SCREW ROTOR COMPRESSOR AND REFRIGERATION PLANT

Abstract: A screw compressor, especially adapted for use in conjunction with a refrigeration plant. The compressor comprises a casing with a working space generally composed of intersecting bores with parallel axes provided with spaced apart inlet and outlet ports, and a pair of intermeshing male and female rotors, each having helical lands and intervening grooves. The compressor is provided with capacity control means comprising a bleed port and a valve member having an axially displaceable control edge defining the actual size of the port and means for additional supply of working fluid at a pressure higher than the normal inlet pressure. The capacity control means is situated peripherally out of the outlet port, and the additional supply means comprises a port within the capacity control valve located at an axial distance from the control edge at least corresponding to the axial distance between two consecutive lands of the facing rotor. The compressor is further provided with separate means for selective adjustment of the outlet port.

Description: SCREW ROTOR COMPRESSOR AND REFRIGERATION PLANT The present invention relates to a screw compressor, espe¬ cially adapted for use within a refrigeration plant with- two stage throttling. Such a compressor comprises a casing with a working space generally composed of intersecting bo- res with parallel axes enclosed by barrel and end walls and provided with spaced apart inlet and outlet ports for com¬ munication with inlet and outlet channels, respectively. Rotors, each having helical lands and intervening grooves, are disposed in the bores and intermesh in pair. One rotor of the pair is of male rotor type formed such that at least the major portion of each land and groove is located outsi¬ de the pitch circle of the rotor. The other rotor is of fe¬ male rotor type formed such that at the major portion of each land and groove is located inside the pitch circle of the rotor. Two communicating groove portions thus form a chevron-shaped chamber having its base end disposed in a plane transverse to the axes of the bores and adjacent to the outlet port. The compressor is provided with capacity control means comprising a bleed port and a valve member having an axially displaceable control edge defining the actual size of the port. The compressor is further provided with means for additional supply of working fluid at a pressure higher than the normal inlet pressure. A compressor of the above type for use in a re rigeration plant is known from Swedish patent 382 663. This compressor is provided with an axially slidable valve of the type shown in Swedish patent 366 375, i.e. a valve member dispo¬ sed in the barrel wall of the working space and including the intersection line between the bores so that the valve simultaneously adjusts a bleed port at one end and the out¬ let port at the other end. This valve member is provided with openings, for additional supply of gas from the first throttling stage, situated so that at maximum capacity the openings are disposed on such a distance from the inlet end wall that they communicate with a chevron-shaped chamber first after the initiation of the compression. By this compressor construction it is possible to achieve a certain "economizer" effect even at part load. However, the construction shown in Swedish patent 382 663 involves problems and disadvantages not discussed in the patent. A valve member of the type shown decreases for a certain displacement thereof the outlet volume, i.e. the volume of the chevron-shaped chamber when opening towards the outlet port, and the inlet volume, i.e. the volume of the chevron-shaped chamber when cut off from the bleed port, with the same quantity. The ratio between the inlet volume and the outlet volume, the built-in volume ratio, thus increases with increasing bleed off. This effect is further accentuated when additional working fluid is supp¬ lied after the cut off from the bleed port. In order to avoid too high built-in volume ratios and over-compression resulting therefrom the bleed port must when opened momen- tarily mean a considerable reduction of the inlet volume eventhough this on the other hand means a low built-in vo¬ lume ratio resulting in under-compression. The possibility to vary the volumetric capacity of a compressor without too high losses in the adiabatic efficiency is thus limited. With additional supply openings in the valve member there is a further problem that those openings in order to main¬ tain a higher pressure therein than the inlet pressure must be located on a considerable distance from the edge thereof controlling the bleed port. Consequently the bleed port in such a compressor must be situated closer to the inlet end wall than in normal compressors. In other words the reduc¬ tion of the inlet volume when opening the bleed port can only be a fraction of that in earlier known compressors and consequently the range for variation of the volumetric ca- pacity will be further reduced. The object of the present invention is to provide a comp¬ ressor of the type specified where the range of capacity variation with high adiabatic efficiency is considerably extended simultaneously as the intermediate, additional supply pressure is on a steady high level required for an optimum process, which means that the over-all efficiency of the plant at part load is kept on a much higher level than that obtained by earlier used compressors. This object is met according to the invention by providing two separate valves, one controlling the bleed port for ad¬ justment of the capacity and one controlling the outlet port for adjustment of the built-in volume ratio, where the means for additional supply is located in the capacity control valve in such a position that it will never direct¬ ly communicate with the bleed port, and where the outlet port valve is selectively adjustable. A further improvement of the efficiency of the compressor may be obtained by providing the outlet port valve with in¬ jection means for a cooling and sealing liquid, said injec¬ tion means being located such that it never communicates with any of the inlet and outlet ports or with the additio¬ nal supply means, at least at low capacity reduction condi- tions. An embodiment of the invention will now be described by way of example with reference to the accompanying drawings, in