Patent Publication Number: US-2013230415-A1

Title: Reciprocating compressor with high freezing effect

Description:
CROSS REFERENCE TO RELATED APPLICATIONS 
     This application is a United States National Phase application of International Application PCT/IT2010/000129, the entire contents of which are incorporated herein by reference. 
    
    
     FIELD OF THE INVENTION 
     The present invention relates to a reciprocating compressor, in particular, although not exclusively, of the hermetic or semi-hermetic type for cooling plants, for example refrigeration plants. 
     BACKGROUND OF THE INVENTION 
     The use of hermetic or semi-hermetic compressors is generally known to feed a refrigerating circuit of cold store facilities comprising a plurality of refrigeration units (such as for example displayers, cold-storage rooms, refrigerated cabinets) for different types of products that must be stored according to different climatic conditions. These facilities can differ also substantially in the required refrigeration capacity and in engineering complexity according to the particular use requirements. 
     Hermetic or semi-hermetic reciprocating compressors comprise, in their main characteristics, a case, inside which an electric engine is housed to move a drive shaft connected to a plurality of pistons sliding in compression cylinders suitable to compress a coolant to perform the refrigerating cycle. 
     Compressors for compression cycles have been developed to improve the yield of the refrigerating cycle, which provide for sucking a first quantity of coolant in the suction chamber and then injecting in the cylinders an additional quantity of coolant at one or more pressure levels intermediate between the suction pressure and the delivery pressure. The refrigerating cycles operating according to this general principle are called “Voorhees cycles” or “steam multiple-compression refrigerating cycles”. 
       FIG. 1  shows a traditional pressure-volume curve, where on the axis of abscissas the position of a compression piston head or crown is indicated, measured from the bottom dead center (BDC), and on the axis of ordinates the pressure inside the cylinder is indicated. In particular, a typical pressure-volume cycle is shown there, where the quadrilateral ABCDA represents the typical trend of a traditional reciprocating compressor, where the segment AB corresponds to the piston expansion phase, the segment BC corresponds to the suction phase, the segment CD corresponds to the compression phase and the segment DA corresponds to the discharge phase. In  FIG. 1  the quadrilateral ABGEA represents also the typical trend of a “Voorhees cycle”, wherein the segment AB corresponds to the expansion phase, the segment BG to the suction phase, the segment GE to the compression phase, and the segment EA to the discharge phase. The ABGEA cycle differs from the ABCDA cycle mainly at the end of the suction phase and in the compression phase CF, as the pressure is increased, as well as the quantity of fluid in the cylinder thanks to the injection of additional fluid. 
     Theoretically, a compressor operating according to a multiple compression cycle therefore increases, with the same construction characteristics and boundary conditions, the refrigerating capacity and the energy efficiency relative to a compressor operating according to a traditional cycle. 
     U.S. Pat. No. 1,821,248, GB 28031 A A.D. 1910, GB190504448, and GB 793864A , all by Voorhees, describe some types of compressors using this “Voorhees Cycle”. In particular, U.S. Pat. No. 1,821,248 and GB 28031 A A.D. 1910 describe systems for modifying a single-stage compressor and adapting it for operation on refrigerating plants with more evaporation pressure levels. GB190504448 and GB793864A describe multiple effect compressors, which provide for an injection system for injection through one or more auxiliary ports controlled by the movement of the main piston and actuated by a rotary valve moved by a gear drive system that takes the motion from the main shaft. 
     These multiple effect compressors have been widely used in the first decades of the past century in big ammonia refrigeration plants; those machines operated with very low rotation speeds, about 100 rpm or even less; the quantity of fluid at intermediate pressure that was possible to inject at these speeds was high, and therefore refrigerating capacity and energy efficiency increase was significant. On the contrary, the modern refrigerating compressors operate at high speeds, typically at  1450  rpm and above, and this entails a high translation speed of the piston inside the cylinder; applying the solutions indicated by Voorhees in the previously mentioned patents on compressors operating at high rotation speeds leads to an extreme reduction of the quantity of fluid at intermediate pressure which can be injected in the cylinder. 
