TELKOM
NIKA
, Vol. 11, No. 8, August 2013, pp. 44
2
2
~4
432
e-ISSN: 2087
-278X
4422
Re
cei
v
ed Ma
rch 7, 2
013;
Re
vised
Ma
y 13, 2013; Accepted Ma
y 22
, 2013
Vibration Analysis of the Steam Turbine Shafting
caused by Steam Flow
Ge Li-juan, Z
h
ang Ch
un-hui, Hao Min, Zhang Yong*
Inner Mon
g
o
lia
Agricultur
al U
n
iversit
y
, 30
6 Z
hao
w
u
d
a
Ro
ad
. Hohhot, Inner
Mongo
lia. P.R
.
Chin
a, 010
018
,
047
1-43
09
21
5
*Corres
p
o
ndi
n
g
author, e-ma
i
l
: gege
glj
@
ya
h
oo.com.cn, 28
972
99
30@
qq.c
o
m*
A
b
st
r
a
ct
T
h
is thesis an
aly
z
e
s
vi
bratio
n test signa
l of
T
U
OKET
UO Pow
e
r Plant 60
0
MW
steam turbin
e un
i
t
throug
h vi
brati
on
monitor
i
n
g
and
sig
n
a
l
a
n
a
l
ysis o
n
th
e b
a
s
i
s of T
N
80
00 S
t
eam T
u
rb
ine
Vibrati
on A
n
a
l
y
s
is
Softw
are. F
ault characteristic
w
h
ich is raise
d
by St
ea
m F
l
ow
Excitations
is reprod
uce
d
by accel
e
ratio
n
constant sp
ee
d
test and lo
ad t
e
st. Steam flo
w
mecha
n
is
m
of excitatio
n
ca
used
by vibr
ati
on fau
l
t and fa
ult-
sensitiv
e para
m
eter
are an
al
y
z
e
d
,
measur
e
s
reduci
ng u
n
it
vibratio
n
has
bee
n pro
pose
d
in lin
e w
i
th the
cond
itions. T
e
s
t
results show
that:
the vibrati
on caus
ed
by the vap
o
r stre
a
m
excit
a
tion
oc
curs main
ly in t
h
e
hig
h
-press
ure
rotor stea
m in
le
t end.
How
e
ve
r, the vi
br
atio
n
sig
nal,
w
h
ic
h occup
i
es a lar
ge perce
ntag
e
o
f
the rotor frequ
ency of a first critical sp
ee
d a
r
e sensit
iv
e to
the cha
nges
in
the loa
d
. Prob
le
ms can
be e
a
rl
y
ide
n
tified;
the
ma
inte
nanc
e
p
r
ogra
m
an
d
mainte
na
nce
me
ans c
a
n
be
d
e
t
ermi
ne
d i
n
th
e p
l
ant
op
erati
o
n
throug
h an
alys
is of vibratio
n mec
h
a
n
is
m an
d sign. Se
c
u
rit
y
and rel
i
a
b
il
ity of the steam turbine ru
nn
in
g
shou
ld b
e
guar
antee
d.
Ke
y
w
ords
: steam tur
b
in
e unit
,
shafting vibrat
ion, sig
nal
ana
l
ysis, steam flo
w
excited vibra
t
ion
Copy
right
©
2013 Un
ive
r
sita
s Ah
mad
Dah
l
an
. All rig
h
t
s r
ese
rved
.
1. Introduc
tion
The turbine
rotor is a rath
er
compli
cate
d structu
r
e, p
o
sse
ssi
ng a
contin
uou
s pl
aster of
mass di
strib
u
t
ion. Due
to such
re
ason
s
as the
man
u
facture, in
stall
a
tion an
d op
eration,
all m
a
y
make b
endi
n
g
vibration an
d torsio
nal vibration
ca
u
s
e
d
by the shaf
t vibration. Vibration p
r
o
b
lem
at No.1
and
2 bea
rin
g
of
No. 6
po
we
r
unit in T
U
OK
ETUO Po
we
r Plant o
c
curs duri
ng tu
rbin
e
operation. Sh
aft vibration v
a
lue flu
c
tuate
s
fro
m
4
0
µm
～
100µm
wh
e
n
the l
oad i
s
more
than
55
0
MW (Figu
r
e 1).
Waveform
of
accid
ent can be
see
n
obviou
s
ly, the re
ason
wh
y shaft vibrat
ion
occur l
a
rge fl
uctuatio
ns li
e
s
in the viol
e
n
t ch
a
nge
of low-f
r
eq
uen
cy compo
nent
s of the vibra
t
io
n
sign
al at an a
pproxim
ately 28.15
Hz
Figure 1(a
)
. Time Dom
a
in
Waveform a
nd
Spectrum of 1X Shaft Vibration
Figure 1(b
)
. Time Dom
a
in
Waveform a
nd
Spectrum of 1Y Shaft Vibration
Evaluation Warning : The document was created with Spire.PDF for Python.
