Purdue University
Purdue e-Pubs
International Compressor Engineering Conference
School of Mechanical Engineering
1994
Development of the High-Efficiency HorizontalType Scroll Compressor
S. Hase
Matsushita Electric Industrial Company
K. Sano
Matsushita Electric Industrial Company
S. Yamamoto
Matsushita Electric Industrial Company
H. Hirano
Matsushita Electric Industrial Company
T. Kohayakawa
Matsushita Electric Industrial Company
See next page for additional authors
Follow this and additional works at: http://docs.lib.purdue.edu/icec
Hase, S.; Sano, K.; Yamamoto, S.; Hirano, H.; Kohayakawa, T.; and Ishii, N., "Development of the High-Efficiency Horizontal-Type
Scroll Compressor" (1994). International Compressor Engineering Conference. Paper 1019.
http://docs.lib.purdue.edu/icec/1019
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Authors
S. Hase, K. Sano, S. Yamamoto, H. Hirano, T. Kohayakawa, and N. Ishii
This article is available at Purdue e-Pubs: http://docs.lib.purdue.edu/icec/1019
Development of the High-F iticienc y Horizontal- Type
Scmll Compressor
by
Shozo Hase1 , Kiyoshi Sano2 , Shuichi Yamamoto3 ,
l:lideo l:liraoo3 , Taisei Kohayakawa1 , Noriaki Jsbii•
1
2
3
4
Engineer , Air Conditioning Research Laboratory,
Manager , Air Conditjoning Research Laboratory,
Chief Engineer , Air Conditioning Research laborator y,
Matsushita Electric Industrial Co.,Ltd. (P ANASON IC),
Noji-cho , Kusatsu 525, Japan
Professor , Faculty of Engineering ,
Osaka Electro-Communication University ,
Neyagaw a, Osaka 572, Japan
ABSTRACT
This paper presents our research into and development of a high efficiency horizonta l-type
scroll compressor
and
its performance. The newly-de veloped horizontal-type scroll compressor was characterized not
only by outstanding high
efficiency, low vibration and low noise characteristics, but it was possible to reduce drastically
the size of the external unit
of the air conditioners. An energy loss analysis using P-V diagrams and visual observatio
n of the wearing
state at the
bearings were perfonne d first, and subsequently various key technologies were developed
to achieve the high pedonnan ce.
As a result, the total losses in the compact horizontal-type scroll compressor were reduced
by 25%, making it possible to
improve the efficiency by 14% over that of the previous horizontal-type scroll compress
or. Moreover, the required floor
space of outdoor units could be reduced, and the image was upgraded, leading the way
for today's compact products.
1. JmRODU CIION
In recent years there are increasing demands for home-use air conditioners which
combine improved energy
efficiency with enhanced comfort and quiet. Early on, the authors recognized the superior
characteristics of the scroll
compressor as far as high performance, low noise and low vibration are concerned
, and have already developed the
small-sized vertical-type scroll compressor which is used in room air conditioners that
were put on the market in 1990.
A small-sized horizonta l-type compressor was developed in 1992. The present research
was carried out in order to
improve even further the efficiency of the horizonta l-type scroll compressor. To accomplis
h this goal, the authors ftrst
made an energy loss analysis using P-V diagrams and visual observation of the wearing
state at the bearings, determining
the sources and levels of energy loss in the previous horizontal-type scroll compresso
r. On the basis of these data, the
compression efficiency was improved significantly by shifting the phase between the two
scrolls, modifying the discharge
system and optimizing oil flow rate and the oil flow conduit to the compression chamber.
Furthermore, the mechanical
efficiency was improved by sm·oothing the surface of the main bearing and adopting
a high precision assembly method.
Simultaneously, the efficiency of the motor was also improved.
