PROCESSES OF FRICTION IN THE BRAKES OF AIRCRAFT WHEELS SELECTION OF FRICTIONAL PAIRS

• Document Type: CREST
• Collection: General CIA Records
• Document Number (FOIA) /ESDN (CREST): CIA-RDP81-01043R002000030004-8
• Release Decision: RIPPUB
• Original Classification: K
• Document Page Count: 194
• Document Creation Date: December 27, 2016
• Document Release Date: May 30, 2013
• Sequence Number: 4
• Publication Date: January 1, 1955
• Content Type: REPORT

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Equation (124) is a canonical Bessel equation. The solution of this equation
has the form (Bibl.26) of
R1= CuJ0 (pr) -1- C2Y0 (pr),
where where Jo(pr) is a Bessel function of zero order of the first kind; Yo(pr) is a Bessel
function of zero order of the second kind.
To solve eq.(123), let us write it in the following form:
Z"-~-p2Z=. 0.
The solution of eq.(126) may be presented in the following manner:
(126)
Z = C3 COS P (Z -1- 70). (127)
Therefore, the partial solution for eq.(115) can then be written as
01part = C. cos (? (z -1- zo) [C1 J0 (pr) + CYO(pr)j.
? Placing the arbitrary invariable C3 into parentheses we get
01 part = i.OS p (z + z0) [C'J0 (pr) + C"YO (pr)]
The general solution will then be
0' _ cos pi (z + zo) [C{J0 (Pir) + CiYo Oil)],
(128)
which also satisfies eq.(115) and the boundary conditions of eq.(116). At z = 0,
we have Of = e*. Therefore,
COS Pizo [00 (P{r) + CYO (pir)1
{
(129)
? Using an artificial way of multiplying both sides by (1 + 1) and expanding the
left-hand side into the following series:
100
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0
1 + = B1Jo (pir) + B2J0 (pir) ; ... -It- B1Yo (pir) +
+ B'Yo(Pir) + B1J o (pir) + BiYo (pir) ? (130)
I
Then, expression (129) may be written as follows:
tj J[BiJo (pir) + B;Yo ((i )] =
+ CtYo(Pir)], (131)
from which we find Ci and C;
Bie?
Ci = 2 cos p;zo
Ci = B;e,
2 cos ptzo
? At z = h, the boundary conditions of eq.(116) yield
[- Pi sin pi (h. + z0) + 3 cos Pi (h + =,~,)1 [C{Jo (pir) +
CtY o (Pis)] 1? = 0. (134)
In expression (134), only the first parentheses can be equal to zero, so that
Cr
t'g Pt (h. --'r 2.0) _ -- ;
Pi (135)
2.0 =
From the
a a
arc t.g -Pt - pile are tg P?
1
Pi Pt
(132)
(133)
(136)
boundary conditions (116), at r = R1 and r = R2, we find
C{ [Pilo (PiR1) - aJ0 (PiRI)] + Ct [P1Y0 (P{R1) -
- aYo (PiRI)] = 0; (137)
101
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? It is clear that this is possible only when each of the terms in eqs.(137)
and (138) in parentheses is equal to zero, i.e.,
PtJo (PtRI) - aJo (PtR1) 0;
PjJo (PiR2) -F aJo (PiR2) = 0;
PiYo (PiR1) - aYo (PjRj) = 0;
PtYo (PiR3) + - QYo (PiR2) = 0.
The relationships of (139) - (142) are possible when pi is the root of the cor-
responding equation.
From the general theory of the Bessel function (Bibl.32, 33) it is known that
rJn (pr) dr = R~ [ J,i (PR) -F C1 -- R'Prl A (PR)] (143)
o
If p is the root of the equations, then
C{,[P{Jo (PtR3) + zJo (PiR2)1+ Ct [ptYo (PRs) `I-
+ aYo (P1R2)] = 0.
PJ,z (PR) 4- :; Jn (PR) = f1
and eq.(143) is simplified and assumes the form
0
\rJ(pr)dr
p [R2 (?2 + p2) - ,i2] Jn (PR).
(138)
(144)
(145)
In our concrete case, when p is the root of eqs.(139) - (142) and the order of
the Bessel functions is n = 0, eq.(145) is written in the following manner:
R
rJo (Pir) dr == 2pa (a2 + p2) Jo (pR).
(146)
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To find the coefficients Bi and Bi, let us take advantage of the expansions
of eq. (130) :
Multiplying the left and right-hand terms of the expansion (147) by rJo(pir)drs_
then integrating within the limits of R1 to R2, and considering expression (146) we
get
R,
rJo (pir) dr =
1 = B1Jo (pir);
1 = ~' B{Y'o (p{r).
{
(a' + p{) [RJ 2 (p{Rs) -
20
{ 2p1
- RiJo (p1R1)] B1.
Sindlarly, we obtain
R,
rJo (p;r) dr =
(Cr [R2Y2 (p1R2) _
2p{
-- Rid o(p{R-)] B
(147)
(148)
(149)
(150)
From egs.(149) and (150) it is easy to compute the coefficients Bi and BI:
B1 =
0
~A?2p{
D
A'2pt
B{ = D,
R,
A = S rJo (p{r) dr;
Ri
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(151)
(152)
(153)
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?
R,
At = S rYo (Pjr) dr;
R,
D =[R 2Jo (P{R:) - RiJo (P{RI)) (a2 +p{2); (155)
D' = IR2 0 (PiR2) - RiY (P{R1)J ((72+p12). (156)
Considering eqs.(151), (152), and (136), the expressions (132) and (133) will
yield ?
O Ape
t
D cos p{ are tg a/p1
Pt
0'A' p~
C;=I arctgQ/-
i D' cos pi
Pt
0
Therefore, the final general solution of the differential equation (115) for
stationary heat conditions with the boundary conditions (116) is written as follows:
arc tg a/p{
cos p, z -}-
- h
P{
(154)
(157)
(i58)
u
{ cos p arc vg a, P{ _ h i
{ p1
X [jA
D Jo (Pr) + D, Yo (Pir) ? (159)
J
Equation (159) is a function of the coordinates z and r and the parameters 6
and pi. The parameter a, again as before, represents the thermal characteristic of
the material of the pair element (a where at is the coefficient of heat emis-
sion and X the coefficient of heat conductivity).
The basic parameter pi of eq.(159) is easiest to compute by a graphic solution
104 STAT
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where i = 1, 2, 3, ... no
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of the equation systems (139) - (142).
For this, we introduce the substitutions
Substituting these in eqs.(139) and (140) we get
or, otherwise,
and by analogy
y3 = aR1Jo (x1i)
nite points xi and x2,.
To compute the parameter p i we reduce egs.(139) and (140), substituting there
the values of the found series xl i and x2i; this yields
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?
In this way, the parameter of the general solution of the differential aqua-
,tion (159) is computed; this parameter, resulting from eqs.(139) and (140), is a
function of the thermal parameters and geometric dimensions of the elements of the
;friction pair.
To study the performance of a friction machine with complete reciprocal contact
overlap, eq.(159) which characterizes stationary heat conditions is quite adequate.
It should be noted that the differential equation of nonstationary pro-
cesses (117) with the initial and boundary conditions (118) and (119) is solved in a
similar way. The general solution of eq.(117) is sought in the form of
0" = ZRT
which, by analogy with the preceding, assumes the form of
\' x
1
zo) [C' .10 ("?) 4- C"Yo 4w-) I -
(166)
(167)
Substituting the general solution (167) in the initial and boundary condi-
tions (118) and (119), we obtain a system of equations, as follows
(_ Il' =-- Jsin z [C'?!
o 0 (?1r) -j C Yo (?i0],
0
) '
it
Si ii r ? {C C?iJo (111R1) -
:;JO (ILiR1)] -r C'., l111}-o (I1iR1) -- aYo (ILiR1)] 1 >
(1 = ~, e~ sin Piz (C' b?iJo (IliR2) -t- I
i l
,J0 (}Lil?-) C l1Lly0 (?1R2) + aYo (?iR2) }
`)
[C'J0 (?1r) + C"Y,) (IL1r)1 x
i ,
i( (p cos 1jh3 + a sin Pjh3).
106
(168)
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The arbitrary invariables C?, CT?,
~_from the system of equations
h !i R ,
Q' = a' [2J?, \J'dz + 21rR2`rr'dz + n (R2 - R;) ~~'dr .
o u R,
A A,
Jo (Par) + fo (par)] It
[~
ty of heat flowing per second through the. material of one
?
characterizing the quanta-`
of the ring cylinders,
/f pvsK (1 - ocrp), (169)
aarc Or a!pi
'OS pi - - h
pi
Adjusting the right-hand terms of egs.(169) and (170) and making the reduc-
? tions, we obtain
E
P{ X
Jf pv3K (1 - oc,) = 21ra'(1' (R, -}- R2)
i
X
?
-F-~i'(i (R~-Ri)
(:OS pi
and the parameters and (3_i are computed
expression for
(170)
Z ? (aretgYp
- -- /l dZ +
Pi '.
81'C tga Era
Cos pi Z -~ I - h
2 pi
Pi are tg a/pi X
COS pi
Pi Iz
R,
X rf J. 1 jy (par) + y Y. (pir)] dr.
Ln
R,
We reduce eq.(171) under the condition that z = 0. In this case,
107
(168).
By analogy with eq.(88), we derive an
where (1 - Z,p) is the coefficient indicating what part of the general quantity of
heat is branched into a given ring cylinder. At stabilized heat conditions, this
quantity of heat passes into the surrounding medium through the inner, outer, and
free end surfaces of the ring cylinder, so that
(171)
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if pvsx (1 -acre) 4r.a'6' (R, + R,) x
X
A, 1 u tg a/Pi h
L Jo (Pir) + D, Yo (Pir) ]Cos P{ Pt - -1.
J
-{- (i' (R -- Ri) p JA CJ0 (PjR2)--J0 (PiR~)1 + III X
i
x [YO'(pjR2) Y'(pjRj)j
COS Pi Pi t h
are tg a/p
Using the following symbols:
=x=arct,galp1.,
0
(172)
(173)
S
K 2 -j J pt. cus (ZH '.- P, ,1) o
i
0
(192)
YO (Pir) dr
.lo(Pir?) + A'
~~ tt (193)
7
T, E'1 cos (z,{ - Pih)
L = a v
i
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(~ -. - - - btain
? From eq.(191) we o
L (RZ - R1) pi sin [piz + (sK - Pih)l .00.
ao?
dz - K + L (B2 - 1{1) cos [Piz + (z1 Fth))
(194)
- I ~ we get the following
symbols in eqs.(192) and (193) for &,p
? _ Using accepted
the
0a
Jf PI'S it (1 - aTF)
K (R1-}- B2) -}- L (R" - Hi) cos [pjZ -}- (ZK - Pih))
(195)
(194); this will yield the following equation for the tem-
which we substitute zn eq.
perature gradient:
L(R2-RI)Pisin[piz+(z,s-Pih)]
R K L (f;s - R1) cos {z + (Zx ._ Pih)1} 2 X (196)
dZ (?1 + 2) {
x J1 Pz sx (1 - aTP)
that the higher the temperature gradient in the
? From eq.(194-) it becomes clear
friction pair, the greater will be the temperature upon contact.
elements of the ,~
shows that the gradient is a function of the coefficient a ?
Equation (196) rature gradient.
Any increase in the coefficient a Tp leads to a decrease in the tempe can be
Y
he temperature gradient, at symmetric points in the elements of the pair,
T
in the case when the friction pair is made of the same material and
the same only
its elements have the same geometric dimensions.
Conditions in Testin Small Specimens
i
n
I?iai_ntainin Certa
is are sufficient-
d
ults of experiments on heating of rubbing bo e
that the res
So
ractical appli-
ly universal as far as their transfer from one model to another or p
cation is concerned, certain conditions must be maintained. the
? as proved by the theory of similitude of physical phenomena,
Generally,
115
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expression:
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uniformity of certain criteria must be preserved (for example, Fourier's, Reynolds)
Strouhal's, Euler's. Pekle. etc. numbers).
The similarity of heat phenomena can be preserved if one or several criteria of
similitude (Bibl.23) are maintained.