TELKOM
NIKA
e-ISSN:
2087
-278X
Vibration Ana
l
ysi
s of the Steam
Turbine
S
hafting cau
s
ed by Steam
Flow (Ge Li-j
uan)
4423
Figure 1(c). Time Dom
a
in
Waveform a
nd
Spectrum of 1X Shaft Vibration
Figure 1(d
)
. Time Dom
a
in
Waveform a
nd
Spectrum of 1Y Shaft Vibration
2. The Relev
a
nt Te
chnica
l Parameter
of Units and
Testing Sy
stem
2.1. The Rele
v
a
nt Technical Parameter
of Units
The fault uni
t is design
e
d
and manufa
c
ture
d by DONG
FANG
steam turbine
plants,
whi
c
h i
s
sub
c
ritical, impul
sive, axial with triplex Fo
ur exha
ust,
the dou
ble
back p
r
e
s
sure
con
den
sing
steam turbin
e and an inte
rmediate
rehe
at modeled
N60
0
-1
6.7/53
8/538. The hi
gh
and mediu
m
pre
s
sure cylinder
po
sse
s
ses co
-cylinde
r, two
-
tier st
ructure
and
th
e lo
w-p
r
e
s
su
re
cylinde
r i
s
di
vided into A,
B, two-cylin
der. Tu
rbi
ne
totals three
whol
e forging
solid
roto
r,
each
sup
porte
d by
two b
eari
ng.
Of whi
c
h,
NO
.1, 2 bea
ri
ng
are tilting
pad
bea
ring
s
with six tilting p
ad
and NO. 3, 4, 5, 6, 7, 8bearing a
r
e elli
ptical bea
rin
g
s
. Thru
st pad
is on the back of exhau
st
o
f
the intermedi
ate pressu
re
and
wo
rki
ng f
a
ce
on th
e
si
de of th
e ge
n
e
rato
r. critical
sp
eed val
ue
of
steam
turbi
n
e rotors
are
sho
w
n
in
Tab
l
e 1 [1]. Ve
rti
c
al
(Y) an
d h
o
rizontal
(x)
doubl
e am
plitude
vibration valu
e mea
s
u
r
ed i
n
any jou
r
ne
y of the stea
m turbin
e uni
t shall n
o
t be
over 0.07
6m
m.
Bearin
g Vibra
t
ion (w) pea
k
shall n
o
t be g
r
eate
r
than 0.
05mm [1].
2.2. Testing
Sy
stem
The main
su
p
e
rvisio
n chart
of steam turbine
is
sh
own
in Figure 2. The monito
ri
ng ch
art
is sh
own in Figure 3 due
to chara
c
te
ristics of
axis oscill
ation. The
sy
st
e
m
is com
p
o
s
ed
of
:
Steam Turbi
n
e Shafting, se
nso
r
, TSI instrume
nt (MMS
6000), T
N
80
00 syste
m
, compute
r
, etc.
Table 1. The
Critical
Speed
of
t
he Turbine Rotato
r
Name of shaft se
ction
First critical spee
d (r/min)
The second critical speed(r/min)
High and mediu
m
pressure rot
a
t
o
r
1692
4000
low
p
r
essure rot
a
tor A
1670
4000
low
p
r
essure rot
a
tor B
1697
4000
Gene
rator rotato
r
933
2691
2.3. The Selection of Sen
s
or
In orde
r to accurately me
asu
r
e the re
al-tim
e situ
ation of turbine
rotor vibrati
on, the
sele
ction of
the sen
s
o
r
should
con
s
id
er t
he follo
wing two a
s
p
e
cts, on the
one han
d, the
cha
r
a
c
teri
stics of the measured
si
gnal,
on the other hand, the perform
an
ce o
f
the senso
r
[2].
PR642
3 di
spl
a
cem
ent sen
s
or
mad
e
by
EPRO with
ra
nge 4
00µm a
nd sen
s
itivity
8mv/µm is u
s
ed
for p
r
ob
e
of
shaft vib
r
atio
n an
d P
R
92
6
8
spee
d
sen
s
or m
ade
by
EPRO
with
range
10
0µm
and
sen
s
itivity 28.5mv/µm for t
he tile
cap
vibratio
n
(W) throu
gh
com
p
aring
the
ch
a
r
acte
ri
stics a
nd
function
s of the oscillation
sen
s
o
r
and
combin
in
g the feature
s
of the turbine
rota
tors.
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e-ISSN: 2
087-278X
TELKOM
NIKA
Vol. 11, No
. 8, August 2013: 4422 –
4432
4424
3. Test Anal
y
s
is of Vibration Signals
Acco
rdi
ng to
the failure
chara
c
te
risti
c
s, t
he cau
s
e
of the malfu
n
ction i
s
ex
ci
tation of
steam
flow at
the first tho
u
ght. Accele
ra
tion con
s
tant
spe
ed te
st a
n
d
the
load
te
st a
r
e
ope
rat
e
d
in orde
r to verify the accuracy of the jud
g
ment
and to
find the sou
r
ce of the fault resp
ectively.
