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SKF Belt Drive Design Calculations tool

SKF Belt Drive Design Calculations tool

The SKF Belt Drive Design Calculations tool (version 3.1.2) allows field engineers and technical sales people to check the quality of an existing belt drive design on their desktops. In addition, the application allows users to propose more than 100 alternative solutions.

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BS 3790:1995 ..Specification for V-belt drives

BS 3790:1995 ..Specification for  V-belt drives                                 BS 3790:1995 ..Specification for  V-belt drives                                             

Standard Handbook of Chains

Standard Handbook of Chains

The first edition of this handbook was a landmark publication. It served the chain industry very

well for more than 20 years. It guided many engineers and technologists through their first chain

drive or conveyor selection. But the first edition was beginning to show its age. It was growing out

of date in several ways. It definitely was time for a major revision.

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Mechanisms and Mechanical devices sourcebook

Mechanisms and Mechanical devices sourcebook 

Mechanisms and Mechanical devices sourcebook e-book by the author Neil Sclater describes proven mechanisms and mechanical devices. Each illustration represents a design concept that can easily be recycled for use in new or modified mechanical, electromechanical, or mechatronic products. Tutorials on the basics of mechanisms and motion control systems introduce you to those Mechanisms.

Mechanisms and Mechanical Devices Sourcebook, Fifth Edition, contains new chapters on mechanisms for converting renewable energy into electrical power, 3D digital prototyping and simulation, and progress in MEMS and nanotechnology based on carbon nanotubes. This practical guide will get you up to speed on many classical mechanical devices as well as the hot new topics in mechanical engineering.

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Belt Drives case study

Belt Drives case study                                      Belt Drives case study                                                                 Belt Drives case study

Belt Drives case study

introduction

A belt drive is a method of transferring rotary motion between two parallel shafts. A belt drive includes one pulley on each shaft and one or more continuous belts over the two parallel pulleys. The motion of the driving pulley is transferred to the driven pulley via the friction between the belt and the pulley.

Belt Drive advantages
- Easy, flexible equipment design, as tolerances are not important.
- Isolation from shock and vibration between driver and driven system.
- Driven shaft speed conveniently changed by changing pulley sizes.
- Belt drives require no lubrication.
- Maintenance is relatively convenient
- Very quiet compared to chain drives, and direct spur gear drives

For belt drives, other than synchronous drives, the belts will slip in a high overload event providing a certain measure of safety.

The belts transferring torque by surface friction need to be in tension. This results in the need for adjustable shaft centres or using tensioning pulleys 

Types of Belt Drives

Flat Belt  transfers torque by friction of the belt over a pulley. Needs tensioner. Traction related to angle of contact of belt on pulley. Is susceptible to slip. Belt made from leather, woven cotton, rubber, balata.

Vee Belt Better torque transfer possible compared to flat belt. Generally arranged with a number of matched vee belts to transmit power. Smooth and reliable. Made from hi-text woven textiles, polyurethane, etc.

Poly-Vee Belt is flat on outside and Vee Grooved along the inside. Combines advantages of high traction of the Vee belt and the use of only one belt.

Timing/ Synchronous Belt toothed on the inside driving via grooved pulleys. This enables positive drive. Limited power capacity compared to chain and Vee belt derivatives. Does not require lubrication. Extensively used in low power applications

Vee Link Belts Linked belts that can be used in place of vee belts. Advantage that the length can be adjusted and the belt can be easily installed with removing pulleys. Expensive and limited load capacity.

Flexible Machine Elements

Belt drives are called flexible machine elements. Flexible machine elements are used for a large number of industrial applications, some of them are as follows.

1. Used in conveying systems Transportation of coal, mineral ores etc. over a long distance

2. Used for transmission of power.Mainly used for running of various industrial appliances using prime movers like electric motors, I.C. Engine etc.

