Purdue University
Purdue e-Pubs
International Compressor Engineering Conference
School of Mechanical Engineering
1986
Single Stage, Oil-Free Screw Compressor with a
Compression Ratio Eight
H. Mori
K. Kasuya
Y. Takahashi
A. Suzuki
M. Aoki
Follow this and additional works at: http://docs.lib.purdue.edu/icec
Mori, H.; Kasuya, K.; Takahashi, Y.; Suzuki, A.; and Aoki, M., "Single Stage, Oil-Free Screw Compressor with a Compression Ratio
Eight" (1986). International Compressor Engineering Conference. Paper 519.
http://docs.lib.purdue.edu/icec/519
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Herrick/Events/orderlit.html
SINGLE STAGE, OIL-FREE SCREW COMPRESSOR WITH A
COMPRESSION RATIO OF EIGHT
Hidetomo Moril, Katsuhlk o Kasuyal,
Mitsuru Fuiiwar al,
Katsumi Matsuba ral, Akira Suzuki2, and Masakazu Aoki2
!Mechan ical Engineer ing Research Laborato ry, Hitachi,
Ltd.,Tsu chiura, Ibaragl, Japan
2Ebina Branch, Narashin o Works, Hitachi, Ltd., Ebina,
Kanagawa , Japan
ABSTRACT
The developm ent of a series of sin~le-stage, oilfree screw air compress ors with a compress ion ratio of
8, previous ly obtainab le only with two-sta~e compress ors,
is describe d.
A new rotor profile , which reduces
leakage loss to achieve higher efficien cy, and desig-ned for
ease of manufac ture, is detailed .
Addition ally, a new
desig-n method for the clearanc e between rotors, to
compens ate for rotor
deforma tion due to thermal
expansio n, is also introduc ed.
This oil-free , screw
air compres sor with a ratin? of 37-55 kW has been
marketed since 1982 and 15-22 kW since 1984.
INTRODUCTION
Oil-fre e air compres sors having a dischar ge
pressure of approxim ately 0.8 MPa (8 ata), and
widely used in electric al, food and chemical industri es
were traditio nally reciproc ating type compres sors.
Recently, however, the advantag es of oi 1-free rotary
screw compress ors have been reco~nized.
These include
mechanic al simplici ty, high reliabil ity,low noise and
low vibratio n. Two-stag -e, oil-free screw air compress or
use has predomi nated in the capacity range above
550 m3/h (motor power above 65 kW).
Neverth eless, for compress or capacity below this range,
where oil-free air is needed, this type of compress or
was not availabl e.
105
tage,
Due to the comp licate d const ructi on of two-s
ght that
oil-f ree screw comp resso rs, it was thou
rocat ing
they would not be econo mical , compa red to recip
range
the
in
nt
opme
devel
if
even
oil-f ree comp resso rs,
There fore, devel opme nt
below 550 m3/h was possi ble.
prefe rable
of singl e-sta ge comp resso rs was consi dered antic
ipate d
and
on
ructi
becau se of their simpl e const
to
ing
relat
ined,
rema
ems
probl
ral
Seve
low cost.
comr
highe
the
as
such
s,
sever e opera ting cond ition
eratu re
press ion ratio and resul ting incre ase in temp
these
to
ions
solut
given
Even
of the disch arge air.
conwere
y
bilit
relia
and
e
rmanc
perfo
es,
culti
diffi
use.
nt
warra
to
poor
sider ed to be too
le,
Addre ssing these probl ems, a new rotor profi
ratio
n
ressio
comp
high
a
at
yield ing highe r effic iency
n metho d
was initi ally devel oped, follow ed by a new desig
nt of
accou
into
·takes
d
metho
This
of the rotor cleara nce.
durin g operrotor defor matio n due to therm al expan sion optim
izati on
This
ation and optim izes clear ance .
menta l
funda
These
.
iency
effic
great ly affec ts comp ressor
e-sta ge,
impro vemen ts enabl ed devel opme nt of a singl
ratio of
oil-f ree screw comp ressor with a comp ressio nattai nable
dered
consi
Such a ratio was previ ously
8.
Appl ying these
only with two-s tage comp ressio n.
comp ressor with
screw
ree
,oil-f
devel opme nts, a pack aged
senti ng the
repre
made,
was
kW
55
37
of
g
ratin
a
comp resso r.
air
screw
ree
oil-f
lest
first and smal
ss of the
proce
ntal
This paper descr ibes the devel opme
new comp resso r.