3.1. The
Ac
c
e
leration
Pr
ocess
o
f
Sta
r
t M
ach
ine
a
nd th
e Vibra
t
ion o
f
the
Constant Spe
e
d
Test
Wavefo
rm of
dire
ction
sh
aft vibration
wh
en Unit sta
r
ts up a
nd a
c
cel
e
rate
s fro
m
No. 1 to
No. 8
watts X
(h
ori
z
ontal
),
Y (verti
cal) i
s
sh
own in
Fig
u
re
4
with th
e
ab
sci
ssa
in t
he figu
re
bei
n
g
the rotator
sp
eed, the verti
c
al axis
bein
g
the
Shaft vibration
amplit
ude value a
n
d
pha
se. From
the Shaft vibration wavefo
rm, the entire
accele
ration
pro
c
e
ss
ca
n be se
en fro
m
con
s
tant spe
ed
to the
rated
spe
ed
300
0r/
m
in
with e
a
ch vibration
b
e
ing
sm
all.
Whe
n
p
a
ssin
g throug
h th
eir
critical spe
e
d
(ea
c
h
rotor orde
r criti
c
al spe
ed
i
s
sho
w
n i
n
Ta
ble1
), the p
e
a
k
v
a
lue i
s
small
(the
critical sp
eed
at a vibration amplitud
e
are s
hown
in Table 2
)
within the all
o
wa
ble ra
ng
e,
indicating tha
t
each tile vibration p
r
op
erl
y
.
Table. 2 Vibration Peak V
a
lue of Each
Bearin
g wh
en
Raisi
ng Spe
ed Over a
Cri
t
ical Value
Shaft s
y
stem of
steam turbine
PR6423 displacement transducer
PR6423 speed t
r
ansducer
TSI instruments
(MMS 6000)
TN8000 s
y
st
ems
computers
Fi
g
ure 3. The
Testin
g
S
y
st
em
Figure 2. The
Main
Monitor Diagram
Speed
dire
ct
ion
NO.1 bearin
g NO.
2 b
earing NO
.3 bearin
g
NO.4 beari
ng NO
.5
bearing N
O
.6 beari
n
g
NO.7 bea
r
ing
NO.
8 bearing
NO.9 be
ari
ng
Eccen
tr
ic
ity
A
x
ial-
disp
lac
e
m
e
nt
Low pres
sur
e
expans
ion
dif
f
er
enc
e
X
:
vi
br
at
i
o
n o
f
X
-
b
e
a
r
in
g
Y: v
i
bra
tion
of Y-
beari
n
g
W: v
i
br
at
ion o
f
ti
le c
a
p
V
i
bra
t
i
on un
it
:
μ
m
K
e
y
phas
e
Hig
h
press
u
re
expans
ion
dif
f
er
enc
e
A
b
solu
te
expans
ion
Shaft
vibration
1X 1
Y
2X 2
Y
3X
3
Y
4X 4
Y
5X
µm
27 31
24 32 20
39 22 32
18
Evaluation Warning : The document was created with Spire.PDF for Python.
TELKOM
NIKA
e-ISSN:
2087
-278X
Vibration Ana
l
ysi
s of the Steam
Turbine
S
hafting cau
s
ed by Steam
Flow (Ge Li-j
uan)
4425
Phase
Ampl
itu
de
Phase
Ampl
itu
de
Phase
Ampl
itu
de
Phase
Ampl
itu
de
Figure 4(a
)
. Wavefo
rm of shaft vibratio
n of No.1 and
2 along x-y directio
n
Figure 4(b
)
. Wavefo
rm of shaft vibratio
n of No.3 and
4 along x-y directio
n
Phase
Ampl
itu
de
Phase
Ampl
itu
de
Phase
Ampl
itu
de
Phase
Ampl
itu
de
Figure 4(c).
Wavefo
rm of shaft vibratio
n of No.5 and
6 along x-y directio
n
Phase A
m
p
l
i
t
ude
（
）
（
°
μ
m
）
Phase A
m
p
l
i
t
ude
（
）
（
°
μ
m
）
Phase
Ampl
itu
de
Phase
Ampl
itu
de
Figure 4(d
)
. Wavefo
rm of shaft vibratio
n of No.5 and
6 along x-y directio
n
Phase A
m
p
l
i
t
ude
（
）
（
°
μ
m
）
Phase A
m
p
l
i
t
ude
（
）
（
°
μ
m
）
Phase A
m
p
l
i
t
ude
（
）
（
°
μ
m
）
Phase A
m
p
l
i
t
ude
（
）
（
°
μ
m
）
Evaluation Warning : The document was created with Spire.PDF for Python.
e-ISSN: 2
087-278X
TELKOM
NIKA
Vol. 11, No
. 8, August 2013: 4422 –
4432
4426
3.2 Vibration
w
i
th L
o
ad
With a loa
d
, watt-axis vib
r
ation ha
s so
me ch
ang
es
to some
deg
ree. On
e of the mo
st
intere
sting thi
ngs i
s
NO. 2
and
NO. 3
-
watt of the high
and m
edium
-pre
ssure rot
o
r a
s
well a
s
NO.
7 and
NO.
8-watt vibratio
n
of the rot
o
r
of gene
rato
r.