3. Replacement of rigid type power transmission system.A gear drive may be replaced by a belt transmission system 

good amount of shock and vibration. It can take care of some degree of misalignment between the driven and the driver machines and long distance power transmission, in comparison to other transmission systems, is possible. For all the above reasons flexible machine elements are widely used in industrial application.

Although we have some other flexible drives like rope drive, roller chain drives etc. we will only discuss about belt drives

Typical belt drives

Two types of belt drives, an open belt drive, (Fig. 1) and a crossed belt drive (Fig. 2) are shown. In both the drives, a belt is wrapped around the pulleys. Let us consider the smaller pulley to be the driving pulley. This pulley will transmit motion to the belt and the motion of the belt in turn will give a rotation to the larger driven pulley. In open belt drive system the rotation of both the pulleys is in the same direction, whereas, for crossed belt drive system, opposite direction of rotation is observed.

fig 1 

Nomenclature of Open Belt Drive

dL- Diameter of the larger pulley

dS – Diameter of the smaller pulley

αL- Angle of wrap of the larger pulley

αS – Angle of wrap of the smaller pulley L

C- Center distance between the two pulleys

Nomenclature of Cross Belt Drive

dL- Diameter of the larger pulley

dS – Diameter of the smaller pulley

αL- Angle of wrap of the larger pulley

αS – Angle of wrap of the smaller pulley L

C- Center distance between the two pulleys

Belt tensions

The belt drives primarily operate on the friction principle. i.e. the friction between the belt and the pulley is responsible for transmitting power from one pulley to the other. In other words the driving pulley will give a motion to the belt and the motion of the belt will be transmitted to the driven pulley. Due to the presence of friction between the pulley and the belt surfaces, tensions on both the sides of the belt are not equal. So it is important that one has to identify the higher tension side and the lower tension side, which is shown in Fig. 3.

fig 3

When the driving pulley rotates (in this case, anti-clock wise), from the fundamental concept of friction, we know that the belt will oppose the motion of the pulley. Thereby, the friction, f on the belt will be opposite to the motion of the pulley. Friction in the belt acts in the direction, as shown in Fig. 3, and will impart a motion on the belt in the same direction. The friction f acts in the same

direction asT₂. Equilibrium of the belt segment suggests that T₁ is higher than T₂. Here, we will refer T₁ as the tight side and T₂ as the slack side, ie, T₁ is higher tension side and T₂ is lower tension side.

Continuing the discussion on belt tension, the figures though they are continuous, are represented as two figures for the purpose of explanation. The driven pulley in the initial stages is not rotating. The basic nature of friction again suggests that the driven pulley opposes the motion of the belt. The directions of friction on the belt and the driven pulley are shown the figure. The frictional force on the driven pulley will create a motion in the direction shown in the figure. Equilibrium of the belt segment for driven pulley again suggests that T₁ is higher than T₂.

It is observed that the slack side of the belt is in the upper side and the tight side of the belt is in the lower side. The slack side of the belt, due to self weight, will not be in a straight line but will sag and the angle of contact will increase. However, the tight side will not sag to that extent. Hence, the net effect will be an increase of the angle of contact or angle of wrap. It will be shown later that due to the increase in angle of contact, the power transmission capacity of the drive system will increase. On the other hand, if it is other way round, that is, if the slack side is on the lower side and the tight side is on the upper side, for the same reason as above, the angle of wrap will decrease and the power transmission capacity will also decrease. Hence, in case of horizontal drive system the tight side is on the lower side and the slack side is always on the upper side.

relationship between belt tensions

fig4 

The centrifugal force due to the motion of the belt acting on the belt segment is denoted as CF and its magnitude is

CF = [m(rdφ)v²]/r = mv²dφ

Where, 

v    is the peripheral velocity of the pulley

 m   is the mass of the belt of unit length,

m = btρ

where,

 b is the width, 

 t is the thickness 

 ρ is the density of the belt material

The final equation for determination of relationship between belt tensions is,

(T₁-mv²)/(T₂-mv²)=e^𝜇𝛼

where
μ is the coefficient of friction between the belt and the pulley.
α should be expressed in radians

Elastic Creep and Initial Tension

Presence of friction between pulley and belt causes differential tension in the belt. This differential tension causes the belt to elongate or contract and create a relative motion between the belt and the pulley surface. This relative motion between the belt and the pulley surface is created due to the phenomena known as elastic creep.