     As a result, the segment GE in  FIG. 1 , representing the compression phase, is very similar to the segment CD related to a traditional compressor, and this significantly reduces the advantages deriving from the refrigerating capacity and energy efficiency increase. In addition to this, these compressors are very complex from a mechanical point of view, and, with the high rotation speeds of the modern compressors, they are somewhat unreliable, difficult to be used and maintained. 
     Substantially, compressors for refrigerating plants operating according to the multiple effect refrigeration cycles, or Voorhees cycles, currently are not widely used, due to the above mentioned difficulties. 
     Therefore, despite the developments of technology the problem currently exists and there is the need for reciprocating compressors, which are more versatile and efficient to be used than the current ones, and which are at the same time sufficiently reliable. 
     SUMMARY OF THE INVENTION 
     According to an aspect, the object of the present invention is to realize some improvements to a reciprocating compressor so that it is more efficient, simple and economical to be constructed and used than the current compressors, overcoming, completely or partially, one or more of the above mentioned disadvantages. 
     These objects and advantages are substantially obtained with a compressor as claimed in claim  1 . Characteristics and particularly advantageous embodiments of the present invention are indicated in the dependant claims. 
     Practically, the invention provides for a reciprocating compressor for a coolant, comprising at least one piston slidable in a compression chamber, actuated by a drive shaft, wherein a supply ports exits in the compression chamber for feeding a coolant at a pressure intermediate between the delivery pressure and the suction pressure of the compressor, and wherein to the supply port a distribution system is associated, to control the supply of the coolant in synchronous manner with the position of said piston. 
     In some embodiments, the distribution system comprises a mechanical distributor, which opens and closes an aperture for supplying the coolant at intermediate pressure towards the compression chamber. In some embodiments the distributor is slidable in a distribution chamber in a synchronous manner with the piston or pistons of the compressor. The distributor can be associated to an arrangement of ducts corresponding, in number and position, to the number and to the reciprocal phase of the pistons of the compressor, so as to control opening and closing of the respective supply ports in the compression chambers. 
     With a distributor slidable in a distribution chamber it is possible to obtain opening and closing of the supply port for the coolant at the intermediate pressure in a manner synchronized with the position of the piston, and therefore the fluid at the intermediate pressure is injected in the compression chamber in a phase of the compression cycle wherein the fluid under compression has a pressure compatible with that of the fluid coming from the supply port. The system is efficient and reliable also at high rotation speeds that are typical of the modern compressors of the refrigerating plants. 
     The distributor can be actuated through a connecting system for cinematic connection to the drive shaft, to obtain, simply and reliably, timing between the distributor and the position of the piston or pistons. In some embodiments the cinematic connection can comprise a cam associated with the drive shaft and a tappet associated with the distributor. The tappet can be provided with a spring, which maintains it into contact with the cam. 
     In other preferred embodiments of the invention, the cinematic connection between drive shaft and distributor can be obtained with a rod-crank mechanism. 
     In possible embodiments, the distributor presents at least one passage duct to put the distribution chamber into fluid connection with the compression chamber or chambers the respective supply port or ports. 
     Advantageously, the distributor is suitable to open gradually the supply port alternatively in both the directions of its stroke when the passage duct crosses the supply port or passes near it; in this way it is possible to put the distribution chamber into fluid communication with the compression chamber. The distributor furthermore closes the supply port alternatively in both the directions of its stroke thanks to its outer surface or shell. 
     In an advantageous embodiment of the present invention, the passage duct presents at least one entrance aperture on the head of the distributor and at least one exit aperture arranged at an intermediate height on the shell. 
     Further embodiments are also possible according to particular use requirements, for example it is possible to obtain one or more exit apertures to put more fluid in a same compression chamber or to inject fluid in different compression chambers, or other else. 
     In an advantageous embodiment of the invention, the exit aperture of the passage duct is arranged in a compartment obtained on the side shell of the distributor, and in this compartment a shaped seal sliding block can be inserted, slidable in radial direction and presenting a through channel in correspondence of the exit aperture. This seal sliding block is designed for being pushed against the wall of the distribution chamber by the fluid distribution pressure, so as to increase at least partially the seal, decreasing the clearance and the leakage of the fluid secondary flows between the distributor and the same chamber. 