TECHNICAL OBJECTIVES
disA screw comp resso r is a rotar y posi tive
rotor
e
femal
and
male
one
g
havin
place ment comp ressor
in a casin g.
mesh ing with each othe r and hous ed rotor
s and the
the
en
betwe
er
chamb
the
into
Air drawn
Air
n.
motio
rotor
to
due
ressed
casin g walls is comp
en the
betwe
ance
clear
of
t
resul
a
as
ing
occur
ge,
leaka
the casin g
rotor s as well as betwe en the rotor s and
nject ion
Oil-i
.
iency
effic
walls , reduc es comp ress ion
an oil
with
ances
clear
these
seal
rs
resso
comp
screw
resso rs
comp
ree
Oil-f
er.
chamb
n
film in the comp ressio
ge.
leaka
this
nt
preve
to
lack seali ng oil in the chamb er
comThis is espe ciall y true with sing le-st age
ly
great
and
s,
ratio
n
ressio
comp
r
highe
g
havin
press ors
t.
reduc ed effic iency as a resul
106
Fig.l shows how the clearance between rotors (interlobe
clearanc e) of an oil-free , screw air compress or
influence s the compressi on efficienc y by computer
simulation .
In this figure, volumetric and overall
adiabatic efficiency at a male rotor tip speed of 100
m/s and a clearanc e ratio of 0.001 are taken as
l. 0.
Increasing rotor rot at ion speed is one solution
to this problem, but not a satisfact ory one.
The
fundament al solution is to minimize clearance so as to
lessen air leakage.
Thus, the first technical problem
was developme nt of a new rotor profile with minimum
clearance for minimum air leakage, along with machining
technology improveme nts enabling a high-prec ision rotor
finish. ·
The second difficulty concerned is the air temper at ur e
increase during compress ion and the resulting thermal
~xpansion
of the rotor.
Air tell'peratu re rises with
Increase in the compressio n ratio.
The lubricant in
oil-inject ion screw compresso rs also functions as a
coolant, permittinf ! the temperatu re rise to be controlled.
In the case of oil-free rotary screw compress ors:
(a) Hot air leaked fro~ the high-pres sure side can heat
the low-press ure, low-tempe rature air.
(b) The temperatu re of this heated air increases further
after compressio n.
(c) This air again leaks and heats up the low-tempe rature air.
Thus, the air temperatu re progressiv ely rises. Heated
by this air, the rotors expand. This reduces the rotor
clearance to a critical degree.
Contact between them
causes compresso r darr.age or failure.
Hence, conventional oil-free screw air compresso rs operating
at a compressio n ratio above 4 had a two-stage design
with an intercoole r.
Fig.2 shovJS a comparison of operating conditions
between a two-stage and a single-st age screw air
compresso r at a discharge pressure of 0.8 MPa (8 ata).
The compressi on ratio of each stage of the
two-stag e co~pressor is 3 and the discharg e air
temperatu re is less than 150°C.
The compressio n ratio
of the sinvlP-sta fe comrresso r is 8 and the discharge
air temperatu re ~av be hipher than 300°C.
The tePlreratu re difference implies very !'evere operatinv condition s for the sin?le-st are compresso r.
Thus, the technical chAllen~e involves rlevelopme nt
of a new sinple-sta ee, oil-free rotary screw compresso r
with reduced internal air leakape and. opti~u~ cl~arance
compensat ion for the theriT'a 1 ex pans ion of the rotors.
107
Furthe rmore , Fig.l indica tes that one perfor mance
involv es
chara cteris tics of oil-fr ee sere"'' compr essors
.
speed
tip
rotor
r
slowe
with
ency
effici
of
lower ing
be
to
need
rotors
essor
compr
There fore, srr,all.
n highe r
obtai
Lo
speed
ion
rotat
r
highe
a
at
n
drive
The result ing rotor vibra tion increa se
effici ency.
The techn ical
must be taken into consi derat ion.
of a sophi stipment
develo
the
in
solved
probie ms to be
cated compr essor are sumar ized in Flg.3 .