The belt always has an initial tension when installed over the pulleys. This initial tension is same throughout the belt length when there is no motion. During rotation of the drive, tight side tension is higher than the initial tension and slack

side tension is lower than the initial tension. When the belt enters the driving pulley it is elongated and while it leaves the pulley it contracts. Hence, the driving pulley receives a larger length of belt than it delivers. The average belt velocity on the driving pulley is slightly lower than the speed of the pulley surface. On the other hand, driven pulley receives a shorter belt length than it delivers. The average belt velocity on the driven pulley is slightly higher than the speed of the pulley surface.

Let us determine the magnitude of the initial tension in the belt.

Tight side elongation ∝ (T₁ – Ti )

Slack side contraction ∝ (Ti – T₂ )

Where, 

Ti is the initial belt tension .

Since, belt length remains the same, ie, the elongation is same as the contraction,

Ti=(T₁+T₂ ) /2

It is to be noted that with the increase in initial tension power transmission can be increased. If initial tension is gradually increased then T₁ will also increase and at the same time T₂ will decrease. Thus, if it happens that T₂ is equal to zero, then T₁ = 2Ti and one can achieve maximum power transmission.

Velocity ratio of belt drive

Velocity ratio of belt drive is defined as,

(N𝑙 /N𝗌) =[(d𝗌+t)/(d𝑙 +t)]×(1-𝑠)

where,

N𝑙 and N𝗌 are the rotational speeds of the large and the small pulley respectively, 

 s is the belt slip 

 t is the belt thickness

Power transmission of belt drive

Power transmission of a belt drive is expressed as,

P = ( T₁ – T₂ )v

where,

P is the power transmission in Watt

 v is the belt velocity in m/s.

Fundamentals of Fluid Film Lubrication bernard

Fundamentals of Fluid Film Lubrication bernard                                                           Fundamentals of Fluid Film Lubrication bernard                

Fundamentals of Fluid Film Lubrication

Fundamentals of Fluid Film Lubrication book by the authors Bernard J. Hamrock, Steven R. Schmid Specifically focuses on fluid film, hydrodynamic, and elastohydrodynamic lubrication, this book studies the most important principles of fluid film lubrication for the correct design of bearings, gears and rolling operations and the prevention of friction and wear in engineering designs thoroughly explaining various theories, procedures and equations for improved solutions to machining problems.

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DESIGN AND PRODUCTION IN ARABIC

كتاب عن التصميم و الصيانه و التركيب 

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water properties pro

water properties pro                                                     water properties pro                                                                     water properties pro     

water properties pro

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Mechanical Design Engineering Handbook

Mechanical Design Engineering Handbook

The aims of this book are thus to present an overview of the design process and to introduce the technology and selection of a number of specific machine elements that are fundamental to a wide range of mechanical engineering design applications. I hope it is useful to you.

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An Introduction To Journal Bearings

An Introduction To Journal Bearings 

Introduction

The term ‘bearing’ typically refers to contacting surfaces through which a load is transmitted.
Bearings may roll or slide or do both simultaneously.The range of bearing types available is
extensive, although they can be broadly split into two categories: sliding bearings also known as plain
surface bearings, where the motion is facilitated by a thin layer or film of lubricant, and rolling element bearings, where the motion is aided by a combination of rolling motion and lubrication. Lubrication is often required in a bearing to reduce friction between surfaces and to remove heat.
illustrates two of the more commonly known bearings: a deep groove ball bearing and a journal
bearing.