     In a particular embodiment of the invention, a backflow valve can be provided in the passage duct of the distributor, in this way the mechanical complexity and the cost of the system is increased, but the probability of a backflow of the fluid at high pressure decreases. 
     When the distributor is actuated by the drive shaft controlling the piston or pistons, preferably the distributor and the piston are mutually out phased so as to open gradually the supply port when the piston is near the bottom dead center or is in the compression phase, to avoid fluid at the distribution pressure being fed during the piston suction phase, but it is also possible to feed the fluid at the distribution pressure also during the suction phase, even if with a decreased energy efficiency. 
     In a possible embodiment of the invention, the compressor is of the two- cylinder type with a pair of compression pistons slidable in respective compression chambers, each presenting at least one supply port for the fluid connection with at least one distribution chamber, where the distributor is arranged. This latter is preferably arranged between the two compression chambers. 
     In this case the distributor passage duct is shaped so as to feed the fluid in the two compression chambers according to the timing of the two pistons and it can be obtained according to different conformations based upon particular use or construction requirements of the compressor. This duct can be obtained for example with a single entrance aperture on the head of the distributor and a plurality of exit apertures corresponding to each supply port, or more exit apertures can be provided for each supply port. 
     These two pistons can be mutually out-of-phase and the distributor can be out-of-phase relative to them by an intermediate angle equal to nearly the half of the timing thereof. 
     Clearly, a different number of pistons and/or distributors can be provided, arranged in various manner (in-line, V-shaped, or other else) in a compressor according to the requirements of the plant to be supplied. 
     An advantage of some embodiments of the compressor according to the present invention is the fact that it has high energy efficiency and refrigerating capacity, as the compressor operates a greater quantity of fluid with less leakages. 
     A further advantage is that it is possible to reduce the end compression temperature of the fluid, as the temperature of the injected fluid at distribution pressure is lower than the temperature of the fluid in the compression chamber; in some use conditions it is therefore possible to use a one-stage compressor according to the present invention instead of a traditional two-stage compressor, thus obtaining a considerable saving both in construction and maintenance 
     The various features of novelty which characterize the invention are pointed out with particularity in the claims annexed to and forming a part of this disclosure. For a better understanding of the invention, its operating advantages and specific objects attained by its uses, reference is made to the accompanying drawings and descriptive matter in which preferred embodiments of the invention are illustrated. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       In the drawings: 
         FIG. 1  is a pressure- volume diagram, on which a traditional refrigerating cycle, a multiple effect refrigeration cycle or Voorhees cycle, and a cycle according to the present invention are indicated; 
         FIG. 2  is a view in vertical section of a reciprocating compressor according to an embodiment of the present invention; 
         FIG. 3  is an enlarged sectional view of the distributor of the compressor of  FIG. 2 ; 
         FIG. 4  is a side view of the distributor of  FIG. 3 ; 
         FIG. 5  is a side view of a piston of the compressor of  FIG. 2 ; 
         FIG. 6  is an enlarged view of the head of the compressor of  FIG. 2  in one position during the operation cycle; 
         FIG. 7  is an enlarged view of the head of the compressor of  FIG. 2  in another position during the operation cycle 
         FIGS. 8 to 10  are diagrams of the timing of some components of the compressor of  FIG. 2  according to some embodiments of the present invention; 
         FIG. 11  is a view of a refrigerating circuit, in which the compressor of  FIG. 2  is used; 
         FIG. 12  is a semi-logarithmic enthalpy-pressure diagram referred to the refrigerating circuit of  FIG. 11 ; 
         FIG. 13  is a view of a further refrigerating circuit, in which the compressor of  FIG. 2  is used; and 
         FIG. 14  is a semi-logarithmic enthalpy-pressure diagram referred to the refrigerating circuit of  FIG. 13 . 
     
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     In the drawing, a compressor according to an embodiment of the present invention is of the semi-hermetic type and is indicated with number  1  (see  FIG. 2 ) and comprises a casing or case  1 A, in which an engine compartment  1 B is obtained, inside which an electric motor  2  is arranged and connected mechanically to a crank drive shaft  3 . 
