NBSIR 73-243

On the Compression of a Cylinder in

Contact with a Plane Surface

B. Nelson Norden

Institute for Basic Standards

National Bureau of Standards

Washington, D. C. 20234

July 19, 1973

Final Report

U.S. DEPARTMENT OF COMMERCE

NATIONAL BUREAU OF STANDARDS

NBSIR 73-243

ON THE COMPRESSION OF A CYLINDER IN CONTACT WITH A PLANE SURFACE

B. Nelson Norden

Institute for Basic Standards

National Bureau of Standards

Washington. D. C. 20234

July 19, 1973

Final Report

U. S. DEPARTMENT OF COMMERCE, Frederick B. Dent, Secretary

NATIONAL BUfCAU OF STANDARDS, Richard W. Roberts. Director

CONTENTS

Page

List of Illustrations

Nomenclature j^-j^-j^

Introduction -j^

General Description of Contact Problem Between Two Elastic Bodies ... 4

Special Case of Line Contact

Experimental Verification 3y

Conclusion

,

/^Q

Summary of Equations

References 44

Illustrations 48
LIST OF ILLUSTRATIONS

Figure 1. Geometry of the contact between elliptic paraboloids

Figure 2. Cross section of two contacting surfaces

Figure 3. Geometiry of deformed bodies

Figure A. Measurement of the diameter of a cylinder

Figure 5. Contact geometry of two parallel cylinders

Figure 6. Relationship for yield stress as function of surface finish

Figure 7. Compression of 0.05-inch cylinder between 1/A-inch anvils

Figure 8. Total deformation of 0.001-inch steel cylinder between

3/8-inch anvils

Figure 9. Total deformation of 0.01-inch steel cylinder between

3/8-inch anvils

Figure 10. Total deformation of 1.00- inch steel cylinder between

3/8-inch anvils

Figure 11. Analysis of total system deformation using equations (55),

(62), (73), (76), and (79)

Figure 12. Analysis of total system deformation using equations (82),

(83), and (84)

Figure 13. Nomograph for computation of maximum stress in cylinderplane contact

Figure 14. Equation for calculation of compression in cylinder-plane

contact
NOMENCLATURE

P maximutn pressure at center of contact zone

o

a major axis of ellipse of contact

b minor axis of ellipse of contact (also half-width of contact in the

cylinder-plane case)

^1' ^l'

PJ^incipal radii of curvature of body 1

^2' ^2* PJ^iiicipal radii of curvature of body 2

distance from a point on body 1 to the undeformed condition

point of body 2

same terms as above for body 2 to body 1

deformation of a point on body 1

deformation of a point on body 2

6 total deformation of bodies 1 and 2

U) angle between planes x^ z and x^z

2 \, parameter equal to 1 - ji where ji is Poisson's ratio for

body i and Ei is Young's modulus for body i

$ potential function at any point on the surface

2 b^

e eccentricity of the ellipse of contact e = 1 - —

j

a

P total load applied to produce deformation

K complete elliptic integral of the first kind

E complete elliptic integral of the second kind

c stress at some point on surface of body

2 h square of the complementary modulus or (1 - e ) V potential function used by Lundberg

M mutual approach of remote points in two plates with cylinder

between the two plates
ABSTRACT

The measurement of a diameter of a cylinder has widespread application

in the metrology field and industrial sector. Since the cylinder is usually

placed between two flat parallel anvils, one needs to be able to apply corrections,

to account for the finite measuring force used, for the most accurate determination of a diameter of the cylinder.

An extensive literature search was conducted to assemble the equations

which have been developed for deformation of a cylinder to plane contact case.

There are a number of formulae depending upon the assumptions made in the development. It was immediately evident that this subject has been unexplored in depth

by the metrology community, and thus no coherent treatise for practical usage

has been developed.

This report is an attempt to analyze the majority of these equations and

to compare their results within the force range normally encountered in the

metrology field. Graphs have been developed to facilitate easy computation of

the maximum compressive stress encountered in the steel cylinder-steel plane

contact case and the actual deformation involved.

Since the ultimate usefulness of any formula depends upon experimental

verification, we have compiled results of pertinent experiments and various

empirical formulae. A complete bibliography has been included for the cylinderplane contact case for the interested reader.
INTRODUCTION

The problem of contact between elastic bodies (male, female or neuter

gender) has long been of considerable interest. Assume that two elastic

solids are brought into contact at a point 0 as in Figure 1. If collinear

forces are applied so as to press the two solids together, deformation

occurs, and we expect a small contact area to replace the point of the

unloaded state. If we determine the size and shape of this contact area

and the distribution of normal pressure, then the interval stresses and

deformation can be calculated.

The mathematical theory for the general three-dimensional contact

problem was first developed by Hertz in 1881. The assumptions made are:

1) the contacting surfaces are perfectly smooth so that the actual

shape can be described by a second degree equation of the form

2 2

z = Dx + Ey +- Fxy.

2) The elastic limits of the materials are not exceeded during

contact. If this occurs, then permanent deformation to the

materials occurs.

3) The two bodies under examination must be isotopic.

4) Only forces which act normal to the contacting surfaces are

considered. This means that there is assumed to be no frictional forces at work within the contact area.

5) The other assumption is that the contacting surfaces must be

small in comparison to the entire surfaces.

Based on the above assumptions and by applying potential theory,

Hertz showed that:

1) the contact area is bounded by an ellipse whose semiaxes can

be calculated from the geometric parameters of the contacting bodies.
-2-

2) The normal pressure distribution over this area is:

[1 - (x/a)2 - (y/b)^]^/^

where = maximum pressure at center

a = major axis of ellipse of contact

b = minor axis of ellipse.

The above assumptions are valid in the field of dimensional metrology

because the materials (usually possessing finely lapped surfaces), and the

measuring forces normally used are sufficient for the Hertzian equations

to be accurate. In the case of surfaces that are not finely lapped, the

actual deformation may differ by more than 20% from those calculated from

equations.

Since the subject of deformation has such widespread impact on the

field of precision metrology, we have decided to publish separate reports -

(1) dealing with line contact and particularly the contact of a cylinder

to a plane, and; (2) which treats the general subject of contacting bodies

and derives formulae for all other major cases which should be encountered

in the metrology laboratory.

An exhaustive literature search was conducted to determine equations

currently in use for deformation of a cylinder to flat surface. The

ultimate usefulness of deformation formulae depends on their experimental

verification and, while there is an enormous amount of information available for large forces, it was found that the data is scarce for forces

in the range used in measurement science. One reason for this scarcity

is the degree of geometric perfection required in the test apparatus and

the difficulty of measuring the small deformations reliably.
-3-

Depending upon the assumptions raade , there are a number of formulae

in use. Various equations will be analyzed along with the assumptions

inherent in their derivations. There are basically three approaches to

the problem for the deformation of a cylinder with diameter D in contact

with a plane over a length L and under the action of force P:

1) the approach where a solution is generated from the general

three-dimensional case of curved bodies by assigning the plane

surface a radius of curvature. This is the same as replacing

the plane surface with a cylindrical surface with a very large

radius of curvature. The area of contact is then a^ elongated

e llipse.

2) The approach where the contact area between a cylinder and plane

is assumed to be a finite rectangle of width 2b and length L

where L »b.

3) The determination of compression formula by empirical means.
GENERAL DESCRIPTION OF CONTACT PROBLEM

When two homogeneous, elastic bodies are pressed together, a certain

amount of deformation will occur in each body, bounded by a curve called

the curve of compression. The theory was first developed by H. Hertz [1].

Figure 1 shows two general bodies in the unstressed and undeformed

state with a point of contact at 0. The two surfaces have a common tangent

at point 0. The principal radii of curvature of the surface at the point

of contact is R^^ for body 1, and R^ for body 2. R^' and R^ ' represent

the other radii of curvature of each body. The radii of curvature are

measured in two planes at right angles to one another. The principal

radii of curvature may be positive if the center of curvature lies within

the body, and negative if the center of curvature lies outside the body.