The term ‘sliding bearing’ refers to bearings where two surfaces move relative to each other
without the benefit of rolling contact. The two surfaces slide over each other and this motion can
be facilitated by means of a lubricant which gets squeezed by the motion of the components and
can generate sufficient pressure to separate them, thereby reducing frictional contact and wear.
A typical application of sliding bearings is to allow rotation of a load-carrying shaft. The portion of the shaft at the bearing is referred to as the journal and the stationary part, which supports
the load, is called the bearing . For this reason, sliding bearings are often collectively
referred to as journal bearings, although this term ignores the existence of sliding bearings that support linear translation of components. Another common term is ‘plain surface bearings’.This section is principally concerned with bearings for rotary motion and the terms ‘journal’ and ‘sliding’
bearing are used interchangeably.A general classification scheme for the distinction of bearings is given in Figure 2.

Journal Bearing Types
The types of journal bearings most commonly found in turbo machinery include axial groove bearing, pressure dam bearing, elliptical bearing, offset half bearing, taper-land and multi-lobe bearings and tilt pad bearing.

FIG2 

 Bearing classification based on type of load carried 
a. Radial bearings
b. Thrust bearings or axial bearings
c. Radial – thrust bearings
 Bearing classification based on type of lubrication 
The type of lubrication means the extent to which the contacting surfaces are separated in a shaft bearing combination. This classification includes
(a) Thick film lubrication
(b) Thin film lubrication
(c) Boundary lubrication
Bearing classification based on lubrication mechanism 
a. Hydrodynamic lubricated bearings
b. Hydrostatic lubricated bearings
c. Elastohydrodynamic lubricated bearings
d. Boundary lubricated bearings
e. Solid film lubricated bearings

1-Radial or Journal
These bearings carry only radial loads.
Plain radial, or journal, bearings also are referred to as sleeve or Babbit bearings. The most common type is the full journal bearing, which has 360-degree contact with its mating journal. The partial journal bearing has less than 180-degree contact and is used when the load direction is constant.

FIG 3

Plain Cylindrical Bearing
The plain cylindrical journal bearing (Figure  3 ) is the simplest of all journal bearing types. The performance characteristics of cylindrical bearings are well established, and extensive design information is available. Practically, use of the unmodified cylindrical bearing is generally limited
to gas-lubricated bearings and low-speed machinery.

FIG 4

Axial Groove Bearing
Axial groove bearings have a cylindrical bore with typically 2 to 4 axial oil feed grooves. (Figure 5) illustrates a two axial groove design. These bearings are very popular in relatively low speed equipment. The stability characteristics of axial groove bearings are primarily controlled by the bearing clearance.
Tight clearances produce higher instability thresholds but tight bearings present other problems that make them undesirable. Due to these, as clearance decreases, the bearing’s operating oil temperature increases.
Furthermore the Babbitt wear during repeated start-ups will increase the bearing clearance thereby degrading stability. In fact, many bearing induced instabilities in the field are caused by bearing clearances that have increased due to wear from oil contamination, repeated starts or slow rolling with boundary lubrication.

FIG 5

Pressure Dam Bearing
The Pressure dam or step journal bearings have long been used to improve the stability of turbo machinery as a replacement for axial groove bearings. In many cases, these bearings provide a quick and inexpensive fix for machines operating at high speeds near or above the stability threshold.
Milling a step in the top pad of the proper size and location may be all that is necessary to eliminate the stability problem. This is much less expensive and faster than installing tilting pad bearings that may require a change in the bearing housing.
The pocket is cut in the upper half of the bearing with one end of the step located in the second quadrant for counter-clockwise shaft rotation.
The pocket has side lands to hold the pressure and flow. A circumferential relief groove or track is sometimes grooved in the bottom half of the bearing.
Both of these dam and relief track effects combine to increase the operating eccentricity of the bearing compared to an axial groove cylindrical bearing.
At high speeds and light loads, the step creates a loading that maintains a minimum operating eccentricity.