     A second compartment  1 C, adjacent to the engine compartment  1 B and divided from it by means of a dividing wall  1 D, is shaped at the bottom so as to produce an oil pan for the lubricant L, in fluid connection with the lower part of the engine compartment  1 B through a passage hole  1 E obtained on the diving wall  1 D. The lubricant L is centrifuged by the disk  1 F integral with the drive shaft  3 , passes from the pan of the compartment  1 C to the tray  1 G, and through the hole  1 H obtained in the drive shaft  3  lubricates the bronze bushes  1 M and  1 N and the connecting rods  9 A,  11 A, and  19 A. 
     In the second compartment  1 C two compression chambers  5  and  7  are obtained, inside which two compression pistons  9  and  11  respectively are housed in a slidable manner, connected to the drive shaft  3  by means of a connecting rod  9 A and  11 A respectively. The pistons  9 ,  11  are suitable for compressing a coolant R in a refrigerating machine or system, in which the compressor is inserted. 
     The compression chambers  5  and  7  end with an upper head  6 , to which are associated a suction system  13 , to supply the coolant R to each of them, and a delivery system  15 , to send the compressed fluid R in the outer refrigerating machine. 
     In this embodiment, the suction system  13  is provided with inlet compartments  13 A and  13 B obtained in the head  6  and presenting respective inlet valves  13 C and  13 D to make the fluid selectively enter in the compression chambers  5  and  7  respectively, see also the description with reference to  FIGS. 6 and 7 . 
     The coolant R at a suction pressure P is supplied to the compressor  1  from a side entrance  1 L in the engine compartment  1 B, passes in an entrance  13 E of a inlet duct and exits from exits  13 F and  13 G of such duct into the inlet compartment  13 A and  13 B respectively. 
     The delivery system  15  is provided with delivery compartments  15 A and  15 B obtained in the head  6  and presenting respective delivery valves  15 C and  15 D to open selectively the exit for the fluid from the compression chamber  5  and  7  respectively. 
     It is clearly apparent that this compressor  1  is described just by way of example, as it may be of any other type suitable to the purpose, for instance it can present a different number of compression chambers arranged in various manner, inline or V-shaped, with one or more drive shafts. 
     According to the present invention a distributor  19  is provided, slidable in a distribution chamber  21  and presenting at least one passage duct  23  (see  FIG. 3 ) inside itself, to put into fluid connection the upper part of the distribution chamber  21  with each compression chamber  5  and  7  through a side supply port  25  and  27  respectively. In the embodiment shown in  FIG. 2  the distribution chamber  21  is arranged between the two compression chambers  5  and  7 ; it is however clearly apparent that this distribution chamber  21  can be obtained and arranged in different manners according to particular construction or use requirements. 
     An inlet system  14  supplies the coolant R in the distribution chamber  21  at a distribution pressure P 2  greater than the suction pressure P 1 , as described in greater detail hereunder. 
       FIG. 3  shows an enlarged section of the distributor  19 , where it should be noted in particular that it presents a substantially cylindrical shape; it is clearly apparent that the shape and the dimensions of the distributor  19  represented herein are not limiting, as it can be produced in any manner useful to the purpose. 
       FIG. 3  furthermore shows the passage duct  23 , which presents an entrance aperture  23  A on the head  19 T and two exit apertures  23 B and  23 C at different heights on the side shell  19 M of the distributor  19 , so as to put into fluid communication the upper part of the distribution chamber  21  with the supply port  25  and respectively  27  (and therefore with the compression chambers  5  and  7 ) according to the position of the distributor  19 , as described hereunder. 
     In the shown embodiment, the passage duct  23  is obtained with a channel  23 D inside, and nearly coaxial with, the distributor  19 , from which the exit apertures  23 B and  23 C extend radially. 
     In this way the distributor is particularly simple and inexpensive to be produced; it is also possible that the passage duct  23  is obtained in a different manner according to particular construction or use requirements, such as for instance with two or more inner channels and/or two or more entrance and/or exit apertures. Furthermore, it is also possible to provide for two or more supply ports  25 ,  27  arranged horizontally, vertically or in other manner for a compression chamber or for more compression chambers according to particular requirements. 