2. STRUcr uRE AND SPECIFICATIONS
The structure of the compact
horizontal-type scroll compressor of 2.4
[kW] refrigerating capacity is shown in
Fig. 1. This is a high-pres sure shell type
compressor. The main shaft is supported
by a sliding bearing on its end on the
orbiting scroll side, and by a ball bearing
on the other end, thus giving a design
that provides double end support.
Oil pump
Motor
Main shaft
Oil pressure reducer
Fig. I Constraction of horizon tal-type scroll compressor
447
Compression Mechanism
The fJXed scroll made of cast iron
The compression mechanism is composed of a fJXed scroll and an orbiting scroll.
is mounted on the discharge
valve
A
plate.
has a suction port on its periphery and a discharge port in the center of the end
orbiting scroll end plate,
the
of
center
the
in
located
port to prevent reverse flow of compressed refrigerant gas. A shaft is
bearing mounted
eccentric
the
of
result
a
as
fashion
orbiting
an
and is driven by the main shaft. The driven shaft moves in
on the vane tip
mounted
is
seal
tip
A
plate.
end
scroll
orbiting
the
in the main shaft and the Oldham ring installed behind
. The orbiting scroll is made of aluminum
of the orbiting scroll to achieve the axial sealing of the compression chamber
to reduce the mass.
alloy so as to withstand the high-sp eed operation under invertor drive, and
Oil Supply Padi
of a trochoid rotor. The oil in the shell
One end of the main shaft is equipped with a volumetric oil pump consisting
assuring a steady flow of lubrication
sm,
is pumped along a path in the center of the main shrft to the compression mechani
orbiting scroll.
the
of
portion
oil to the orbiting shaft bearing, the main shaft bearing and each sliding
3. RESULTS OF LOSS ANALYSIS FOR PREVIOUS SCROl L COMPRESSOR
3-1 Identification of Drive Losses from P-V Diagram
results of the loss analysis based on
The P-V diagram for a previous scroll compressor is shown in Fig. 2, and the
the two scrolls are referred to as Chamber
this P-V diagram are shown in Fig. 3. The compression chambers formed by
port), as in Fig. 2. From the loss analysis
A (on the suction port side) and Chamber B (on the opposite side of the suction
sor, the mechanical loss comprises the
compres
results shown in Fig. 3a, it can .be concluded that for the previous scroll
second largest loss. The compression
the
32%,
largest percentage, 44% of the total loss, and the compression loss, with
and heating during the suction
leakage
are
causes
loss can be divided into various losses as shown in Fig. 3b. The major
that the pressure in Chamber
2
Fig.
in
diagram
P-V
the
and compression processes. It can be seen, in particular, from
a significant imbalance of
creating
thus
process,
sion
compres
the
A is significantly higher than that in Chamber B during
caused by the difference in the thermal
leakage and heating between the two. It is thought that this imbalance was
scroll. Another possible reason is the
expansion coefficient of the materials used for the fJXed scroll and the orbiting
e in heating of the suction gas, since a
difference in suction path length between the two, which causes the differenc
discharge loss is caused by the discharge
relatively large amount of oil is needed to seal the compression chamber. The
COMPRESSION
resistance due to the valve mounted on the discharge port.
CHAMBER A.
rements : 90Hz
2.0
Leakage & Healing
~
:.:g
\ ,
1.5
Adiabatic
a.. 1.0
0.5
0
/ \\ \
L.os9
{COinprel~Son)
Suction Loss
Compression
Loss
~\\',
'<::
lheorellca l
Leakage & Heating
(suction process)
Mechanical
Loss
Suction L.os9
Leakage & Heating
(compression process)
-~
Discharge Loss
Motor Loss
0.5
1
V/Vo
Fig.2 P-V Diagram for a previous design
Total loss %
b)
Compression loss %
a previous desi~n
Fig.3 Result of loss analysis for
(JIS-A requirements : 90Hz)
448
3-2 Sliding States of Fach Pair
Each sliding surface was observed carefully in order to detennine the level
of mechanical loss. Fig. 4 shows a
cross-se ctional view of the main bearing. An electric insulator is fitted to the
outer surface of the bearing bush to examine
the contacting state between the bearing bush .and the main bearing. Since the
bearing bush is electrically insulated from
the bearing block, a contact signal can be detected when the bearing bush contacts
the main bearing. A similar device was
-r
I'
(a) Main Bearing
]1[~1::~;
I
'I
~
"'"-.,~._
i
..... ,.._ 'T·r .• ,.,......., ...,L..,..._
I
(b) Orbiting bearing
(c) Thrust bearing
Fig.5 Contacting signals showing
sliding state at each pair
( s)
Fig.4 A device for examining the sliding state
used to examine the contacting states at the orbiting bearing and at the thrust
bearing between the scrolls, respectively.