Also planes x, z and x_z should be chosen such that

The angle w is the angle between the normal sections of the two bodies

which contain the principal radii of curvature R^^ and R^.

Figure 2 shows a cross-section of the two surfaces near the point of

contact 0. We must limit our analysis to the case where the dimensions

of the compressed area after the bodies have been pressed together are

small in comparison with the radii of curvature of bodies 1 and 2, We

also assume that the surface of each body near the point of contact can

be approximated by a second degree equation of the form

Z = Dx + Ey +2 Fxy

where D, E and F are arbitrary constants.
-5-

If the two bodies are pressed together by applied normal forces (Figure

3), then a deformation occurs near the original point of contact along the

Z-axis. Here again, we consider only forces acting parallel to the z-axis

where the distance from the z-axis is small.

The displacements at a point are w^^ and W2 where w^^ is the deformation

of point ?^ of body 1 and w^ is the deformation of point P2 for body 2,

plane C is the original plane of tangency; is the distance from P^^ to

the undeformed state, and z^ is the distance from to the undeformed

state. For points inside the contact area, we have

(z^ + w^ + (z^ + w^) =6 (1)

where 6 is the total deformation which we are so diligently seeking.

The equation for surface 1 may be written as:

2 2 = X + y + 2F^ xy

and for surface 2,

2 2 z^ = X + E^ y + 2F2 xy.

Since the sum of z^ and z^ enter into the equation we obtain

+ ^2 " (D^ + 02^''^ ^^1 ^2^^^ ^ ^^^1 ^l^""^-

Now Hertz showed that the axis can be transformed so that F^^ = "F2> ^"^"^

hence, the xy terra vanishes. To simplify the above equation further we

replace the constants (D^ + D^) with A and (E^ + E^) with B thus giving,

z^ + z^ = Ax + By

From equation (1) we obtain:

2 2 Ax + By + Wj^ + w^ = 6 (3)
-6-

The constants A and B are expressible in terms of combinations of

the principal curvatures of the surfaces and the angle between the planes

of curvature. These combinations have been derived by Hertz and are as

follows:

B- A = I

2 / , , X 2 + (^i"r;) "(^2"r')

(^1 ' R')(^2 " R')

1

^ ^ "2

ll/2

cos 2w

1 ' ' ' "2

(5)

Since the points within the compressed area are in contact after the

compression we have:

2 2 w^ + w^ = 6 - Ax - By

and since 5 is the value of w^ and at the origin (Figure 3, x = y =0),

we must evaluate w^ and w^.

The pressure P between the bodies is the resultant of a distributed

pressure (P ' per unit of area), over the compressed area. From Prescott [2]

the values of the deformations w^ and w^ under the action of normal forces

are :

w^ = \^ $ (x, y) (6)

and,

.2

where =

i \ ^Ei

J
-7-

= Poisson's ratio for the 1th body

= Modulus of elasticity for the ith body,

and 5(x, y) ~ jj ~ dx'dy' which represents the

A

potential at a point on the surface. Here r is the distance from some

point (x, y) to another point (x', y') and P' is the surface density.

By substitution in Equation 3 we obtain,

A

where the subscripts 1 and 2 represent the elastic constants for bodies

1 and 2.

One important fact should be observed from Equations 6 and 7 and this

is : ^

2 2., 2 (^)2

which means if the two bodies are made of the same material w^^ = w^.

The integral equation 8 allows one to compute the contact area, the

pressure distribution, and the deformation of the bodies.

The problem is new to find a distribution of pressures to satisfy

equation 8. Since the formula for $^ is a potential function due to matter

distributed over the compressed area with surface density P' we see the

analogy between this problem and potential theory. Hertz saw the analogy

since the integral on the left side of equation 8 is of a type commonly

found in potential theory, where such integrals give the potential of a

distribution of charge and the potential at a point in the interior of a

uniformly charged ellipsoid is a quadratic function of the coordinates.
2 2 2 X V z If an ellipsoid — +2 ^ ~2 ^ ^ ^ uniform charge density P with

a b c mass npabc, then the potential within the ellipse is given by Kellog , 3] as

y, z) =

npabc H1 - -f-

^ 5^ ]

^

(10)

V

a .V b c + , / (^^2 ^,^^^^2 ^^^(^2 ^^^^1/2

If we consider the case where the ellipsoid is very much flattened (c -« 0)

then we have

4(x, y) = npabc 1 - ^- ^

(11)

The potential due a mass density is

a b

distributed over the ellipse + 2~ = 1 in the plane Z = 0, where the

a b

total load P is given by,

P = 4/3 npc ab. (13)

By substituting into Equation 11 we obtain,

2

rt('-^ A) ^

V a b^ . v/ ,.32 2 ^

4(x, y)=^/ 5^1 (14)

,,2 , ,^2 ,

. . ,1/2

((a +f)(b +f)(Y))

From Equations 8, II and 14 we obtain,

^ » "*/((a2.-,)(b2+Y)(.))l/2

Thus we have

,

/• /?_/> 9 \

dY

((a2 +^^)(b2 +Y)(Y){^^ ^^^^
-9-

We now substitute into Equation 8 to obtain,

^

^ ^ \ a+Y b+Y/ 2 2 1/2

((a^^ + Y)(b^ + Y) (Y))

6 - Ax^ - By^ (17)

2 2 Since the coefficients of 1 , x , and y must be equal in Equation 17,

we have

,

= ^ P a + ^

(18)

J 2 2 1/2

•'o ((a^ + Y)(b^ + Y)(Y))

3 /So

A = r P (>-, + >-,) /

^^"^ (19)

2 1/2 2 1/2

'o (a^ + Y) ((b'' + Y)(Y))

3 dY

B = f

P (\ + \ ) / (20)

2 3/22 1/2

o (b^ + Y) ((a + Y)(Y))

Equations 19 and 20 determine a and b (major and minor axis of the

ellipse of contact) and equation 18 determines the total deformation 6,

(or the normal approach) when a and b are known.

Since the integrals in Equations 18, 19 and 20 are somewhat cumbersome,

they may be expressed in terms of complete elliptic integrals where tables

are readily available. Since the eccentricity (e) of any ellipse may be

expressed in terms of the major and minor axis as,

2 2

2 1 b^,

e e = (l - -2") a a

we may express Equations 18, 19 and 20 in terms of the eccentricity of

the contact ellipse, • .

From Equation 19 we obtain,
10-

A 4 f + s) t

1 5/2

By multiplying the numerator and denominator by (—2) and making the

2 ^ change of variable a ^ = Y we obtain,

•1 ^ v/ 3 d£

A = f

p a — '

or

Aa^ = ^ P {K. + \.) r

^

(21)

3/2 2 1/2 1/2

(1 + C) (1 - e^ + ^)^^^ C

From the same analysis we obtain for Equations 18 and 20,

Ba^ = I

P (X^ + (22)

(1 - e^ +C)^^^[4 (1 +C)]^^^

and

3

/oo d C 6a = - P (.^ . / — <23)

[C (1 - +e )]

2 2 By making the substitution C= cot e [3;, and d^ =-2 cot e esc 6 d 9

where q: j to 0 and C :0 to qd we obtain.

Aa^ = I

P a, + X ) r -2 cot e csc^ ^ d ^

' / 2., 3/2 , .2 2 . 2 .1/2.1/2

(1 + cot e) (cot 9(1 - e + cot 9) )

2 2 and since (1 + cot 9) = esc

J i 2 ,3/2, 2^ 2.1/2

(esc 9) (esc 9 - e )
-11-

3 /° -2

7" p (\. + \.) / de

/o . 2 2,1/2 n/2 (cscQ) (esc 9 - e )

4

' ^"1•1 -" -2> '^2' /J 2

rr/2 (csce)i csc 9(1

J"^ ^ CSC 9

= f

P (Xj + / — 172

and since sinp

CSC

1 CSC9 '

2. Aa^ = I

P (.^ . / 1/ 2

By the same analysis we find,

B

3 3 _ ^ ^ . f° -2 sin^e

and.