which Fig. 1 shows a transverse section of a screw compressor ta¬ ken along line 1-1 in Fig. 2, Fig. 2 shows a developed view of the barrel wall of a screw compressor taken along line 2-2 in Fig. 1, Fig. 3 shows diagrammatically a refrigeration plant with two stage throttling and a compressor according to Figs. 1 and 2, Fig. 4 shows diagrammatically the built-in volume ratio of a compressor as shown in Figs. 1 and 2 as a function of the position of the capacity valve, and Fig. 5 shows diagrammatically the thermodynamical process within a plant as shown in Fig. 4. The compressor shown in Figs. 1 and 2 comprises a male ro¬ tor 10 provided with five helical lands 12 and intervening grooves 14 and a female rotor 16 intermeshing therewith and provided with six helical lands 18 each enclosing a rotor and intervening grooves 20. The rotors 10, 16 are mounted in a housing having low and high pressure end wall members and a barrel member 22 comprising two intersecting bores each enclosing a rotor. The housing is provided with an in- let port 24 at the low pressure end and extending as a re¬ lieved portion into the barrel wall and an outlet port 26 at the high pressure end. In the barrel member 22 there are further two circular grooves intersecting with the barrel wall and each enclosing an axially adjustable valve member 28, 30 sealingly cooperating with the facing rotor 10, 16. One of the valve grooves forms the outlet port 26 and a pressure adjusting valve member 28 enclosed therein is pro¬ vided with a control edge 32 defining the actual shape and size of the outlet port 26. The other valve groove is loca- ted in the wall of the bore enclosing the male rotor 10, peripherally out of the area of the outlet port 26 and en¬ closes a capacity adjusting valve member 30 provided with a control edge 34 facing the inlet port 24 and at full capa¬ city defining a part of the edge thereof, whereas at redu- ced capacity this edge defines the size and shape of a bleed port 36 communicating with the inlet port 24. The ca¬ pacity valve 30 is further provided with an opening 38 for supply of additional working fluid at an intermediate pres¬ sure higher than that of the fluid supplied through the in- let port 24. This additional supply opening 38 is located at an axial distance from the control edge 34 corresponding to the axial distance between two consecutive lands 12 of the facing rotor 10. The pressure valve 28 is similarly provided with an opening 40 for supply of a cooling and sealing liquid, which in a refrigeration compressor may consist of liquefied working fluid preferably having a tem¬ perature below the condensing temperature, which will im¬ prove the efficiency of the compressor. In Fig. 3 a refrigeration plant is shown, which comprises a compressor 42 of the type specified above, a condenser 44 connected to the compressor outlet port 26, a first thrott¬ ling valve 46 located between the condenser 44 and a flash chamber 48, which in turn is connected on the gas side to the intermediate pressure supply opening 38 of the compres¬ sor and on the liquid side to a second throttling valve 50, and an evaporator 52 connected to the second throttling valve 50 and the low pressure compressor inlet 24. In Fig. 4 the built-in volume ratio (V^) is shown as a function of the position (L) of the control edge 34 when moving from the position for maximum capacity to the posi¬ tion for minimum capacity, where a fixed edge of the axial portion may decide the size of the outlet port 26. As can be seen from the continuous line the built-in volume ratio will be kept practically constant at all capacities which means that the efficiency of the compressor can be kept on a high level over a wide field of part load capacity, si¬ multaneously as the so called "economizer" effect is main- tained at an optimum level. As a comparison are also the corresponding function for compressors as shown in Swedish patent 382 663 and in US patent 3 568 466, supplemented by a conventional capacity valve as shown in US patent 3 314 597, respectively. The conditions in a compressor according to Swedish patent 382 663, where the additional supply openings are located in a conventiional capacity valve, is shown by a dotted line. As the intermediate pressure at maximum capacity is fixed and due to the fact that the supply openings must be located on a considerable distance from the edge of the valve controlling the bleed port which in turn means that the position "a" of this edge at full load must be situated close to the low pressure end of the compressor. At the mo- ment when the bleed port is opened the inlet volume will be abruptly reduced whereas the outlet volume at the outlet port will be practically constant which means that the built-in volume ratio decreases momentarily. As the valve is moved further on towards lower compressor capacity the inlet and outlet volumes will be reduced by the same amount. This means, however, a relatively low reduction of the inlet volume percentagewise and a high reduction of the outlet volume percentagewise which means that the built-in volume ratio increases rapidly and the more the valve is moved the faster the built-in volume ratio increases up to the point where the edge of the axial portion of the outlet port takes over as deciding factor for the outlet volume. After that point the inlet volume continues to decrease, whereas the outlet volume is kept constant and consequently the built-in volume ratio starts to decrease once more. Over the range where the built-in ratio thus is higher than that at full load the working fluid will be overcompressed resulting in considerable increase of the power required and consequently a decreased efficiency. Owing to the fact that a too high* built-in volume ratio cannot be accepted the field for capacity variation will thus be rather limited simultaneously as the part load efficiency will be poor. The conditions in a capacity controlled compressor according to US patent 3 568 466, where the additional supply openings are fixed and located in the wall of the compressor housing, is shown by a point-dotted line. In order to obtain an acceptable width of the capacity variation range the bleed port controlling edge of the valve is located on a considerable distance from the low pressure end of the compressor. The position of this edge at full load is in the diagram indicated by "b" . The variation of the