As shown in Fig. Sa, the contacting signals can be occasionally observed
at the main bearing, thus indicating a slight
contact between the bearing bush and the main shaft. On the other hand, as
shown in Fig. Sb, no large contacting signal
can be observed at the orbiting bearing. Furthennore, as shown in Fig. Sc, no
contacting signal can be observed at all at
the thrust bearing. The results in Figs. Sb and Sc suggest that a thick oil ftlm
is well developed at both the orbiting bearing
and the thrust bearing, resulting in a fluid lubrication state. Thus one may conclude
that the mechanical efficiency can be
improved by reducing the mechanical loss at the main bearing.
4. IMPROVEMENT OF EFFICIENCY
4-1 Improvement of Compression Efficiency
(1) Opdmiz ation of Radial Oearanc e
It was confrrmed from the P-V diagram in Fig. 2 that in the previous scroll
compressor, the pressure increase in
Chamber A on the suction port side is too large, thus producing an additional power
loss. It is thought that one of the major
.--_
(a) when Cold:
assemble d in
ideal position
0
(b) when Hot :
clearance" i5 " caused
by thermal expansion
(c) additional phase
offset" a"
Fig.6 Mating of the fixed scroll and the orbiting scroll,
with an additional phase offset angle
449
ftxed scroll and the orbiting scroll, which causes the gap to widen in the radial direction. Normally the fiXed and orbiting
scrolls, which have exactly the same shape, are assembled with a phase offset of 180 degrees, whereby the compression
chamber has no leakage. However, this is an ideal case for two scrolls made of the same material. It is clear that thermal
expansion can be ignored when assembling two scrolls made of the same material. When the two scrolls are made of
different materials, however, as in the present study, that is, the fiXed scroll is made of cast iron and the orbiting scroll
of aluminum alloy, the orbiting scroll with a large thermal expansion coefficient expands more than the fiXed scroll. As
a result, even if the two scrolls are assembled in an ideal position when cold, the radial clearance increases, as shown
schematically in Fig.6.
In order to solve this problem, the authors assembled the two scrolls with an extra phase offset (" a ") in addition
to the ideal phase offset of 180 c:!egrees, as shown in Fig. 6c. The temperature distribution of the ftxed scroll was measured
during normal operation ftrst, and the optimum angle for the extra phase offset was determined. The test results of the
leakage loss during the compression process are shown in Fig. 7, in which the horizontal axis is the additional phase offset
angle between the two scrolls. It can be seen that there is an optimum offset angle which minimizes the leakage loss; in
fact the leakage loss was reduced by 25% with the best offset angle. Furthermore, it was confirmed that the optimum offset
angle found experimentally was in good agreement with the theoretically calculated offset angle. The effect of the optimum
offset angle is shown in the P-V diagram in Fig. 8. A comparison with the data for the previous compres!ior design, shown
in Fig. 2, shows that there is a vast improvement in the imbalance of pressure across the compression chambers. The
pressure curve in both the compression chambers approaches the ideal pressure curve:
1.8
.9 1.6
~
.....
1.4
~
..9
i
~
1.2
1.0
Optimum angle
by calculation
I'
«<
Previous I
a.. 1.5
______
IT___
/
'
:::2:
I
a..
