CA1051399A - Method for improving the efficiency of helical screw type compressors - Google Patents
Method for improving the efficiency of helical screw type compressorsInfo
- Publication number
- CA1051399A CA1051399A CA202,839A CA202839A CA1051399A CA 1051399 A CA1051399 A CA 1051399A CA 202839 A CA202839 A CA 202839A CA 1051399 A CA1051399 A CA 1051399A
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- oil
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- gaseous medium
- efficiency
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Abstract
A B S T R A C T
A method for improving the efficiency of helical screw rotor type compressors working on a gaseous medium of the hydrocarbon or halocarbon compound type, wherein sealing of the clearances is effected by circulating an oil through the compressor, which oil is chosen to fit said gaseous medium in such a way that said oil with said gaseous medium dissolved in said oil has a viscosity which is sufficiently high to maintain a high overall efficiency of said compressor.
A method for improving the efficiency of helical screw rotor type compressors working on a gaseous medium of the hydrocarbon or halocarbon compound type, wherein sealing of the clearances is effected by circulating an oil through the compressor, which oil is chosen to fit said gaseous medium in such a way that said oil with said gaseous medium dissolved in said oil has a viscosity which is sufficiently high to maintain a high overall efficiency of said compressor.
Description
~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 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
. 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 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.
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 . .
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 . , . .
¦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.
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.
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 . .
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 . , . .
¦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.
Claims (4)
PROPERTY OR PRIVILEGE IS CLAIMED ARE DEFINED AS FOLLOWS:
1. A method for improving the efficiency of helical screw rotor 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, said oil being chosen among available oil qualities to fit said gaseous medium in such a way that the relative capacitivity of the oil, .epsilon.roil , and that of the liquified gaseous medium, .epsilon.rgas, apply to the formula ¦ 1n .epsilon.rgas - 1n .epsilon.roil ¦ ? 1 where 1n is the natural logarithm, said oil having a kinematic viscosity, v, at 50°C that amounts to or exceeds v = 25 ? e where P1 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 if P1 is measured in kp/cm2 and u is measured in m/sec.
2. A method as claimed in claim 1, in which the vis-cosity index of the pure oil according to ASTMD 2270 is at least 90.
3. A method as claimed in claim 1 or 2 in which an oil of the synthetic hydrocarbon type is used in combination with refrigerant R22.
4. A method as claimed in claim 1 or 2 in which an oil of the synthetic polyglycol type is used in combination with refrigerant R12 or hydrocarbon gases.
Priority Applications (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| CA202,839A CA1051399A (en) | 1974-06-19 | 1974-06-19 | Method for improving the efficiency of helical screw type compressors |
Applications Claiming Priority (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| CA202,839A CA1051399A (en) | 1974-06-19 | 1974-06-19 | Method for improving the efficiency of helical screw type compressors |
Publications (1)
| Publication Number | Publication Date |
|---|---|
| CA1051399A true CA1051399A (en) | 1979-03-27 |
Family
ID=4100431
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| CA202,839A Expired CA1051399A (en) | 1974-06-19 | 1974-06-19 | Method for improving the efficiency of helical screw type compressors |
Country Status (1)
| Country | Link |
|---|---|
| CA (1) | CA1051399A (en) |
-
1974
- 1974-06-19 CA CA202,839A patent/CA1051399A/en not_active Expired
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