Note: Descriptions are shown in the official language in which they were submitted.
~i~
. j
; ~ -
~ 1051399
.. ~ .
This invention relates to a method for improving the
efficiency of helical screw type compressors working on a gaseous
medium of the hydrocarbon or halocarbon compound type, wherein
clearances present in the compressor are apositively sealed at
working conditions by means of an oil circulating through the
compressor.
The invention is particularly, but not exclusively, concerned
with a method for improving the low tip speed characteristics of
- compressors of the above-mentioned type.
Moreover, the invention is particularly, but not exclusively,
concerned with an air conditioning method employing a refrigerant,
for example the medium pressure refrigerant widely known as R12,
or the high pressure refrigerant widely known as R22, and including a
`.j helical screw compressor having oil injection facilities, an oil
~ separator, a condenser, an expansion valve and an evaporator. The
i~ invention may also be concerned with methods for storage of liquid
propane gas or pipeline transport of natural gas.
Air conditioning apparatus including screw compressors have
been widely used with satisfatorily results in relatively large re-
frigeration and air conditioning plants. However, up to now it has
not been feasible to use screw compressors in air conditioning plants
having a refrigeration capacity of less than 300 000 kcal/hour and,
in certain instances for refrigeration plants less than 100 000
kcal/hour.
This lower limit of refrigeration capacity depends upon the
specific characteristics of the screw compressor. It is well known
. ~
. ~ , .
~05~399
~ 2
,............................................................................ .
that a screw compressor is a positive type of machine in which a
certain internal leakage always takes place through clearances
existing between the rotors and surrounding casing walls. For this
reason the rotors must be run at a high peripheral speed in order
to reduce the internal leakage and achieve a sufficiently high
volumetric efficiency which is necessary for obtaining an acceptable
overall efficiency. In dry running compressors acting upon air or
other gases as the working fluid it has been found that the tip speed
of the male rotor should not fall below about 80 m/s. In compressors
acting on air or other gases and fitted with means for injecting
oil during compression, the cooling of the rotors and the housing
is improved and smaller clearances can therefore be accepted between
the relatively movable members. In such so-called wet compressors
the tip speed of the male rotor can then be reduced to about 25-30 mts
with retained satisfactory overall efficieny. Owing to the fact that ~-
refrigeration compressors are usually directly driven by electric -~
motors, the maximum speed of which is normally 3000 rpm or 3600 rpm
depending on the frequency of the current, the diameter of such
compressors is normally no less than 160 mm, which compressor size
corresponds to the minimum refrigeration capacity 100,000 - 300,000
kcal/hour mentioned above.
One particular application of an air conditioning apparatus
in which it has hitherto not been suitable to use a screw compressor ~
is in automotive air conditioning plants, where the required refri- ~ -
geration capacity is in the region of 3,000 kcal/hour. A solution
of the performance problems at low tip speeds, that could make it
possible to use the screw compressor for this purpose, would be
highly desirable, owing to the small bulk, the low weight and the
vibration free operation of the screw compressor compared with
1051399
conventional reciprocating piston compressors now used for such
purposes. The underlying reasons for the unsuitability of the screw
compressor in automotive air conditioning installations are, on
one hand, that the refrigeration capacity required is so small and
amounts only to about 3 - 1 % of the normal minimum capacity, re-
ferred to previously, and, on the other hand, that the compressor
in such a plant is normally driven from the engine of the car via
a belt transmission, which means a very low compressor speed at
motor idling.
The capacity of the compressor will of course increase with
the speed. However, the capacity needed for cooling the air in a
car has to be available already at low engine speed. This means
that the compressor will give more refrigeration capacity than
needed at higher engine speeds. This is aproblem which is common
for all types of compressors, but use of the screw compressor offers
a very attractive way of regulation in that a slide valve or other
efficient means for capacity control may be used, which is an addi-
tional reason for the suitability of the screw compressor for use
in automotive air conditioning systems.
Assuming a step up gear ratio between the compressor input
speed and the engine speed of 2:1, the compressor input speed
at 1400 rpm engine speed will be 2800 rpm. If the refrigerant R12
is used and if the refrigeration capacity demand at this speed
is 3,000 kcal/hour at 70C condensing temperatur and 0C evaporating
temperature the requisite compressor displacement volume to obtain
this refrigeration capacity is round 15 m3/h, which corresponds to
a rotor diameter of round 50 mm, giving a tip speed at 2800 rpm of
round 11 m/s.
lOS1399 "
However, in an automotive air conditioning apparatus the
compressor has to operate at engine speeds from 700 rpm to 7000 rpm.