? In particular, similarity of heat flows, which is sufficiently accurate in
practice for a stationary temperature field of rubbing bodies, is ensured by the
constancy of the coefficient a Tp which is computed from egs.(100) and (179). Ilow-
ever, in view of the complexity of these equations for a preliminary clarification
of the conditions of similitude, the solution of Fourier's equation for linear heat
flow can be used.
The solution for linear heat flog.-has the form of
(1 - aTV) Jipv$,(
(197)
Ir UaC S
where W = perimeter of nominal contact;
S = nominal area of contact.
All other symbols as earlier.
In particular, for prismatic specimens with contact dimensions of h1 and h2, we
have
6) =2(hl-}-h2), S-h1h2.
In this case, eq.(197) assumes the form
(1 - TO Jf PI'S 1(
0 =
0
(198)
*See the article by 9.S.Shchedrov, (Bibl.39) ItTemperature on a Sliding Contact".
-collection Friction and Lear in Machines, published by Kh.Izd., Academy of Sciences
of the USSR, 1955-
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If, for the second model, all terms in the equation are indicated'by the-same
letters but with the addition of primes, we will have
(1 - ate,) J/' p'v'
if 2 (hi + hz) hlh2Aa',
(199)
The similarity of the heat phenomena in the two described cases is equivalent
to the condition
(200)
0'
0,
(201)
if (hi -}- h2) /111,2); a~ Y(/tl -}- h2) hl/t0Xa'
where 13 = 1 and and (3T = 1 - a' Tp'
The equality (201) makes it possible to study the influence of different groups
of parameters for two different models.
As an example, let us examine the influence of shape and area in a nominal
contact.
For this, we presume the following:
;
1) Same materials of the rubbing bodies, XT =X
2) Same sliding speeds, vT = v;
3) Load changes by n times, pT = pn;
4) Area of nominal contact changes, by n times:
case a - at a change in form of the nominal contact:
hi = hl, h2 = h2n;
case b - at geometric similitude of form of the contact:
hi . hi j/n and h; = ht Yn.
117
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In case a, the equality (201) assumes the form
POP,
Y(hl+h2)a' v nl +h,) all
c
(202)
From here, we obtain the unknown ratio of Of and Off? which assumes equality of
?
temperatures in both cases
(h1 T h2)a'
V(+h2)e
(203)
From this it follows that the ratio of similitude criteria (3f depends not
(3' f
only on the area but also on the shape of the nominal contact.
In case b, the equality (201) assumes the form
4
Pf _ f' Yn
(204)
Pf = PT 1/n vr;l.
?
(205)
This ratio shows that, in geometrically similar nominal contacts, their dimen-
sions do not at all affect the ratio of the similitude criteria.
In particular, if the coefficient of friction does not depend on pressure, we
will have ft = f for the same materials. If, in addition, the same conditions of
heat exchange in the ambient medium are assumed, then a" = at. In this case, the
equality (205) is considerably simplified:
(206)
118
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The correlation (206) shows that a change in load and area has little effect)
since the quantity n appears in a ratio of 1/4. Therefore, the same conditions for
heat exchange with the ambient medium are'not the best for minature-scale operations
of heat phenomena. It is easy to see that, if all = at V-n, the equality (206) will
assume the form
Therefore, the coefficient a.1pis a necessary and sufficient criterion of similitude.
The same results are obtained with pt = p, vsk = vsk n, other conditions being
equal.
Some studies for determining the criteria for miniature-scale operations were
made by Parker (Bibl. 35), but the value of these studies is considerably reduced
? because he did not take into account the process of heat emission to the sur-
rounding medium. In connection with this and also as a result of the admitted mathe-
matical error, Parker's formula is not correct.
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STUDY OF FRICTIOI'TAL PROPERTIES OF I AT MIALS ON RIPG
FRICTION i1ACHII:E I-47
chines (Fig. 39) give real values of specific
rExfisting standard friction ma
ssure and sliding speed (criteria pud vsk) but do not take into consideration the
pre
cially
rinciples of heat in model operation's, espe geometric similitude; this re-
p
sults in the fact that the friction temperatures, developed on these machines, are
very limited (175-20(:0C ).
The following experiment shows the extent to which the ratio of friction areas
affects heat conditions and consequently the intensity of wear. Samples of the
friction material Ts-h-52, similar in volume and area of friction but different in
geometric configuration, were tested on a Itminiature" machine and on the friction
hine I-47 with complete reciprocal contact overlap (for description of the machine
mac The
sec below). In both cases, friction was produced on brand ChI11,1h cast iron. ?
speed in both setups was vsk = 3.31 m/sec and the pressure was pud 15 leg/cm .
slidin~ The test was conducted under stabilized temperature conditions. The results of the
experiment are shoim in Table 1.
From the data in Table 1 it is evident that the intensity of wear of the sample
is about 30 times higher on the machine 1-47 than on the "miniature" machine.
Figure 1hO shows samples tested on the "miniature" machine (b) and a ring ma-
chine (a) under the same conditions, i.e., under the same values of pud and vsk'
the first case, the friction surface absolutely was not damaged at all, while in
In
the second case the samples showed intense wear.
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Type of Friction Machine
?
Standard "miniature" machine
I-47 with complete reciprocal
contact overlap
0
Table 1
Ratio of
Friction
Areas
0. lit
0.9
e ?C
/4 00
Intensity of
Wear
iB . 10 ling/m . cm2
13.5
As is generally known, actual aircraft brakes work with nearly complete or suf-
ficiently adequate reciprocal overlap of rubbing pairs. That is the reason why con-
siderably higher temperatures are developed in these actual units than the tempera-
tures developed under the same pud and vsk on "miniature" friction machines. There-
fore, test according to GOST 1786-42 on these machines used to give faulty results,
Fig. 39-Diagram of Standard Friction Machine
1 - Disk; 2 - "Miniature" samples
?
especially for frictional materials which had been intended for forced braking con-
ditions when the temperature reaches 700-800 and even 1000?C (Bibl. 10).
The desire to obtain simultaneously on a friction machine the actual values of
specific pressure, speed of sliding, and temperature is based on insufficient knowl-
121
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ledge of small-scale modeling principles and results in extreme enlargement of the
dimensions of the machine and the power of its drive. In the last analysis, such a
laboratory machine for testing samples would grow to the dimensions of powerful in-
ertia stands.'
Ts-4-52 after Testing.
a - Tested on Ring Machine; b - Tested on ":Miniatures"Machine
Thus, to create a small power laboratory setup it is necessary to single out a
basic parameter which can be considered essential and practically adequate for de-
termining frictional characteristics of tested materials.
We assume temperature (Dibl. 36) as such a basic, controllable parameter.
On the basis of these principles, the special friction machine I-L.7 was con-
structed in the Laboratory of Frictional :,aterials EtASh, Acadeny of Sciences, USSR.,
for testing small samples of experimental types of frictional materials. 'T:hc friction
machine I-L7 is completely subject to the ratios derived above for the case of
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their bases [ see eqs ? (i79),- (183-)p friction of two ring cylinders osculating along
(196)]. this machine simulates actual working conditions sufficiently
,
0
0
Fig. 41 - Design of Working Unit of Friction Machine I-47.
1,2 - Tested Samples, 3 - Torsion shaft, measuring
friction moment. 1~ - Spindle turning the upper sample.
5 - Pipe for supply of air
closely, i.e., it creates an actual thermal field in the samples to be tested.
? ,- ~.
Figure U shows the design of the central assembly for _ the friction machine
I-47.
123
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I-07 is built on the base of a common drill press. The
The friction machine
drive is furnished by a DC motor of 5 kw which, by means of a field rheostat in com-
bination with removable drive pulleys, permits smooth variation of the number of
revolutions of the spindle shaft within very broad limits (100-5000rpm).,`
Pressure of the required magnitude on the test specimens is created by suspend-
ing a load from a lever over a cogvrh.eel gear.
Q gives the values and records the brake moment Produced by the
The oscillograph torque of the calibrated brake shaft. The temperature is recorded by a galvanometer
of one of the two therocouples built into the metal specimen at a depth of 1 mm from
the friction surface.
The test specimens consist of ring cylinders with the standard dimensions of
-2g mm 2R =20mm, h=11-20nun.
As can be seen from Fig. 1,1, the spec
2R2 - 1, 1
ds are ressed together, resulting in complete overlap of the couple or overlap as
en p
required by the particular conditions.
? The lower test sample is stationary and rigidly connected to the torsion shaft,
whose angle of torsion determines the friction moment.
e u er test sample is fixed in a special spindle head which makes it possible
Th pp
to shift the load along the axis of rotation, thus causing the specimen to rotate to-
gether with the spindle.
Since the upper and lower samples have similar dimensions they are interchange-
so that the tester if he wishes, can check the performance of the couple di-
able,
rect or reversed.
It should be specially emphasized here that tests are conducted on the machine
This is
I-h7 of real friction pairs since both members of the pair are removable.
case w?riith a standard friction machine where usually one element of the fric
not .the
tion pair (the metal disk) is permanently fixed.
? the control is accomplished by a
-~ Removable pulleys are not necessary if
Leonard circuit.
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The operating conditions of the actual aircraft brake, for which the test
materials are intended, are usually known. The specific pressure and speed of slid-
ing are also know,- but, most important, the temperatures and the ratio of friction
areas are known. In other words, the reciprocal overlap, under which the tested
S materials should perform, is known.
For this reason, testing conditions are selected on laboratory friction machine
i-47 which correspond to actual heat conditions.
Usually, the tests are conducted under constant specific pressure -rith an over-
lap corresponding to the actual unit, and at various gradually decreasing speeds of
sliding ( from 0.125 to 5 m/sec). i:oreover, the duration of testing at every stage
of speed is such that the heat conditions become practically stabilized.
Thus, each stage of speed gives a definite degree of temperature. Usually, the
duration of an experiment at a given stage of speed is 15 min, since the heat con-
ditions are practically stabilized within 3-h min.
Accordingly, the oscil.lograph readings for the brake moment ( and consequently
? for the friction coefficient) as well as the galvanometer readings for the tempera-
ture are taken every 5, 10, and 15 min after the start of the experiment, for a
given stage of speed. If, at a given stage of temperature, no deformation of materi-
al takes place, then all three readings on the oscillograph tape are identical.
In this way, moving from one thermal stage to a higher one, we reach the temp-
eratures which occur in the actual friction unit.*
Temperatures of the order of 100-600?C are usually reached at sliding speeds of
vsk = 1.6 - 2 m/sec. Higher temperatures (900-1000?C and higher) are formed in the
tested samples at sliding speeds of vsk = 3.5 - 5 m/sec.
,~13efore the basic tests, the specimens are fitted together at a low tempera-
40 ture so that the effect of the high temperatures on material properties is eliminat-
ed when friction takes place. The fitting is done approximately and is considered
complete when the entire friction surface of the sample shows traces of wear.
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0
The resistance of the test pair to wear is rated according to the wear of the
samples in terms of .weight. The specimens are weighed before and after each test.
Knowing the length of the friction path traveled and the friction area of each ele-
ment of the pair, it is easy to compute the intensity of wear JB (mg/m . cm 2 ) as a
function of temperature. Moreover, the linear intensity of wear J is sometimes com-
puted.
To approximate the heat conditions of the experiment conducted with a labora-
tory setup (friction machine 1-47) as closely as possible to the heat conditions of
an actual brake (chamber or disk type), i.e., to get actual thermal gradients in
materials of a pair - artificial electric heating or cooling by compressed air was
used, or else conditions of heat transfer from the samples into the surrounding
medium were varied by means of insulating gaskets and coatings.
All this work was conducted with small test samples to find the best types of
frictional materials for brakes of high-speed jet aircraft. In the course of work
0 on the machine I4.7, there was an opportunity not only to rate materials according
to a given index but to develop suggestions for making the friction materials them-
selves.
Results of Rating Certain Friction laterials on the Friction i Lachine I-l47
in the past, about 70 friction materials of different types have been tested.
Some materials were tested in several different ways.