     In a particularly advantageous embodiment, the exit apertures  23 B and  23 C exit in respective compartments  19 A and  19 B obtained on the side shell  19 M of the distributor  19 , in which can be inserted shaped seal sliding blocks  29  slidably in radial direction, presenting a through channel  29 A close to the exit aperture  23 B and  23 C. These sliding blocks  29  are pushed by the distribution pressure P 2  of the coolant R against the wall of the distribution chamber  21 , so as to decrease the clearance, the leakage and the secondary flows of the coolant R and to increase the seal between the distributor  19  and the chamber  21 . 
     Advantageously, the through channel  29 A of each seal sliding block  29  can present a diameter slightly smaller than that of the corresponding exit aperture  23 B and  23 C, so as to hinder the backflow of the coolant R and to increase the thrust on the seal sliding block. 
     In  FIG. 3  it should be furthermore noted that the distributor  19 , in this advantageous embodiment, provides for at least one upper seat  19 C for a dry seal obtained between the head  19 T and the seal sliding blocks  29 , and at least one lower seat  19 D for another seal obtained below the seal sliding blocks  29 . 
     In this embodiment, the distributor  19  is actuated by the drive shaft  3  and is connected to it by means of a connecting rod  19 L (see  FIG. 2 ) using a plug, inserted in a through hole  19 S below the lower seat  19 D; however, it is clear that the distributor  19  can be actuated by any other suitable mechanism, provided that in synchronous manner with the motion of the pistons  9 , 11 . 
       FIG. 4  shows a side view of the distributor  19  of  FIG. 3 , where it should be in particular noted the upper and lower seats  19 C and  19 D, the through hole  19 S and one of the sliding blocks  29  of substantially rectangular shape and slightly concave, to follow the perimeter of the shell  19 M. 
       FIG. 5  shows the compression piston  9  (completely similar to the piston  11 ), which presents a head  9 T and three upper seats  9 S near its head  9 T for respective scrapers, a through hole  9 F for a plug of the connecting rod  9 A and, below the hole  9 F, a lower seat  9 B for a lower seal. 
       FIGS. 6 and 7  show a sectional enlargement of the head  6  of the compressor  1  of  FIG. 2  in two different positions during the functioning cycle, wherein it should be noted in particular the suction system  13 , comprising the inlet compartments  13 A and  13 B with the inlet valves  13 C and respectively  13 D to make the fluid selectively enter in the compression chamber  5  and respectively  7 , and the delivery system  15 , comprising a delivery compartment  15 A and  15 B with the delivery valves  15 C and respectively  15 D to open selectively the exit for the fluid from the compression chamber  5  and respectively  7 . 
     Advantageously, the inlet system  14 , for supplying the coolant in the distribution chamber  21  at a distribution pressure P 2  greater than the suction pressure PI but lower than the delivery pressure P 3 , comprises a distribution compartment  14 A obtained in the head  6 , this compartment presenting a first entrance aperture  14 B for the fluid connection with the distribution chamber  21  and a second entrance aperture  14 C for the fluid connection with an external high pressure supply circuit, see the description below with reference to  FIG. 11 . 
     It should be noted that the first and second aperture  14 B and  14 C of the distribution compartment  14  do not present valves, but they are always open. 
     In  FIGS. 6 and 7  it should be furthermore noted that the supply ports  25  and  27  present entrances  25 A and  27 A arranged in intermediate positions along the stroke of the distributor  19  in the distribution chamber  21 . The exits  25 B and  27 B are arranged in intermediate positions between the positions taken by the head or crown of the pistons  9  and  11  corresponding to the positions of bottom dead center (BDC) and of top dead center (TDC). 
     In the illustrated embodiment, the compression chambers  5  and  7  and the distribution chamber  21  are arranged adjacent to each other substantially at the same height and present a similar shape, and the supply ports  25 ,  27  open (entrances and exits  25 A,  27 A, and  25 B,  27 B) at an intermediate height in the respective compression chambers of the distributor. 
     It is clear that the entrances  25 A,  27 A and the exits  25 B,  27 B of the supply ports  25 ,  27  can be in different number, shape, and dimensions, and they can be furthermore arranged at different heights in the chambers  5 ,  7 , and  21  according to particular construction or use requirements; the position and the number of the compression and supply chambers  5 ,  7 , and  21  can vary too, according to particular requirements, as explained above. The supply ports  25 ,  27  can be obtained, for instance, at different heights from each other, whilst the exit apertures  23 B and  23 C of the passage duct  23  can be obtained at the same height. 