TECHNICAL SOLUTIONS
In the light of the s e d i f f 1 c u l t i e s , inten sive R&D
fundaeffor ts to develo p new techno lovy were made by
these
of
Some
ts.
imen
exper
ment al analy sis and
are outlin ed below .
The rotor profi le of an oil-fr ee screw compr essor
achiev ing
is one of the most influe ntial factor s in
such a
for
nts
reme
reaui
The
.
highe r effic iency
profi le are as follow s.
(l)
Seal lines to preve nt air leakag e betwe en rotors
made.
and betwe en rotors and casing walls should be
should
nce
cleara
the
and
short
be
These lines should
e.
be of a shape that funct ionall y preve nts leakag
more
is
g
sealin
e
surfac
to
e
surfac
As an examp le,
effici ent than sharp edge sealin g.
and
(2) The cleara nce betwee n rotors arid betwee n rotors ndth
casing walls should be smalle r than one thousa
for
of the rotor diame ter, resul ting in a need the
ally,
Basic
ng.
acturi
super precis e rotoi:- manuf
for
rotor profl le shoul d be a shape suita ble
manuf acturi ng.
pin~
Giv~ri ~he time ~nd expen~e invol~~d in develo
(CAD)
design
aided
ter
compu
a
a screw rotor profi le,
The CAD system used to
ststem can be v~ry us~ful.
This
shown in Fig.4 .
is
le
profi
rotor.
develo p the new
estim ates
system is compo sed of two sub-sy st'ems , one
lates
screw compr essor perfor mance , and the other calcu
les.
rotor and cuttin g hob profi
les
Apply ing this CAD system , many rotor profi
ee screw
accep table for high compr ession ratio, oil-fr
The most promi sing new
compr essors were inves tigate d.
108
rotor profile is shown in Fig.S.
This profile has the
following advantage s over a conventio nal profile.
(a) The respective lobe combinatio n of five for male
and six for female rotors is sui table for h i g her
compressi on ratio applicati ons,
creating one
more working chamber than the conventio nal profile.
The addition al chamber reduces the pressure
differenc e and compresse d air leakage between
chambers.
(b) The blow hole area is only 27 % of the conventio nal
profile, and the seal line between rotors is 20 %
shorter than the conventiona l one. Additiona lly, the
new profile is created by continuou s, smooth
curves.
The smaller blow hole, shorter seal line
and surface to surface sealing results in less air
leakage.
(c) The continuou s, smooth curves of the profile made
it possible
to machine rotors with higher precision by hobbing, resulting in smaller clearance
between the rotors.
As a result of computer simulatio n, the new rotor
profile showed 7 %higher volumetric efficiency and 17 %
higher overall adiabatic efficiency compared with the
conventio nal rotor profile (Fig.6).
Compensat ion for Rotor Thermal Expansion
It is neccessary to know rotor temperatu res for
optimizat ion of rotor clearance s.
Fig. 7 shows rotor
temperatu res of a prototype single-sta ge, oil-free air
compresso r measured at a discharge pressure of 0.8
MPa (8 ata), compared with a caiculated rotor temperature.
The discharge side rotor temperatu re was approximately 300 °C.
This temperatu re was much higher than
predicted beforehan d.
A new design method was neccessary to minimize the ·rotor clearance along the whole
rotor profile at such a high temperatu re.
The fundament al desip:n concept of the new method
is shown in Fig.S.
® is an initial profile of a male rotor at room
temperatu re, from wh lch the heat-expan ded profile @
is obtained.
Hith allowances for timing gear backlash
and rotation reliabilit y, profile
© is determine d.
The heat-expan ded female rotor profile Q» is generated
b_y_ © . The female rotor prof lle at room
temperatu re
© is obtained by contractin g @. _CuttinA hob profiles
are transferr ed from rotor prof1les
® and @ .
Applying this method, the rotor cleRrance is
optimized at a higher operating temperatu re.
The new
design produced a 40-50 % smaller clearance compared
with a conventio nal design.
This is approxim ately
equal to a 13 % higher overall adiabatic efficiency at
a
rotor tip
speed of 100 m/s according to Fig.l.
109
PROTOTYPE COMPRESSOR
, seve ral
Follo wing this fund amen tal rese arch all being
~ade,
were
rs
resso
comp
proto type sing le-st age
Some of them_ had
appl icabl e to the 37-55 klf' rat infs.
the conv entio nal
had
rs
othe
while
ile
the new rotor prof
one.
