sponsored by ASTM Committee G-2 on Erosion and Wear, the National Association of Corrosion Engineers, the Slurry Transportation Association (now the Slurry Technology Association), and the American Society for Metals. John E. Miller, White Rock Engineering, and Frederick Schmidt, E.I. du Pont de Nemours & Co., presided as symposium chairmen and were coeditors of this publication.

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TENNESSEE EROSION AND SEDIMENT CONTROL HANDBOOK

TENNESSEE
EROSION AND SEDIMENT CONTROL HANDBOOK
A Guide for Protection of State Waters
through the use of Best Management Practices
during Land Disturbing Activities

This handbook serves as the primary reference for the development and implementation of Stormwater Pollution Prevention Plans (SWPPP), as required per the Tennessee General NPDES Permit for Discharges Associated with Construction Activities and individual NPDES permits. These permits allow the use of innovative or alternative BMPs or other controls, whose performance can be shown to be equivalent or superior to BMPs indentified in this handbook.

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An Introduction To Erosion Types , Causes ,Models And Prevention

An Introduction To Erosion Types , Causes ,Models And Prevention 


Introduction 
Erosion is used as a general term to refer to several types of mechanical processes in which a structure or component incurs physical damage. These processes generally remove some of the base material of the affected structure or component, structurally weakening the component.

 If the component has a protective coating or an oxide film, erosion processes may remove that coating. Erosion processes can affect many different materials; within the context of nuclear plant systems, they act primarily on metal structures and components. The following types of erosion processes are known to cause mechanical damage to structures and components, and are generally considered when evaluating metallurgical failures:
· Fluid Erosion
· Cavitation
· Solid Particle Erosion
· Impingement
· Flashing Induced Erosion
· Fretting
Fluid Erosion
Fluid Erosion is the mechanical removal of material by a moving fluid. This occurs when the shear stress (𝜏f) for a fluid constrained by a given structure or component exceeds the maximum allowable shear stress (Sv) of the material used for that given structure or component.

FIG 1

The surface of that structure or component will break down and be subject to wear by erosion (figure 1). The surface shear stress for a flowing fluid can be calculated from following relationship for turbulent flows:

where:
𝜌  is the fluid density (lbm/ft³)
f   is the Darcy friction factor
V  is the average flow velocity (ft/sec)

The material removed by fluid erosion can be either base metal or the protective film formed from corrosion products. Soft metals such as copper and specific copper alloys are very susceptible to erosion damage. Brass, aluminum brass and cupronickel are more susceptible to erosion damage than steel.
Cavitation
Cavitation occurs when the local pressure in a flowing fluid drops below vapor pressure(figure 2) , resulting in the formation of vapor cavities. When the local pressure rises above vapor pressure, the vapor cavities collapse. The collapse of the bubbles results in mechanical damage to the surrounding material due to the impingement of high velocity microjets and shock waves. Estimates of the microjet velocities range from 300 ft/sec to 3,000 ft/sec. For example in the Heater Drains system, cavitation could occur downstream of a level control valve.

FIG 2

Cavitation can occur in a variety of components, including pumps, turbines, valves, orifices and elbows. The four broad categories of cavitation are 
bulk, flow curvature, surface roughness and turbulence.
· Bulk cavitation may occur in locations where the flow velocity is increased – such as at the vena contracta (the contracted portion of a liquid jet at or near the orifice from which it issues) of a valve, or in the fluid stream near the impeller in a pump.

FIG 3

· Flow curvature cavitation occurs when a surface curves away from the direction of the flow or the flow curves to attach to a surface.

FIG 4

· Turbulence cavitation is usually associated with low recovery valves where high velocity, low pressure eddies can easily be formed.

FIG 5

· Surface roughness cavitation is caused by low static pressures in the wakes that form downstream of a surface protuberance or obstruction, such as a mound of weld metal or the presence of a backing ring at a weld.

FIG 6

All of these types of cavitation are the result of a reduction in static pressure due to changes in flow conditions. Cavitation can also have dramatic impacts on pump performance
Cavitation levels can be summarized as follows:
1· Incipient cavitation – Occurs intermittently over a restricted area. There is no objectionable noise or vibration and is considered acceptable.
2· Critical cavitation – Continuous light cavitation. Noise and vibration are acceptable and only minor damage is expected after long periods of operation (months to years).
3· Incipient damage – Continuous moderate to heavy cavitation. The onset of pitting will occur after short periods of operation; the noise levels may be objectionable.
4· Choking cavitation – Cavitation is severe enough to cause fluid flow to become choked. Reduction in pressure downstream of cavitation location does not cause a corresponding increase in flow. Noise and vibration levels reach maximum values.

Solid Particle Erosion
Solid particle erosion (figure 7) is caused by solid particles entrained in a fluid stream (usually liquid) impacting on the surface of a structure or component. In nuclear plants, solid particle erosion is most commonly seen in service water systems due to the entrainment of sand or silt.