There i
s
a b
i
g fluctuatio
n
in the first an
d
se
con
d
Z
W
Z
vibration
whe
n
full ca
pa
city is 600
MW.
There is a m
u
ch
more in
crea
se in
No. 7,
No. 8 ZWZ
vibration tha
n
that withou
t load.
Given the spee
d and load
wat
t
vibration-p
a
ss
freque
ncy dat
a are sho
w
n i
n
Table 3. we
will analyze in the followin
g
para
g
raph
s.
Table. 3 Vibration Peak V
a
lue of Each
Bearin
g wh
en
Raisi
ng Spe
ed Over a
Cri
t
ical Value
Shaft vibration
Constant speed
（
3000 r/min
）
100MW 300MW 450MW 600MW
1x
40
38 40 40
24~126
1
y
40
34 40 36
22~132
2x
25
27 28 30
30~124
2
y
30
30 35 36
40~149
3x
43
46 44 48 58
3
y
52
54 43 37 55
4x
28
28 32 36 34
4
y
35
32 39 47 39
5x
43
33 36 34 46
5
y
50
39 36 30 49
6x
16
18 25 28 28
6
y
20
22 38 38 33
7x
32
32 30 32 61
7
y
57
58 59 66 76
8x
12
18 22 22 28
8
y
40
45 47 51 57
9x
41
45 58 56 60
9
y
43
50 55 56 64
3.2.1. The Analy
s
is of the Vibration for the High a
nd Medium-pressu
re Ro
tor
Whe
n
No. 1
and
No.
2 th
e Z
W
Z vib
r
at
ions are 6
00
MW, the
r
e
are large flu
c
tu
ations.
Take
2Y a
s
an exampl
e, the minimum
value is o
n
l
y
about 40
μ
m and the
m
a
ximum valu
e of
more
than
12
0
μ
m. The
vib
r
ation
sp
ect
r
u
m
s of th
re
e t
y
pical
situatio
ns
are
sho
w
n
from Fi
gu
re
5
to Figure 7.
Figure 5 is a
vibration sp
ectru
m
witho
u
t
low-frequ
e
n
t comp
onen
ts, of which t
he 1X
vibration
-
pa
ss fre
que
ncy
value is
abo
ut 24
μ
m a
n
d
one
octave
amplitude
of
about 1
6
μ
m. The
sign
al amplit
ude of 28.13
Hz i
s
about 2
μ
m. Pass-f
r
e
quent value o
f
1Y vibration is about 22
μ
m,
one o
c
tave a
m
plitude of 1
4
μ
m and si
g
nal amplitud
e
of 28.13Hz is app
roximat
e
ly 3
μ
m, pass
-
freque
nt value of 2X vibration is abo
ut 30
μ
m, one octave am
plitude of 21
μ
m and sign
al
amplitude
of
28.13
Hz i
s
a
pproxim
ately 4
μ
m. Pass-f
reque
nt value
of 2Y vibratio
n is
abo
ut 40
μ
m,
one o
c
tave a
m
plitude of 3
3
μ
m and
sig
nal amplitu
d
e
of 28.13Hz i
s
ap
proximat
ely 4
μ
m, and
the
NO. 1 and
NO. 2 watt vibration focu
s o
n
octave ing
r
edient
s.
Figure
6 i
s
a
vibration
spe
c
trum
with
evident lo
w-freq
uen
cy compo
nents of which th
e
1X vibration-pass freq
uen
cy value is ab
out 28
μ
m an
d one octave
amplitude of about 16
μ
m. The
sign
al amplit
ude of 28.13
Hz i
s
about 1
1
μ
m pass-fre
quent value
of 1Y vibration is about 35
μ
m,
one o
c
tave a
m
plitude of 1
5
μ
m and
sig
nal amplitud
e
of 28.13Hz i
s
app
roximat
e
ly 16
μ
m,
pa
ss-
freque
nt value of 2X vibration is abo
ut 35
μ
m, one octave am
plitude of 21
μ
m and sign
al
amplitude
of
28.13
Hz is a
pproxim
ately
16
μ
m p
a
ss-freque
nt value
of 2Y vibratio
n is ab
out 6
0
μ
m,
one o
c
tave a
m
plitude of 3
3
μ
m and
sig
nal amplitu
d
e
of 28.13Hz i
s
ap
proximat
ely 21
μ
m. At this
time, the low-freque
ncy co
mpone
nts a
r
e
close to or o
v
er one o
c
tave comp
one
nt.