0.8
0.6L----...... L....--...J...._ --_.J
1•0
llleoredal / ' - , ,
compression , ,
curve
',,
0.5
------;;.+
-«---Aditional phase offset angle
0
a deg
Fig. 7 Leakage loss ratio vs.
aditional phase offset angle
0.5
V/Vo
Fig.B P-V diagram for the
newly-designed compressor
(2) Optimization of Oil SuppJy'into tb: Co~sion Chamber
It was confrrmed from the loss analysis
Back pressure
using the P-V diagram that the leakage and
chamber
heating loss during the suction process makes
loss.
up a large percentage of the compression
It is thought that this leakage and heating loss
was caused by the heating of the refrigerant gas
by the compression chamber walls and the
sealing oil injected into the compression
chamber, and the gas leakage from the high
compression chamber. The latter cause, namely,
gas leakage, can be reduced through the offset
angle optimization described above. Thus, the
authors turned their attention to the oil injected
Fig.9 Path of oil supply
into the cylinder. The oil supply path is shown
in Fig. 9. The high-temperatur e and
high-pressure oil pumped by the oil pump passes through the center of the main shaft, with most of it returning to the shell
after lubricating the rotating shaft bearing and main shaft bearing. Some of the oil, however, passes through the oil pressure
reducer in the orbiting scroll into the back pressure chamber, flowing into the suction chamber through the two notches
(A, B) in the fiXed scroll. The oil is essential for sealing the compression chamber, but the high temperature of the oil is
450
a major factor heating the intake refrigerant gas. To clarify the effect of this oil upon the compressio
n loss, the amount
of oil passing through the oil pressure reducer was varied step by step and its effect upon volumetric
efficiency was
examined. As shown in Fig. 10, the volumetric efficiency increases as the amount of oil flow1ng into
the suction chamber
is reduced. When the oil supply is reduced too far, however, the.oil sealing perfonnanc e drops and gas
leakage during the
compression process increases, thus leading to a lower volumetric efficiency. The tendency of the volumetric
efficiency
to decrease becomes greater as the operation speed decreases. Thus the oil supplying path into the suction
chamber was
examined, in addition to detennining the optimum amount of oil in the suction chamber. As shown
in Fig. 9, in the
conventional design, the oil was injected into the suction chamber through the two notches on the fixed
scroll, shown by
"A" and "B." In the conventional design, the oil passing through notch A is split to flow almost equally
into the two
compression chambers by the motion of the orbiting scroll, but the oil passing through notch B flows
almost entirely into
compression chamber B. Therefore, an imbalance in the amount of oil injected into the two compressio
n chambers appears.
As a result, there is also. an imbalance in the heating of the refrigerant by the oil, and oil sealing
is different between the
two compression chambers. On the basis of the above study, we decided not to use notch B and the
oil was injected into
the suction chamber only through notch A Thus the amount of oil injected into the suction chamber
was split equally
between the two compression chambers, and the optimization of the amount of oil was carried out again,
resulting in the
test results shown in Fig. 11. Compared with the previous design, the volumetric efficiency was improved
significantly
across the entire range of operating frequencies. The improvement during low-speed operation is especially
remarkable.
* +2 .0r-------'J,.I:.:.S-_.A......,.r=ui""re::..:m.:::e::..:;nl=s._:
*1
~90~H....,z
s
~ +1.0
"!J
!BQl
s
=
Previous model
Ql
j
~ 1.04
=·2l'
·c
ri Q)e 1.00
e
~ -1.0
-2.0!:~___.___.__...._~:--'--~.....___~
-u
Previous
-
--------"I-
"'=
1
E'o
i
8 :> 0 · 9 f\J:--'--.:t..--~.t;::--'---;tgo:;:--~120
Ql=
.E :>
·-
~Ql
.....
g
o_
tra
.0Sr---------~J~IS~-~A~r~u~ir~em~en~ts~
100
200
Oil supply amount mtio ·%
Operating speed Hz
Fig.IO Volumetric efficiency vs. percentage of oil
supplied to the compression chambers
Fig.ll Comparison of volumetric
efficiency vs. operating speed
4-2 Bearing Surface Rouglmess and Mechanical Efficiency
Observations on sliding states (see Fig. 5) suggested that
an improvement in the sliding state of the main bearing would
enhance the mechanical efficiency. For this reason, the effect of
the surface roughness of the bearing bush upon the input power
was studied experimentally. The test results are shown in Fig.