FIG 6

Elliptical Bearing
The elliptical or lemon bore bearing is a two-pad-fixed geometry bearing which is preloaded in the vertical direction, as shown in Figure 7 . This bearing can be manufactured by inserting a shim in the
split line before boring. When the shim is removed, the vertical clearance will be less than the horizontal clearance. The centers of curvature of the top and bottom halves are not coincident with the true bearing center.

FIG 7

Offset Half Bearing
This is a unique type of plain journal bearing which is a very simple and yet very effective bearing generally used in some applications. It is a 2-pad multi-lobe bearing with 100% offset. The pads are generally always preloaded. It is easy to make this bearing by boring a hole in plain bushing. It
should be noted that this type of bearing will not tolerate reverse rotation. While more stable than the axial groove journal bearing, there is still a tendency for instability. An offset half bearing is shown in Figure 8

FIG 8

Taper-land and Multi-lobe bearings
The taper land bearing is shown in Figure 9  and multi-lobe bearing is shown in Figure 10 . These are very popular sleeve bearings that are successful in increasing the instability threshold speed when compared to cylindrical sleeve bearings. The taper land bearing has side lands similar to a pressure dam bearing. Care must be taken when using bearings with this design because they are not suited for heavy load applications. Taper land bearings are very frequently utilized in small, light rotors operating at high speeds such as small turbo-expanders and turbochargers.

FIG 9

FIG 10

Multilobe bearings are similar in concept to lemon bore bearings. (i.e., there is a separate pad machined bore and bearing set bore). The centers of the pad arc for each of the lobes form a circle, referred to as a preload circle. Therefore, a wide range of preloads is possible by changing the pad

and set bores. Under certain conditions, the cross-coupling can cause the bearing to be unstable and an oil whirl will result.
Tilt Pad Bearing
In the tilt pad design, the cylindrical bearing element is divided into a number of pad arc segments, depending on the shaft load. Each pad segment is supported and held in circumferential position by an outer housing.
This bearing includes pad load orientation, pivot offset, pad preload and pad axial length. The load between pivot configurations is shown in Figure 10. The load between pads provides more symmetric stiffness and damping coefficients.

FIG 11

The offset pivots are very popular with thrust bearings as offsetting the pivot increases the operating film thickness thereby decreasing the operating temperature. Further, offset pivots increase bearing stiffness, especially when compared to centrally pivoted pads.
For this tilt pad bearing it can be seen that as the preload decreases, the bearing damping increases while bearing stiffness remains constant. Both of these trends help in increasing the bearing effective damping. Normally, as bearing damping increases, bearing stiffness also increases. However, there is
a disadvantage to low preload pads. Light preload is the loss of damping



2- Thrust bearings or axial bearings

The thrust load is transferred through lubricant film between thrust collar on rotor and thrust collar on
housing.

FIG 12

3- Radial thrust bearings

Radial thrust bearings are subjected to combined radial and thrust loads. These bearings carry both radial and thrust loads.

FIG 13

Bearing classification based on type of lubrication

There are three regimes of lubrication for sliding bearings (figure 14):
1. boundary lubrication
2. mixed film lubrication
3. full film lubrication

FIG 14

1- Boundary lubrication
Here the surface contact is continuous and extensive. The lubricant is continuously smeared over the
surfaces and provides a continuously renewed adsorbed surface film which reduces the friction and wear. The typical coefficient of friction is 0.05 to 0.20.
2- Thin film lubrication (Mixed lubrication)
Here even though the surfaces are separated by thin film of lubricant, at some high spots Metal-to-metal contact does exist. Because of this intermittent contacts, it also known as mixed film lubrication. Surface wear is mild. The coefficient of friction commonly ranges from 0.004 to 0.10.
3- Thick film lubrication (Hydrodynamic lubrication)
The surfaces are separated by thick film of lubricant and there will not be any metal-to-metal contact. The film thickness is anywhere from 8 to 20 μm. Typical values of coefficient of friction are 0.002 to 0.010. Hydrodynamic lubrication is coming under this category. Wear is the minimum in this case.