•1 ^ s> /

6a =

I" P a, + \J /

^-^ :r—77o de

By rearranging the above equations we obtain,

rr Aa^ = I

P (K^ + / r" ^ Wo de

1

•1 * S> /

/2(1

2 , 2 .3/2

- e sin e)

2 /2(1

2 ,2,1/2 - e sin 9)

obtain,

/2

sm 9

(1 - e^ sin e) ^ ^

n/2

I* , 2

sin 9

2 . 2 .3/2

- e sm e)

/2

1

Ba^ = IP (\, + \J / Wo de

3 /* 1
-12-

Now the Legendre forms of elliptic integrals of the first and second

kinds are from Boas 4],

d4) ^ f\ 1 . 2 ,1/ 2 [0 ^ - -

1 o (1 - e sin (}))

e = sinR

E (e, 4)) =y^(l - e^sin^(})) dcj) | 0 < - s n/2

where e is the modulus and <^ the amplitude of the elliptic integral. e' =

2 1/2

(1 - e ) and is called the complementary modulus.

The complete elliptic integrals of the first and second kinds are the

values of F and E for <J) = tt/2 so that,

K = K(e) - F(e, a/2) -

E = E(e) = E(e, t/2)

:/2

I-

•:/2

2 2 (1 - e sin 9)

2 2 1/2

(1 - e sin e) d

There are numerous ways to evaluate the above integrals. Hastings [5],

g has polynomial approximations accurate to 2 parts in 10 which are of

the form,

K(e) = ^a + a.m. + ... + a.m. ^ +

o 11 4 1

4.

) oil + b.m. + ... + b.m, 1 ln(l/m,) 41 1

a =

a, =

1.38629 436112

.09666 344259

.03 590 092383

.03742 563713

.01451 196212

b = o

^2

=

"4

'

. 5

. 12498 593597

.06880 248576

.03328 355346

.00441 787012
-13-

2 2 where e = sin a and = cos a

and the approximation for the elliptic integral of the second kind is

given by,

E(e) = n + am, + . . . a,m,^" + ^b,ra, + . . . b,m,^nn(—

)

11 41' 11 41^ mj^

a^ = .44325 141463

a^ = .06260 601220

a^ = .04757 383546

a, = .01736 506451

4

b^ = .24998 368310

b^ = .09200 180037

b^ = .04069 697 526

b, = .00526 449639

4

New since the complete elliptic integral of the first kind is,

n/2

K = / (1 - e^sln^e)'^^^ de -I

we can obtain.

n/2

dK f

. 2 sin 9

, 2

^ 2^,3/2

% (1 - e sin e)

de (30)

and in a similar manner for the elliptic integral of the second kind

we obtain,

tt/2

dE r

. 2 sin 9

2 . 2 .1/2 ^9

(1 - e sin e)

(31)

and we obtain from Equations 27, 28 and 29,

Aa^=|p(.^..p(f)(-I) (32)

Ba^ = |P (.^..p (f)(1) (33)

5a = I

P (\^ + K^) (K) (34)
-14-

We may rewrite Equation 32 as,

n/2

3 3 Aa^ = - P ~ 2 , 2.1/2

(1 - e sin 9)

2.2,1/2

2 , 2 e sin 9

(1 - e sin e)

r/2

(1 - e sin e)

do

=

2 P (^-1 + >^2^ (

= - P + k^) ( Hi

-(1 - e^sin^^) 1 7! 2 , 2-1/2 2 , 2.1/2

(1 - e sin 9) (1 - e sin e)

1

de

2 . 2^.1/2

(1 - e sin 9)

d9

Aa^ = I

P + A-^X^) Lk - E]

e

(3 5)

Equating equations 32 and 35 we obtain.

f 4 <E - K)

de e (36)

From equation 30

n/2

de 7

e sin

2.2 .3/2

o (1 - e sin 9)

d9

rT/2

/- (1 - e 2^2 sin ,1/2, 22, 9) (e sin p)

, 2,2 ,1/2,, 2 , 2 ,3/2 e(l - e sin e) (1 - e sin 9)

d9

n/2

/- 2 2 .1/2,, 2.2,- (1 - e sin 9) [1 - (1 - e sm 9) . 7, 2 y 2 .1/2, ' 2 , 2.3/2 e(l - e sin 9) (1 - e sin 9)

d9

tt/2

/[;

(^ 2 , 2 ,1/2 2 . 2-1/2,, 2.2, (1 - e sin b) (1 - e sm 9)~ (1 - e sm 9) ~

2 , 2^,1/2,, 2 , 2 ,3/2 " ~~ 2 . 2 ,1/2," 2 ^ 2,3/2 e(l - e sin 9) (1 - e sin 9) e(l - e sm 9) (1 - e sin )

d9
-15-

/ ~ 2 ^ 2 1/2 ^ ~ 2 ^ 2 ,3/2

y L e(l - e sin 9) e(l - e sin e) ^ o

de e e / 2 2 ,3/2

(1 - e sin 9)

The integral above may be reduced to the form

tt/2 J/

thus

,

2 2 3/2 2 (1 - e sin e) e(l - e ) o

e(l-e) e(l-e)^

By dividing Equation 32 by 33 we obtain,

(38)

-dE

A ^ de

B dK

de

and substituting the values in Equations 36 and 37 into Equation 38, we

obtain,

° E - (1 - e^)K

We now see that for any value of the eccentricity of the contact ellipse

we can obtain values for ^, K, and

^

which will allow us to compute the

B e de

normal approach 5 (deformation).
-16-

SPECIAL CASE OF LINE CONTACT

In the metrology field one fundamental measurement occurs frequently,

i.e., the measurement of the diameter of a cylinder. Figure 4a shows a

cross-section of the typical measurement between plane parallel anvils

of a measuring machine. Since it is infinitely difficult to measure the

object with zero force applied, Figure 4b shews the exaggerated resultant

shape of the anvils and cylinder after a measuring force is applied. Since

the customer usually desires to know the "unsquashed" diameter, certain

corrections must be applied to account for the measurement process.

To solve this problem we shall make use of the expressions already

developed to solve for the "pressure distribution" and size of the area

of contact by allowing one axis of the ellipse of contact to become infinite.

To determine the deformation, the contact area will be taken as being a

finite rectangle with one side very much greater than the other.

The derivation will be for the case of a pair of cylinders with their

axes parallel and is based on the works of Thomas and Hoersch [7], and

Love [8]. The solution for a cylinder to plane contact can easily be

obtained by allcwing the radius of one of the cylinders to become infinite.

Line contact occurs when two cylinders rest on each other with their

axes parallel (figure 5a), and when a cylinder rests on a plane. As the

two cylinders are pressed together along their axes, the resulting pressure

area is a narrow "rectangle" of width 2b and length L (assuming no taper

in the cylinders). In other words, the area of contact is an elongated

ellipse with the major axis of the ellipse equal to L and the eccentricity

approaching unity.

The distribution of compressive stress along the width 2b of the surface

of contact is represented by a semi -ellipse (figure 5c). The stress
-17-

distribution over the ellipse of contact for the three-dimensional case

will be remembered from Equation 12 as,

(x', y') = 2nab yj ^ 2 ,2

3P ./ ...2

where ct(x', y') represents the stress acting at any point (x', y').