13. There is a close correlation between the surface roughness of
the bearing bush and the input power, and the mechanical loss
can be reduced by improving surface roughness. In a test
operation, when the surface roughness was under IS, no
contacting signal could be observed on the previous compressor
(see Fig. Sa), suggesting a significant improvement of the sliding
state at the main bearing.
Surface roughness S
Fig.12 Increase ofpower input vs.
surface roughness of the main bearing
5. PERFORMANCE COEFFKJENT AND RESULTS OF WSS ANALYSIS FOR 1HE NEW DESIGN
In addition to the improvements described in the previous section, the authors also reduced the discharge
loss by
refming the discharge valve and increasing the ntotor efficiency. The new design incorporated all
the improvements
described herein, resulting in the perfonnance coefficient shown in Fig. 13. The perfonnance coefficient
of the new
compressor is superior across the entire operation frequency range: about 14% improvement over the
previous compressor
at 90Hz and about 30% at 35 Hz in low speed operation. Room air conditioners are often used at low
cooling capacity,
451
in performance coefficient
under constant operation both for heating and cooling. Therefore, the significant increase
standard operation but in a
especially in the low-spee d range results not only in an improvement in efficiency under
analysis for the new design
loss
the
significant improvement in the seasonal energy efficiency ratio (SEER). The results of
also indicated. Both the
is
design
previous
the
at 90 Hz are shown in Fig. 14, in which each power loss reduction from
of the total loss.
reduction
25%
about
in
resulting
reduced,
compression and mechanical losses have been significantly
JIS-A r
uirements
1.2.--- ----=... .....,"'- """=== .
1.1
0
-~
Mechanical
1.0
Loss
~ 0.9
Previous
u
Motor Loss
0.8
0.70
30
60
90
120
Total Loss
150
Operation speed Hz
0
50
100
Fig.l4 Power loss reduction
in the new compressor
Fig.13 Performance coefficient
(JIS-A requirements : 90Hz)
6. CONCLUSION
compression loss was
Losses were reduced in the previous compact horizontal-type scroll compressor. The total
ion of the matching angle
reduced by 28% by reducing the leakage loss during the compression process through optimizat
ion of the oil supply amount
between the scrolls, by reducing the heating loss during the suction process through optimizat
The mechanical losses were
and path, and by reducing the discharge loss through modification of the discharge valve.
As a result, the total
motor.
the
refming
reduced by 24%, by improving surface roughness of the main bearing and by
90Hz.
at
14%
of
losses were reduced by 25%, resulting in an increase in compressor efficiency
7.REFERENCES
Compressor", Proc.Intemational
1. Ishii ,N.,et al., "On the Superior Dynamic Behavior of a Variable Rotating Speed Scroll
the Japan SocietY of Mechanical
Compressor Engineering Conference at Purdue,Vol.l,July,l988,pp.75-82; Transactions of
Engineers (in Japanese),Series C,Voi.55,No.517 ,1989,pp.2436-2445
Comditioners",Proc.Intemational
2. Sawai ,K,et al,"A Compact Horizontal Scroll-Type Compressor for Room Air
69-575
1992,pp5
Compressor Engineering Comference at Purdue,Vol.2,July,
Compressor Engineering
3. Kawahara ,S.,et al.,"Compact Type Scroll Compressor for Air Conditioners" ,Proc.Intemationl
Conference at Purdue,Vol.l,1990,p.140
452