FIG 15

Bearing classification based on lubrication mechanism

1- Hydrodynamic lubricated bearings

In these bearings the load-carrying surfaces are separated by a stable thick film of lubricant that prevents the metal-to-metal contact. The film pressure generated by the moving surfaces that force the lubricant through a wedge shaped zone. At sufficiently high speed the pressure developed around the journal sustains the load.

FIG 16

2- Hydrostatic lubricated bearings

In these bearings, externally pressurized lubricant is fed into the bearings to separate the surfaces with thick film of lubricant. These types of bearings do not require the motion of the surfaces to generate the lubricant film. Hence they can operate from very low speed to high speed.

FIG 17

3- Elasto hydrodynamic lubricated bearings

Rolling contact bearings come under this category. The oil film thickness is very small. The contact

pressures are going to be very high. Hence to prevent the metal-to-metal contact, surface finishes are to be of high quality. Such a type of lubrication can be seen in gears, rolling contact bearings, cams etc.

4- Boundary lubricated bearings

When the speed of the bearing is inadequate, less quantity of lubricant is delivered to the bearing, an

increase in the bearing load, or an increase in the lubricant temperature resulting in drop in viscosity  any one of these may prevent the formation of thick film lubrication and establish continuous metal-to-metal contact extensively. Often bearings operating in such situations are called boundary lubricated bearings.

FIG 18

5- Solid film lubricated bearings

For extreme temperature operations ordinary mineral oils are not satisfactory. Solid film lubricants such as graphite, molybdenum disulfide or their combinations which withstand high operating temperature are used. These types of bearings are common in furnace applications, or trunnion bearings of liquid metal handling systems, hot drawing mills etc

SLIDING CONTACT BEARINGS - ADVANTAGES AND DISADVANTAGES 

These bearings have certain advantages over the rolling contact bearings. They are:

1.The design of the bearing and housing is simple.

2.They occupy less radial space and are more compact.

3.They cost less.

4.The design of shaft is simple.

5.They operate more silently.

6.They have good shock load capacity.

7.They are ideally suited for medium and high speed operation. 

The disadvantages are:

1.The frictional power loss is more.

2.They required good attention to lubrication.

3.They are normally designed to carry radial load or axial load only

Decide lubricant for the bearing
The performance of a sliding bearing differs markedly depending on which type of lubricationis physically occurring.This is illustrated in Figure 19, which shows the variation of the coefficient
of friction with a group of variables called the ‘bearing parameter’ which is defined by:

K =𝛍 ×N/P

where
𝛍 viscosity of lubricant (Pa s)
N speed (for this definition normally in rpm)
P, load capacity (N/m²) given by:

P = W/(LD)

where
W, applied load (N);
L, bearing length (m,mm );
D, journal diameter (m, mm).

FIG 19

The bearing parameter, 𝛍N/P, groups several of the bearing design variables into one number
Normally, of course, a low coefficient of friction is desirable. In general, boundary lubrication is
used for slow speed applications where the surface speed is less than approximately 1.5 m/s. Mixed
film lubrication is rarely used because it is difficult to quantify the actual value of the coefficient of
friction (note the steep gradient in Figure .19 for this zone).