       FIG. 6  shows a configuration, wherein the piston  9  is substantially at the bottom dead center and is going to move upwards (arrow F 1 ); the exit  25 B of the supply port  25  is open, whilst the entrance  25 A is going to be put into communication with the distribution compartment  14 A through the duct  23  of the distributor  19  moving downwards (arrow F 2 ). 
     Instead, the piston  11  is substantially at the top dead center, the exit  27 B of the supply port  27  is closed by the shell of the piston  11 , whilst the entrance  27 A is closed by the shell of the distributor  19  which is going to move downwards (arrow F 3 ). 
       FIG. 7  shows the position of the pistons and of the distributor after a 180° turn of the drive shaft  3  relative to that of  FIG. 6 , wherein the piston  9  is substantially at the top dead center and is going to move downwards (arrow  4 ), the exit  25 B of the supply port  25  is closed by the shell of the piston  9 , whilst the entrance  25  A is closed by the shell of the distributor  19 . 
     Instead, the piston  11  is substantially at the bottom dead center and it is going to move upwards (arrow F 6 ), the exit  27 B of the supply port  27  is open, whilst the entrance  27 A is going to be completely opened by the distributor  19  moving upwards (arrow F 5 ). 
       FIG. 8  shows a diagram, in which on the axis of abscissas the crank angle is indicated and on the axis of ordinates the distance of the piston  9  (curve B 9 ), of the piston  11  (curve B 11 ) and of the distributor  21  (curve B 21 ) from the respective bottom dead center is indicated. In this diagram it should be noted in particular that the pistons  9  and  11  are advantageously mutually displaced by 180°, whilst the distributor  21  is displaced by 90° relative to each of them. 
     It is clear that this displacement can be varied according to the number of pistons and/or distributors or according to particular construction and use requirements of the compressor and of the refrigerating plant. 
       FIG. 9  shows a diagram, in which on the axis of abscissas the crank angle is indicated and on the axis of ordinates the passage opening or area (in square millimeter, mm) of the entrance  25 A and of the exit  25 B of the supply port  25  as a function of the position of the aperture  23 B on the distributor  21  and respectively of the head or crown of the piston  9 . The trend of the net flow sections or passage areas are represented by the curves U 25 A and U 25 B respectively for the entrance and the exit  25 A and  25 B. A similar diagram can be drawn with reference to the compression chamber  7  as a function of the given phasing angle. 
     In the case of the graph of  FIG. 9 , the entrance  25 A is obtained in an intermediate position relative to the stroke of the distributor  19  in the chamber  21 , whilst the exit  25 B is arranged so as to be completely uncovered when the piston  9  is in the bottom dead center. 
     It should be noted in particular that the exit  25 B starts to open (point N 1  of the curve U 25 B with a nearly 130° crank angle) when the piston  9  is sliding towards its BDC, and it opens completely when the piston  9  has achieved the BDC (point N 2  at about 180°). By furthermore increasing the crank angle, the piston  9  starts its stroke upwards towards the TDC, gradually closing the exit  25 B, which remains closed from the point N 3  at about 230° until the subsequent downstroke of the piston. 
     Instead, the entrance  25 A starts gradually to open (point N 4  of the curve U 25 A at about 180°) close to the exit  23 B of the distributor  21 , which slides towards its BDC, it completely opens (point N 5  at about 210°) and then gradually closes (point N 6  at about 270°). 
     The injection of fluid from the distribution chamber  21  to the compression chamber  5  therefore occurs between the crank angle of about 180° and the crank angle of about 230°, i.e. when both the passage opening of the entrance  25 A and that of the exit  25 B of the port  25  are at least partially open. The net passage opening is maximum at the point N 7 . It should be noted that the entrance section  25 A starts again to open gradually (point N 6  at about 270°), it opens completely again (point N 8  at about 330° and closes gradually (point N 9  at about 360°). However, during this second opening phase the exit opening  25 B is null, as it is completely closed by the shell of the piston  9 . The seal segment mounted in the lower seat  9 B (see  FIG. 5 ) limits the entity of fluid leakages between the shell of the piston  9  and the walls of the compression chamber  5 . 