FIG 7

Impingement
Impingement erosion (figure 8) is caused by liquid droplets entrained in a fluid stream (usually vapor) impacting on the surface of a structure or component. The impact of droplets can produce craters by plastic deformation of the component surface. The surface roughness caused by these deformations can increase the localized shear stress on the material and as a consequence can accelerate the degradation process. This form of material degradation is also referred to as droplet impingement erosion or liquid impact induced erosion. 
Impingement damage most commonly occurs in systems that contain wet steam or when water is injected into a steam filled system. Impingement can also occur as the result of partial blockage of a tube, resulting in deflection of the flow stream against the tube wall. Components that are commonly damaged by impingement include condenser tubes, turbine blades, valve seats and valve disks in nuclear plant systems as well as piston rings in engines.

FIG 8

Flashing Induced Erosion
Flashing induced erosion (figure 9)  is the result of spontaneous vapor formation caused by sudden pressure changes. This commonly occurs in drain and vent lines downstream of valves in liquid systems where the fluid is near saturation pressure. As some of the liquid flashes to vapor, it undergoes a rapid expansion of volume that increases the fluid velocity and accelerates the remaining liquid phase, liquid droplets and/or liquid film, leading to erosion. Typical examples are found downstream of Feed water Heater shell level control valves.

FIG 9

Fretting
Fretting occurs when tight fitting metal surfaces experience cyclic relative motion that causes them to impact on or rub against each other. The relative motion abrades one or both surfaces, producing debris from base metal or corrosion products. The debris may remain in contact with the original components, further increasing the abrasive effects of fretting. The debris can also prevent precision devices from operating due to fouling. Fretting has been a concern in steam generator tube bundles and control rods.

FIG 10

Mitigation and Prevention of Erosion
Erosion damage can generally be mitigated by using more resistant materials. Reduction in flow velocity will also reduce the amount of damage caused by most erosion mechanisms. The following sections discuss methods that may be used to prevent or mitigate some of the erosion type damage mechanisms discussed in this training guide.
Fluid Erosion Damage
is generally combated by using more resistant material and reducing flow velocities.
Cavitation
Cavitation Damage may be mitigated by using more resistant materials. Design alteration, including use of cascading orifices or low recovery valves, is a method for preventing cavitation.
Solid Particle Erosion 
Eliminating or reducing the solid particles is the preferred method of preventing or mitigating solid particle erosion. This could be done by installing strainers, or preventing unintended ingress. Use of a more resistant material and velocity reduction will also mitigate solid particle erosion.
Impingement
Eliminating or reducing moisture will combat impingement. Use of a more resistant material and velocity reduction will also mitigate impingement. Design solutions such as installing turning vanes, waste plates and improving flow conditions (such as using long radius elbows or bends) should be considered.
Flashing Induced Erosion
Use of resistant material will combat flashing induced erosion
Fretting 
Use of a more resistant material may help mitigate fretting. Elimination of the source of the relative motion will eliminate fretting. Improved mechanical fitup or use of lubricants may also mitigate or prevent fretting. Varying the location of metal to metal contact will mitigate the consequences of fretting by distributing the damage more widely.

THEORIES OF EROSION
Erosion is commonly measured in terms of a parameter W which is equal to the mass of material removed from the surface divided by the mass of the eroding material. Occasionally it is more convenient to refer the parameter to the volume loss divided by the volume of eroding material. In either case the parameter is dimensionless.

In most cases W > 0, a condition which indicates that material is removed during erosion but under certain circumstances W < 0.

DUCTILE MATERIAL MODELS
The trajectory of a particle cutting and removing material was calculated, and the eroded volume, V, was determined to be given by the expression (Finnie  1960):

where
𝜎f is flow stress,
m is the particle mass,
v𝗈  is the impact velocity,
K  is the ratio of vertical force to horizontal force on the particle,
d   is  the depth of cut.
g(𝛼) is a function describing the effect of attack angle 𝛼.

Sheldon and Kanhere formula

where
d is the (spherical) particle diameter,
𝜌p is the particle density,
H is the Vickers hardness value of the material.
This theory leads to a greater velocity dependence than expected from energy arguments (proportional to v² )

BRITTLE MATERIAL MODELS
The load at which crack propagation occurs is related to the distribution of surface flaws through the Weibull statistics. The approximate area. A, of cracked material is calculated for a particle penetration depth of h, and the volume removed per impact is set proportional to Ah. The final equation for the erosion rate , W, is expressed in terms of the particle size, r, the particle velocity, vo  and Weibull constants, m and 𝜎o  Sheldon and Finnie (1966):

where the exponents a and b are given by:
a = 3(m-0. 67) / (m-2) for round particles
a = 3 . 6 (m-0. 67) / (m-2) for angular particles
b = 2. 4 (m-0. 67) /(m-2) for either shape
For particles much stiff er than the target, the constant k1 , is given

where E is the modulus of elasticity of the target and 𝜌 is the density of the particle.