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TELKOM
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e-ISSN:
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Vibration Ana
l
ysi
s of the Steam
Turbine
S
hafting cau
s
ed by Steam
Flow (Ge Li-j
uan)
4427
Figure 5. Vibration sp
ectru
m
di
agram of No.1 an
d 2 b
earin
g
when lo
w fre
quen
cy com
p
onent is
small
Figure 6. Vibration sp
ectru
m
di
agram of No.1 an
d 2 b
earin
g
when lo
w fre
quen
cy com
p
onent is o
b
vious
Figure 7. Vibration sp
ectru
m
di
agram of No.1 an
d 2 b
earin
g
when lo
w fre
quen
cy com
p
onent is g
r
eat
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TELKOM
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Vol. 11, No
. 8, August 2013: 4422 –
4432
4428
Figure 7 is a
vibration sp
ectru
m
with
evi
dent low-freque
ncy com
pone
nts. of whi
c
h the
1X vibration-pass freq
uen
cy value is ab
out 70
μ
m an
d one octave
amplitude of about 16
μ
m. The
sign
al amplit
ude of 2
8
.13
H
Z
is a
bout 4
7
μ
m. Pass-freque
nt value
of 1Y vibratio
n is a
bout 9
0
μ
m,
one o
c
tave a
m
plitude of 1
7
μ
m and
sig
nal amplitud
e
of 28.13Hz i
s
app
roximat
e
ly 62
μ
m,
pa
ss-
freque
nt value of 2X vibration is abo
ut 95
μ
m, one octave am
plitude of 22
μ
m and sign
al
amplitude of
28.13Hz i
s
approximatel
y 70
μ
m. Pass-frequ
ent value of 2Y vibration i
s
ab
out
120
μ
m, one
octave amplit
ude of
3
2
μ
m
and
sig
nal
a
m
plitude
of 2
8
.13Hz i
s
ap
proximately
8
4
μ
m.
At this time, the low-fre
que
ncy co
mpon
e
n
ts ARE far
much m
o
re th
an one o
c
tav
e
comp
one
nt.
Thro
ugh
spe
c
tral analy
s
is,
we
can find t
hat the
re
aso
n
why the 1
s
t
and the 2
nd
tile of X,
Y-dire
ction
shaft vibration
occu
rs large
fluctuation
s
l
i
es in
violent
vibration
of a
n
app
roximat
e
ly
28Hz lo
w fre
quen
cy com
p
onent
s. As you ca
n se
e,
however 1, 2
corrug
ated frequ
en
cy value
cha
nge
s, the
1 o
c
tave
am
plitude
and
p
hase a
r
e ve
ry stable
with
few
ch
ang
es (Ta
b
le
4). T
he
axis tra
c
k of
the vibration
(Figu
r
e
8) i
s
in di
sord
er,
whi
c
h i
s
cau
s
ed
by the
complexity of the
freque
ncy co
mpone
nt of the vibration si
gnal [3, 4], prece
s
sion di
re
ction is p
o
siti
ve
.
Table 4. Amp
litude and Ph
ase of 1 O
c
ta
ve Vibration
Signal of No. 1 and 2 Bea
r
i
n
g
Test
point
1x 1
y
2x 2
y
Amplitude(µm)
16 15
21 33
Phase
（
°
）
161 49
196 81
The
ba
sic ch
ara
c
teri
stics of
malfunctio
n
are a
s
follo
ws from the a
bove analy
s
is:
(1) Malfu
n
cti
on occu
rs to
the inlet steam end
of
high pressu
re turbine, be
cau
s
e vibration
amplitude flu
c
tuation
s
very instantly, th
ere will b
e
a severe vibratio
n.
(2)
Whe
n
P is no less tha
n
500M
W, vibration is inte
n
s
ified. Once th
ere is a threshold load val
ue,
this malfun
ction occu
rs.
(3) In th
e vib
r
ation
sign
al, t
he first
criti
c
al
sp
eed
(f
req
u
ency) of th
e
signal
of the
m
a
lfunctio
n
rot
o
r
is the mai
n
compon
ent of
and am
plitud
e occu
rs
vol
a
tility, accom
panie
d
by a l
a
rge
num
ber of
low-f
r
eq
uen
cy compon
ent and hig
h
-o
rd
er ha
rmoni
c compon
ents.
(4) Failure occurs,
even if t
he fre
quency
of vibration fluctuates vo
l
a
tility, but the
basi
c
frequency
in vibration si
gnal re
main
s
stable.
(5) T
he fre
q
u
ency comp
on
ents of the vibration m
a
lfu
n
ction a
r
e a
b
unda
nt with volatile vibrati
o
n
amplitude; th
erefo
r
e, orbit
track is in
di
sorde
r
in po
siti
ve precessio
n
.
(6) T
he vibrat
ions have
a fine reproducibility.