Comp resso r Struc ture
e-sta f!e
Figur e 9 shows a cutaw ay viet.J of a nel.\1 singlradia lly
s are
rotor
male
and
le
Fema
r.
resso
comp
and axia lly by
supo rted by cylin drica l rolle r beari ngs of timin g gears
pair
A
ings.
comb ined angu lar ball bear
the clea ranc es
at the end of roto r shaf ts adju sts
efore , these
Ther
rs.
roto
the
of
betw een the lobe s
n ?ear at the
lobes do not conta ct one anoth er .. A pinio a large gear
by
drive end of the male rotor is drive n
the comp resso r is
cont aine d in a gear casin g on which
The male rotor was teste d
£lang -mou nted horiz onta lly.
at up to 22,00 0 rpm.
oil from
Visc o-typ e seals preve nt the lubri cant
n rings
carbo
while
ber
enter ing the comp ressio n cham
roto r
the
in
ent
pres
is
hole
oil
An
seal the air.
is
ng
coati
f
proo
heatial
spec
A
shaf t for cooli n)l:.
g of the
casin
The
.
aces
surf
r
roto
the
to
appl ied
comp resso r is coole d by a wate r jack et.
r
Perfo rman ce Qi the Proto type Comp resso
type comPerfo rman ce test resu lts of the proto
adia batic
all
over
The
O.
Fig.l
in
shown
press ors are
incre ase
the
and
y
ienc
effic
etric
effic ienc y, the volum
conv entio nal
the
of
re
eratu
temp
air
e
harg
disc
in
of 8 (desi gn
prof ile comp resso r at a comp ressio n ratio
Volu metr ic
1.0.
as
comp ressi on ratio ) are taken
y obtai ned
ienc
effic
batic
adia
all
over
and
y
effic ienc
impro ved, and
with the new rotor prof ile are grea tly re is reduc ed.
eratu
the incre ase in disch arge air temp
the high er the
The high er the comp ressi on ratio ,
nt redu ction in
ifica
sign
This indic ates
impro veme nt.
rotor prof ile.
new
the
of
s
mean
by
loss
air leaka ge
the rela tive
(8),
ratio
on
ressi
At the desig n comp
entio nal to the
conv
the
from
nts
veme
impro
y
ienc
effic
etric effic ienc y
new rotor prof iles are 27 'loin volum
y.
ienc
effic
and 38 % in over all adia batic
new prof ile
The perfo rman ce char acte risti cs of the
show n in
is
ds
spee
tion
rota
us
vario
at
r
comp resso
y curve
ienc
effic
batic
adia
Fig. ll. A very flat over all
l speed .
iona
rotat
rotor
of
range
wide
a
is seen over
llO
These test results allowed develop ment of a new
machine able to deliver 0.8 MPa (8 ata) air by a
single -stage compre ssion.
Table 1 lists standar d
specifi cations of a series of single- stage, oil-fre e
screw air compre ssors.
CONCLUSION
The develop ment of a single- stage, oil-fre e screw
air compre ssor was describ ed.
This is the world' s
first and smalles t single- stage compres sor operati ng at
a compre ssion ratio of 8, previou sly obtaina ble only
with two-sta ge compre ssion.
As a result of intensi ve R&D effort, reliabl e and
low cost oil-fre e screw compre ssors became availab le in
the capacit y range from 120 to 455 m3jh.
Formerl y,
oil-fre e recipro cating compre ssors had predom inated in
this range.
ACKNOWLEDGMENTS
The authors would like to thank their colleag ues
at the Mechan ical Enginee ring Researc h Laborat ory and
the Ebina Branch Works for their concern and coopera tion.
REFERENCES
[ 1] Trulsso n, I., "A New Develop ment in Rotary Screw
Compre ssor Design ", Purdue Compre ssor Techno logy
Confere nce, 1970.
[2] Fujiwar a, M., Mori, H. and Suwama, T., "Predic tion
of Oil-fre e Screw Compre ssor Perform ance Using Digital
Comput er", Purdue Compre ssor Technol ogy Confere nce,
1974.
[3] Fujiwa ra, M., and et al, "Comp uter Modeli ng
for Perform ance Analys is of Rotary Screw Compre ssor",
Purdue Compre ssor Technol ogy Confere nce, 1984.