This means a compressor input speed of 1400 to 14000 rpm, corre-
sponding to a male rotor tip speed of 5.5 m/s to 55 m/s. The
corresponding tip speeds for the female rotor are 3.7 to 37 m/s.
In this example we have assumed a "4 + 6 lobe combination" and
"female rotor drive". This means that the male rotor has 4 lands
and the female rotor 6 grooves and that the compressor is driven -
on the female rotor. The speed ratio between the male rotor speed
and the female rotor speed is consequently 1.5 ; 1.
On cars fitted with automatic transmission the engine speed
seldom exceeds 3000 rpm, corresponding to the male rotor tip speed
of 23 ID/S. This means that the compressor will operate at tip
speeds below 20 m/s 95 % of the time.
We have now unexpectedly found that by circulating oil of
a special quality in the compressor, which oil is limited dissoluble
in R12 and which has a considerably higher viscosity, compared to
other refrigeration oils used for such purposes, the volumetric
efficiency of the compressor and thereby the capacity of the compres-
sor will show a decided improvement. At the same time the compressor
input torque and thereby the compressor input power will remain
equal or even decrease. This means that the overall efficiency of
the compressor will increase to a level which is acceptable for
use in automotive air conditioning plants as well as in other air
conditioning and refrigeration plants. After further investigations
of this amazing effect we have found general and optimal relations
between the oil quality, the oil viscosity, the gaseous media on
which t-he compressor is working, the compressor (male rotor) tip
speed and the actual working conditions. By using these relations
for compressor- plants, operating on hydrocarbon or halocarbon
10513g~
compounds, it is possible, in air conditioning plants, to use
the screw compressor for capacities down to round 1,000 kcal/
hour and/or 5 m3/h and generally to improve the overall
efficiency of oil injected screw compressors of all sizes
operating on halocarbon or hydrocarbon compounds to a superior
efficiency level.
According to the invention this improved efficiency
is obtained by the method defined above being characterized -
in that said oil is chosen among available oil qualities to
fit said gaseous medium in such a way that the relative
capacitivity of the oil, ~r ~ and that of the liquified
gaseous medium, Er ~ apply to the formula
gas rOil
where ln is the natural logarithm, said oil having a kinematic
viscosity, v, at 50C that amounts to or exceeds
c p
v = 25-e u
where Pl is the discharge pressure of the compressor, u is
the tip speed of the male rotor, e is the base of the natural
system of logarithms and c is a constant equal to l cm m
2 kp sec
if Pl is measured in kp/cm and u is measured in m/sec.
Thus, the invention eliminates the use of very small clearances,
very small rotor diameters in com~ination with internal step up
gears and other unsuitable measures, previously suggested in
order to improve the overall efficiency of helical screw
compressors. - ~
Preferably the viscosity index of the oil (according -
to ASTM D 2270) is at least 90 so that effective sealing is
maintained at working temperature up to at least 150C. The
kinematic viscosity may be higher than the value obtained
from this formula especially when the difference defined in
the first mentioned formula is close to l.
1051399
The invention will be further described with
reference to the accom~anying drawings in which:
Fig. la is a graph of volumetric efficiency in a
helical screw rotor type compressor in an air conditioning
apparatus using refrigerant R12 operating at 40C condensing
temperature and 0C evaporating temperature for various
refrigerant oils against compressor tip speed and RPM of the
male rotor.
Fig. lb is a similar graph to Fig. la for input torque.
Fig. lc is a similar graph to Fig. la for coefficiency
and performance.
Fig. ld is a similar graph to Fig. la for coefficient
of performance.
Figs. 2a through 2d are similar graphs respectively
- to Figs. la and ld for refrigerant oils at its 70C condensing
temperature and 0C evaporating temperature.
Fig. 3 is a graph of volumetic and adiabatic
efficiency for two compressors operating at 60C condensing
' temperature and 0C evaporating temperature and refrigerant
R12 with various refrigerant oils with tip speed of the rotor
and
, Figs. 4a and 4b are graphs of kinematic viscosity
of the various refrigerant oils with the tip speed of the
rotor of the compressor using refrigerant R12 and R22
respectively.