All test materials can be divided into two basic groups: 1) nonmetallic frict-
ion materials and 2) cerot friction materials. Both these groups were tested in a
pair of alloyed cast iron, type Ch'_"T h, ~.,:hich is used in m~rn'facturing brake drtuns.
Tested nonmetal friction niateria are divided according to their cohesive
median into asbestos-ba::el_t^, ^ sbestos-rubber, and resins. Cornet friction test materials wore made on copper and iron bases.
Plastic To. 22, which had also been tested on machine I-L 7, was the standard
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for all-test materials.
Besides finding the thermostability of friction mcterials, i.e., the dependence-
of the friction coefficient and intensity- of wear on the temperature during these
Fig. 1F2 - Friction Coefficient as a Function of Temperature
for Various"Plastics (the Temperatur- in the Cast Iron was
Measured at a Depth of 1 TMm from the Friction Surface)
1 - Plastic No. 22; 2- Plastic ";;"; 3 - Plastic 6 YKh-l;
4 - Plastic PA-2; 5 - Plastic 6FP; 6 - Plastic 6FS
tests, particular attention bras paid to the quality of the friction surface and to
the interaction of the individual elements of the -air, one another (phenomena of
abrasion, fouling, folding over, etc.).
In the first stage of testing nonmetallic friction materials, attention was di-
rected to the friction material Plastic rrnr, taken from the brakes of an excavator
hoist.
As can be seen from the graphs (Fig. 42 and 1,3 ), the Plastic ~ri;tr retained a
sufficiently high friction coefficient up to temperatures of 800-900?C, at which
point its resistance to wear was considerably higher than the resistance of the
serial friction material No. 22. While the Plastic No. 22, above temperatures of
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600-700?C, showed very high ,.rear and also, as in an actual chamber brake, first be-
came coated with the cast iron (Fig. 141.) and then charred, the Plastic "H" at the
,My101//. CM1
3
1000
600
200
100
60
20
(0
4
2
a)
100 ?00 300 400 500 600 700 800 900 1000 O C
?
?
STAT
Fig. / 3 - ;Dear Intensity as a Function of Temperature, for
Different Plastics (Temperature was asured at a depth
of 1 mm from the Friction Surface) 1 - Plastic 11o. 22;
2 - Plastic 111 "71; 3 - Plastic 6P:Y_h-l; i. - Plastic PA-2;
5 - Plastic FY.-21~a; 6 - Plastic 6F a) Charring and
destruction
same temperatures became somewhat brittle, but did not char or disintegrate. It
was noted that the Plastic trrTt displayed some tendency of becoming coated with the
cast iron.
On this basis, it was possible to recommend the Plastic "H" for actual tests on
an inertia stand in an actual brake.
Tests made on the machine I-47 and on an inertia stand permitted conclusions on
the following principles of manufacturing friction materials:
1) Friction materials should consist of components which change their mechan-
ical properties very little when used within broad temperature limits..
2) Friction materials should contain a component which prevents the appearance
of scratches and fold-over.
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Fig. h/L-Fold-Over of Cast Iron on Plastic No. 22
a),and Damage to a Drake Drum ?:ade of Cast Iron
ChDMKh b),after Testing in a Chamber Drake
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100 200 300 400 500 600 700 800 900 O?r
Fig. L5 - Friction Coefficient f as a Function of Temperature,
for Different Experimental Variations of Ccrmets Coupled
with Cast Iron Chhr1h (Temperature of Cast Iron Sample at a
Depth of 1 mm from the Friction Surface). 1, 2, 3 - Cerm-t
samples; h?- Plastic ITo. 22
Simultaneously with tests of nonmetallic friction materials on the machine I-h7-
tests were conducted with different types of cermets. Very food results (Fig. !45,
curve 1) were obtained by one of the experimental versions of the cermet type
with an iron base.
bases merited less
strength, and some
X
2
X
xX
x
X-
x
3
%
x`x
x
x
x
X
x
x
4 n
'
attention,
experimental versions of cermets with iron or copper
since they showed greater wear and less mechanical
displayed a tendency to link
with the, cast iron. For some cermets
very small linear wear with
pointed out that
r-"ocesseo '?rl~en tiroduc n-; production Ict s.
;`c,.'r typcs of nonmetallic materials 6F am 3 ^ were developed by ,pork 'rs
the transition from pilot-scale
o_`' victual i roducts present _ considerable difficulties.
Therefor C, particular atten 4.on should b~> given to working out technolo-lcal
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negative wear by weight was observed. It should be
for si::1i lap types of materials
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research organizations, in cooperation ?vrith factory workers. In these materials,
which have high resistance in the contact zone during the braking process, a special
?
xxx
0
4
b)
v o
3
xxx
x
l+c x
C
_ x
2 _x
x
x
_
100 200 300 400 500 60
0 70
0 800 90
Fig. 1,6 - Intensity of Wear JB as a Function of the Temperature 0
for Different 1cperimental Versions of Cermets Coupled with
Cast Iron ChM.'~(h (Temperature of Cast Iron Sample at a Depth
of 1 mm from the Friction Surface) 1, 2, 3 - Cermet samples;
L - Plastic 1o. 22 a) Start of considerable wear, charring,
and destruction of sample; b) Start of considerable wear,
abrasion of friction surface, and peeling; c) Smooth friction
surfaces of a pair; d) Smooth friction surfaces of a pair
131
STAT
e' c
0
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friction layer forms which has special properties, namely: increased capacity for
secondary deformation and increase in strength of the material in this layer with
increasing distance from the friction surface. The latter characteristic improves
the quality of the friction surface and practically eliminates fold-over since the
possibility of depth tearing is eliminated. The test results for the friction
material variants 6F (6FP and 6FS) showed considerably better friction properties
than the Plastic io. 22.
of
0,001
0,0008
0,0005
0,0004
0,0002 200 300 600 2 900
0
-0,0002 100 400 500 700 800 1000 eC
_0
0004
-0.0006
-0,0008-
-0,001
0f
Fig. /7 - Temperature Coefficient of Friction, According to
G. I. Troyanovskiy for Plastics 6FP and !o. 22
1 - Plastic 1:o. 22; 2 - Plastic 6FP
The material 6F had a very high friction coefficient even at temperatures of
900-1000?C and high heat resistance. At high temperatures, the wear of the material
6F was less than the ~vrear of the Plastic 1o. 22, and the phenomena of abrasion of the
cast iron surface and its slidin: over the plastic were absent. The material 6F did
not char and was not destroyed. However, it showed an increase in brittleness.
As shown in the graphs (Fig. 47), the material 6FF had a considerably better
temperature friction coefficient than the Plastic No. 22.
This was the reason for reconuncnding the friction material 6F for actual tests.
132 STAT
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showed good results when tested on the machine I-I7 and are recommended for full
scale tests.
veloped which satisfy the principles for making friction materials. These materials
The material "Retinax", which was developed later, belongs to this group.
Lately a number of new experimental cermet friction materials have been de-
The experience of the Laboratory for Friction and Friction Materials I'SASh,
Academy of Sciences, USSR and of the basic research organizations of the country,
working on the development of metal and nonmetal friction materials, makes it poss-
ible to recomment the ring friction machine 1-47 for laboratory rating of various
types of friction materials, in small sample lots.
Project for Methods of Rating Heat Resistance in Friction Materials on the Ring
Friction Machine I-h
Tests were conducted on a machine of the type-!L7, which has two structural
variants for mounting the specimens and which allows samples to be tested:
1) when friction takes place between their bases at a different
ratio of contact surfaces.
2) when friction takes place along the gcneratrix a1 a different
ratio of contact surfaces.
For the first variant the shape of the specimens is shorn in Fig. 48. In the
lower specimen, usually metallic, narrow notches are cut to remove wear particles.
The upper specimen may be provided with notches whose size is determined by the re-
quired ratio of contact surfaces. The Sizes of samples for the first variant coin-
cide w th the dimensions of the inner sample for the second variant (Fig. L.9).
For the second variant, in which the inner sample coincides in dimension with
the specimens of the first variant, the outer specimen is built up of several
separate small blocks; every block is subjected to the sane pressure, produced by a
? hollow rubber chamber.
The loner specimen in 'the first variant and the outer specimen in the second
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i
?
a b
Fig.'i8 - Specimens for Testing on the Ring Friction Machine 1-47
a - Upper specimen; b - Lower specimen
The upper specimen in the first variant and the inner specimen in the second
are mounted to the torsion shaft. The torsion angle of the shaft is used for com-
puting the friction moment and consequently the friction coefficient.
are connected to the revolving spindle.
? In the first variant, the spindle should move freely in the guides and produce
the necessary pressure on the specimen. In the second variant, the spindle has no
free axial movement, and to attain this the spindle is fixed by a special pin. The
pressure on the specimens in the second variant is achieved by changing the inside
pressure in the hollow rubber chamber (pressure controlled by a manometer).
Preparation of Specimens for Testing. For both variants, specimens are cut out
?
of a finished lot of friction products.
For each test, not less than three specimens are cut from three different pro-
ducts of the same lot.
The metal counterpart body is made of the metal actually used in a given brake
Before testing, the specimens are fitted (lapped) at small sliding speeds which
provide for an increase in temperature in the friction unit of not more than
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60-70?C. Lapping is considered completed when the entire friction surface shows
traces of wear.
A thermocouple is built into the stationary specimen, with the cold junction
located at a distance of 1 mm from the friction surface. The safe margin of linear
? wear is up to 0.5 mm, after which the thermocouple should be attached in a new place.
The friction surface of the metal sample is cleaned before lapping.
Technique for Conducting the Experiment
?
Fig. 49 - Specimens for Testing on
the Ring Friction Machine I-47.
1 - Inner specimen (connected
with the revolving spindle);
2 - Outer specimen; 3 - Brake
Chamber; 4 - Outside Head (Con-
nected with the torsion shaft)
To determine the heat resistance in a
friction pair to be tested, a load on a
lever is used which provides specific pres-
sure on the test specimens corresponding to
the actual specific pressure of that fric-
tion unit for which the test pair is in-
tended.
This specific pressure is maintained
constaht during the entire test. Also, the
I
specimens are prepared for testing in such
a way that, in the contact plane (contact
is achieved by having the bases touch) re-
ciprocal contact overlap of the specimens
would be analagous (or close) to recip-
rocal~contact overlap in the actual fric-
tion unit
After this the specimens, which are
under a given pressure, are made to slip
relative to each other.
The sliding speed vsk increases gradually by stages from 0.125 to 5 m/sec, and
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at each stage a definite temperature is formed for the pair under test. Usually the
test is conducted at 7 - 9 speed stages.
The duration of the experiment for each stage is 15 min. Normally, stabiliza-
tion of the heat conditions is reached after three or four minutes. Therefore, the
? oscillograph reading for the friction moment and the galvanometer reading (or oscil-
lograph reading) for temperature are made every 5, 10, and 15 min after starting the
experiment at a given step.
If, at a given temperature, no changes occur in the materials of the friction
pair then all three readings on the oscillograph tape are identical.
Thus, passing from one stage to another, temperatures as they exist in the act-
ual friction unit are obtained. Usually, for many friction materials, temperatures
of 400-600?C are reached at sliding speeds of vsk - 1.6 - 2.11 m/sec. However, in
general the testing of a given friction pair is not stopped at the actual tempera-
ture stage but is continued to determine the limits of friction properties in the
test pair. On the friction machine I-47, testing can be conducted at a temperature
? of 1000?C and higher. Such temperatures are created in the samples when the sliding
speeds reach vsk = 11.0 - 5.0 m/sec.
The wear resistance of a friction test pair is rated according to the wear in
terms of weight of the cooled specimens, which are weighed on analytical scales be-
fore and after each experiment at a given temperature stage. If, in some cases, it
is more convenient to have a linear intensity of wear, several readings must be used
for fixing a mean value for the difference in linear dimensions of specimens, after
they have been in rubbing contact for fifteen minutes at a given temperature stage.
Basic changes (change in color, charring, sparking, fouling etc.) are usually
recorded by the experimenter in his records. The dependence on temperature of both
the friction coefficient and the intensity of wear are found as a mean value during
? the testing of three specimens.