Lubricants

As can be seen from Figure 19 bearing performance is dependent on the type of lubrication occurring
and the viscosity of the lubricant.The viscosity is a measure of a fluid’s resistance to shear.
Lubricants can be solid, liquid or gaseous, although the most commonly known are oils and greases.
The principal classes of liquid lubricants are mineral oils and synthetic oils.Their viscosity is highly
dependent on temperature as illustrated in Figure 20.They are typically solid at 35°C, thin as paraffin at 100°C and burn above 240°C. Many additives are used to affect their performance. For

example, EP (extreme pressure) additives add fatty acids and other compounds to the oil, which attack the metal surfaces to form ‘contaminant’ layers, which protect the surfaces and reduce friction even when the oil film is squeezed out by high contact loads. Greases are oils mixed with soaps to form a thicker lubricant that can be retained on surfaces. The viscosity variation with temperature of oils has been standardized and oils are available with a class number, for example SAE 10, SAE 20, SAE 30, SAE 40, SAE 5W, SAE 10W, etc.The origin of this identification system developed by the Society of Automotive Engineers was to class oils for general purpose use and winter use, the ‘W’ signifying the latter.The lower the numerical value, the thinner or less viscous the oil. Multigrade oil, e.g. SAE 10W/40, is formulated to meet the viscosity requirements of two oils giving some of the benefits of the constituent parts.An equivalent identification system is also available from the International Organisation for Standardisation (ISO 3448).

FIG 20

Journal Bearing Design

The simplest form of journal bearing is a plain circular bushing with an internal diameter that is slightly larger than the shaft outer diameter. The following parameters are considered for the design of a journal bearing.

Length to diameter ratio

The length to diameter ratio is one of the first things a bearing designer considers. This factor is tuned to give a sufficiently steady and alternating load capacity.An L/D ratio less than 0.3 have

poor damping and an L/D ratio greater than 0.75 usually shows a little gain in the effective damping. Heavily loaded bearings such as gears often exceed an L/D ratio of 1.0 and may need special care in aligning the bores to the shafting.

FIG 21

Bearing Pressure

The pressure over the projected area in the bearing is assumed uniform for the purpose of expressing the load capacity. 

FIG 22

In steadily loaded bearings, the load is mainly due to the weight of the rotating element, in addition to the dynamic loading imparted by the unbalance in the rotating components. The bearing pressure is given by the Equation :

Where, 

P – Bearing pressure in N/mm²

L- Length of the bearing in mm

D- Diameter of the bearing in mm

W = Bearing Radial Load (N)

Bearing clearance

Bearing clearance is one of the most important parameters in the operation of a bearing. Therefore, it is important to determine the installed clearance along with the bore contour and concentricity to the outside fit diameter.

FIG 23

The final bearing clearance is influenced by the contour of the housing into which it is installed and also by the amount of interference or crush between the bearing shell and the support housing. Proper crush is crucial in the operation of high speed and critical machinery. Improper crush can lead to either a hot bearing or a loose bearing fit. A loose bearing may contribute to a synchronous or a sub synchronous type of vibration. It is critical that a metal to metal fit to 0.002 inches of crush on diameter be maintained for a proper bearing installation. This criterion applies to most thick shell bearing and linings. For thin shell bearings, and in situations where the rigidity of the housing is much greater than that of the bearing liner, the values for the crush should be arrived at by referring with the bearing and equipment manufacturer. Excessive crush or a contact stress could cause local yielding of the material. This might lead to the collapse of the liner and the loss of bearing clearance.

Sommerfield number

Dynamic stiffness and damping coefficients are generally presented as a function of dimensionless parameter known as the Sommerfield number.

The more common way of expressing is by the Equation

Where, 

μ is the viscosity in (pa.s)

N is the running speed in revolutions/second

L is the pad length in (inches,mm)

R is the journal radius in( inches ,mm)

W is the bearing load in pounds

C is the pad bore radial clearance in (inches.mm)

Higher loads and lower speeds will result in a lower Sommerfield number for a given bearing configuration. On the other hand, lighter loads and higher speeds result in a higher Sommerfield number.

The table below provides some typical diametrical clearances for journal bearings under steady loads and for hydrodynamic lubrication

Materials Used for sliding contact bearings

1. Metallic bearings

             . Metallic bearings

             . Babbitt metal

             . Bronze – Alloy of copper, tin and zinc

             . Cast iron

             . Silver

2. Non metallic bearings