Now the integrated pressure over the surface of a finite rectangle

across the minor axis of the ellipse in the plane x = 0 (figure 5c), is

P = / %(0, y') dy'

•b ^

New as a D, let P in a manner so that P/a remains equal to a finite

constant. Then the value of P is the force per unit length of the contact

area. For a = , the compressive stress at any point y is given by,

2 1/2

-(y) =2^ <1

-f2>

<^1>

b

and by substitution of Equation 40 into Equation 41 we obtain, since

P = P/L,

op 2 1/2

b

We can new see that the maximum value for the stress within the area

of contact will be at the center where y = 0 and is.

= 2P max nLb (43)
-18-

The maximum compressive stress is important in any contact problem

because the surface area of contact is small and very high stresses may

easily be obtained with relatively light loads. If a exceeds the max microplastic yield point of the material, then permanent deformation

will occur. A relationship for hardened steel relating yield stress to

surface finish is shewn in Figure 6.

We new need to develop an expression for the width of contact b, and

by knowing b, the stress at any point can be computed. The surface of

each cylinder may be represented by the equations,

\ - \

From Equation 20 and using the expression developed in Equation 40

we obtain,

and since the one axis of the ellipse, a - t=, the expression becomes,

By using the expression in Equation 42, Thomas and Hoersch evaluated the

above equation to give,

B = 2(\^ + \^) P/Lb^ or

2(K + K )?

^ - 13

where b is the ha If -width of contact,

P is the total force,

L is the contact length.
-19-

For a pair of cylinders with their axes parallel and radii and r^,

respectively, we have

„ 2 1 2

o

2 1 2 ' ^2 = '2>' = 17^ y

Thus, B = B + B =

i z zr^^r^

^° = 4(\^ + xpPCr^r^) (47)

For a cylinder on a plane surface,

1 2 ^1 = 27 y

Z- = - = 0

1 thus , 2r

° ^-^^H " S^** (48)

L

2 Now since our cylindrical surfaces are described by By forms, we obtain

from Equation 8,

(K + ^o) / / dx'dy' = 6 - By^ (49)

2 2 1/2 where r = ((y - y') + x' ) and the integral applies to the finite

rectangle of contact with one side (L) very much longer than the other (2b)

We now have

,

$(0, y) = f^f^^2 ^^y') dx'dy'

-b •'-L/

'2
-20-

r+h rL/2

-b

((y - y') + x'^)^^^

'+b2p(y') In

L , // 1x2 _^ ,L. 2- 1/2

2 + ((y - y ) + (j) )

|y -y'l

dyIf we assume (j) is very large in relation to (y - y') then.

and

P

-b

HO, y) = / 2p(y') In

y - y

dy'

$(0, 0) = [ 2p(y') In f^) dy

Let us pause a moment to recap what we are doing. Equation 49 gives

us the relation between the total deformation <5 to the potential for

points within the contact area. Since we are interested in the maximum

deformation which occurs, we need to evaluate the integral in Equation 49

at the center line or where y = 0. Thus, Equation 51 represents that

maximum potential.

Continuing, we may rewrite Equation 51 as,

$(0, 0) = 2 In L / p(y')dy' - / p(y') InCy'^) dy

Since the force per unit length — =J

p(y')dy' and from Equation 41 we

may substitute into Equation 52 to obtain,

*(0, 0) = 2 2 In L -

To evaluate the integral in the above relationship, we make the substitution ^

= sin 9 and then dy = b cos 5 d to obtain.

(50)

(51)

(52)

(53)
-21-

''^ 2 , 1/2 _ „ 2 . 2 (1 - sin e) sin e)](bcosfl) de

r/2

,-tt/- W2

) / cos e de + 2b /

/2

2 2 2 cos 9 [ln(b sin 9)] d9

= b(21nb) / cos e de + 2b / cos 0 ln|sine|de

-n/2 .•;t/2

we know that,

and the integral

111

cos e de =

2

1=1 cos^A Inlsinelde

-n/2

has been evaluated by Birens de Haan to give

I = (1 + InA)

So we obtain the value for the total integral as.

= b(21nb) ^

+ 2b (- J

- J

InA),

= nb (Inb - J

+ ~)

Thus by substituting into Equation 53 we obtain

(0, 0) = 2 I

In

$(o, o) = 2 ^

^InL - Inb + 1.193 14

sj

Since

6 = + K)n dx'dy' + By2
-22-

and Ho, o) gives us the potential at the center of pressure zone we have

6 = (A.^ + k^) 2 7- |lnL - Inb -f 1.193145

as the total deformation of a pair of cylinders with their axes parallel

or a cylinder on a flat surface. The value of b (half -width of contact) is

substituted into equation 54 to obtain the appropriate answer for any

particular case.

Since equation 48 gives us the value of b as,

1/2

we have

r4r(\ —+ \ )Pl

I—

6 = (\^ + \p 2 ^ InL - j

In I ^

-

/ ^^^^^^

r /4r(>. +A.)P\

] (X.^ + 21nL - ln( ^

—-— 1 + 2.38629

+ K ) 12 InL + In / L \ + 2.38629]

6 = f

(X^ H- K^)

where P = total measuring force

L = length of contact between plane and cylid

r = radius of cylinder

6 = total deformation of cylinder and plane.

Another form of Equation 54 can be obtained by not evaluating the

/ l + . term , u .

I I which gives.
-23-

l3 1.00 + In (55)

Remember this formula is for the case of a cylinder to plane on only one

side. If both sides are desired, such as the case in Figure 4, then

If the lengths of the lines of contact are not equal or if the material

of either cylinder or plane are different, then the total deformation

will be,

n 2

A similar approach for the computation of the deformation of a

cylinder to plane has been obtained from correspondence with Bob Fergusson [10].

The basis of the work is a paper by Airey [ll]in which formulae are given

for the solutions of elliptic integrals when the eccentricity (e) approaches

unity.

From Equations 4 and 5 we have,

where aid D^' are the two principal diameters of body 1,

and ' are the two principal diameters of body 2.

I.i the case of a cylinder to plane, if body 1 is the cylinder a id body 2

is the plane, then

= Diameter of cylinder
-24-

D2 = °° if the plane is truly a flat surface

D^'= 00 Since we have shown for the case of a cylinder to plane

B . -^^

A A

Thus the parameter — = —z— = AD^

B —

"1

From the previous Equation 38, we know

dE

de

dK

^ dl

where E is the complete elliptic integral of the second kind

K is the complete elliptic integral of the first kind

e is the eccentricity of the "ellipse" of contact which approaches unity

In Airey's work [11], expressions were given for evaluation of the

elliptic integrals when the eccentricity approaches one, viz

K(e^) = In — • - K (56)

2 ^

where K(e ) = complete elliptic integral of first kind

e^ = square of modulus e

2 h = square of the complementary modulus or (1 - e )

!

2 2 2 ) 1 1*32 f 1 + -y - h + ^ ^ • h + • • • [ 2 2-4 J

2 .\ ,2 ^2 ^2 2^2 2 \ ^2 ^ + — - T- )• h +

0- \- - / „2 ,2 \ 1 2 3 4

2 \ / 2 • 4 \

and E(e ) = In — • E, + E. (57)

1 2

2 where E(e ) = complete elliptic integral of second kind !2

I

• h + ^2 ' ^

• h^ + • •

•

2 P _ 1/1 l\ , 1^ ' 3 /2 2^1 l\ ,2 E._U-2(^r-2j-h-^-(l-2^3-4)-^ -

I
-25-

Thus as e - 1

k(e^) = In . ^

(58)

- e2)

From Equation 37 and using Airey's expressions when e

1

dK ^ _ 1 ^ E dE e X, 2v

e(I - e )

which reduces to

dK 1

and

dE-- 2, (59)

(1 - e )

i ^ ^ _ j, dE (60)

h B " e de

Since from previous derivations, namely Eq. 32 where

We substitute (6) into the above to obtain

A a-^ = |p (A^ + K^)

and

3 3

A a = |P (X^ + X^) ^

(AD^)

a

3 3 '''l

or

L

Now since a =

2 (the ellipse of contact approach where the major axis of the

ellipse is a) we have

PD

h = 12 (A

1

+ A ) -T

2 1}
From Equation 34

-26-

5a = fP (X^ + X^) (K)