Figure 8. Vibration Axis pat
h diagram of No.1 be
arin
g
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TELKOM
NIKA
e-ISSN:
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-278X
Vibration Ana
l
ysi
s of the Steam
Turbine
S
hafting cau
s
ed by Steam
Flow (Ge Li-j
uan)
4429
3.2.2. The Analy
s
is for Reason o
f
Vibration in the
High and Me
dium-pres
s
u
r
e Rotators
Acco
rdi
ng to
analysi
s
of th
e me
ch
anism
of rotor dyna
mics , u
n
it vibration
ph
en
omeno
n
is clo
s
ely rel
a
ted with critical sp
eed freque
ncy
of the high pressure roto
r, which me
an
s the
vibration freq
uen
cy doe
s n
o
t match
with
the roto
r
freq
uen
cy, but co
mplies
with th
e criti
c
al
spe
e
d
of the rotor
and contain
s
a low-frequ
ency ha
rmo
n
i
c [5, 6, 7, 8, 9]. There a
r
e othe
r fact
ors
cau
s
in
g in
sta
b
le
spo
r
ts vib
r
ation f
r
eq
ue
ncy 2
8
.13
H
z
is
clo
s
e to
cri
t
ical
spe
ed 1
692r/mi
n
of t
he
high p
r
e
s
sure rotor
by co
mpari
ng fault
unit in low-freque
ncy com
pone
nts of th
e high p
r
e
s
sure
rotor the
r
e a
r
e not obviou
s
low-fre
que
ncy co
mpon
e
n
ts No. 1 an
d No. 2Z
WZ
vibration bef
ore
450p
w by tra
cki
ng the p
r
o
c
e
ss of the lit
er loa
d
. A sig
n
ificant lo
w freque
ncy com
pone
nt occu
rs
after 50
0kw,
up to
10
μ
m,
a si
gnificant i
n
crea
se
hap
p
ens in
th
e lo
w-frequ
en
cy
comp
one
nt, u
p
to
20
μ
m, the v
a
lue of
whi
c
h ca
n n
o
t be
overlo
oked.
Values
in the low-freq
uen
cy vibratio
n
and
occurre
n
ce frequ
en
cie
s
have gre
a
tly incre
a
sed i
n
the full load 60
0MW. The vibration
phen
omen
on
is difficult to define as a rand
om
fricti
on forced vibration with e
x
clusi
on of low-
freque
ncy
oscillation
ca
used by motio
n
of oil whirl
(oil whi
r
l vibra
t
ion sig
nal freque
ncy i
s
0
.
5
times sl
ower
than unit op
e
r
ation fre
que
ncy). All of
this sugg
est
s
that instable
vibration of the
high a
nd me
dium-pre
ssu
r
e roto
r bel
o
ngs to th
e e
x
citation of the ste
a
m flo
w
. Steam Fl
ow
Excitations b
e
long
s to a
self-ex
c
ited
vibr
ation (o
r neg
ative damping vib
r
a
t
ion). Dam
p
i
ng
gene
rated
m
o
vement of vi
bration
itself
exace
r
bate
the movem
e
n
t
[5, 9, 10] ra
ther tha
n
sto
p
it.
Steam Flo
w
Excitations
g
enerally o
c
cu
rs to
hig
h
-p
re
ssure
(o
r
hig
h
and
medi
u
m
-pressu
r
e
)
rotors
of the turbin
e in high po
wer u
nde
r hi
gh load, the
main feature
s
of whi
c
h is that vibration is
sen
s
itive to load. Sudd
en
vibration ha
s
a thre
shol
d load, whi
c
h
st
imulates th
e excitation of the
steam flo
w
whe
n
vibratio
n exce
ed
s th
e load.
Wh
il
e the lo
ad lo
wers to
som
e
certain
values,
vibration
i
s
re
sume
d with
g
ood rep
eatab
ility.
T
he mai
n
re
ason
ca
u
s
ing
the m
a
lfunctio
n
is ste
a
m
flow excitatio
n
throug
h the
analysi
s
of vibrant
si
gnal
s
and features
of
steam flow excitation.
3.2.3. The Analy
s
is of Mechanism
o
f
the Ste
a
m
Flo
w
Ex
cita
tio
n
It must have
two
conditio
n
s to
re
sult i
n
st
ea
m flow excitation, o
ne is
uneve
n
radi
al
distrib
u
tion
of the p
r
e
s
sure
within
se
al g
ap of th
e h
i
gh
p
r
e
s
s
u
r
e
r
o
to
r
;
th
e
o
t
he
r
is
imba
la
nc
e
in
rotor to
rque
radial [5].The type of excitation force:
The excitin
g
force
ca
used
by wee
k
ly chang
es of th
e steam
seal
cham
ber
pressure.
Vapor p
r
e
s
su
re of the hig
h
-p
re
ssure
ro
tator of
the l
a
rge tu
rbin
e i
s
high
with la
rge a
m
ount
s
of
leakage i
n
th
e stea
m seal
of blade
s.
Whe
n
the
te
mperature
is con
s
tant, th
e pressu
re o
f
the
shaft
seal
ch
ambe
r is p
r
o
portion
al to t
he flow rate o
f
the ch
amb
e
r
. The
ra
dial
clea
ran
c
e
(Fi
gure
9:
δ
1 =
δ
2
)
of the front and re
ar teeth
is equal
wh
en
roto
r is in
the rest po
si
tion. The ste
a
m
inflow i
s
eq
ua
l to the outflo
w
with
out
circulation
in th
e
cham
be
r. On
the premise o
f
the outlet ga
p
δ
2 le
ss than
the inl
e
t gap
δ
1, When
the
roto
r radial
d
i
spla
ce
s
(whi
ch i
s
th
e p
r
e
r
equi
site of
se
lf-
excited vibra
t
ion), the rel
a
tive chan
ge
in expor
ts t
eeth flow area is large
r
than that in the
entran
c
e of the Tooth flow area if
the radial displ
a
ce
ment of the rotor makes t
he shaft se
al gap
increa
se, ste
a
m outflows outnumb
e
r inf
l
ows with re
d
u
ce
d pre
s
sure in the shaft seal cham
ber. In
the other
wa
y round, the result
s are o
n
the contra
ry. Rotor p
r
e
s
sure an
d displ
a
cem
ent are not
synchro
n
ized
due
to the
i
nertia
of the
roto
r,
i.e. when th
e
rotor is
displa
ced
up
ward to
the
highe
st positi
on, the uppe
r gap is the smallest. Ho
wever the pre
s
sure in the chambe
r ro
om
is
not the highe
st.