       FIG. 10  shows a further diagram, in which the axis of abscissas and the axis of ordinates show the same variables as the diagram of  FIG. 9 . The difference from the diagram of  FIG. 9  is that the exit  25 A is obtained in the chamber  5  in an intermediate position relative to the stroke of the piston  9 , and therefore the exit  25 B remains open longer. This substantially corresponds to the configuration of the compressor of  FIGS. 2 ,  6 , and  7 . Furthermore, on the diagram of  FIG. 10  the opening and closing curves are indicated of the entrances and the exits  25 A,  25 B and  27 A,  27 B, i.e. for both the pistons  9  and  11 . The curves are indicated with the letter U, followed by the reference of the opening (ex.  25 A;  25 B) to which they refer. The curves for the openings  25 A,  25 B are represented with a continuous line, whilst the curves for the openings  27 A,  27 B are indicated with a dashed line. 
     In particular (and with initial reference to the compression chamber  5 ), according to this configuration the piston  9  starts to open the exit  25 B (point M 1  of the curve U 25 B with a crank angle of about 120°) during its stroke downwards from the TDC. Subsequently, the piston  9  continues to go down, completely opening the exit  25 B (point M 2 I at about 130°). By further increasing the crank angle, the piston  9  continues its descending stroke towards the BDC and the exit  25 B remains open. The piston  9  arrives at the bottom dead center, and then it inverts its stroke until it meets again the exit  25 B, closing it gradually (point M 2 II at about 220° until M 3  at about 240°). 
     Instead, the entrance  25 A starts to open (point M 4  of the curve U 25 A at about 180°) close to the exit  23 B of the distributor  19 , which slides towards the bottom dead center until the maximum opening (point M 5  at about 220°) and then gradually closes (point M 6  at about 240°). 
     In this case, the fluid injection from the distribution chamber  21  to the compression chamber  5  therefore occurs between the crank angle of about 180° and the crank angle of about 240°, i.e. when both the passage opening of the entrance  25 A and that of the exit  25 B of the port  25  are at least partially open. The passage section of fluid injection is maximum in the point M 5 , where at the same time the passage opening of the entrance  25 A and of the exit  25 B are maximal. 
     It should be furthermore noted that in this case the exit  25 B remains completely open much more longer (from the point M 2 I to M 2 II), and the quantity of injected fluid is therefore greater than in the configuration described in  FIG. 9 . 
     In  FIG. 10  with dashed lines are furthermore indicated the curves U 27 A and U 27 B, representing the passage opening or area of the entrance  27 A and of the exit  27 B of the supply port  27  as a function of the stroke of the distributor  19  and respectively of the piston  11 , not described in detail for the sake of simplicity. 
       FIG. 1 , together with the cycle of a traditional compressor and that of a “Voorhees cycle” compressor, described above with reference to the prior art, also shows the diagram of the compressor in the configuration shown in  FIGS. 2 ,  6 , and  7  and represented by the quadrilateral ABCFA, wherein the segment AB corresponds to the piston expansion phase, the segment BC to the suction phase, the segment CF to the compression phase and the segment FA to the delivery phase. The position of the end compression point F is determined by the quantity of injected fluid and therefore both by the dimension of the entrance sections  25 A and  25 B and exit sections  27 A and  27 B and by the distribution pressure P 2 . Therefore, by acting on these parameters it is quite easy to vary the ratio between the flow rate sucked by the compressor at the pressure P 1  and that of fluid injected at pressure P 2  so as to optimize both the increase in the refrigerating capacity and the energy efficiency of the refrigerating machine on which the compressor with the distributor is installed. 
       FIG. 11  schematically shows a refrigerating circuit using a compressor of the type described above.  FIG. 12  shows the respective transformation cycle of the coolant on a pressure-enthalpy diagram. The illustrated example refers to a cycle with CO 2  as coolant, but it should be understood that other adequate refrigerating fluids can be used. It is clearly apparent that the numerical values (of temperature and pressure) indicated below are indicated just by way of example, as they can vary according to the type of coolant R used and to the desired use conditions. 