In the drawings and the description relative thereto
the oils referred to are as follows:
Oil Viscosity Indexv(50C) cSt~r(50C)
A 154 360 2.1
B 241 278 5.7
C 148 41 2.1
D 14~ 114 2.1
--6--
. .
10513~
The viscosity defined by the above mentioned formula
(and shown in Figure 4a and 4b) will give an overall efficiency
at any condition which is maximuD 10% lower than the optimal
overall efficiency achievable.
According to the invention an oil of the synthetic
hydrocarbon type is preferably used in combination with
refrigerant R22 and an oil of the synthetic polyglycol type
is preferably used in combination with refrigerant R12 or
hydrocarbon gases.
An example of a synthetic hydrocarbon oil is Mobil
SHC oils and the improved efficiency obtained by using one
type of thi.s oil in combination with refrigerant R12 is shown
in Figure 1 (condensation temperature 40C, oil A) as compared
with a naphtenic mineral oil, Mobil Gargoyle Arctic 300.
An example of a synthetic polyglycol oil is Mobil
Glygoyle oils and the improved efficiency obtained by using
one type of this oil in combination with refrigerant R12
is shown in Figure 2 (condensing temperature 70C, oil B)
as compared with a naphtenic mineral oil, ~obil Gargoyle
Arctic 300.
The use of SHC and Glygoyleoils has a marked effect on
-6a-
- . ~ ,. . .
~051399
the volumetric efficiency (nvOl), the total adiabatic efficiency
(nad ) and the coefficient of performance (COP) as shown in
Figures la, 2a; lc, 2c and ld, 2d respectively. The input torque tT)
is equal to~as shown in Figure lb, or lower than,as shown in Figure
2b, the torque obtained with a standard refrigeration oil of the
naphtenic base type such as Mobil Gargoyle Arctic 300.
The improved volumetric (~vol) and adiabatic (nad
efficiencies are also shown in Figure 3 for two compressors,
operating at 60C condensing temperature, namely compressor A
(rotor diameter = 47 mm, rotor length (L)/diameter (D) ratio
= 1.7, Vs = 0.0825 l/rev), using a polyglycol oil (FV~ 62~1~7 J~
and compressor B (rotor diameter = 102 mm, L/D ratio = 1.0,
Vs = 0.516 l/rev), using a standard mineral oil (Arctic 300).
It is believed that injection of the polyglycol oil into compressor
B would give even better results than that indicated for compressor
A due to the scale factor, and consequently, the improvement would
be even greater than that shown in Figure 3.
The preferred method defined by the formula
¦ln r ln r . ¦ 1
gas oll
is illustrated by following examples.
It is possible, according to the formula, to combine R22
(er = 6.00) with SHC oil (r = 2.10) but not with Arctic 300
(r = 2-.30) because
¦l`r. 6.00 - ln 2.10¦ = 1.05 ~ 1 and
lln 6.00 - ln 2.30¦ = 0.96 < 1, respectively.
Similarly, it is possible to combine R12 (r = 1.80) with
_F~r ~12~ 1) (r = 5-70), and propane (r = 1.30) with Mobil
Glygoyle 30 (r = 5-00) because
. ,
1051399
. .
¦ln 1.80 - ln 5.70 ¦ = 1.15~ 1 and
lln 1.30 - ln 5.00 1 = 1.35~1, respectively,
but not with paraffinic mineral oils such as Mobil Arctic 300,
because
¦ln 1.80 - ln 2.30 ¦ = 0.25< 1
¦ln 1.30 - ln 2.30 ¦ = 0.57< 1.
As mentioned above an oil is suitably chosen having, -
additionally or as an alternative, a kinematic viscosity v that
must not drop below a value obtained from the formula
C-Pl : :
v = 25 e u
This formula is illustrated in Figures 4a and 4b for refrigerants
R12 and R22. The pressure curves corresponding to different con-
densating temperatures are shown as well as the viscosity values
of some oils.
From Figures 4a and 4b it is evident that common mineral
oils such as Mobil Gargoyle Arctic 300 are excluded.
With apositive sealing, in contrast with positive sealing,
such as that effected by the piston rings of a positively sealed .
reciprocating piston compressor, volumetric efficiency of the -~
machine is dependent upon the extent of the pressure rise in
the compression chambers during any one cycle, or in other words,
the value of the compression ratio, since the leakage from the
apositively sealed compression chambers will obviously increase
with increase in the pressure rise in a single stage.