Whe
n
the rot
a
tor go
es b
a
ck to the vicini
ty
of the rest positio
n from
the uppe
r, the uppe
r
cham
be
r p
r
e
s
sure
is th
e h
i
ghe
st. Thu
s
, top and
botto
m of roto
r
wi
ll form a
pre
s
sure differenti
a
l
urgin
g
the rotor continui
ng
its
downward movement
from the rest
positio
n so th
at the roto
r can
not stay
in
a
stationa
ry p
o
s
ition .thi
s
i
n
ertia
hy
steresis
effe
ct makes the pre
ssure in the lo
wer
cham
be
r ro
o
m
increa
se
whe
n
the rotor
contin
u
e
s its do
wn
ward moveme
nt. And this va
po
r
pre
s
sure differen
c
e will e
n
able the rotor to produ
ce a
displa
cem
e
n
t
and forms
a
eddy due wh
irl.
Becau
s
e it is
cau
s
e
d
by the vapor sy
ste
m
, ther
efore terme
d
as the
steam flow ex
citation.
The
stre
ss a
nalysi
s
of
rot
o
r e
ddy
cau
s
ed by p
r
e
s
su
re
cha
nge
s in
sh
aft seal
ch
ambe
r is
sho
w
n
in
Fig
u
re
10.
pre
ssure
differe
ntia
l in the
cham
ber ma
ke
s th
e an
gle
90
be
tween
the
ste
a
m
forc
e F
1
and
rotor el
asti
c restori
ng force
F
2
, both of which e
nabl
es
the rotor
disp
lace. (0<
Φ
<9
0)
At this
time,
t
he s
t
eam forc
e F
1
can b
e
decompo
se
d into a rotor el
astic resto
r
in
g force
with the
same
directio
n of the force F
11
and a
n
o
ther o
ppo
sit
e
to the di
re
ction of the da
mping fo
rce
F
3
,
w
h
ic
h
is F
12
functioni
ng
a
s
the
n
egativ
e da
mping.
Whe
n
F
12
i
s
greate
r
th
an
F
3
, the rotor
wil
l
gene
rate self-excited vibrati
on (F
4
is cent
rifugal force
)
[11].
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e-ISSN: 2
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TELKOM
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Vol. 11, No
. 8, August 2013: 4422 –
4432
4430
The exciting f
o
rce ca
used
by imbalan
ce
of the torque
of the rotor:
Due to u
n
it installatio
n
, cylinder d
e
viation
in op
erati
on and
ra
dial
displa
ce
men
t
of the
rotor, roto
r wi
ll
deviate co
mpari
ng with cylinde
r, cau
s
ing
uneve
n
radial
di
stribu
tion when
ste
a
m
acts
on the
rotor an
d the
rotor
whi
r
l. The de
co
m
p
o
s
ition of the
eddy mome
n
t
um is sho
w
n
in
Figure 11. Be
cau
s
e th
e lo
ss is sm
aller o
n
the radi
al
g
ap si
de, the
p
o
we
r a
c
ted
b
y
steam i
s
la
ger
than that on
the large
r
gap si
de. Where
b
y,
torq
ue forme
d
b
y
rotor is u
nbala
n
ced. T
h
e
unbal
an
ced t
o
rqu
e
force
can b
e
divid
ed into force cau
s
in
g rotation of the roto
r in the
circumfe
ren
c
e and
un
bal
anced fo
rce
Ft rotating
wi
th the rotor
and o
ppo
site
to the d
a
m
p
ing
force.
Ft a
c
ts o
n
the
rot
o
r
cente
r
an
d fun
c
tion
s a
s
a
ne
gative
dampi
ng.
When th
e force is
greate
r
than t
he dampi
ng force of the system, the
ro
tor will gen
erate self-excit
ed vibration [
12].
The unb
alan
ced force is:
F
t
=F
t1
-F
t2
(1)
As sho
w
n in
Figure 12, Ft is proj
ecte
d o
n
the ax
is wit
h
x, y as the centre po
sitio
n
of the
whe
e
l.