The coefficient of friction f is calculated according to the following formula:
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Al
rP '
where M = measured friction moment, in kg - cm;
S r = mean radius of the specimen (1.2 cm) for the first variant;
where A g = difference in weight, in mg;
for the second variant, r is the outside radius of the inner specimen
(1.4 cm);
p = total pressure on the sample in kg.
The intensity of wear JB is calculated according to the formula-:
JB SL'
0
(209)
(210)
S = area over which rubbing of the sample takes place, in cm2.
It should be noted that, when the ratio of the touching surfaces is not equal to
unity, the sizes of areas by which wear intensity is measured are different for the
upper and lower specimens in the first variant and for the outer and inner-specimens
in the second variant.
L which denotes the friction path (in meters) in the first variant is determined
from the mean diameter of the specimen dmean = 0.021 m, and in the second variant
from the outer diameter of the inner specimen dmean = 0.028 m.
Recommended Testing Conditions. It is necessary to maintain different condi-
tions for friction materials intended for different friction units.
Approximate mean values for these conditions are given in the following Table.
* Intensity of wear may also be found in accordance with the formula J - Ag
r
where AT is the friction work produced at a given temperature stage.
137
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?
?
?
Type of Materials
Specific Pressure
kg/cm2
Temperature
?C
For
aircraft brakes
15 -
18
800
900
For
drill hoists
6 -
10
600
For
excavators type ESh-11./40
7 -
12
320 - 360
For
automobile power brakes
5
250
For
ordinary automobile brake
3
170
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LABORATORY TESTING OF FRICTIONAL PAIRS IN ACTUAL
BRAKES ON AN INERTIA STAND
?
Under nonstationary heat conditions, the rating of friction properties in small
samples cannot give exact results. As a result of this, actual tests of brakes on
an inertia stand operating under definitely nonstationary conditions because of de-
cisive importance. These tests have the drawback in that they require much power
and very powerful and bulky equipment.
Short Description of an Inertia Stand
The underlying principle of tests on an inertia stand is as follows: The fly-
wheel and drum are set in motion and the speed is increased up to the required number
of revolutions; in this manner, the energy of the flywheel masses is, stored. This
energy is equal to the kinetic energy applied to the aircraft brake being tested, and
it is absorbed and dissipated by braking the stand to which the test brake is mounted
until the flywheel masses and drum come to a complete stop with the wheel pressing
against the latter. Figure 50 shows a general view of an inertia stand and control
panel. In the process of braking, any changes in the brake moment, the pressure in
the brake, and the revolutions of the stand are automatically recorded according to
the braking time.
The wheel is pressed against the drum with a force equal to the radial stress
transmitted to the given wheel by the weight of the aircraft when at a standstill.
The compressive force is controlled by contraction of the tire. Pressure is created
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is
?
is
140
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in the test brake (by pneumatic or hydraulic means), with the magnitude of the pres-
sure being so selected that the required brake moment is attained. In existing-cham-
ber brakes, a pressure of PT - 15-40 kg/cm. 2 is
used. Multidisk brakes have pressures
in the cylinders differing little in magnitude from the pressures in a chamber brake,
? while the pressure in a unidisk brake cylinder is considerably higher than the pres-
sure in chamber and multidisk brakes. With some unidisk brakes cylinder, the pres-
sure reaches 180 kg/cm2 and more.
According to standards, the peripheral speed of the stand drum and therefore the
speed of the wheel forced against it by the test brake, at incipient braking, should'
Vper = 0, 8v 05
?
(211)
When testing a brake isolated from a wheel, the shaft revolutions should be e-
qual to the wheel revolutions, guaranteeing a speed equal to 0.8 vpos0
The number of flywheels connected to the stand shaft should be such that, at a
given peripheral speed, the kinetic energy of the flywheels is equal to
AT = 0,03GPST vP05 + U, (212)
where AT - kinetic energy,-standardized for a given brake;
Pst = radial stationary stress on wheel;
Vpos = landing speed of the aircraft for which the given wheel and brake are -
intended;
U - energy loss in the stand due to overcoming friction in the shaft bearings,
aerodynamic resistance to rotation of flywheel masses, and resistance to
rolling of the wheel over the drum.
?
After the stand is stopped, the wheel is disengaged from the drum and cooled to
room temperature. Then the cycle is repeated, i.e., the stand is brought to the re
quired speed, the wheel is pressed to the drum, and the test brake,is applied to
bring the stand to a full stop.
lhi
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The inertia stand (Fig. 51) consists of a shaft, a set of heavy flywheels (2)
and a drum (3) installed on the stand shaft, an electric motor (5) connected to the
end of the stand shaft, a brake (1,.) for stopping the stand in emergency cases, a pres-
sure device (6), a device (1) for testing isolated brakes, and the base.
Basic Data of the Stand: Motor Pn-1320, N = 150 kw, n = 1500 rpm; maximum peri-
pheral speed of drum v 400 km/hr, inertia moments of revolving masses: Jmax - 211
kg/m-sect, Jmin = 28 kg/m-sect, maximum size of tested wheel 850 " 250, Pst = 3800 kg.
The Main Shaft of the stand, carrying the flywheels (2) and the drum (3), is
set on four antifriction bearings, one on each side of the drum and two next to the
flywheels. An electric motor is connected to the right end of the shaft over an
elastic coupling. A brake drum for testing isolated brakes is connected to the left.
Between the stand drum (3) and the flywheels, a special coupling is installed which
makes it possible, when necessary, to disconnect the shaft and thus to disengage all
flywheels at once. The shaft with the system of flywheels, drum and bearings, rests
on supports set on a common solid frame of welded girders connected by anchor bolts
? to the base.
The Flywheels of the Stand are made in the form of solid steel disks of various
thicknesses. Each flywheel can be easily connected to the shaft and, when necessary,
disconnected. When flywheels have to be connected to the shaft, they are bolted to
counterdisks which are rigidly connected to the shaft. Disconnected unused flywheels
rest on special cylindrical footings fastened on supports. The number and weight of
flywheels is selected so that, at a given peripheral speed vokr the required kinetic
to A with an accuracy of
energy can be obtained in a range from Astand min stand max,
+ 5%.
The Drum of the stand (3) consists of a steel cylinder 600 mm wide and 106 mm
in diameter, welded to two ribbed steel
? tightly to the stand shaft.
disks which are connected with a hub, fitted
The Emergency Brake (1.) consists of a double chamber brake with a diameter of
11.2
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?
r
?
990
IT
I V
0
?
1I.3
>~
4-1
?w
0
0
9
bo
co
U
.rl
.rl
H
0
N
U
?rl
U
Co
.rq
4-3
N
U1
C14
u
4.3
u
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of 507 mm and a width of 140 mm, rigidly mounted to a bracket set on the stand base,
and a steel brake drum connected to a flange solidly set on the main shaft. The
weight of the brake drum of the emergency brake is so selected that, when the maximum
energy of the stand is absorbed in emergency braking, its temperature reaches 500?C.
? The maximum brake moment of the emergency brake equals 250,000 kg/cm when pT = 17 `
kg/cm2. The brake stops the stand at a speed of vokr = /L00 km/hr in 6 sec. The e-
mergency brake is switched on from the control panel.
The Electric DC Motor (5) of the stand is fed from a separate motorgenerator
system. With all flywheels connected, the motor brings the machine to the maximum
number of revolutions in 2 min. Starting and stopping and controlling the speed of
the motor is done, through the control panel.
The Pressure Device serves to press the wheel and the tested brake to the stand
drum. It consists of two frames in one horizontal plane. The first of these frames
is hinged to the stand frame; the second is fastened to the cylindrical guides of the
first frame. The second frame can slide along guides of the first frame and so per=
form reciprocating motion under the action of the hydraulic cylinder. A wheel and
its axle are set on the second frame. The wheel axle is rigidly fastened in clamps
to the second frame.
The test brake is set freely on the wheel axle so that, under the action of the
brake moment, it can easily turn around the axle center. A lever is fixed to the.
brake, with its free end resting on the piston of a hydraulic dynamometer which
transfers the pressure along a pipe to a manometer recorder on the control panel. _
The pressure device has a hydraulic system to help press the wheel against the
drum and disengage it. The rate of pressing and the speed of disengaging are constant
and are equal to 0.5 m/sec. The compressive force is regulated by a reduction valv-c
switched into the line between pump and hydraulic cylinder and is controlled by the
? magnitude of radial contraction of the tire pressed to the drum of the wheel. Control
of the hydraulic system of the pressure device is carried out from the control panel.
11+4
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The Setup (1) for Isolated Test of brakes consists of a support to which the
test brake is mounted. The setup has a dynamometer connected by a pipe to a record-
er on the control panel; the free end of the brake lever rests on this dynamometer.
A removable brake drum is connected rigidly to the end of the shaft and revolves to-
0 gether with the shaft.
The Base has a depth of 3.5 m. It is furnished with an iron trussing and anchor
bolts which connect the stand frame to it. To eliminate vibrations, the area of sup-
port and the volume of the base are designed in such a manner that the amplitude of
vertical vibrations of the base is less than its elastic contraction under active
loads. At the same time, the frequency of the base vibrations, when the stand is
operating, is 3Q% higher than the frequency of vibrations induced by the acting loid.
The Control Panel for the stand has one vertical and one slanted panel. On the
vertical panel are the manometer recorder and the manometers which show the air pres-
sure in the storage balloon, reduced pressure, pressure in the test brake and pres-
sure of the hydraulic system of the pressure device. In addition, the vertical panel
? carries an ammeter for the electric drive of the stand, a galvanometer to measure the
temperature of the heated wheel and brake during testing, an electric tachometer show-
ing the peripheral speed of the stand drum, and a system of signal lights used when
the stand is being started.
The slanted panel carries pushbuttons and knobs for controlling the stand, the
electric drive, the hydraulic system of the pressure device, the emergency brake, and
a toggle switch for switching the tape recorder mechanism on and off. Above the ma-
chine-controlknob is a manometer to show the pressure in the emergency brake. The
panels of the control panel are lighted by a fluorescent tube mounted to the top of
the vertical panel. When tests are conducted, an operator standing at the control
panel faces the test object. The control panel is.1300 mm high. This height was
? selected to permit the operator to watch not only the panel instruments but also the
test object.
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Power test on the stand were preceded by selective tests of friction materials
on small samples. These tests had been conducted in the Laboratory for Friction and
Frictional Materials of the Machine Studies Institute, Academy of Sciences, USSR.
? Testing-Conditions on an Inertia Stand
As mentioned above, after preliminary tests on small samples several types of
new experimental frictional materials which have shown better properties when com-
pared to the mass-produced friction plastic No. 22 were recommended for actual brake
tests on the inertia stand. Let us see how the tests of these suggested materials
were conducted.
The characteristics of these experimental frictional materials and mass-produc-
ed materials are shown in Table 2.
Tests were conducted on wheels, size 660 x 160, equipped with two brakes of
basically different design. In one case, the wheel had a unidisk brake with a part-
ially open and ventilated friction surface; in the other case, the wheel had a cham-
ber brake whose friction surface was mostly covered.
The brake disk and brake drum of the wheel were made of ChNMKh cast iron.
The tests were conducted under the following conditions:
Radial stress on wheel pst, kg - 2250:
Pressure in pneumatic tire po, kg/cm2 - 7.5:
Stationary contraction of tire b st, mm - G.O.
When tested on a disk brake, the peripheral speed of the stand drum was vow, _
0.8 vpos = 160 km/hr. This speed corresponded to the sliding speed of the disk brake
pair
d =0 I (213)
,39 Lokh =17,3 M sec . 1
v
? Tests with a chamber brake under calculated conditions were conducted at two
speeds : vokr = 160 Ian/hr and vokr = 300 km/hr; this corresponded to sliding speeds
of 21.1 and 39.6 m/sec, calculated by the formula
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vstc = 0,475 Vokr.
seeds of v = 211, 254, and 3 00 kni/hr? The weight of the brake disk was equal to
p okr
half the calculated weight. Therefore, the kinetic energy of the flywheel masses of
the stand, when tested on a disk brake, were assumed equal to one half of the energy
normal for a wheel of 660 X 160: AT 2 = 125,000 kgm.