The erosion model developed by Evans  . assumes that the erosion rate is proportional to the amount of material removed by each impact event. The volume, V, lost per impact is calculated from the depth, h, of penetration and the maximum size of the lateral cracks formed during impact. Since the lateral crack size is proportional to the radial crack size cr , V is given by the following equation:

cr is crack formed during impact
you can follow that  it is wide range 

Erosion Wear Models
Finnie Model’s
A sharp edged particle moves into the specimen surface causing deformation and removal of
surface material

where,
Q is the volume of material removed,
M is total mass of the abrasives,
V is velocity of the abrasive particles,
p is constant horizontal component of the contact stress on the surface,
𝜑 is the ratio of depth of contact (l) to the depth of cut (yt),
K is the ratio of vertical to horizontal force components of the total force acting on the particle (considered constant and equal to 2)
α is the angle of approach (impact angle) for the particles.
Bitter’s model:
In Bitter’s model represented, deformation wear by WD and cutting wear by WC. Cutting wear has two component depends on the impact angle. Deformation and Cutting wear equation as follows:

where
WC1 = cutting wear loss if impinging particle stays in the target after collision,
WC2 = cutting wear loss if particle leaves the target surface after collision,
𝛜 & Q = wear factor of deformation and cutting wear, respectively,
V = incident velocity of erosive particle,
M = total mass of impinging particles,
K & K1 = constants determined by the elastic properties of target,
C = constant and
𝛼 = impact angle
Hutching’s Model:
Stack studied the hutching’s model of erosion and reviewed that, when spherical particle strikes on
the ductile material at normal, erosion wear occurs by the formation and subsequent detachment of
platelets of metal lying parallel to the eroded surface. Model gives information the about
detachment of platelets is only possible when the accumulated plastic strain within the fragment,
after many cycles of plastic deformation, reaches a critical value.

Where
Hd - Dynamic hardness
Dt - Density of the target material
Dp - Density of erodent particles
Ui - Erodent particle velocity.
The term 𝛼r/𝜖c ² cannot be measured independently. Hutchings assumed the value of 𝛼r/𝜖c ²  is equal to 0.7
Sundararajan and Shewmon erosion model
The basic idea of the model is that erosion loss is processed under high-strain-rate and hence adiabatic deformation conditions, and therefore, the mechanical response of the target is dynamic.
The criteria for the removal of such lip are assumed to be based on a critical strain which the lips attain after a number of particle impacts.

Where
Ui - velocity of impacting particles.
Dp – density
Cp - specific heat
Tm - melting temperature
Hs - static hardness of the target material

Rolling Bearing Analysis Tedric A. Harris

Rolling Bearing Analysis

Tedric A. Harris

One of the most well-known experts in the field brings cutting-edge research to practitioners in the new edition of this important reference. Covers the improved mathematical calculations for rolling bearing endurance developed by the American Society of Mechanical Engineers and the Society of Lubrication and Tribology Engineers. Updated with new material on Condition-Based Maintenance, new testing methods, and new bearing materials.

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ROLLING CONTACT FATIGUE TESTING OF BEARING STEELS ASTM SPECIAL TECHNICAL PUBLICATION 771

ROLLING CONTACT
FATIGUE TESTING OF
BEARING STEELS
ASTM SPECIAL TECHNICAL PUBLICATION 771

The symposium on Rolling Contact Fatigue Testing of Bearing Steels was held on 12-14 May 1981 in Phoenix, Ariz. Sponsoring the event was ASTM Committee A-1 on Steel, Stainless Steel, and Related Alloys and its Subcommittee A01.28 on Bearing Steels. The chairman of the symposium was J. J. C. Hoo, Acciaierie e Ferriere Lombarde Falck and Acciaierie di Bolzano, who also served as editor of this publication.

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SKF rolling bearings catalogue

SKF rolling bearings catalogue

SKF is a leading global supplier of bearings, seals, mechatronics, lubrication systems, and services which include technical support, maintenance and reliability services, engineering consulting and training. SKF is represented in more than 130 countries and has around 15,000 distributor locations worldwide. Annual sales in 2012 were SEK 64,575 million and the number of employees was 46,775. www.skf.com

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Rolling bearings — Explanatory notes on ISO 281 Part 2: Modified rating life calculation, based on a systems approach to fatigue stresses

Rolling bearings —
Explanatory notes on
ISO 281
Part 2: Modified rating life calculation,
based on a systems approach to fatigue
stresses

Since the publication of ISO 281:1990 , more knowledge has been gained regarding the influence on bearing life of contamination, lubrication, fatigue load limit of the material, internal stresses from mounting, stresses from hardening, etc. It is therefore now possible to take into consideration factors influencing the fatigue load in a more complete way.

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Residual stress modelling in laser welding marine steel EH36 considering a thermodynamics-based solid phase transformation

Residual stress modelling in laser welding marine steel EH36 considering a thermodynamics-based solid phase transformation