111
1.2
0.6
SUCTION PRESSURE - lOlkPa (1.03ata)
DISCHARGE PRESSURE - 0.8 MPa (8 ata)
-~1.4( X\IJ~J)
-~1.2
~~0~£-0.8-~--~I0
Fig.l :
INTERLOBE CLEARANCE TO RDTOR
DIAMETER RATio
obe Clearance and Compressor
lnterl
en
Relat ion betwe
metric effici ency
Effic iency by Camputer Sfinulation (Volu
rotor tip
male
a
at
ency
effici
atic
adiab
ll
and overa
ratio of
nce
cleara
obe
interl
an
and
m/s
speed of 100
1.0.)
as
taken
are
0.001
~
AFTERCOOLER
0
~'
~
r.n
r.n
~
~
Fig.Z :
'""'
ou
~
r.n
I
~
~
!n
I
g
""'
r.n
~i
li
jOQ
JDO
2>0
200
l:;o
~' ~
~
I
<.::
.....
~
r.n
'""
en TWo-stage and
Ganparison of Operating Conditions betweressor
s (Discharge air
Singl e-stag e, Oil-f ree, Screw Air Comp
atic effici ency
temperature is calcu lated assuming an adiab
of 0.8.)
112
SINGLE-STAGE OIL-FREE I
IROTARY
SCREW OJMPRESSOR
HIGHER OJMPRESSION
RATIO
I
'HIGHER DISCHARGE
AIR TEMPERATURE
I
LARGER LEAKAGE
LOSS
1HERMAL
EXPANSION OF ROTORS
I IJ..<I.RGER
LOWER EFFICIENCY
I
--
I
ROTOR
ILARGER
VIBRATIONS
II.WER RELIABILITYj
(TECHNICAL PJIDBLEMS)
f---
- - - - - - ---~~,
MINIMIZING AIR LEAKAGE
•DEVELOPMENT OF HIGHLY EFFICIENT
NEW ROTOR PROFILE
•HIGH-PRECISION ROTOR
MANUFAC1URING
Fig.3:
!HIGHER ROTATIONAL
SPEED
OPTIMUM ROTOR CLEARANCE
• <XlMPENSATION FOR TERMAL
EXPANSION OF ROTORS
Technical Problems in the Development of a Single-stage
O::mpressor
Fig.4:
CAD System to Develop a New Rotor Profile
113
I
I
I
I
CaY CONVENTIONAL ROToR PROFTLE
(b)' NEw ROTOR PROFtLE
Fig.S:
NeW and~
Conventi"onal Rotor Profi!les
SUCTION_ PRESSURE - iOlkPa (LCnata)
SuCTION TEMPERATURf: = 2b""C
MALE ROTOR TIP SPEED' - SO in! s
INTERuDBE CLEARANCE RATIO
~
0.67xlo-3
-
-- -- -- --~
---
.......
NEW PROFILE
t..:>
,......_0,~
~
E::
~~
12
mNIJENnbNAL
~~
~~
:$U
__ .... __ _;_ PROFILE
10
----~-
~---
- ........
~ ~ ~!---+--+----:~:---+-COMPRESSION RATIO
Efficiency between New and COnventional
Fig.6: pdnparison of Cdmpressot
ROtor Profil~s by Computer Sfinulation (Caripressor efficiencies
with the conventional profile at the cowpression ratio of 0.8
are taken as 1.0.)
114
SUCTION
SIDE
MEASURED TEMPERATURE
AT THE TOP OF THE IDBE
AT 1HE OOTTCM OF THE rnooVE
I
CALCULATED SURFACE
TEMPERATURE
I
I
~
I
SUCTION PRESSURE
~
DISGlARGE PRESSURE
lOlkPa (l.03ata)
~
0.8 MPa (8 ata)
0~--------~~-----------~------0
1.0
DIMENSIONLESS ROTOR LENGHT
Fig.7:
Measured and Calculated Rotor Tanperatures of a Single-sta ge,
Oil-free Screw Compressor (Rotors are non-cooled .)