Under conditions of different pressures in the brake, tests were conducted at
Table 2
ials
t
M
Type of
Friction
Material
Compos
er
a
Cohesive Some Chara cteristics of the
Medium Specific Heat Con- Heat
ition ductivity Capacity
Weight
kcal/m kcal/kg oC
gm/
hr oC
0.218
No. 22
Asbestos-r
compositio
ubber Rubber 1.87
n and dark
factice
0.52
0.345
6FS
Plastic co
sition
mpo- 3.05
6FP
Plastic Co
sition
mpo- 2.97 1.2 0.327
Tests in a chamber brake were conducted with full kinetic energy AT = 250,00
kg - m.
A pressure PT = const was intermittently applied to the brakes and was maintain-
ed until the flywheel forces of the stand came to a complete stop. At the end of
each individual braking, 15-20 sec after the wheel had been stopped, the temperature
was measured by means of the thermocouple. This was the temperature of the brake
? drum (or brake disk when a disk brake was being tested).
The wear was figured by measuring the thickness of the brake lining before and
l1.7
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braking. Beam compasses served for measuring.
Before the tests, the test lining was fitted to the friction surface of the brake
disk and the friction surface of the drum. The brakes were worked on until the selec-
ted magnitude of pressure in the brake resulted in a calculated magnitude of the
? brake moment MT max , which was recorded at three consecutive brakings.
Considering that the landing speeds of jet aircraft are equal to 50-100 m/sec,
the sliding speeds of a friction pair in wheel brakes of modern jet aircraft are made
equal to vsk = 25-50 m/sec.
The power absorbed by a brake is found from the formula
NT = IPudST r r VPos = IAudSTUstc0
(215) -
and this power characterizes the effectiveness of braking. The magnitude of the pow-
er NT shows how effectively a given brake absorbs the given kinetic energy of an air-
craft. The more kinetic energy is absorbed by a given brake in unit time, i.e., the
becomes, the more in-
higher (other conditions being equal) the mean power NT mean
? tensive will be the braking and the higher the brake efficiency. Equation (215) in-
dicates that the power absorbed by a given brake depends not only on its geometric
dimensions ST, rT, specific pressure pud and rate of sliding vsk, but also on the
magnitude of the friction coefficient f of the pair. Therefore, we may state that
the more power is absorbed by a given brake at a given pud and vsk the higher will be
the efficiency of a given frictional brake pair. The specific powers used in modern
brakes are in the range 30 - 70 kg/cm2 sec.
The design of a brake has a noticeable effect on the performance of the brake
-lining. As shown by experiments, the friction plastics 6FP, 6FS, and 11141' combined
with cast iron ChMNKh and operating under conditions of a unidisk brake with a part-
ially open friction surface, have a lower friction coefficient than when operating
? under conditions of a chamber brake. This is explained by the different heat con-
ditions on the friction surfaces of chamber and disk brakes. This also explains the
148 STAT
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fact that the efficiency of a pair, the stability of its friction coefficient at
braking, and its resistance to wear are different for disk and chamber brakes.
Braking efficiency, stability of friction coefficient, and resistance to wear of
the tested frictional pairs were studied as functions of the sliding speed vsk and
? the pressure pT in a brake whose size, in the case of chamber and unidisk types, was
proportional to the specific pressure pud?
The kinetic energy of the stand in these tests was made constant and equal to the
standardized kinetic energy for a given wheel. Additional tests were made with the
plastic No. 22, combined with cast iron ChNNKh, in a chamber brake. This was done at
calculated MT max and vokrO i.e., at calculated NT and for different magnitudes of
kinetic energy. The dependence of the wear of the brake lining on the value of the
kinetic energy was determined in these tests.
The speed of fitting of brake shoes is an important feature in braking. The
best friction pair, other conditions being equal, is one which requires a minimum a-
mount of braking for a calculated brake moment, under calculated pressure.
? After mounting unidisk and chamber brakes, ten brakings were made for each plas-
tic under calculated conditions which were the same for all tested plastics, i.e., at
the same magnitudes of max., vokr and AT. As long as these ten brakings for every
sample were made at similar MT max and vokr' it can be said that tests were made un-
der similar maximum power
NT. T. rndx - PIT. max Op
where co is the angular velocity of a wheel.
(216)
The mean value of the power NT mean in the process of braking depends'.-on the
value of the mean brake moment. Since various test plastics may have a different
stability of the friction coefficient, the value of the mean specific brake power will
? be different for different materials. Moreover, in every one of the tested plastics _
the value of the absolute maximum friction coefficient differed from the absolute
149
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?
The results of the tests are shown in Table 3: The numerical values of the max-
for each material at pT = 8, 11, 12, 13, 16, 20, 2L., 26, 28, and 30 kg/cm2. Using a
chamber brake, two brakings were made at speeds of vokr = 160, 211., 251, and 300
km/hr with brake pressures of pT = 1~, 8, 10, 12, 15, l8, 20, 22, 25, and 28 kg/cm2.
speed in a unidisk brake at a speed of vokr = 160 km/hr, two braking tests were made
ducted to determine the resistance to wear of each given pair and also to determine
the general comparative evaluation of the friction characteristics of the tested
pairs at various sliding speeds.
To determine the dependence of the braking efficiency and the stability of the
friction coefficient of the tested plastics on the specific pressure and 'sliding
er and greater sliding speeds. All brakings, under calculated conditions, were con-
chamber brake, but at speeds of vokr = 214, 251, and 300 km/hr, i.e., at higher pow-
this, five more brakings were conducted under the same calculated conditions on a
were conducted on disk and chamber brakes, for the plastics 6FP and No. 22. Besides
plastics was not maintained.
Ten brakings for each plastic under calculated conditions of vokr = 160 km/hr
culated brake moment. Consequently, similarity of specific pressures for the tested
the pressure PT in a brake was selected separately for each plastic to secure a cal- -
value of the maximum friction coefficient of other plastics. Therefore the value of'.-
imum calculated brake moment MMT max , of the mean brake moment IL , of the length
mean L of the brake path of the wheel, the time t of braking, and the linear wear 6 of
ing under the calculated conditions shown above, at a speed of vokr = 160 km/hr
when tested on a unidisk brake and at speeds of vokr = 160 and 300 km/hr when using
the shoes for one braking - all were obtained as mean values of ten brakings operat-
.a chamber brake.
Table 3 shows that when the braking is performed under the.same calculated con-
ditions for all tested pairs, at practically equal MT max , AT, and vokr, the disk
brake equipped with linings made of plastics 111111 and 6FS guarantees a shorter brake
150
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STAT
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path for a wheel, compared with a brake equipped with linings made of plastic No. 22
At the same time, to get the calculated brake moment, the first brake requires the
lowest pressure.
A chamber brake equipped with linings made of plastic No. 22, when braking un-
? calculated conditions at a sliding speed of the pair of 21.1 m/sec, also gives a:
der
longer brake path of the wheel than a chamber brake quipped with linings made of the,
experimental plastics 6FS, 6FP, and "M". In a chamber brake with linings made of the
plastics No. 22, the pressure in the brake required for MT max 0-4 at this speed is
no longer lower but higher than in brakes with linings made of the plastics 6FP and
6FS.
?
At a higher sliding speed of the pair equal to 39.6 m/sec, corresponding to a
peripheral speed of vokr = 3 00 Ian/hr according to eq. (214), the chamber brake with
linings made of the plastic No. 22 gives a shorter brake path length than a brake
with linings made of the plastics "M", 6FP, and 6FS. However, the pressure in a
chamber brake with facings made of the plastic No. 22 was double at this speed com-
pared to a brake with the plastics 6FP and 6FS. This is explained by changes in the'
surface of the friction pairs, which take place in the braking process.
Surface Changes of 2?iaterials in Braking
Figure 52 gives a photomicrograph of a polished cross-section specimen of a
brake shoe friction surface made of the plastic 6FP, while Fig. 53 gives the same
for the plastic No. 22.
As can be seen from Fig. 47, the specific feature of the plastic 6FP is a well-
defined film on the friction surface of the shoe. This film prevents the metal of
the brake jacket from folding over on the shoe.
Figure 54 shows a brake drum and a brake with shoes made of the plastic 6FP.
? The friction surfaces of the drum and.shoes have a smooth polished appearance. Fig-
ure 55 shows a brake drum and a brake with shoes made of the plastic No. 22.
151
STAT
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d
.~ O c C CO
t17.V IC'MN^~
MM MMc MM
-- ors
to O -~ 00 tO to M ca
t` :&) ti . 1 CO - Lt Co
Co Loc.: :off
CO aQCO00
M .-4 CO v : 1 :n G'~
t~ ,M In to O CJ 00
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-
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MLnt-M00 ,~ c~ ti W t- sv O 0
c:~ L^:~c~-n
f- MM MMMM.
00 N
4
M
+f
111a4A4
..O 4VTd "t&
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co -4 N
0
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d to ?M -TLOt-
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u1o . bI .
NOOv D ~.,O "MCA"O
?
- ~-+ :V V V to
C\3 :,4 NN.'1N
ds?pnAr = 'o
I,. V
jq-~pq ac4. aM and}e{a~
11~ _
? gas ,~a,IM'fl _
v
N %amod asaoM
uo.}~-a~ ~1~i~ads %02W
~o `e buNVAq co
pua }~ (swp) uanup
a~jvjq ;o a.In Ve.%ad uul
uaiu'p bu:Mvaq aUO .to(
bu-uil4o ).V3 H a~?au11
-J25 t} sdo~s pint}s
1, Uvn bud)JVaq 10 awl
U.ta - bx ? ~'o "e "N
~ao;GH Pnd
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tnC.O cc co IT cc N
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1r.
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CDs nos--o
cc=
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N +e Q ~
~- t0 cC
STAT
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?.
The metal of the brake drum, folded over the shoes, formed a hard crust
bands on the friction surfaces of the_ shoes. - --'
its Eric
During folding over, the friction surface of a pair is destroyed and -
tion qualities deteriorate. When brake shoes are subject to fol ding over,-a partof
?
Fig. 52 - Friction Surface of a Brake Shoe Made of Plastic 6FP
after operation in a Chamber Brake at vokr s 3 00 km/hr. Cross
section of polished specimen, not etched, x 140 (V. M.
Gudchenko)
1 - Woodfs alloy, 2 - Film on shoe surface; 3 - plastic; 4 - Brass.
the friction surface of the brake forms a homogenerous pair: cast iron plus cast iron
with a friction coefficient considerably lower than that of a pair made of plastic
?
lus cast iron. Folding over disrupts the contact of pair elements; this, in turn,
p
153
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decreases the friction force and the uniformity of specific pressure distribution
over the contact surface of the shoes and drum. Therefore in the case of folding
?
?
Fig. 53 - Friction Surface of Brake Shoe made of Plastic No. 22
after Operation in a Chamber Brake. Cross Section of Polished
specimen, not etched, x 200 (V. M. Gudchenko).
1 - Friction surface; 2 - Wood's alloy; 3 - Metal layer, applied
to shoe (white layer not etched). 1, - Crumbled layer of burned
plastic; 5 - Plastic No. 22 of changed composition
over a higher pressure is required in the brake to reach a calculated brake moment.
? When folding over takes place, in addition to deterioration of the frictional
qualities of the pair, burs appear in the form of deep, ring-like, small grooves
STAT
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_(Fig. 55a). Strips of metal on the surface of a shoe do not remain intact. The met--
_al--films, during _brakings, _ is either_detached--in_ places_from the_ friction-surface--of
Fig. 54 - Friction Surfaces after Braking.
a - Brake drum (cast iron ChNMKh), b - Shoes (6FP)
i the shoes, tearing off pieces of plastic as in the case of linking, or metal appears
.in another place. This leads not only to considerable destruction of the brake fric-~
tion surface, but to instability of the brake moment. At the same pressure in the
brake, the brake moment changes its value from one braking to another within rather
wide limits.
It should be noted that a unidisk brake with shoes of plastic No. 22, during
--braking under calculated conditions did not show folding over of the disk metal over
the friction surface of the linings.
When working under calculated conditions, the unidisk brake with linings made of
1
'i
plastics 6FP and "M" showed no folding over of metal either. At all sliding speeds,,'
the friction surface of the- frictional pair in a unidisk brake retained- its,-smooth,--)
clean appearance.