Low alloy steels are one of the most used materials in structural applications owing to their excellent properties such as ease of manufacturing, good toughness, and high strength. Among the available material joining methods, welding is undoubtfully the most suitable. This has led to significant research amongst scientists over the past years with the aim of improving the manufacturing processes. Generally, there are different types of welding whose application depends on the types of material and the intended application. Consequently, the microstructures and mechanical properties of these welds affect the properties and functionality of the materials and their entire structures in general.
Recent studies have shown significant improvements in the investigation of the joint performance of high strength steels using the residual stress, microstructure and weld pool flow. Unfortunately, considering the different nature of the arc welding and marine welding, the aforementioned criteria are insufficient to assure high-quality welding in marine manufacturing.
To this note, Huazhong University of Science and Technology scientists: Dr. Youmin Rong, Ting Lei, Dr. Jiajun Xu, Professor Yu Huang, Professor Chunming Wang assessed the distribution of residual stresses in laser welding in marine high strength steel EH36. In particular, a finite element model was designed by taking into consideration the solid transformation based on thermodynamics. Their research work is currently published in the journal, International Journal of Mechanical Sciences.
Briefly, the research team assessed the phase transformation and its effects of the residual stress by further taking into account the response of the microstructure to strain in laser welding marine high strength steel. Next, the distribution of the temperature was investigated using a heat source model while on the other hand, thermodynamics of the solid phase transformation was used in determining the microstructure fractions. To actualize their study, the research team experimentally verified the prediction accuracy of the designed model based on the residual stress, microstructure and weld profile.
The authors uncovered the usefulness of the index increment double cone method in fitting the penetration resulting from laser welding. As such, they recorded prediction errors of 11.06%, 10.24% and 6.69% in UW, MW, and BW respectively. On the other hand, a prediction error of 10.372% and 5.6435 were observed in the microstructures of the laser-welded EH36 and in particularly for martensite and ferrites. This was attributed to the influence of the weld microstructures on the strain and residual stresses of the material. However, it was worth noting that the heat affected zone, and not the center of the fusion zone, produced the maximum stress.
Therefore, the Huazhong University scientists successfully proposed a finite element model for not only accurately predicting the residual stresses in laser welded EH36 steels but also providing a basis for minimizing the welding associated stresses. Furthermore, considering the stability if the plastic strain without the need for external forces, the study will advance marine manufacturing through high-quality welds.

About the author
Dr. Youmin Rong, Assistant researcher/Postdoctor, State Key Lab of Digital Manufacturing Equipment and Technology, School of Mechanical Science and Engineering, Huazhong University of Science and Technology, China.
Prof. Yu Huang, State Key Lab of Digital Manufacturing Equipment and Technology, School of Mechanical Science and Engineering, Huazhong University of Science and Technology, China.

Ting Lei is a doctor of Mechanical and electrical engineering. He received his MS and BS in Mechanical Engineering from Huazhong University of Science & Technology and Changchun Institute of Technology in 2006 and 2010,

SOURCE LINK

Effect of Steel Manufacturing Processes on the Quality of Bearing Steels STP 987

Effect of Steel Manufacturing
Processes on the Quality of
Bearing Steels
STP 987

The fatigue life of rolling bearings has experienced a significant increase in the past several
years. The improvement is attributed to the lowering of oxygen content in the bearing steels to
the level of less than 10 parts per million by weight. This accomplishment is the direct result
of new ladle degassing practices adopted in the steel making processes. Associated with this
progress the methods used to evaluate the quality of bearing steels have been also improved,
refined, and in many instances newly developed.
It is very timely for the American Society for Testing and Materials Committee AOl,
Subcommittee AOl.28 on Bearing Steels to sponsor an international symposium on the theme
of effect of steel manufacturing processes on the quahty of bearing steels. The symposium was
held on 4 to 6 Nov. 1986 in Phoenix, Arizona. This is the third symposium sponsored by
ASTM Subcommittee AOl.28 since May 1974 on bearing steels. We have set up a target that
for every five to six years we will provide a forum for bearing steel producers and users to get
together to present their latest research results, the newest state of the art, and to discuss the
direction for future development. The response from the scientific and engineering community
in the world has been very enthusiastic. This symposium in Phoenix received papers from
Canada, France, Grermany, Italy, Japan, Netherlands, Sweden, United Kingdom, and the
United States. Almost all the major rolling bearing and bearing steel manufacturers in the
developed countries participated. We are extremely gratified for both the quaUty and quantity
of the papers received.

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Rolling bearings — Radial bearings, retaining slots — Dimensions and tolerances BS ISO 20515:2012

Rolling bearings — Radial bearings,
retaining slots — Dimensions and
tolerances
BS ISO 20515:2012

This International Standard specifies dimensions and tolerances of retaining slots to be used for outer rings of single-row angular contact ball bearings, four-point-contact ball bearings and radial cylindrical roller bearings.
The retaining slots are not suitable for use in the outer rings of sealed and shielded radial ball bearings, nor in the outer rings of radial cylindrical roller bearings without ribs.

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Rolling bearings — Linear motion rolling bearings — Part 2: Static load ratings ISO 14728-2

Rolling bearings — Linear motion
rolling bearings —
Part 2:
Static load ratings
ISO
14728-2