built-in volume ratio with the capacity will in principle be the same as that discussed above with regard to the compressor according to Swedish patent 382 663. However, as the valve control edge is located further away from the low pressure end of the compressor the abrupt reduction of the inlet volume, at the moment when the bleed port is opened, is considerably larger the drop in the built-in volume ratio will also be larger which on the other hand means that the increase of said ratio, as the valve is moved further on, will be less steep. Over a considerable capacity range the built-in volume ratio will thus be within reasonable limits from the ideal value. During the first part of the capacity range the ratio will be lower than the ideal, which means undercompression and consequently a completion of the compression will be by full back pressure compression, which means a loss and consequently a decreased efficiency. Already from this point of view the invention thus means a considerable improvement over the prior art. The thermodynamical process within the plant shown in Fig. 3 is indicated in Fig. 5 as a diagram showing the relation between the pressure (log p) and the enthalpy (i). In the diagram the different points thereof related to specific details indicated in the plant shown in Fig. 3 has been denoted by the same reference numbers. Thus the line 42 corresponds to the compressor, with its inlet port, additional supply opening, and outlet port at point 24, 38 and 26, respectively. Line 44 corresponds to the condenser, point 46 corresponds to the first throttling valve, the point 48 to the flash chamber, point 50 corresponds to the second throttling valve, and line 52 corresponds to the evaporator. The process drawn in continuous lines indicates the process at full load and at part load according to the invention. In this process the flash gas from the first stage thrott-ling is only compressed from the pressure in point 38 to point 26 which means a considerable reduction of the power consumption and thus an improvement of the process compared with a system with single stage throttling where all flash gas has to be compressed from point 24 to point 26. This is the so called "economizer effect". When using a compressor of the type shown in Swedish patent 382 663 at part load the intermediate pressure drops which in combina¬ tion with overcompression results in a change of the process as indicated by dotted lines and points 48a, 50a, 38a, and 26a. In comparison with the ideal process there are thus losses relating to the "economizer effect", owing to de¬ creased intermediate pressure, and to the compressor effi¬ ciency, owing to the overcompression. When using a compres¬ sor of the type shown in US patent 3 568 466 supplemented by a standard valve at part load the intermediate pressure drops down to evaporator pressure, simultaneously as the compressor acts with over- or undercompression. The process including undercompression is indicated by point-dotted lines and points 48b, 50b, 38b,and 26b. The invention thus results not only in an improved compres¬ sor efficiency but also when used in a refrigeration plant in an improved process efficiency, which means a considerab¬ ly higher total efficiency of the plant.

Claims: Patent Claims 1. Screw compressor, especially adapted for use in a refri¬ geration plant with two stage throttling, comprising a casing with a working space generally composed of inter¬ secting bores with parallel axes enclosed by barrel and end walls and provided with spaced apart inlet and out¬ let ports for communication with inlet and outlet chan¬ nels, respectively, and rotors, each having helical lands and intervening grooves, disposed in said bores and intermeshing in pair, one rotor of the pair being of male rotor type formed such that at least the major por¬ tion of each land and groove is located outside the pitch circle of the rotor, the other rotor being of fe¬ male rotor type formed such that at least the major por¬ tion of each land and groove is located inside the pitch circle of the rotor, whereby two communicating groove portions form a chevron-shaped chamber having its base end disposed in a plane transverse to the axes of the bores and adjacent to the outlet port, said compressor being provided with capacity control means comprising a bleed port and a valve member having an axially dis- placeable control edge defining the actual size of the port and means for additional supply of working fluid at a pressure higher than the normal inlet pressure, cha¬ racterized in that said capacity control means is situa- ted peripherally out of the outlet port, that said addi¬ tional supply means comprises a port within said capaci¬ ty control valve located at an axial distance from said control edge at least corresponding to the axial distan¬ ce between two consecutive lands of the facing rotor, and that separate means are provided for selective ad¬ justment of the outlet port. 2. Compressor as defined in claim 1, in which two spaced axial grooves are provided in the barrel wall of the working space extending from one end wall to the other, a first groove communicating with the outlet port is lo- cated within an area adjacent to the related line of in¬ tersection between said bores and encloses an axially adjustable valve member matching the barrel wall and having a control edge defining the size and position of the outlet port, the second groove communicating with the inlet port is located to the first groove and enclo¬ ses an axially adjustable valve member matching the bar¬ rel wall and sealingly passing through the end wall of the outlet port. 3. Compressor as defined in claim 2 , in which said valve member for the outlet port is provided with an opening for injection of liquified working fluid, said opening being located at an axial distance from said control ed¬ ge at least corresponding to the axial distance between two consecutive lands of a rotor. 4. Compressor as defined in claim 2 or 3, in which said first groove has at least its major portion facing the female rotor, whereas the second groove faces the male rotor. 5. Refrigeration plant with two stage throttling comprising a compressor as defined in any of the preceeding claims.

Inventors: SJOEHOLM LARS, SJOEHOLM, LARS
Applicants: SVENSKA ROTOR MASKINER AB
Application year: 1986
Main classification: F04C18
IPC codes: F04C18, F04C28