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a chamber brake changes somewhat. In the chamber brake at vsk s 21.1 m/sec,
When working under ci1culated&conditions, the behavior of the plastics'6FP-and -:
linings made of plastic "qq'' showed fold-over of a very thin layer of metal folded
over the friction surface in the form of separate ring-like bands. However, even
? with light folding over the metal, the plastic "Mf when tested in a chamber brake
Fig. 55 - Friction Surfaces after Braking
a - Brake Drum (ChNMKh); b - Shoes (plastic No. 22).
-showed improved frictional properties: In the chamber brake less pressure was re-
__:quired to produce the calculated moment than in a unidisk brake.
Chamber brake linings ?made of plastics 6FP and 6Fs showed no fold over on their!
--friction surface. Their friction surfaces were smooth and polished, with traces of
-?surface film in the form of molten, softened brass. The brake drum, coupled with
.'shoes made of plastics 6FP and 6FS, had a perfectly smooth and polished friction
surface, just as in braking at an initial speed of vow, = 160 km/hr at a correspond-
'ing vsk = 21.1 m/sec, and just as in braking at an initial speed of vow, - 300 kin/hr
?~
---at a corresponding vsk = 39.6 m/sec. On the friction surface of the drum there was
156
STAT
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__a thin smooth layer of brass transferred from the friction surface of the linings.
This folding of brass over the friction surface of the drum-did not result. in .the
_: _ _'
of the lining. On the contrary, this folding over of brass together with other ele-!
-type of destruction usually caused by folding of cast iron over the friction surface
chamber brake than in a unidisk brake, and less than the plastics No. 22 and t1lPt in
chieve a calculated brake moment MT max the plastic 6FP requires less pressure in a
and 6FS increase their friction qualities when operating in a chamber brake. To a-
a F
lubricating and polishing action. Evidently this is the reason why the plastics 6FP
ments of the plastic - barite and brass resins - evidently forms a film which does
not destroy the softened friction surface of the cast iron drum and moreover has a
a chamber brake. In a chamber brake, the plastic 6FS requires considerably less pres+
the plastic No. 22, allow the use of a brake with smaller dimensions but of the same
sure than No. 22 and "Mtt. In other words, the plastics 6FP, 6FS, and "M" compared to
7
force.
Brake Characteristics of a Wheel
Diagrams of brake moments determined in tests on disk and chamber brakes are
shown in Fig. 56, 57, and 58. Each diagram is based on data from ten separate brak-
ings.
S.
completeness of a brake moment diagram characterizes the stability of the
friction coefficient of a pair: the more
ue of brake moment is decreased from its
plastic 6FP has the greatest stability of friction coefficient in the process of
..ditions at sliding speeds of vsk - 0-39 and vokr ' 17.3 m/sec (Fig. 56) show that the
The diagrams of brake moments?of a disk brake in braking under calculated con-
complete the diagram, i.e., the less the valL
maximum in the process of braking, the high-i
._..er will be the stability of the friction coefficient of a brake pair.
I
braking, when combined with cast tiron ChNMKh. The brake moment of a unidisk brake
with linings made of the plastic 6FP shows practically no
change in value all along
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its friction path. The plastics No. 22 and 'Ti" have less stable friction coefficients:
The brake moment decreases at the end of the braking (for the plastic Plo. 22 - 36~ ~'ll
H, 115cm
10000
400 SM
10000
'OD SM
300
100 200 300
400 fm
Fig. 56 - Diagram of Brake Moments of a Disk Brake at a Sliding
Speed of 17.3 m/sec
a) For the plastic 6FP at pT = 16.5 kg/cm ; b) For the Plastic
2
It? at p T= 22 kg/cm2; c) For plastic No. 22 at p T = 13.5 kg/cm .
As a result of the instability of the friction coefficient of a frictional pair,
the brake-path length of a wheel equipped with a brake with plastic "Mn linings and
the brake-path length of a wheel equipped with brakes with plastic No. 22 linings are
longer than the brake-path length of a wheel equipped with a brake with plastic 6FP
linings.
The diagrams of the brake moment of a chamber brake (Fig. 57), equipped with
linings made of the same plastics No. 22, 6FP, ''M", plotted in brake tests at a speed
of vsk = 21.1 m/sec differ in nature from diagrams of unidisk brake moments.
Each curve presented in Fig. 57 has a saddle-like depression (flexure in the
?
middle).
The nature of the instability of the friction coefficient of the plastics No.22
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 Declassified in Part - Sanitized Copy Approved for Release 2013/05/30: CIA-RDP81-01043R002000030004-8
and rNrr in the case of a chamber brake differs from the nature of the instability in
a disk brake. 'The diagram of the brake moment for a unidisk brake with plastic "M"
MIX
30000
20000
10000
my K Crn
3000,"
20000
10000
Mr' Cm
30000
20000
10000
20000
10000
Mr K CM
300001
200
100 200
300
` Sm
100
200
300
400 S~
100
200
300
400 S M
Fig. 57 - Diagrams of Brake Moments of a Chamber Brake at a
Sliding Speed of 21.1 m/sec
a) Plastic 6FP at pT = 11.5 kg/cm2; b) Plastic 6Fs at
PT = 8 kg/cm2; c) Plastic rrj,rn at pT = lh kg/cm2; d) Plastic
No. 22 at pT = 15 kg/cm2.
linings has a hump (rise in the middle) instead of a depression.
The diagrams in Fig. 57 indicate that, for a chamber brake with linings of plas-
tics No. 22, 6FP and r'Mrr, there is no decrease in the value of the brake moment at
the end of braking under calculated conditions. Only a chamber brake equipped with
4) plastic 6FS (this plastic was not tested on a disk brake) has - besides a depression-
a sharp decline of the brake moment at the very end of braking.
159
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 Declassified in Part - Sanitized Copy Approved for Release 2013/05/30: CIA-RDP81-01043R002000030004-8
The divergence of the brake moment value as a result of instability of the fric-
tion coefficient of a pair, represented on the diagram by a depression, considerably
?
MiKgCM
JOAOV
20000
10000
MTKgCm
30000
20000
10000 i`
100
200
300
400 SM
/00 200 300 400 SM
!00 200 300 400 sM
Fig. 58 - Diagrams of Brake Moments of a Chamber Brake at a
Sliding Speed of 39.6 m/sec
a - Plastic No. 22 at pT = 20 kg/cm2; b - Plastic 6FS at
PT 10 kg/cm2; c - Plastic 6FP at pT = 10 kg/cm2
decreases the braking efficiency: the length of the brake path of a wheel equipped
with a chamber brake made of plastics 6FP, ?Wt, and No. 22 producing a depression in
the moment diagram, is longer than the length of the brake path of a wheel equipped
with a disk brake made of the same plastics but not giving a depression in the dia-
gram.
Figure 58 shows diagrams of the brake moment of a chamber brake with linings
made of plastics 6FP, 6FS, and No. 22. These diagrams were plotted during brake
tests under calculated conditions at considerably higher speeds vsk = 39.6 m/sec,
corresponding to vpos = 3 00 km/hr).
The diagram shows that an increase in the initial sliding speed of a pair from
21.1 to 39.6 m/sec considerably affects the stability of the friction coefficient of
160 STAT
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 Declassified in Part - Sanitized Copy Approved for Release 2013/05/30: CIA-RDP81-01043R002000030004-8
these plastics when they operate in combination with cast iron_ChNMKh. All three
plastics produce a more sharply expressed delay in reaching the calculated brake mo-
ment at the start of braking and produce depressions in the middle of braking. At
the very end of braking, all three plastics produce a sharp rise in the value of the
brake moment; the diagram concludes with a peak at the end of the braking path. The
moment diagram for a brake with linings of plastic No. 22 has the sharpest peak.
The peak-like nature of the growth of the moment at the end of braking is dang-
erous because the nature of such a change in the value of the brake moment shows a
tendency to jamming; besides, the nature of such a change in the brake moment during
the process decreases the braking efficiency most. _
The diagrams shown in Fig. 58 indicate that the plastic No. 22 plus the cast
iron ChNMKh react most sharply to an increase of the initial sliding speed, i.e., to
the action of the surface temperature on the friction coefficient of a pair.
MT,x9crrt
15000
0
10000
IllllfiNL3
5000
? 50
100 150 `
vokfKi%lhr
Fig-59 - Dependence of the Brake Moment MT on the Peripheral Speed vokr
for a Disk Brake of a Wheel 660 x 160 B under Different Pressures, Giving
the Same NIX Braking Conditions: A - 125,000 kg-m.
1 - Plastic 6FP at PT = 16.5 kg/cm2; 2 - Plastic "M" at
PT = 22 kg/cm2; 3 - Plastic No.22 at PT = 13.5 kg/cm2
A change in the brake moment as a function of the peripheral speed vokr is
shown in Fig-59 for a disk brake and in Fig.60 and 61 for a chamber brake.
? The brake moment of a unidisk brake, equipped with linings made of the plastic
6FP changes little with speed. This shows that a friction pair consisting of plastic
161
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 Declassified in Part - Sanitized Copy Approved for Release 2013/05/30: CIA-RDP81-01043R002000030004-8
6FP and cast iron ChNMKh, under conditions of a disk brake and in the range of slid-
ing speeds from 160 to 5 km/hr, operates stably and that its friction coefficient de-
pends little on the temperature, produced on the friction surface of a pair. Plastics
'M1' and No. 22, under these conditions, produce a less stable friction coefficient.
At the start of braking, when the sliding speed and consequently the surface
temperature have the highest values, the brake moment of a unidisk brake with linings
made of the plastic "M" has the lowest value. Then, as the speed decreases, the
brake moment gradually increases and reaches a maximum at a speed of 75-80 km/hr.
Then, regardless of a decrease in sliding speed, the brake moment decreases. This
decrease in the moment apparently is the result of a temperature influence during the
second half of the brake path, the temperature in this case being not surface but
volume temperature of the brake disk. The plastic No. 22 shows approximately the
same change in the frictional properties as a function of the speed.
HT,KgCm
30000
20000
/0000
1
2
4
5
0 50 /00 150
UoknK^Vhr
Fig. 60 - Dependence of the Brake Moment MT on vokr for a Chamber Brake
with an Initial Peripheral Speed of 160 km/hr
;
1 - Plastic 6FS at pT = 8 kg/cm2; 2 - Plastic 6FP at pT - 11.5 kg cm 2
3 - Plastic "M" at pT = lL kg/cm2; 4 - Plastic 3FS at pT = 8.5 kg/cm2;
5 - Plastic No. 22 at PT = 15 kg/cm2
The plastic 6FP, 6F3, "M", and No. 22 have a somewhat different nature of change
in the frictional properties as a function of speed when operating in a chamber
brake, combined with cast iron ChN!1Kh. When braking at a speed of 160 km/hr, the
change in the brake moment of a chamber brake equipped with linings of plastics
STAT
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 6FP, 6FS, ttn, and No. 22 differs as follows: At the start of braking, when the speed
is near 160 km/hr, the brake moment decreases; then it begins to increase gradually
and reaches a maximum at the speed of 40-15 km/hr. Later, at the end of braking, the'!_
Declassified in Part - Sanitized Copy Approved for Release 2013/05/30: CIA-RDP81-01043R002000030004-8
value of the brake moment decreases again.,
? The rate at which the greatest decrease in the brake moment during the braking
process takes place and at which the curve; Mr - f (v) forms a depression is not the
same for all tested plastics. Speed has the greatest effect on the value of the {
brake moment when a brake is equipped with the plastic No. 22. This type of brake,
at the start of braking at v - 160 km/hr has a brake moment of 27,000 kg/cm; then
okr
it decreases and, in the depression of the curve at vokr - 70 Ion/hr, reaches a value
of 17,500 kg/cm. Later at vokr - 15 Ian/hr the brake moment is restored and again
reaches the value of 27,000 kg/cm. At the very end of braking, at a speed approaching
zero, the brake moment decreases again from 27,000 kg/cm to 22,500 kg/cm. A chamber
brake with the plastic No. 22 shows fluctuations in the value of the brake moment
during the process of braking which, in this case, are between the limits of + 3 and
? - 33%.