It is often impractical to establish the suitability of a linear motion rolling bearing selected for a specific application by testing. The following procedures have proved to be an appropriate and convenient substitute for testing:
— life calculation with dynamic load (ISO 14728-1);
— static load safety factor calculation with static load (ISO 14728-2).
Permanent deformation appears in rolling elements and raceways of rolling bearings under static loads of moderate magnitude and increases gradually with increasing load.
It is often impractical to establish whether the deformation appearing in a bearing in a specific
application is permissible by testing the bearing in that application. Other methods are therefore
required to establish the suitability of the bearing selected.
Experience shows that a total permanent deformation of 0,000 1 of the rolling element diameter, at the centre of the most heavily loaded rolling element/raceway contact, can be tolerated in most bearing applications without the subsequent bearing operation being impaired. The basic static load rating is, therefore, given a magnitude such that approximately that degree of deformation occurs when the static equivalent load is equal to the load rating.
Tests in different countries indicate that a load of the magnitude in question may be considered to
correspond to a calculated contact stress of
— 5 300 MPa for recirculating linear ball bearings, sleeve type,
— 4 200 MPa to 4 600 MPa for recirculating linear ball bearings, linear guideway type 
— 4 200 MPa to 4 600 MPa for non-recirculating linear ball bearings, and
— 4 000 MPa for linear roller bearings,
at the centre of the most heavily loaded rolling element/raceway contact. The formulae and factors forthe calculation of the basic static load ratings are based on these contact stresses.
The permissible static equivalent load may be smaller than, equal to or greater than the basic static
load rating, depending on the requirements for smoothness of operation and friction, as well as on
actual contact surface geometry. Bearing users without previous experience of these conditions should consult the bearing manufacturers.

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Rolling bearings — Linear motion rolling bearings — Part 1: Dynamic load ratings and rating life ISO 14728-1

Rolling bearings — Linear motion
rolling bearings —
Part 1:
Dynamic load ratings and rating life
ISO
14728-1

It is often impractical to establish the suitability of a linear motion rolling bearing selected for a specific
application by testing. The following procedures have proved to be an appropriate and convenient
substitute for testing:
— life calculation with dynamic load (ISO 14728-1);
— static load safety factor calculation with static load (ISO 14728-2).
The life of a linear motion bearing is given by the distance which one of the raceways moves, in relation
to the other raceway, before the first evidence of fatigue develops in the material of one of the raceways or one of the rolling elements.
The formula for calculating the basic dynamic load ratings are derived from the theory of Lundberg
and Palmgren

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ROLLING CONTACT BEARINGS COMMON PARTS

ROLLING CONTACT BEARINGS  COMMON PARTS 

An Introduction To Rolling Contact Bearings , Types , selection , And Applications

An Introduction To Rolling Contact Bearings , Types , selection , And Applications 

Introduction

The fundamental purpose of a bearing is to reduce friction and wear between rotating parts that are in contact with one another in any mechanism. The length of time a machine will retain its original operating efficiency and accuracy will depend upon the proper selection of bearings,
the care used while installing them, proper lubrication, and proper maintenance provided during actual operation.

Rolling-contact bearings are designed to support and locate rotating shafts or parts in machines. They transfer loads between rotating and stationary members and permit relatively free rotation with a minimum of friction. They consist of rolling elements (balls or rollers) between an outer and inner ring. Cages are used to space the rolling elements from each other. Figure 1 illustrates the common terminology used in describing rolling-contact bearings.

FIG 1

COMPONENTS AND SPECIFICATIONS

Rings The inner and outer rings of a rolling-contact bearing are normally made of SAE 52100 steel, hardened to Rockwell C 60 to 67.The rolling-element raceways are accurately ground in the rings to a very fine finish (16 min or less).Rings are available for special purposes in such materials as stainless steel, ceramics, and plastic. These materials are used in applications where corrosion is a problem.

Rolling Elements Normally the rolling elements, balls or rollers, are made of the same material and finished like the rings. Other rolling-element materials, such as stainless steel, ceramics, Monel, and plastics, are used in conjunction with various ring materials where corrosion is afactor.

Cages Cages, sometimes called separators or retainers, are used to space the rolling elements from each other. Cages are furnished in a wide variety of materials and construction. Pressed-steel cages, riveted or clinched and filled nylon, are most common. Solid machined cages are used where greater strength or higher speeds are required. They are fabricated from bronze or phenolic-type materials. At high speeds, the phenolic type operates more quietly with a minimum amount of friction.

Bearings without cages are referred to as full-complement.

Seals—Standard materials used in bearing seals are generally nitrile rubber. The material is bonded to a pressed steel core or shield. Nitrile rubber is unaffected by any type of lubricant commonly used in antifriction bearings. These closures have a useful temperature range of -70° to +225°F (-56° to 107°C). For higher operating temperatures, special seals of high temperature materials can be supplied. 

Lubricant—Prelubricated bearings are packed with an initial quantity of high quality grease which is capable of lubricating the bearing for years under certain operating conditions. As a general rule, standard greases will yield satisfactory performance at temperatures up to 175°F (79°C), as long as proper lubrication intervals and lube quantities are observed. Special greases are available for service at much higher temperatures. Estimation of grease life at elevated temperatures involves a complex relationship of grease type, bearing size, speed, and load. Volume 4 of this series can provide some guidance, although special problems are best referred to the product engineering department of major bearing manufacturers.

FIG 2

PRINCIPAL STANDARD BEARING TYPES

The general types are usually determined by the shape of the rolling element, but many variations have been developed that apply conventional elements in unique ways. Thus it is well to know that special bearings can be procured with races adapted to specific applications, although this is not practical for other than high volume configurations or where the requirements cannot be met in a more economical manner. “Special” races are appreciably more expensive. Quite often, in such situations, races are made to incorporate other functions of the mechanism, or are “submerged” in the surrounding structure, with the rolling elements supported by a shaft or housing that has been hardened and finished in a suitable manner.