Since In

4 1I

in 16

= r - (X, + X ) In -pr- 2 a 1 2 ^

I

In h = ^

In

3 P . . fl / 16 ^

^^2 Vl2(X^ + xpPD^

and since a = L/2

= I

(X^ + ^2^

^ L )

^ ^ ^ \3(\^ + X2^PD /

3

^ ' I + ^2^ I

^" 3(A^ + ^2^^^

(61)

since D = 2r

8 = T + X„) 7 In

1 4 L ^ 4 ^ , 1/ In "TTT — = In "7 + In 7 :

t-t— 6(A^ +X )Pr 6 (X^ + X2)Pr

In 7 = In 4 - In 6 = 1.38629 - 1.79176 - -.40547

o

- .40547 + In

S = (X^ ^X^) - -.608205 + 1.5 In

(A^ + X2)Pr

(62)

This expression is for the contact case of one side of the cylinder pressed
-27-

against a plane. If the normal approach (deformation) for the cylinder is

desired then

\otal = ^-'«2

If the materials of body 1 (cylinder) are the same as body 2 (plane)

then = and Eq. 61 becomes

« = 3 <V I TTtS o

where = ( + X^) = 2X,

P = Total load applied

L = Length of contract

D = Diameter of cylinder

Another approach for the cylinder-plane case was undertaken by

Lundberg [l2] in 1939 in which the contacting bodies were treated as elastic

half-spaces. Lundberg obtained approximations for the deformation by assuming

an elliptical pressure distribution in the narrow dimension (2b in Figure 5b)

and constant in the length of contact dimension (L in Figure 5b).

Recalling Eq. 1 and Figure 3 we have

Wj^ + w^ + + = 8

where 8 is the approaching distance of the points at the surface of the bodies.

4(1 - ',^) 4(1 - v ^)

Z = V Z = — V

^ E E

where V is the potential function
-28-

where 9 . =

1

1 - v/

1

Nov F(x,y) = ~ </) (x,y) = ~•li

dx'dy'

'.2

where r =(x-x) +(y-y 1/2

and if w^^ = Fj^(x,y)

= F2(x,y)

that is some function of x and y,

Then we have

Do not confuse the potential F(x,y) with Fj^(x,y) and F2(x,y). Since the

total deformation is

(63)

where (xQjy^) represents the point where the deformation is largest

we have

Wj^ + w^ = F^(x,y) + F2(x,y) = 4^^^ + |J F(xo,yo) - F(x,y)j (64)

Thus we see the relationships between the pressure distribution and the total

deformation is expressed by Eqs. 63 and 64,

The interested reader is invited to follow Lundburg's analysis but for

brevity the results will only be given here.

The relationship for the displacement of the surface of the plane under

the action of an elliptical pressure distribution when b<<a is
(x.o) (1

-29-

^ \fi (1.1932 + In ^) (65)

where F (x,o) + F (x,o) = depression of the center of the contact area (y « o)

E. e. = —

P = Total Load

L = Length of contact

b = Hertzian half-width

1. = Poisson's ratio

= Modulus of elasticity

The relationship for the total deformation 8 is

/l ^ 1 \ P 2

1.8864 + In ~i (66)

f2b This equation is based upon an approximation as a function of lnl~|and

is accurate for all values of I— lexcept when !• In this case the accuracy

is approximately four percent.

Since we have shown that

'4r(A

b =

1/2

we may substitute into Eqs , 65 and 66 to obtain

^2^ T [1.8864 + In L - In 2b]

+ k^) ^

[2(1.8864) + 2 In L - 2 In 2b]

= ^ v! 3.7728 + 2 In L - 2 In 2 - 2 Inb]

= (A, 3.7728 - 1.3863 + 2 In L - 2(-2) In

= ^\ ^ v! 2.3865 + In L^ + In ^ 1

4r(A^ + X^)^

j
-30-

2.3865 + In (.25) + In

1.0000 + In (67)

This S gives us the normal approach for one side of a cylinder-plane contact.

and

F = (A + A ) I

[1.1932 + In L - In 2b]

J. ^ Li 4r(A^ +

(X^ + A^) ^{[|1.1932 4- In L - In 2 -| In

= (Al + f[l. 11932 -.69315 ^ In ^

]

F = (A^ + f 4r(A^ + Xpp]

F = + A^) 7 (68)

This equation gives us the deformation of the plane surface.

In 1933 Weber [l3] considered the deformation of a circular cylinder

loaded on a side by pressure distributed elliptically in the narrow dimension

and constant along the longitudinal axis of the cylinder.

The assumption made here was that the Hertzian half-width b « radius of

cylinder rc The expression for the deformation of the cylinder (one-sided) is

2

(69)

where b is Hertzian half-width

This equation assumes that both contacting bodies are of the same material

>

If they are not we may rewrite the above relationship as
-31-

+

1 --2

Now

4r

In — = In 4 + Inr - Inb

b

= ln4 + lnr-'^ln ~

= In 4 + Inr + j In

4r(A^ + \^)?

= 1.38629 4 inr^ 4 in

, )p

= 1.38629 + r In

Lr

2 H\ + }^2^^

= 1.38629 4 In (.25) 4 in ^^^^^ ^^^^

= -693145 + fin

So we have

c I 1 2/ L

.193145 + T In

Lr

2 + X^)? (70)

Weber also obtained an expressed for the deformation of the plane surface as

(71)

Using the same logic as before in the case of the cylinder

2^
-32-

In —r = In L - Inn - Inb

= In L - InTT -

= \lnL - In. + ^ In ^^^^ ^

1 , L -lnn+

2 4r(A^ + X^^P

3

-1.1447 + f

In (.25) -h ^

In ^\pp^

1 -1.1447 -.69315 -f- in ^ ,^)p.

In ~- = -1.83787 + In . Trb 2 (A, + X )Pr

So

p 1 2 L

-1.33787 + T In

2

" + X2>P^ (72)

To obtain the total deformation for the case of a cylinder between two

anvils we have

8 = 2(V + V^)

°Total c

= 2

= 2

8 = 2

(A^ + Xp

(A^ +

(A^ +

193145 4 in

1 "2-

,4

5[- -1.33787 + T In

2 (X^ + X2)Pr

14473 + IIn

2 2

(Xj + X^) V J

(73)

Various other relationships for deformation of circular cylinders have

been derived by D6*rr [l4] , FCfppl [is] , Kovalsky [l6] , and DinniK [l7] . Only

the pertinent results will be given here.

D6*rr calculated the displacement of the points of initial contact by

assuming a smooth circular cylinder of radius r was compressed between smooth

"rigid" planes from the re lationship
•33-

2(1 - ^)

S = —

one-sided

1/2

where b =

or

2(1 - p

S = i -

ttE L

.193145 +

Thus the change in diameter of the cylinder according to Dorr would be

Total

According to Roark [iS] a similar expression for the total compression of

of a cylinder between two rigid planes was derived by Foppl, It is

2>

where b is Hertzian half -width

2

one-side

Roark also gives an expression for the mutual approach of remoted points in

two plates as

When a cyl inder is compressed between two parallel planes.

Kovalsky, in 1940, derived an expression for the deformation of a circular

cylinder of finite length loaded from two sides by pressure distributed

elliptically across the contact width. The change in diameter parallel to the

direction of the force was
-34-

^> I

[in ^ . O..O;]

Again V = Poisson ratio for the cylinder

E = Modulus of elasticity for cylinder

P = Total load

L = Length of contact

r = Radius of cylinder

b = Hertzian half width

Dinnik also calculated the change in diameter of a circular cylinder but

by assuming the pressure distribution across the contact width was parabolic

in nature. His expression was

Since Lundberg calculated the displacement of the surface of an elastic

half-space under the action of an elliptical pressure distribution from the

relation

where A = depression at the center of pressure zone.