MT,#gem
30000 l i i
200W
1000b
50
?
100
150 200
250 ZO,,KPjhr
Fig. 61 - Dependence of the Brake Moment MT on vokr for a Chamber
Brake with an Initial Peripheral Speed of 300 km/hr
1 - Plastic 6FS at pT - 10 kg/cm2, 2 - Plastic 6FP at
p = T 10 kg/cm2; 3 - Plastic No. 22 at pT = 20 kg/cm2
Chamber brakes with the plastics 6FP and tTtt have a more stable moment. When
braking at a speed of 300 km/hr (Fig. 61), the value of the brake moment of brakes
equipped with plastics 6FS, 6FP, and No. 22 is less than the calculated value at the
163
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is
very start of braking; then it decreases still more and, having reached a minimum at
,a speed of 150 km/hr, begins to increase slowly again. Only near the end of braking,
at a speed of vokr = 15 3 0 kaa/hr, does the brake moment reach its calculated value
and after that drops sharply.
The phenomenon of sharp decrease in the brake moment at the end of braking, just
before the wheel stops, had been noted by us before. It occurs in brakes equipped
with all tested plastics. The explanation should be looked for in the prolonged
(10-15 sec) action of the volume temperature on a given plastic.
The decrease in'the brake moment at high speed (300 - 100 km/hr) apparently can
be explained as the action of high temperatures on the friction surface.
Figure 61 shows that a brake equipped with the plastic No. 22, at nearly all
speeds, has a brake moment less than the brake moment of brakes equipped with plastics
6FP and 6FS. As mentioned before, this is explained by the fact that a brake with
the plastic No. 22, at speeds higher than 200 km/hr, is subject to fold-over which de-
creases the friction coefficient of a brake pair.
Numerous inertia-stand tests on chamber brakes equipped with shoes made of the
plastic No. 22 show: If the specific power absorbed by the brake does not exceed
1L0 kg/cm2.sec, then no fold-over occurs; however, as soon as the value of the specif-
ic power exceeds 40 kg/cm2 . sec, then fold-over appears. The more intensive the
fold-over, the more specific power is absorbed by the brake. These data refer only
to a pair of plastic No. 22 plus cast iron ChNNKh, a chamber brake.
Figure 62 shows the change in the mean and maximum friction coefficients of the
plastics No. 22 and 6FS on cast iron ChNMKh, as a function of the sliding speed.
.These friction coefficients are found by calculation according to the equation
R11,
/T = (217)
P Y STT
In calculating, the brake moment MT was taken from the diagram of moments and
the specific pressure pud was found from the equation
161.
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The peripheral speed vokr for a chamber brake with a wheel 660 x 1608, is re-
lated to the sliding speed of a pair by the equation (214):
vsK = 0,475 Vokr,
When the sliding speed increases, the friction coefficients of the plastics No.
22 and 6FS decrease. The absolute values of the friction coefficient is less for the
P
8K cm:
E8K cm? IOKdcm
200 300
voktr,Kn/hr
Fig. 62 - Change of Mean (Broken Line) and Maximum (solid line)
Coefficients of Friction for the Plastics 6FS and No. 22 over
Cast Iron ChNMKh, as a Function of the Sliding Speed at the
Specific Pressure pud = 8 kg/cm2
1 - Plastic 6FS; 2 - Plastic No. 22
plastic No. 22 than for the plastic 6FS. The curves of changes, in the friction co-
efficient for the plastic 6FS are steeper than for the plastic No. 22. The differ-
ence in values of maximum and mean friction coefficients for the plastic 6FS over the
entire range of speeds is greater than for the plastic No. 22. This shows that the
stability of the frictional properties of the plastic No. 22 is higher. Comparing
these two plastics, it must be remembered, however, that in starting at a speed of
200 km/hr, the plastic No. 22 folds over, introducing an uncertainty in calculating
the friction coefficient value.
The data in Fig. 62 were found at a specific pressure on the friction
165 STAT
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 Declassified in Part - Sanitized Copy Approved for Release 2013/05/30: CIA-RDP81-01043R002000030004-8
,i surface of the brake shoes of pud - 8 kg/cm2. For other values of pud, the nature of-
the curves should not change, since the dependence of the friction coefficient on the
specific pressure is practically. linear. Only the absolute values of the friction
coefficient will change. Figure 63 shows curves for changes in the friction coef-
ficients of the plastics 6FP, 6FS and No. 22 over cast iron ChNMKh, as a function of
the brake chamber pressure.
These curves are the result of chamber brake tests at an initial peripheral
speed of vokr = 300 km/hr.
f
0.5
44
0,3
0,2
41
/5
1-022
20 Aftew
Fig. 63 - Curves for Changes in the Friction Coefficient of the
Plastics 6FP, 6FS, and No. 22 over Cast Iron ChNMKh in a Chamber
Brake, as a Function of the Pressure at the Initial Speed of
Braking v = 3 00 km/hr (Solid Lines f ; Broken Lines finean) max ? okr
The curves show that, in braking at the speed vokr = 3 00 km/hr, the plastic
No. 22 has the highest stability of friction properties. In the plastic No. 22, the
curves of maximum and mean friction coefficients are closest. In this respect, the
plastics 6FS and 6FP are lagging behind the plastic No. 22. The curves of change in
? their friction coefficients are rather far apart and diverge still more as the pres-
sure pT rises. This means that, under higher pressures, the stability of the Eric
166
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 Declassified in Part - Sanitized Copy Approved for Release 2013/05/30: CIA-RDP81-01043R002000030004-8
Results of-tests with the plastics No. 22, 6FP, "M" and 6FS under calculated
conditions show that for stability of frictional properties when used in unidisk and
chamber brakes, the best are the plastics 6FP and 6FS. The plastic "N" is next in
tion coefficient for these plastics will decrease,
stability.
Table 3 gives data on the wear of the plastics No. 22, 6FF, 6FS, and "W'. These
data were obtained in actual tests with unidisk and chamber brakes on a stand, under!
calculated conditions, with brakings at initial speeds of vokr = 160 km/hr and
vokr = 300 km/hr:
creased slightly in wear for one braking. The latter is explained by fold-over of
? insignificantly in wear. The plastic No. 22, when used in a chamber brake even de-
In testing the plastic 6FP in a chamber brake, the results were worse than in
testing it in a unidisk: The wear of chamber brake linings, made of this plastic,
for one braking under calculated conditions, is equal to 0.075 mm, i.e., the wear
increased nearly three times. The plastic "M", used in a chamber brake, increased
with linings made of the plastic No. 22, under similar conditions.
plastic 6FP will be serviceable more than three times longer than a unidisk brake
equal to 0.100 mm. Judging by these data, a unidisk brake, with linings made of the
mm when used in a unidisk brake. The wear of linings made of the plastic No. 22 is
The wear of linings made of the plastic 6FP for one braking is equal to 0.0258
comparative rating of tested frictional pairs with respect to their wear resistance.
? ed by the brake at a given braking. Data for the relative wear are necessary for a
linings, i.e., the wear for one. braking, as it pertains to the specific power absorb-
linear wear for one braking, the Table gives data for the relative linear wear of
necessary thickness of the linings in a given brake. Besides the absolute values of
ings for one braking and the required time of brake service, a designer can find the
Absolute values of linear wear of linings for one braking are given in the Ta-
ble. They are necessary for designers of brakes. Knowing the degree of wear of lin-
STAT
167
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? 4
calculated brake moment was maintained.
Figure 611 shows curves for the variation in the shoe wear d for the above-men-
tioned brakes for one braking, as a function of the coefficient C in eq. (2) of the
kinetic energy
?
AT = CsaMP$TV?pos.
168
..metal on the brake linings, which results in wear of the metal counterbody of a pair;
i.e., the cast iron ChNMKh.
When evaluating the wear resistance of a given plastic, it is necessary to con-
sider not only its own wear resistance but also the extent of wear of the brake drum
or disk. The plastic No. 22, when operating in a chamber brake, although showing
.less wear than in a disk brake, results in a considerable wear of the brake drum.
The plastics 6FP and T"N" produce practically no wear of the drum.
To find the effect of the sliding speed on the wear of brake linings, brake
tests were made at a speed of vokr = 3 00 m/sec with a chamber brake. These tests
showed that at an increase in sliding speed, the wear of the linings increases. How-
ever, despite the fact that the sliding speed was raised rather sharply (from 21.1
to 39.6 m/sec), the wear of the lining increased only slightly.
To obtain a comparative rating of the wear resistance of tested plastics under-
calculated conditions, their relative wear should be compared.
Table 3 shows that the plastic 6FP has the lowest relative wear in a unidisk
brake. In a chamber brake, the plastic III has the lowest relative wear.
Tests made with the plastics No. 22, 6FP, and 6FS at an increased speed showed
that the plastics 6FP and 6FS have the lowest relative wear at a sliding speed of
vsk = 39.6 m/sec.
The effect of the value of the kinetic energy on the wear of brake linings was
studied in chamber brakes of wheels 1260 X 39CB and 660 x 16CB. Braking was perform-
ed at a speed of vokr = 160 - 200 km/hr. The pressure in the brake was such that the
STAT
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 Declassified in Part - Sanitized Copy Approved for Release 2013/05/30: CIA-RDP81-01043R002000030004-8
The curves show that the wear of the brake lining has a-linear dependence on-the --
value of the kinetic energy.
The stability of the friction coefficient and the braking efficiency of the test-
ed plastics 6FP, 6FS, No. 22, and "M" as a function of the pressure PT in the brake -
?
51, w
0,10
1
I
I
405
0.01
0.02
403
0,04 -
? Fig. 61. - wear of Shoes for one Braking as a Function of the 2
Kinetic Energy (Coefficient C in the Formula Astand CPst 2os).
The circles denote wear of linings under smooth pressure on the
brake from p T = 1.5 + 2 kg/cm2 at the initial braking to pT calc
at the end of braking; the triangles denote lining wear at sudden
application of calculated pressure to the brake and keeping it
constant until the end of braking.
1 - Wheel 1260 X 390 B; 2 - Wheel 660 x 160 B
was studied while operating in chamber and unidisk brakes, with brakings at speeds
of 160, 211, 251 and 300 kin/hr.
When we change the value of the pressure in the brake at a given speed, we
? vary the power absorbed in the brake. This power, at a given speed and friction
radius, is proportional to the specific pressure on the sliding surface and to the
169
STAT
1t'
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 Declassified in Part - Sanitized Copy Approved for Release 2013/05/30: CIA-RDP81-01043R002000030004-8
friction coefficient of the frictional pair. The coefficient of friction, in turn,
depends (Fig. 63) on the specific pressure.
specific pressure on the friction surface of a frictional pair, doubly affects the
Consequently, the pressure in a chamber or disk brake, being proportional to the`
value of the power absorbed in the brake; on one hand, the pressure in the brake of
fects the value of the brake power directly, and on the other hand, it affects it in
directly through the friction coefficient of the brake frictional pair.
NgAgm/cm1sec
40
20
/I lee I
667:5% FP
?22
1140" Le
fool,
16 20 24 28 32
PT,xq/crt
Fig. 65 - Specific Power of the Chamber Brake of a Wheel 660 x 160 B,
Equipped with Linings of Different Plastics, as a Function of Pres-
sure in the Brake System. Conditions of braking:
A = 0.036v2 Pst = 250,000 kgm, vow = 160 km/hr.
pos
?
Figure 65 shows curves for the specific power of friction for the tested plas--
the plastics 'q-P' and No. 22. This shows that the frictional efficiency of the plas-
tics combined with cast iron ChNNKh, as a function of the pressure in a chamber
brake. _
The plastics 6FS and 6FP develop the greatest specific power of friction at a
given pressure in a brake. The curves for a change of specific power, as functions
of the brake pressure for the plastics 6FS and 6FP are steeper than the curves for
170
STAT
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 Declassified in Part - Sanitized Copy Approved for Release 2013/05/30: CIA-RDP81-01043R002000030004-8
tics 6FS and 6FP must be higher than the efficiency of the plastics '~fi' and No. 22.