In general, most ball, spherical roller, and cylindrical roller bearings made to metric boundary dimensions have standardized boundary plans, dimensions, and tolerances according to the International Standards Organization (ISO). Therefore, bearings from all subscribing manufacturers throughout the world are dimensionally interchangeable. Most taper roller bearings are made to inch dimensions and have standardized boundary dimensions and tolerances according to the Anti-Friction Bearing Manufacturers Association (AFBMA), a U.S. standards organization. Metric taper roller bearings utilizing ISO boundary plans are also made. Dimensionally interchangeable taper roller bearing components are thus available from several manufacturers. 

FIG 3

Ball Bearings

Single-Row Radial (Fig. 4) This bearing is often referred to as the deep groove or conrad bearing. Available in many variations—single or double shields or seals. Normally used for radial and thrust

loads (maximum two-thirds of radial).

Maximum Capacity (Fig. 4) The geometry is similar to that of a deep-groove bearing except for a filling slot. This slot allows more balls in the complement and thus will carry heavier radial loads. However, because of the filling slot, the thrust capacity in both directions is reduced drastically.

Double-Row (Fig. 4) This bearing provides for heavy radial and light thrust loads without increasing the OD of the bearing. It is approximately 60 to 80 percent wider than a comparable single-row bearing. Because of the filling slot, thrust loads must be light.

Internal Self-Aligning Double-Row (Fig.4) This bearing may be used for primarily radial loads where self-alignment (64°) is required. The self-aligning feature should not be abused, as excessive

misalignment or thrust load (10 percent of radial) causes early failure).

Angular-Contact Bearings (Fig. 4) These bearings are designed to support combined radial and thrust loads or heavy thrust loads depending on the contact-angle magnitude. Bearings having large contact angles can support heavier thrust loads. They may be mounted in pairs (Fig. 3.6) which are referred to as duplex bearings: back-toback, tandem, or face-to-face. These bearings  may be preloaded to minimize axial movement and deflection of the shaft.

Ball Bushings (Fig. 4) This type of bearing is used for linear motions on hardened shafts (Rockwell C 58 to 64). Some types can be used for linear and rotary motion.

Split-Type Ball Bearing (Fig.4) This type of ball or roller bearing has split inner ring, outer ring, and cage. They are assembled by screws. This feature is expensive but useful where it is difficult to install or remove a solid bearing.

FIG 4

Roller Bearings

Cylindrical Roller (Fig. 5) These bearings utilize cylinders with approximate length/diameter ratio ranging from 1 : 1 to 1 : 3 as rolling elements. Normally used for heavy radial loads. Especially useful for free axial movement of the shaft. Highest speed limits for roller bearings.

Needle Bearings (Fig.5) These bearings have rollers whose length is at least 4 times their diameter. They are most useful where space is a factor and are available with or without inner race. If shaft is

used as inner race, it must be hardened and ground. Full-complement type is used for high loads, oscillating, or slow speeds. Cage type should be used for rotational motion. They cannot support thrust loads.

Tapered-Roller (Fig. 5) These bearings are used for heavy radial and thrust loads. The bearing is designed so that all elements in the rolling surface and the raceways intersect at a common point on the axis: thus true rolling is obtained. Where maximum system rigidity is required, the bearings can be adjusted for a preload. They are available in double row.

Spherical-Roller (Fig.5) These bearings are excellent for heavy radial loads and moderate thrust. Their internal self-aligning feature is useful in many applications such as HVAC fans.

FIG 5

Thrust Bearings

Ball Thrust Bearing (Fig. 6) It may be used for low-speed applications where other bearings carry the radial load. These bearings are made with shields, as well as the open type.

Straight-Roller Thrust Bearing (Fig.6) These bearings are made of a series of short rollers to minimize the skidding, which causes twisting, of the rollers. They may be used for moderate speeds and loads.

Tapered-Roller Thrust (Fig 6) It eliminates the skidding that takes place with straight rollers but causes a thrust load between the ends of the rollers and the shoulder on the race. Thus speeds are limited because the roller end and race flange are in sliding contact

FIG 6

FIG 7

LOAD RATINGS

All manufacturers of rolling bearings establish a dynamic and static load rating for each bearing

produced. An ISO method for calculating this method exists, but not all manufacturers adhere to

the method. The unfortunate situation therefore exists that two almost identical bearings produced

by different manufacturers can have different published load ratings. Ratings are expressed as a

load which will provide a basic rating life of a defined number of revolutions. Basic rating life is

the number of revolutions (or the number of operating hours at a given constant speed) which the

bearing is capable of enduring before the first sign of fatigue occurs in one of its rings or rolling

elements. The basic rating life in millions of revolutions is the life 90 percent of a sufficiently large

group of apparently identical bearings can be expected to obtain or exceed under identical operating

conditions. In other words, this is a reliability or statistical rating, the only mechanical component

so rated. The ISO definition of the basic rating life is the most common and is at 1 million

revolutions. Some taper roller bearings are rated on the basis of 90 million revolutions, or 500

rev/min (rpm) for 3000 hours. Hence it can be easily seen that comparing manufacturers’ ratings

as published in catalogs can be misleading if appropriate adjustments are not made to published

values.