(2)
K
1
(u
n
it:N/m
)
is the co
effici
ent of gap exciting fo
rce, the cal
c
ul
ation
of which i
s
:
i
1
ii
95
49
P
b
K=
3
DH
n
（）
F
t2
n
F
t1
O
1
O
F
t
O
O
1
F
t
F
3
F
4
F
2
Figure 11. Ro
tor Eddy cau
s
ed by Torq
ue
Imbalance of
the Rotor
F
tx
=-K
1
x
F
ty
=K
1
y
Figure 9. Rot
o
r eddy cau
s
ed by
pre
s
sure ch
a
nge
s in sh
aft seal
Figure 10. Pressure analy
s
is in
shaft
seal
Evaluation Warning : The document was created with Spire.PDF for Python.
TELKOM
NIKA
e-ISSN:
2087
-278X
Vibration Ana
l
ysi
s of the Steam
Turbine
S
hafting cau
s
ed by Steam
Flow (Ge Li-j
uan)
4431
Formul
a:
D
i
: impeller
section
circula
r
diamete
r
, m;
H
i
: leaves, m;
β
: Gap excita
tion factor;
P
i
:
class p
o
w
e
r,
K
W
;
n: a rotor spe
ed, r/min.
As Equatio
n
3 sh
ows th
at K
1
stage
po
wer P
i
is
pro
portion
al to,
disp
rop
o
rtion
a
l to the
blade heig
h
t H
i
when
certa
i
n sp
eed
n, a
s
a
re
sult, po
wer is i
n
cre
a
sed, K
1
in
cr
e
a
se
s,
t
he
rot
o
r is
easy to i
n
stability. There i
s
not vi
bration malfunction in the units
when the l
o
ad po
wer stay
s low.
Only
on
the con
d
ition of power of
the
gen
erat
in
g u
n
its b
e
ing
greater
450
MW, can
the vib
r
ant
malfunctio
n
o
c
cur. Furth
e
rmore, the un
stable condi
ti
on of units is more se
riou
s with a furth
e
r
increa
se of t
he load
of the ran
dom g
r
oup. The
hi
g
h
pre
s
sure rotor ste
a
m in
the first ord
e
r of
bendi
ng n
a
tu
ral fre
que
ncy
will b
e
subj
ected to
a l
a
rge
amplitu
de when
the
total in a
cut
perp
endi
cul
a
r to the ecce
n
t
ric di
re
ction
of the
high
-p
ressure rotor
betwe
en the
exciting force
of
the steam
seal an
d excit
i
ng force of
unbal
an
c
ed t
o
rqu
e
excee
d
s the
damp
i
ng force of
the
beari
ng oil film.
4. Conclusio
n
Steam
flow e
x
citation
p
r
o
c
essing progra
m
of the hig
h
and m
edium
-pressu
r
e to
rque i
s
clo
s
ely lin
ke
d with
its vi
brant
me
cha
n
ism. A
c
co
rd
ing to
analy
s
is for th
e
situation of
ro
tor
vibration an
d mech
ani
sm o
f
the flow excitation as
well
as in-depth rese
arch of the vibration da
ta
and vibration
grap
h, re
du
cing
excitatio
n
forc
e an
d
increa
sing th
e system
da
mping a
r
e t
w
o
method
s to eliminate the vibration. Th
e measures
red
u
cin
g
steam f
l
ow excitatio
n
are as follo
ws:
(1)
The
adju
s
tment of the
cylinde
r a
nd t
he cente
r
of t
he rotor to
avoid the
evide
n
t displ
a
cem
ent
of the rotor a
nd the cylind
e
r ce
nter in o
peratio
n.
(2) Chan
ge
open
proced
ure
s
of
the
g
o
verno
r
va
lv
e, thus avoid
i
ng u
nbal
anced to
rque
on
the
rotor
cau
s
ed
by radial disp
lacem
ent of unilate
ral ste
a
m force. The adjustm
ent
methods
ca
n be
adopte
d
: inlet steam way is chan
ged
into a single valve, reducin
g steam
paramete
r
s and
cha
ngin
g
the valve sequ
en
ce;
(3) in
cre
a
si
n
g
bu
sh
dam
ping. Th
e
st
eam flo
w
ex
citation
belo
ngs to the
n
egative d
a
m
p
ing
vibration
whi
c
h plays
a po
sitive role in i
m
provin
g its
dampin
g
sup
p
re
ssi
on to
ward lo
w-freq
u
ency
vibration. Me
asu
r
e
s
can
b
e
taken: imp
r
oving
l
ubri
c
a
n
t tempe
r
ature, red
u
ci
ng
cl
eara
n
ce of th
e
top bea
ring,
adjustin
g
coordi
nate
s
of
the bear
i
n
g
,
and increa
sing the l
e
n
g
th of the b
u
sh,
adoptin
g lubri
c
ant with, a hi
gh viscosity ,etc.;
(4) Bush with
fine stability;
(5) T
o
improv
e the criti
c
al speed of the rotor.
Ackn
o
w
l
e
dg
ements
Sponsored b
y
National
Natural S
c
ien
c
e Fou
ndatio
n
of Chi
na
(NO. 1
1262
0
15)
and
Natural Scie
n
c
e Fou
ndatio
n of Inner Mo
ngolia (NO. 2
012MS0
732
).
y
x
O
t
F
ty
F
tx
F
1
O
Figure 12. Th
e Unb
a
lan
c
e
d
Force
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