- f (p) have a practically linear nature-
The curves N
T u
Rating of the stability of the friction coefficient and of the braking efficiency
il
f
it.V coe
of a given frictional pair can be done conveniently by means or the at+b
ficient
? i
Rf Max
ac AIT. sp I
(219)
representing the ratio of the maximum brake moment developed by the brake to the mean
I
brake moment as well as by the efficiency factor -
rr~~ a5+ (220)
ell Qkol '
representing the ratio of the stability coefficient to the brake path of the wheel.
Consequently, the most stable coefficient of friction will be obtained by a friction
pair whose coefficient a st equals 1, i.e., MT max and MT sP are equal in magnitude.
The most effective pair is one which has the maximnn coefficient Beff, i.e.,
? one whose wheel-brake path, expressed by the total angle " kol of wheel revolution
during braking, is the smallest.
Data referring to the stability of the friction coefficient and the frictional
efficiency of the plastics 6FP, 6FS, "14", and No. 22, obtained in tests of the above,
plastics in unidisk and chamber brakes under different brake pressures and brakings
at different speeds are shown in Table 1..
Figure 66 shows curves for the change of maximum MT
max and mean M sP brake
moments as functions of the pressure PT in a unidisk brake.
The more closely spaced the curves MT max and M sP are, the higher will be the
stability of the friction coefficient of a given plastic combined with cast iron
ChNMKh in a unidisk brake. When the curve MT max coincides with MT sp, absolute sta-
bility is established.
The curves show that, during operation in a unidisk brake, the plastics 6FP,
171
STAT
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 Declassified in Part - Sanitized Copy Approved for Release 2013/05/30: CIA-RDP81-01043R002000030004-8
i x4
1111
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STAT
4J
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 Declassified in Part - Sanitized Copy Approved for Release 2013/05/30: CIA-RDP81-01043R002000030004-8
?
??MI?~ and No. 22~ with an increase in pressure pT in the brakes i.e., with an increase
in specific pressure on the friction surface, cause a decrease in the stability of the
MT KgCr
friction coefficient; the curves MT
30000
max and MT sp diverge.
A' T Kgon
20000
10000
12
0 5 10 15 20 25 30 35 40
MT Kgcm
30000
20000
10000
0 5
?
2
P1,Kx/cm
30000
20000
1
10000
PTKy/Cm2 l 1 t 1 1 1 1 '---
0 5 10 15 20 25 30 35 40
PT, KgJcnT2 -
Fig. 66 - Dependence of Maximum (Curve i)
and Mean (Curve 2) Brake Moments on the Pres-
sure in Brake System (Disk Experimental Brake`
for a 600 x 160 B Wheel with Brake Linings of
Different Plastics Combined with a Cast Iron
Disk of the ChNMKh Type. Testing Conditions:
A - 125,000 kgm, vokr - 160 km/hr
The plastic No. 22 shows the greatest divergence in the curves 1?L max and T sp up to
pressures of pT = 30 kg/cm2.
The change in the stability of the coefficient a st, as a function of the pres-
sure p in a unidisk brake of 6FP, "M", and No. 22 plastics is shown in Fig. 67.
T
The "M" and 6FP plastics have rather high (0.9) stability of the friction coef-
ficient in a range of pressures pT from 8 to 32 kg/cm ? The coefficient of stability
in these plastics decreases with higher temperatures. The change in the stability
coefficient of the plastic No. 22 takes place differently than in the plastics 6FP
and "M"; starting from the smallest pressures and up to pressures of PT a 20 kg/cm,.
? the stability coefficient of the plastic No. 22 decreases and then depending on the
extent of the increase in pressure PTO the coefficient increases reaching a greater
173
STAT
A~-
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 Declassified in Part - Sanitized Copy Approved for Release 2013/05/30: CIA-RDP81-01043R002000030004-8
value than the plastic 6FP or 111"1 at pressures of pT > 32 kg/cm2?
One can draw the conclusion that, in actual performance, a unidisk brake made of
the plastic No. 22 combined with cast iron ChNMKh under specific pressures of
pud > 20 kg/cm2, and with brakings at speeds of vokr 160 km/hr, has are advantage-
18 20
22
24
26 28 30 32 34
36 38 40
Pr. Ky/Cm2
Fig. 67 - Dependence of the Stability Coefficient on Pressure in the
Brake System (Experimental Disk Brake for a 660 x 160 B Wheel with
Brake Linings of Different Plastics Combined with a Cast Iron
Layer of a ChNI?Kh Disk). Braking conditions : A = 250.00 kgm,
vokr = 160 km/hr.
over the 6FP and I'M plastics in its coefficient of stability.
Figure 68 shows the curves for changes in the efficiency factor Beff of 6FP, 1"W'
and No. 22 plastics, depending on pT when these plastics perform in a unidisk brake.
It is clear from these curves that the functional efficiency of all three plas-
tics grows with an increase in pressure pT from 6 - 12 to 32 - hO kg/cm2. In this
range of pressures pT, the efficiency factor Beff of all three plastics changes
practically linearly. The factor Beff reaches its maximum during the pressures of
PT = 40 - 45 kg/cm2, corresponding to the specific pressures pud = 25 - 28 kg/cm2.
In the range of pressures of pT = 6 - 40 kg/cm2, plastic No. 22 has the highest
174
STAT
ht3
r
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 Declassified in Part - Sanitized Copy Approved for Release 2013/05/30: CIA-RDP81-01043R002000030004-8
coefficient Beff? -In a unidisk brake, the plastic No. 22 is superior to the plastics'
9
e _
6FP and "N", not only in_ stability of the !.friction -coefficient a t,__.but _?in its
_'braking efficiency Beff
The study of the stability in the friction coefficient and of the frictional
?
0..000/8
0,016
0,0014
0,0012
0.0010
0.0008
0,0006
ffN
0 2 4 6 8 10 12 14 16 18 20 ?2 24 26 28 30 32 34 36 38 40 42
Porg/cw
Fig. 68 - Dependence of Brake Efficiency Beff on the Pressure in the
Brake System. Braking conditions: A = 250,000 kgm,
vokr = 160 Ion/hr
efficiency in 6FP, 6FS, limit, and No. 22 plastics was conducted in a chamber brake
at braking speeds of 160, 211, 251, and 300 km/hr. The stability and efficiency
characteristics of 6FP, limit, and No. 22 plastics, during performance in a chamber
brake, differ greatly from the stability and efficiency characteristics of these plas-
tics when performing in a unidisk brake.
In brakings at speeds of 160 km/hr with 6FP, 6FS, "NP", and No. 22 plastics, the
brake moment curves max and MT s (Fig. 69) diverge more sharply depending on in-
'T
? crease of pressure in the brake, than the MT max and 'IT sP curves of these plastics
when performing in a unidisk brake.
175
STAT
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This indicates that, under the conditions of a chamber brake, the stability of
'the friction coefficient in these plastics deteriorates.
the change in the stability coefficient
F
T
T
N'22
2
?
1 4 6
8 10 12 /4 16 18
20 22 24 26 28 3
No. 22 when performing in a chamber brake.
MTKgcA
5X00
?
MTKgCI
50000 -
E
0 2 ! 6 -8 19 12 14 16182 20
PT/roc"? PT 41c *YICM
Figure 70 shows curves for
a st of the plastics 6FP, 6FS, 'EM"T, and
2 4 6 8 10 12 14 16 18 20 22 24 26 28 30
PTKg/Cm=
J9 .'2 14 16 18 20
P_vq/Cm2
Fig. 69 - Dependence of the Maximum (Curve 1) and Mean (Curve 2)
Brake Moments on the Pressure in the Brake1 System of a T 132 Chamber
Brake in a 660 X 160 B Wheel with Brake Linings of Different
Plastics. Braking conditions: A = 250,000 kgm, vokr 160 km/hr
In a unidisk brake, the stability coefficient of the 6FP and "M" plastics in
the range of pressures pT = 6 - 30 kg/c2 maintains its value and, starting at pres-
sures of pT = 30 kg/cm2, decreases. In a chamber brake, the coefficient of the plas-
tic "MTT is stable over the whole range of pressures pT, while the stability coef-
176 STAT
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M"
 Declassified in Part - Sanitized Copy Approved for Release 2013/05/30: CIA-RDP81-01043R002000030004-8
ficient of the plastic 6FP, starting with pressures of PT - 9 kg/cm. , sharply de-
creases with an increase of pressure in the brake. Under conditions of a chamber
oG ST
/,0
?
0,9
0,8
47
0,6
0,5 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30
PT,lfg/rm2
Fig. 70 - Dependence of the Stability of the Friction Coefficient in
a Pair on the Pressure in the Brake System (in a T 132 Chamber
Brake of a 660 X 160 B Wheel with Brake Linings of Different
Plastics). Braking conditions: A = 250,000 kgm, v okr = 160 km/hr
? brake the plastic No. 22 also has a stability coefficient which decreases sharply
with an increase of pressure in the brake.
Any increase in the initial braking speed in a brake lowers even more the sta-
bility of the friction coefficient of all plastics tested, which do not fold over -
metal on the friction surface of the shoes.
Figure 71 shows curves for the change in the brake moments MT max and M
s p
depending on PT during brakings in a chamber brake with a speed of ,.. 3 00 km/hr.
It is clear from Fig. 71 that only the curves for the plastic No. 22 MT max
and M diverge somewhat less. However, this decrease in divergence of the moment
T sP
curves is due to the fold-over of metal on the friction surface of the blocks, and
although the decrease makes the friction coefficient more stable up to a pressure of
? 16 kg/cm2, it is accompanied by a sharp decrease in the absolute value of the fric-
tion coefficient of the pair, reducing the braking efficiency.
177
~
6
N?22
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When braking in a chamber brake with a speed of vokr 300 km/hr, the coef-
ficient of stability st in 6FP and 6FS plastics decreases more sharply, depending
on PT (Fig. 72), than in braking with a speed of vow, 160 km/hr.
As for the plastic No.'22, its stability coefficient during brakings in a chin-_
X10-15
? ber brake with a speed of vokr 300 km/hr in a range of pressures of pT
30000
10000
30000
20000
6 8 /0 12 14 16 18
P, KSlchz
Fig. 71 - Dependence of the Maximum (Curve 1)
and Mean (Curve 2) Brake Moments on the Pres-
sure in the Brake System (T 132 Chamber Brake
of a 660 x 160 B Wheel with Brake linings of
Different Plastics). Braking conditions:
A = 250, 000 kgm, v okr = 299 km/hr
The explanation for the nature of such changes in the value of the stability coef-
ficient for the plastic No. 22 lies in the effect of fold-over, which appears in
plastics under these braking conditions, on the friction coefficient of the pair.
The magnitude of the braking efficiency Beff changed when working in a chamber
brake. The slope of the curves for braking in a chamber brake at a speed of _
v = 160 km/hr (Fig. 73), as compared with the slope of these curves when braking
okr
at the same speed in a unidisk brake (Fig. 68), showed a decrease. The dependence _
of the change in the efficiency factor Beff on the pressure pT in a chamber brake
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kg/cm2 increases sharply; on further increase in the pressure
 Declassified in Part - Sanitized Copy Approved for Release 2013/05/30: CIA-RDP81-01043R002000030004-8
,produced a clearly expressed nonlinearity.e In addition, the plastic No. 22, during 3
operation in a. chamber brake, decreased the value of its efficiency-coefficient for--!
the whole range of pressures PT as compared with the values of the efficiency coef-
ficient (Figs. 73 and 68) produced in this' plastic during operation in a unidisk
f `1
- brake.
aST
0,9
0,8
0.7
0,6
0,5 2 4 6 8 10 12 14 16 18 20 22 24
OTKq/Cm2
Fig. 72 - Dependence of the Stability Coefficient of the Friction
Pair on the Pressure in the Brake System (T 132 Chamber Brake in
a 660 x 160 B Wheel with Brake Linings of Different Plastics).
Braking Conditions: A 25,000 kgm, vokr 299 Ian/hr
An increase in the initial braking speed considerably lowered the efficiency of,
the brake with linings of 6FP and No. 22 plastics. In braking at a speed vokr = 3001
k