There are several other “bearing lives,” including service life and design or specification life.

Service life is the actual life achieved by a specific bearing before it becomes unserviceable. Failure

is not generally due to fatigue, but due to wear, corrosion, contamination, seal failure, etc.

The service life of a bearing depends to a large extent on operating conditions, but the procedures

used to mount and maintain it are equally important. Despite all recommended precautions, a bearing

can still experience premature failure. In this case it is vital that the bearing be examined carefully

to determine a reason for failure so that preventive action can be taken. The service life can

either be longer or shorter than the basic rating life.

Specification life is the required life specified by the equipment builder and is based on the hypothetical load and speed data supplied by the builder and to which the bearing was selected. Many times this required life is based on previous field or historical experiences.

ROLLING-CONTACT BEARINGS’ LIFE, LOAD, AND SPEED RELATIONSHIPS

An accurate knowledge of the load-carrying capacity and expected life is essential in the proper selection of ball and roller bearings. Bearings that are subject to millions of different stress applications fail owing to fatigue. In fact, fatigue is the only cause of failure if the bearing is properly lubricated, mounted, and sealed against the entrance of dust or dirt and is maintained in this condition. For this reason, the life of an individual bearing is defined as the total number of revolutions or hours at a given constant speed at which a bearing runs before the first evidence

of fatigue develops

Definitions

Rated Life     : The number of revolutions or hours at a given constant speed that 90 percent of an apparently identical group of bearings will complete or exceed before the first evidence of fatigue develops; i.e., 10 out of 100 bearings will fail before rated life. The names Minimum life and  life are also used to mean rated life.

Basic Load Rating C The radial load that a ball bearing can withstand for one million revolutions of the inner ring. Its value depends on bearing type, bearing geometry, accuracy of fabrication, and bearing material. The basic load rating is also called the specific dynamic capacity, the basic dynamic capacity, or the dynamic load rating.

Equivalent Radial Load P Constant stationary radial load which, if applied to a bearing with rotating inner ring and stationary outer ring, would give the same life as that which the bearing will attain under the actual conditions of load and rotation.

Static Load Rating   : Static radial load which produces a maximum contact stress of 580,000 lb/in² (4,000 MPa).
Static Equivalent Load      Static radial load, if applied, which produces a maximum contact stress equal in magnitude to the maximum contact stress in the actual condition of loading.
Bearing Rated Life
Standard formulas have been developed to predict the statistical rated life of a bearing under any given set of conditions. These formulas are based on an exponential relationship of load to life which has been established from extensive research and testing.

(1)

where = rated life, r; C = basic load rating, lb; P =equivalent radial load, lb; K = constant, 3 for ball bearings, 10/3 for roller bearings.
To convert to hours of life , this formula becomes

 (2)  

where N = rotational speed, r/min. Table 1 lists some common design lives vs. the type of application. These may be altered to suit unusual circumstances

Load Rating
The load rating is a function of many parameters, such as number of balls, ball diameter, and contact angle. Two load ratings are associated with a rolling-contact bearing: basic and static load rating.
Basic Load Rating C This rating is always used in determining bearing life for all speeds and load conditions
Static Load Rating  This rating is used only as a check to determine if the maximum allowable stress of the rolling elements will be exceeded. It is never used to calculate bearing life.
Values for C and  are readily attainable in any bearing manufacturer’s catalog as a function of size and bearing type. Table 2 lists the basic and static load ratings for some common sizes and types of bearings.

Equivalent Load
There are two equivalent-load formulas. Bearings operating with some finite speed use the equivalent radial load P in conjunction with C [Eq. (1)] to calculate bearing life. The static equivalent load is used in comparison with in applications when a bearing is highly loaded in a static mode.
Equivalent Radial Load P All bearing loads are converted to an equivalent radial load. Equation (3) is the general formula used for both ball and roller bearings.

(3)

where P = equivalent radial loads, lb; R = radial load, lb; T = thrust (axial) load, lb; X and Y 5 radial and thrust factors (Table 3). The empirical X and Y factors in Eq. (3) depend upon the geometry,
loads, and bearing type. Average X and Y factors can be obtained from Table 3. Two values of X and Y are listed. The set X1 Y1 or X2 Y2 giving the largest equivalent load should always be used.

Static Equivalent Load  The static equivalent load may be compared directly to the static load rating . If  is greater than the
rating, permanent deformation of the rolling element will occur. Calculate  as follows:

  (4)

where  = static equivalent load, lb; X0 = radial factor (see Table 4); Y0  thrust factor (see Table 4); R = radial load, lb; T = thrust (axial) load, lb.

Required Capacity

The basic load rating C is very useful in the selection of the type and size of bearing. By calculating the required capacity needed for a bearing in a certain application and comparing this with known capacities, a bearing can be selected. To calculate the required capacity, the following formula can be used:

 (5)

where C = required capacity, lb;  = rated life, h; P= equivalent radial load, lb; K =constant, 3 for ball bearings, 10/3 for roller bearings; Z =constant, 25.6 for ball bearings, 18.5 for roller bearings;

 N =rotation speed, r/min.