We may write the normal approach of a distant point in the elastic plane to

the axis of the cylinder as the sum

where

2 Tot

Another relationship was obtained from a Canadian report which gives the

compression at the point of contact of a cylinder and a plane as

(78)
s =

a

-35-

where F = Applied measuring force

a = Length of contact

w = Diameter of the cylinder

1 + I +

if we replace L = a and 2r = w then

J

=. |i + i„ (79)

Thus the total deformation for a cylinder between two planes would be

'Total ' \* h

[19]

In a paper by Loo the problem of line contact between two infinitely

long isotropic, elastic circular cylinders of radius and under the

action of a compressive force P (where P a load per unit longitudinal length)

is discussed.

1/2

8 = P In 2R^ + In ZR^ + (K^ + K^) In

- |(K^ + K^)] (80)

where 5 = Normal approach of the two cylinders

P = Load per unit length P^^^^^^ /contact length

2(1 - .-.^)
= Poisson's ratio for the i^*^ body

t h

= Modulus of elasticity for the i body

B = 2( ^ + k^) (81)

In the case of a cylinder to a plane we may replace the plane surface

with a cylindrical surface with a very large radius of curvature. In the

real world of metrology this may well be the case anyway because any "flat"

measuring surface is not perfectly flat. In Figure 4, if one of the contacting

anvils has a very slight curvature and its surface can be approximated by the

arc of a circle, then we can solve for the radius by the sagitta formula.

For example, if the anvil surface is 0.5^" higher at the center than the

ends and the anvil width is 0.375 inch then

t

Anvil Width

i

Anvil Width

Radius =

2 ' A

2

2(.5)(10"^)

Radius = 35156 inches

Thus we may solve for the deformation by plugging this value back into

Equation 80. To obtain the total defoirmation for the case of a cylinder

between two planes

8 =8+6

Total 1 2
-37-

EXPERIMENTAL VERIFICATION

Since the validity of any theoretical foraulation is dependent upon experimental verification, we have performed a literature search for verification of

the formula describing the deformation of a cylinder between two planes.

Bochman [20] performed his dissentation on this subject along with measurement

of the deformation of steel balls. He began by assuming the oblateness of a

cylinder between two planes increases in proportion to the load and is

influenced by the size of the measuring surfaces. An elaborate lever system

was devised for amplifying the deformation and a Zeiss optimeter was used as

the readout device. The contacting anvils were hardened and finished to

"end-standard quality". The optimeter was calibrated with end standards but

could only be estimated to 0,1/x (4m") with a measuring accuracy of approximately

0.2^ (8m") • The cylinders used in the test were steel wires (used to measure

the pitch diameter of external threads) with nominal diameter 0,18mm to 5.70mm.

The applied load varied from 1 Kg to 10 Kg in some cases. Anvil pairs used

were 5.92, 8.03, and 14.95mm in length.

The contact case where the wires were shorter than the measuring surfaces

was also examined. The test results proved that the length of the wire exceeding

the size of the measuring surface has no effect on the amount of deformation.

In other words it is immaterial whether the length of the wire exceeds that of

the measuring anvil or whether the measuring anvil is larger.

The values for oblateness for the 1 Kg load were plotted against the cube

root of the curvature (reciprocal cylinder radius) with the result being linear.

Thus he concluded the proportionality equation was

M = C X lO"^ f nJ^Td mm (82)
-38-

where M = oblateness in mm

C = Constant of proportionality

P = Applied load in Kg

L = Contact length in mm

D = Cylinder diameter in ram

The factor C was determined as 0.9228 from 35 test series. The mean

error of the individual value of C was +1.997o and the maximum deviations from

the mean were + 3.6% and - 5.l7o.

By rewriting the equation we obtain

26 = 0.2207 I

d"-^''-^ microinches

where 26 = oblateness in microinches (10~^ inches) of a

steel cylinder when compressed between two flat

V , steel planes

P = Measuring force in pounds

L = Contact length in inches

D = Cylinder diameter in inches

In a 1921 report by the National Physical Laboratory a test was conducted

on the compression of a 0.05 inch steel cylinder between a pair of flat

measuring anvils 0,25 inch in diameter. The results of that experiment are

shown in Figure 7 along with the results of Equations 55,62 and 82 frc«n this

paper. There is no mention of the uncertainty in the NPL experiments. We

see from Figure 7 that the data from NPL agrees quite closely with Equation 62,

In a more recent study of line contact deformation by Fergusson [6] the
-39-

results of a series of tests on hardened steel rollers between parallel plates

was discussed. From Equation 61 the equation for two bodies of the same

material Is

2 8 = 3(Vfln 3K PD

o

Thus the total deformation for the case of a cylinder between two planes would be

«Iotal = 28

From communication with Fergusson a number of tests were performed and

an empirical relationship was found which agreed within 7% of those values

given by the theoretical equation above for values of force from «2 lb to

16 lb, values of diameter .01 inch to 10 inch, and values of L from ,01 to

1 inch. The relation is

§ = 0.00534 LDP (10 ) inch (83)

where g = Defoimation

L = Contact length in inches

D = Cylinder diameter

P = Total force in lbs

Again the total deformation would l)e 26 -

The experimental results for a 1.0 inch cylinder between flat, parallel

planes 0.118 inch in width were summarized by still another empirical formula.

6 = 0.0091 L-0-" D-O-O^ pO-^^'^lQ-*) inch (84)

Many factors have an effect upon the deformation in a cylinder to plane

case such as the type of surface finish and the parallelism of the plane

surfaces.
-40-

In the most recent measurements on the compression of a roller between

parallel planes Thwaite [21] devised an apparatus consisting of a cylindrical

shaft with an air bearing for loading coupled with capacitance probes to sense

the compression. The test cylinder and flat were made of Cr - Mn hardened

tool steel. The plane was flat within 0.025um (1m") and the cylinder was

straight to 0.025Mm. Both plane and cylinder possessed a surface roughness

of O.Olpm center line average.

The entire system possessed a resolution of 0.003iJm (.12(j"). Three series

of measurements were made on a 6.35mm diameter cylinder in contact over a length

of 9.525mm for the load/contact length range of 0.05 to 0.4 Kg/mm. The mean

observed slope differed by approximately 5% from the equation based on the

finite rectangle, that is Equation 55

while the values predicted by the Bochman and Berndt [20] relationship differ

from Thwaite 's slope by 45%.

Thwaite concludes that within the load range tested the equation based on

the finite rectangle gives the best approximation to compute the compression

involved in the case of a cylinder between two planes. That is, the finite

rectangle relation seems to be the one that would be of more practical use in

the metrology field.

CONCLUSIONS , Since the contacting case of a cylinder between two planes has such

widespread application in the science of metrology one needs to be able to

apply elastic corrections for the most precise determination of a diameter of the

cylinder. There are many prediction formulae depending upon the assumptions made.

2« =^ (X^+ X^)
-41-

This paper has assembled the majority of the equations in existence in the

literature and a master computer program was written to evaluate their

agreement over the range normally used in precision metrology. Figures 8,9

and 10 give the results of three equations 62,55 and 82 which are representative

of the three basic types of cylinder-plane equations discussed on page 4c,

Eight of the equations were evaluated over the range 0.5 - 2.5 lb force

and cylinder diameter varying from .001 to 1.00 inch. The slopes (microinches/lb) of five theoretical equations are plotted in Figure 11, The slopes

of three empirical equations over the same range are shown in Figure 12c To

obtain the deformation for any one equation

a) Find the diameter of the cylinder,

b) Read the corresponding slope from the vertical axis, and

c) Multiply that slope by the measuring force in pounds to obtain the

total deformation.