     The refrigerating circuit comprises the compressor  1  supplying the coolant R at a pressure P 3  of about 90 bar and temperature of about 85° C. towards a main heat exchanger X 1  (through the point I), which is in turn connected (point II) to an exchanger-economizer X 2  for cooling the fluid R. The fluid R passes (point III) from the exchanger X 2  to the expansion valve X 3 , where the pressure is reduced to the pressure PI at about 25 bar, and it is subsequently supplied (point IV) to an evaporator X 4  for evaporating the remaining part of the fluid still in the liquid state. The fluid in the form of steam at pressure P 1  of about 25 bar is supplied (point V) from the evaporator X 4  to the entrance  1 L of the compressor  1 . 
     According to a particularly advantageous embodiment of the present invention, an economizer circuit or auxiliary circuit is connected downstream of the condenser X 1  (in the point II) to deviate part of the fluid R towards a secondary expansion valve X 5  designed for decreasing the pressure P 2  to about 50 bar. From here, the coolant R is supplied (point VI) to the exchanger-economizer X 2 , where it is heated and then completely evaporated. The main flow of the coolant flowing in the other branch of the exchanger-economizer X 2  is cooled between the point II and the point III. The fluid R in the auxiliary circuit is subsequently supplied (point VII) to the entrance  14 C of the distribution chamber  14 A of the compressor  1  at a pressure P 2  at about 50 bar. The fluid R coming from the suction  1 L is compressed in the compression chamber and at the same time mixed with the fluid R coming from the connection  14 C on the auxiliary circuit (point VIII). The fluid R is then further compressed, until it achieves the pressure P 3  (point I). 
     The distributor  19  provided inside the compressor  1  supplies the fluid R at 50 bar in the compression chamber or chambers  5 ,  7  according to a preset phase, as described above. 
       FIG. 12  shows an usual semi-logarithmic diagram enthalpy-pressure, on which is represented the transformation cycle to which the fluid R is subject in the plant of  FIG. 11 . In the diagram of  FIG. 12  on the axis of abscissas is shown the heat content of the fluid (the enthalpy) in kJ/kg and on the axis of ordinates the values of absolute pressure in bar. Briefly, this diagram is conventionally divided into three areas by a saturation bell: an area of undercooled liquid L (on the left of the bell), an area of overheated steam G (on the right of the bell) and an area L+G, where steam and liquid coexist in different percentages (within the bell). 
     On the curve representing the cycle the points I to VIII of the circuit of  FIG. 11  are indicated. From this diagram it should be noted in particular that during the compression phase from the point V to the point I, passing through the point VIII, the injection of further fluid R at a pressure P 2  entails a decrease in the temperature at the end of the compression (point I) relative to a traditional compressor: the temperature decrease will be the greater the greater the ratio between the rate of injected fluid from the connection  14 C and that sucked from the connection  1 L. Furthermore, this diagram shows how the heating of the fluid R in the main circuit from the point II to the point III involves an increase in the refrigerating effect. 
       FIG. 13  schematically shows another type of refrigerating plant using advantageously a compressor of the type described above.  FIG. 14  shows the respective transformation cycle of the coolant on a pressure-enthalpy diagram. 
     Unlike that of  FIG. 11 , in the plant of  FIG. 13  the pick-up point of the coolant R at the secondary expansion valve X 5  is performed downstream of the exchanger-economizer X 2  (point III) and therefore with a lower enthalpy content, as the coolant R has been already undercooled in the exchanger-economizer X 2 . The coolant passes from the point III in the secondary expansion valve X 5 , where the pressure is reduced to the pressure P 2  (point IV). From here, the fluid R is supplied to the exchanger-economizer X 2 , where it is heated and completely evaporated (point VII), and conveyed to the entrance  14  C of the distribution chamber  14 A of the compressor  1 . The description of the diagram in  FIG. 14  is completely similar to that of  FIG. 12 , except for the pick-up point of the coolant R at the secondary circuit, and it is therefore omitted for the sake of conciseness. 
     While specific embodiments of the invention have been shown and described in detail to illustrate the application of the principles of the invention, it will be understood that the invention may be embodied otherwise without departing from such principles.