LIFE ADJUSTMENT FACTORS

Modifications to Eq. (2) can be made, based on a better understanding of causes of fatigue. Influencing factors include

1. Reliability factors for survival rates greater than 90 percent

2. Improved raw materials and manufacturing processes for ball bearing rings and balls.

3. The beneficial effects of elastrohydrodynamic lubricant films Equation (2) can be rewritten to reflect these influencing factors:

 (6)

where A1= statistical life reliability factor for a chosen survival rate, A2 = life-modifying factor reflecting bearing material type and condition, and A3 = elastohydrodynamic lubricant film factor.

Factor A3

This factor is based on elastohydrodynamic lubricant film calculations which relate film thickness and surface finish to fatigue life. A factor of 1 to 3 indicates adequate lubrication, with 1 being the minimum value for which the fatigue formula can still be applied. As A3 goes from 1 to 3, the life expectancy will increase proportionately, with 3 being the largest value for A3 that is meaningful. If A3 is less than 1, poor lubrication conditions are presumed. 

Speed Limits

Many factors combine to determine the limiting speeds of ball and roller bearings. It depends on several factors, like bearing size, inner- or outer-ring rotation, contacting seals, radial clearance and tolerances, operating loads, type of cage and cage material, temperature, and type of lubrication. A convenient check on speed limits can be made from a dn value. The dn value is a direct function of size and speed and is dependent on type of lubrication. It is calculated by multiplying the bore in millimeters (mm) by the speed in r/min.

dn = bore (mm) × speed (r/min)

Friction

One of the assets of rolling-contact bearings is their low friction. The coefficient of friction varies appreciably with the type of bearing, load, speed, lubrication, and sealing element. For rough calculations the following coefficients can be used for normal operating conditions and favorable lubrication:

        Single-row ball bearings               0.0015

        Roller bearings                             0.0018

Excess grease, contact seals, etc., will increase these values, and allowances should be made.

Selection of Ball or Roller Bearing

The selection of the type of rolling-contact bearing depends upon many considerations, as evidenced by the numerous types available. Furthermore, each basic type of bearing is furnished in several standard ‘‘series’’ as illustrated in Fig. 8. Although the bore is the same, the outside diameter, width, and ball size are progressively larger. The result is that a wide range of load-carrying capacity is available for a given size shaft, thus giving designers considerable flexibility in selecting

standard-size interchangeable bearings. 

FIG 8

Selection of the type of rolling-element bearing is a function of many factors, such as load, speed, misalignment sensitivity, space limitations, and desire for precise shaft positioning. However, to determine if a ball or roller bearing should be selected, the following general rules apply:

1. Ball bearings function on theoretical point contact. Thus they are suited for higher speeds and              lighter loads than roller bearings.

2. Roller bearings are generally more expensive except in larger sizes. Since they function theoretically on line contact, they will carry heavy loads, including shock, more satisfactorily, but are limited in speed.

Bearing Size Selection

Known type and series:

1. Select desired design life (Table 1).

2. Calculate equivalent radial load P [Eq. (3)].

3. Calculate required capacity Cr [Eq. (5)].

4. Compare Cr with capacities C in Table .2. Select first bore size having a capacity C greater than Cr .5. Check bearing speed limit [Eq. (7)].

Bearing-Type Selection Known bore size and life:

1. Select ball or roller bearing

2. Calculate equivalent load P [Eq. (3)] for various bearing types (conrad, spherical, etc.).

3. Calculate Cr [(Eq. (5)].

4. Compare Cr with capacities C in Table 3, and select the type that has a capacity equal to or greater than Cr .

5. Check bearing speed limit [Eq. (7)].

Bearing-Life Determination Known bearing size:

1. Select ball or roller bearing

2. Calculate equivalent radial load P [Eq. (3)].

3. Select basic load rating C from Table 2.

4. Calculate rated life  [Eq. (1) or (2)].

5. Check calculated life with design life.

Use Fig. 9  as a general guide to determine if a ball or roller bearing should be selected. This figure is based on a rated life of 30,000 h.

FIG 9

BEARING MOUNTING

Correct mounting of a rolling-contact bearing is essential to obtain its rated life. Many types of mounting methods are available. The selection of the proper method is a function of the accuracy, speed, load, and cost of the application. The most common and best method of bearing retention

is a press fit against a shaft shoulder secured with a locknut. End caps are used to secure the bearing against the housing shoulder (Fig.10). Retaining rings are also used to fix a bearing on a shaft or in a housing (Fig. 10). Each shaft assembly normally must provide for expansion by allowing one end to float. This can be accomplished by

FIG 10

allowing the bearing to expand linearly in the housing or by using a straight roller bearing on one end. Care must be exercised when designing a floating installation because it requires a slip fit. An excessively loose fit will cause the bearing to spin on the shaft or in the housing. Table 8 lists shaft and housing tolerances for press fits with ABEC 1 precision applications (pumps, gear reducers, electric motors, etc.) and ABEC 7 precision applications (grinding spindles, etc.).