115
~
MALE ROTOR PROFILE AT ROOM TEMPERATURE
HEAT-EXPANDED MALE ROTOR PROFILE
PROFILE WITH AN ALLOWANCE FOR GEAR BACKLASH
@ HEAT -EXPANDED FEMALE ROTOR PROFILE GENERATED BY
@ FEMALE ROTOR PROFILE AT ROOM TEMPERATURE
Fig.8:
©
Compensation for Rotor Thermal Expansion
PINION GEAR
MALE ROTOR
TIMING GEARS
WATER JACKET
Fig.9:
Prototy pe Compressor
116
r•
'-'t
I0
-o-o--a--o-o_o_
__ 6 _ _.t.___ o.
---n.---~---0.--
tO
,,
u
o
I-.........
<(f-
l
~
cc«
"""'
0 >u
""
~t;
""-
12
<(
....!"-'
10
...... lL
E; tti
us
NEW PROFILE (WITH COMPENSATION
FOR THERMAL EXPANSION)
I.,.,....-- - o - o - - o-0
/0
-
/0~-"'-"'-,-..
/~
\
CONVENTIONAL PROFILE (WITHOUT
COMPENSATION FOR THERMAL EXPANSION)
COMPRESSION RATIO
SUCTION PRESSURE ~ lOlkPa (1.03ata )
SUCTION AIR TEMPERATURE ~ 20 °C
Fig.lO: PerfoTinance Test Results (Efficie ncies and discharg
e air
tempera ture increase of the compressor with the conventi onal
profile at the compression ratio of 8 are taken as 1.0.)
117
'[
I0
oe
~
1.2
>-
~
10
"'
1---'
.....
O.B
""
:r:
"" o•
ROTATIONAL SPEED OF MALE ROTOR (rpm)
SUCTION PRESSURE - lOlkPa (1.03 ata)
SUCTION AIR TEMPERATURE - 20 °C
DISCHARGE PRESSURE - 0.8 MPa (8 ata)
Fig.l l:
the New Profi le Compressor
Performance Char acter istics ofchara
cteris tics at 15,000 rpm
versu s Rotat ional Speed (The
are taken as l.O.)
Tabl e 1:
MODEL
MOTO!l POWER(kW)
3
CAPAC inim /h)
Stand ard Spec ifica tions
I\ DSP-SOWA DSP-60W A OSP-7SW A
PSP-20C A PSP-JO CA DSP-201 /A DSP-30W
15
)20
22
]5
22
37
45
55
J 80
J 20
)80
280
365
•ss
"
D.lSCHARGI!: PRESSURE (a ta)
COOLING MET"OO
DIMENSION (LxWx"J (nun)
NET W£IGHT (k91
NOISE LEVEL dB (AI
AIR COOLED
J ,)OOx)
8. 7
,]00><) ,)00
515
625
70
n
118
WATER COOLED
1,1 00x700 x)
1 )
1, SSOx770M]
00
, ;i!OO
500
550
750
780
aoo
68
70
72
n
12
PERFORMANCE ANALYSIS OF OIF SINGLE SCREW COMPR
ESSOR
Tetsu o Hirai 1 ,
Sadaf umi Noda2 , Taiic hi Sagar a 2 , Kiyoh aru
Tsuzi 2
1 centr al Resea
rch Lab., Mitsu bishi Elec tric Corp .,
8-1-1 , Tsuka guchi honm achi, Amag asaki, Hyogo
, 661 JAPAN
2Nag asaki Works , Mitsu bishi Elec tric Corp .,
517-7 , Ramad a-Go, Togit su-Ch o, Nishi sonog i-Gun
,
Naga saki, 851-2 1 JAPAN
ABSTRACT
This paper prese nts the perfo rman ce analy sis
and
the inter nal press ure measu remen t of the oil
free singl e screw comp resso r. The geom etric injec tionshape of
the singl e screw comp resso r and its theor etica
l
perfo rmanc e with the slide valve have been
analy zed.
The inter nal press ure has been measu red with
piezo type
press ure senso rs, and the press ure-v olum e
diagr am has
been obtai ned. The exper imen tal resul ts agree
fairl y
well with the theor etica l predi ction s, and
the
volum etric and adiab atic effic iency can be
estim ated by
this analy sis.
SYMBOLS
A
cross secti onal area
Dsg
dista nce betwe en screw axis and gate rotor
axis
F
coeff icien t of flow rate
Gc
mass of gas in the groov e
G·l
mass of gas which enter s the groov e
Go
mass of gas which comes out of the groov
e
119