Note: Keep in mind that these slopes are for the total deformation of a

cylinder between two plane surface. For example the values are

twice the values given by the equations listed. Also the values for

Poisson's ratio and Young's Modulus are for 52100 bearing steel

in which

V = 0.295 and E = 29.0 x 10^ psi.

The length of contact between cylinder and plane is 0,375 inches

in all cases ^

One important fact to remember in all contact problems is to keep the

measuring force less than that force which will cause permanent deformation to

the object under test. For the case of a cylinder against a plane a

nomograph has been developed for easy computation of the stress (Figure 13).
-42-

In certain cases where the compression needs to be known to better than

57o then the only practical way to account for surface finish, the actual

material parameters, and the exact geometry of the cylinder and flat would be

to determine the compression experimentally. For the load range normally

used in the metrology field it appears that Equation 55, based on the finiterectangle approach, gives the best approximation to the compression obtained

between a cylinder and flat (Figure 14).

On a practical basis, since most line compression is small an error of

10 - 307o is not enough to give one insomnia because the measurement uncertainty

(standard deviation) will likely be the same order of magnitude if not larger

than the actual deformation of the system.

SUMMARY OF EQUATIONS

b = Hertzian half-width of contact

|4R(A

b =\ L

Equations for Compression of Cylinder and Plane in Contact

.3

(55) 8 ^ ^

(X^ + X^)

(62) 8 = - (A^ + k^)

1.00 + In

(A^ + ^2^^^

-.608205 + 1.5 In

P r ^ '1 ^ ^ "2 (79) S=T^—^ +

L \ ttE,

(A. + AJPR

^ L

1.00 + In PR

1 + 1^ 1 + V2 ^E, ^E,

Empirical (Steel Cyl inder--Steel Flat)

(82) 2 8 = 0.2207 | d'^^"^ (lO'^) inches

(83) 8 = 0.00534 l"'^^^ d"*^^^P '^^^ (lO"^) inches

(84) 8 = 0.0091 l"'^^ d"-^^ P-^^^^ (10""^) inches
-43-

Equations for Deflection of Elastic Half-Space

(68) F = f (^1 + [^-.193145 + j

In 1

1 + ^2^ ''"J

(72) 8^ = I ^

1-1.33787 + f

1. ^ X^PR

• Equations for Compression of A Cylinder Between Rigid Planes

(70) 8^ = 2 f (^) [.193145 . i m f,^)p]

(75) 8. = 2

(77) 5. = 2

(78) 6, = 2

c

L \ ^E /

L ( -E )

P /l - V L \ ;rE

.3333 + 1 1 LR 1

2 (A^ + A^)!*]

.407 + In

2R

9R

.333 + In

^b

• Equations for Normal Approach of Two Planes with Cylinder Between the Planes

. 2 (73) ^ = I

+ -1.14473 + In

(A^ + V2)Pj

Equation 73 is the sum of Eqs. 70 &. 72
-44-

REFERENCES

[l] He Hertz: Uber die Beriihrung fester elastischer Kbrper, 1881; Uber

die Beriihrung fester elastischer Korper und iiber die Harte, 1882,

English translations in H. Hertz, "Miscellaneous Papers," pp 146-183,

Macmillan, New York, 1896.

[2] J. Prescott: Applied Elasticity , Dover, New York, 1961.

[3J 0. D. Kellogg: Foundations of Potential Theory , Murray, New York, 1929.

[4] Mary L. Boas: Mathematical Methods in the Physical Sciences . John Wiley

6c Sons, Inc., New York, 1966.

[5] C. Hastings, Jr.: Approximations for Digital Computers , Princeton Univ.

Press, New Jersey, 1955.

[6] Bob Fergusson: Unpublished correspondence on deflection at point and line

contact, School of Engineering, University of Zambia.

[7] H, R. Thomas and V. A. Hoersch: Stresses due to the pressure of one elastic

solid upon another. Bulletin No. 212 Engineering Experiment Station, University

of Illinois, Urbana, Illinois.

[8] E. R, Love: Compression of elastic bodies in contact. Unpublished report,

Defense Standards Laboratory, Australia.

[9] D. Bierens de Haan: Nouvelles Tables d'integrales Ddflnies. Hafner,

New York, 1957.

B. Fergusson: Elastic Deformation Effects in Precision Measurement,

Microtecnic, Vol. 11, pp 256-258, 1957.

[11] J. A. Airey: Toroidal Functions and the Complete Elliptic Integrals,

Phil. Mag, p 177, Vol. 19, No. 124, Jan. 1935.

[12] G. Lundberg: Elastische Beriihrung zweier Halbraume, Forschung auf dem

Gebiete des Ingenieurwesens , Vol. 10, September /Oc tober 1939, pp 201-211.

[13] C. Weber: Beitrag zur Beriihrung gewolbter Oberflachen beim ebenen
-45-

Formanderungszustand, Z. angew Math. u. Mech., Vol. 13, 1933, pp 11-16.

[14] J. Dorr: Oberf lachenverf ormungen und Randkr^fte bein runden Rollea

der Stahlbau, Vol. 24, pp 204-206, 1955.

LI5] A. Foppl: Technische Mechanik , 1907.

L16] Kovalsky: Quoted in H. Rothbart's Mechanical Design and Systems Engineering ,

1964.

[17] Dinnik: Quoted in H. Rothbart's Mechanical Design and Systems Engineering ,

1964.

[18] R. J. Roark: Formulas for Stress and Strain . McGraw Hill, New York, 1965.

[19] To T. Loo: Effect of curvature on the Hertz Theory for Two circular cylinders

in contact, J. Applied Mechanics, Vol. 25, pp 122-124, 1958.

[20] I, H. Bochman: The Oblateness of Steel Balls and Cylinders due to Measuring

Pressure, Z. fiir Feinmechanik , Vol. 35, 1927.

L21] E. G. Thwaite: A Precise Measurement of the Compression of a Cylinder in

Contact with a Flat Surface, Journal of Scientific Instruments, (Journal

of Physics E), Series 2, Volume 2, pp 79-82, 1969.

l22] T. F. Conry and A. Seireg: A Mathematical Programming Method for Design of

Elastic Bodies in Contact, Journal of Applied Mechanics, pp 387-392,

June 1971.

[23] D. H, Cooper: Tables of Hertzian Contact-Stress Coefficients, University of

Illinois, Technical Report R-387, August 1968.

[24] H. Poritsky: Stresses and deflections bodies in contact with application

to contact of gears and of locomotive wheels. Journal of Applied Mechanics,

Vol. 17, pp 191-201, 1950.

[25] P. R. Nayak: Surface Rough ness Effects in Rolling Contact, Journal of

Applied Mechanics, pp 456-460, June 1972.
-46-

[26] J. A. Greenwood: The Area of Contact of Rough Surfaces and Flats, Journal

of Lubrication Technology, Trans. ASME, Vol. 89, Series F, 1967, pp 81-91.

[27] S. Timoshenko and J. N. Goodier, Theory of Elasticity , McGraw Hill,

New York, 1951.

[28] S. D. Ponomarev, V. L. Biderman, et al . : Material Strength Calculations in

Machinery Design, Vol. 2, Mashgiz, Moscow, 1958,

[29] F. C. Yip and J, E. S, Venart: An Elastic Analysis of the deformation of

rough spheres, rough cylinders and rough annuli in contact, J. Phys. D:

Appl. Phys., 1971, Vol. 4, pp 1470-1486.

[30] C. A. Moyer and H. R. Neifert: A First Order Solution for the Stress

Concentration Present at the End of Roller Contact, ASLE Transactions 6,

pp 324-336, 1963.

[31] J. P. Andrews: A Simple Approximate Theory of the Pressure between Two

Bodies in Contact, The Proceedings of the Physical Society, Vol. 43,

Part 1, No. 236, pp 1-7, January 1, 1931.

[32] C. F. Wang: Elastic Contact of a Strip Pressed between Two Cylinders,

Journal of Applied Mechanics, pp 279-284, June 1968.

[33] J. Dundurs and M. Stippes: Role of Elastic Constants in Certain Contact

Problems, Journal of Applied Mechanics, pp 965-970, December 1970.

[34] W. A. Tupl in: Commonsense in Applied Mechanics, The Engineer, June 26, 1953.

[35] J. B. P. Williamson and R. T. Hunt: Asperity Persistence and the real

area of contact between rough surfaces, Proc . R. Soc, London, Vol. 327,

pp 147-157, 1972.

[36] G. Berndt: Die Abplattung von Stahlkugeln und zylindern durch den

Messdruck, (The Compression of Steel Spheres and Cylinders by Measurement

Pressure), Z. Instrumentenkunde , Vol. 48, p. 422.

[37] L, D. Landau and E. M, Lifshitz: Theory of Elasticity , Pergamon, London,

1959.
-47-

[38] I, Ya. Shtaerman: The Contact Problem of Elasticity Theory . MoscowLeningrad, 1949.

USCOMM-NBS-DC

z

Geometry of the contact between elliptic paraboloids

with principal axes of body I (Xi^yi, z) and the

principal axes of body 2 (xa.yeyZ). The angle is

the angle between the x, z and XgZ planes, i.e.

the principal radii of curvatures R| and R2.

FIGURE I
Cross-section of two surfaces near the point of

contact 0*

FIGURE 2

Geometry of deformed bodies. Broken lines show the

surface as they would be in the absence of deformation.

Continuous lines show the surfaces of the deformed bodies

FIGURE 3
(a)

Measuring

Force

Measuring

Force

(b)

MEASUREMENT OF THE DIAMETER OF A

CYLINDER

FIGURE 4
Contact geometry of two parallel cylinders.

FIGURE 5
o o o o o o o lo o in o lO

ro fO cvj OJ

—
10

8

U 4

'Eq. 62

Compression of 0.05 inch diameter

steel cylinder between 0.25 inch

flat anvils . NPL 1921 data ° °

0.5 1.0 1.5 2.0 2.5 3.0 3.5

LB FORCE

FIGURE 7
0.5 1.0 1.5 2.0 2,5 3.0
1

[

Eq. 62

/

using equations 55, 62» and 82.

0.5 1.0 1.5 2.0 2.5 3.0

LB. FORCE FIGURE 9
Total deformation of 1.00 inch

diameter steel cylinder between

0.375 inch flat steel contacts

using equations 55,62, and 82 /Eq.62

/

/

/

/

/

/

0.5 1.0 1.5 2-0 2.5 3-0

LB. FORCE

FIGURE 10


FIGURE 13 (a)

NOMOGRAM FOR COMPUTATION OF MAXIMUM STRESS

ENCOUNTERED IN CONTACT PROBLEM OF A

CYLINDER AGAINST A PLANE

Purpose: To determine whether the force one uses in the measurement of

the diameter of a steel cylinder will exceed the yield point

and cause permanent deformation to the cylinder.

Governing Equation

where

max

2P

TTLb

measuring force

contact length between cylinder and plane

Hertzian half-width of contact

0

1/2

R =

X. =

1

Vi =

Ei =

Radius of cylinder

1 - vi^

nEi

Poisson ratio for one material

Modulus of elasticity for one material.

The nomograph was developed for 52100 steel in which

Example

:

= 0.295 and E = 29.0 x 10° psi

If Force = 1 lb, Contact Length = 0.375 inch, Radius = 0.01 inch--Find maxi

mum stress -r

.OlO-"-.

stress

37,000 psi

Radius

LO

Force

0.375

Contoct Length

(1) Locate measuring force and length of contact on the appropriate scales

and connect these points until the turning axis T is intersected; (2) Locate the radius and connect with the intersected T axis; (3) Read maximum

stress to be expected within the contact area; (4) Compare maximum stress

read from nomograph with curve in Figure 6.
T

-r.OOl

J

400

"300

.-200

--.005

--.010

-- 100

.05

.100 10

STRESS

XIO' PSI

50

--.50

-- 1.00

RADIUS

IN.

-r3.0

2.5

--2.0

1.5

1.0

.5 - .4

.3 - .2

-L .1

FORCE

LBS.

-r .5

4

35 — .3

— .25

— .2

— .15

-J-.IO

CONTACT

LENGTH

IN.

FIGURE 13 (b)
Equation For The Calculation Of Compression Between

A Cylinder And Plane Surface

Force

'////

8= f

(X, + X2)[l.00 + In

(X,+ X2)PR

where P« measuring force

L"Contact length between cylinder and plane

R » radius of cylinder

X,«

uj * Poisson's ratio

Ej • Young's modulus

When material of cylinder and flat are same, Xj^Xg

Cylinder Between Two Parallel Planes

£

FIGURE 14
FORM NBS-114A (1-71) U.S. DEPT. OF CCMW.

BIBLIOGRAPHIC DATA

SHEET

1. PUBLICA t ION OR REPOKl NO.

NBSIR 73-243

2, Gov't AcceP«'icr? No.

3. Recipient's Accession No.

4, TITLE AND SUBTITLE

ON THE COMPRESSION OF A CYLINDER IN CONTACT Willi A PLANE SURFACE

5, Publication Date July 19, 1973

6. Performing Organization Code

7. AUTHOR(S)

B. Nelson Norden

8. Performing Organization

NBSIR 73-243

9. PERFORMING ORGANIZATIO.N NAME AND ADDRESS

NATIONAL BUREAU OF STANDARDS

DEPARTMENT OF COMMERCE

WASHINGTON, D.C. 20234

10. Project/Taslc/Work Unit No.

11. Contract/Grant No.

•

12. Sponsoring Organization Name and Address

Same

13. Type of Report & Period

Covered

Final

14. Sponsoring Agency Code

IS. SUPPLEMENTARY NOTES

16. .\l)SrKACT (A 200-word or less factual summary of most significant information. If document includes a significant

_, biblio^raphv or literature survey, mention it here.) , . , , The me.^sUreraent of a diameter of a cylinder has widespread application in the metrology

field and industrial sector. Since the cylinder is usually placed between two flat

parallel anvils, one needs to be able to apply corrections, to account for the finite

measuring force used, for the most accurate determination of a diameter of the cylinder.

An extensive literature search was conducted to assemble the equations which have been

developed for deformation of a cylinder to plane contact case. There are a number of

formulae depending upon the assumptions made in the development. It was iirjnediately

evident that this subject has been unexplored in depth by the metrology community, and

thus no coherent treatise for practical usage has been developed.

This report is an attempt to analyze the majority of these equations and to compare theii'

results within the force range normally encountered in the metrology field. Graphs have

been developed to facilitate easy computation of the maximum compressive stress encountered in the steel cylinder -steel plane contact case and the actual deformation involved,

Since the ultimate usefulness of any formula depends upon experimental verification, we

have compiled results of pertinent experiments and various empirical formulae. A complete bibliography has been included for the cylinder -plane contact case for the

interested reader.

17. KEY WORDS (Alphabetical order, separated by scmic ilons)

[! Contact stresses; cylinder measurement, deformation; Hertzian contact problem; line con- tact; mechanical measurement; metrology.

18. AVAILABILITY SI ATEMENT

UNLIMITED.

[Xj FOR OFFICIAL DISTRIHUI ION. DO NO I RELEASE

TO NTIS.

19. SECURI TY CLASS

(THIS REPORT)

UNCL AS'^IFIFI)

21. NO. OF PAGES

67

20. SFCURl l-Y CLASS

Ci Ills PAGE)

22, Pru

UNCL ASSIFIFD

USCOMM-DC Cl^2.»4.o;l