Abstract:
A compressor includes a screw rotor and a gate rotor. The screw rotor has a plurality of spirally extending groove portions disposed radially outwardly from the center axis of the screw rotor. The gate rotor has a plurality of tooth portions circumferentially arranged on an outer circumference to engage the groove portions. Preferably, an inclination angle of a groove portion side face contacting the tooth portions is inclined relative to a circumferential direction of the gate rotor varies. Alternatively a first plane contains the screw rotor center axis, a second plane orthogonally intersects the screw rotor center axis, a third plane orthogonally intersects the first and second planes, the gate rotor center axis is on the third plane, and the tooth portions do not overlap the first plane as viewed orthogonally relative to the third plane.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This U.S. National stage application claims priority under 35 U.S.C. §119(a) to Japanese Patent Application No. 2006-299227, filed in Japan on Nov. 2, 2006, the entire contents of which are hereby incorporated herein by reference. 
     TECHNICAL FIELD 
     The present invention relates to a compressor to be used in, for example, air conditioners, refrigerators and the like. 
     BACKGROUND ART 
     Conventionally, there has been provided a compressor including a disc-shaped screw rotor which rotates about a center axis and which has, in its end face in a center-axis direction, a plurality of spirally extending groove portions radially outward from the center axis, and a gate rotor which rotates about a center axis and which has a plurality of tooth portions arrayed circumferentially on its outer circumference, the groove portions of the screw rotor and the tooth portions of the gate rotor being engaged with each other to form a compression chamber (see JP 60-10161 B). 
     That is, this compressor is a so-called PP type single screw compressor. The term “PP type” means that the screw rotor is formed into a plate-like shape and moreover the gate rotor is formed into a plate-like shape. 
     Then, as viewed in a direction orthogonal to the screw rotor center axis and the gate rotor center axis, all the tooth portions of the gate rotor overlap with the screw rotor center axis. That is, the tooth portions of the gate rotor are engaged with the groove portions of the screw rotor along the radial direction of the screw rotor. 
     With a view to preventing interferences between the screw rotor and the gate rotor, side faces of the gate rotor tooth portions are given a maximum angle and a minimum angle each of which is formed by a gate rotor tooth-portion side face and a screw rotor groove wall surface on a plane which orthogonally intersects with the gate rotor plane and which contains a rotational direction of a tooth center line extending radial direction of the gate rotor (hereinafter, angles formed between the maximum angle and the minimum angle will be referred to as edge angles of the gate rotor; see edge angles δ 1 , δ 2  of  FIG. 20 ). 
     SUMMARY OF INVENTION 
     Technical Problem 
     However, with the conventional compressor described above, since all the tooth portions of the gate rotor are aligned with the screw rotor center axis as viewed in a direction orthogonal to the screw rotor center axis and the gate rotor center axis, angles formed by side faces of the screw rotor groove against side faces of the gate rotor tooth portions on the plane orthogonally intersecting with the gate rotor plane and containing the rotational direction of the gate rotor tooth center line involves a larger difference between a maximum value and a minimum value. 
     As a result of this, edge angles of gate rotor seal portions to be engaged with the side faces of the screw rotor groove portion become acute, so that a blow hole (leak clearance) present at an engagement portion between the screw rotor groove portion and the gate rotor tooth portion becomes larger. This would result in a lowered compression efficiency. 
     Accordingly, an object of the present invention is to provide a compressor in which the blow hole is made smaller so as to improve the compression efficiency. 
     Solution to Problem 
     In order to achieve the above object, there is provided a compressor comprising: a disc-shaped screw rotor which rotates about a center axis and which has, in at least one end face thereof in a direction along the center axis, a plurality of spirally extending groove portions radially outward from the center axis; and a gate rotor which rotates about a center axis and which has a plurality of tooth portions arrayed circumferentially on its outer circumference, the groove portions of the screw rotor and the tooth portions of the gate rotor being engaged with each other to form a compression chamber, wherein 
     a variation width of an inclination angle to which a side face of a groove portion of the screw rotor to be in contact with the tooth portions of the gate rotor is inclined against a circumferential direction of the gate rotor, the variation being over a range from radially outer side to inner side of the screw rotor, 
     is made smaller than 
     a variation width resulting when all the tooth portions of the gate rotor overlap with a plane containing the screw rotor center axis. 
     With such a compressor, the variation width of the inclination angle to which the side face of the groove portion of the screw rotor to be in contact with the tooth portions of the gate rotor is inclined against the circumferential direction of the gate rotor, the variation being over a range from radially outer side to inner side of the screw rotor, is made smaller than a variation width resulting when all the tooth portions of the gate rotor overlap with a plane containing the screw rotor center axis. Therefore, edge angles of the seal portions of the gate rotor to be engaged with side faces of the groove portion of the screw rotor can be made obtuse, so that the blow holes (leak clearances) present at engagement portions between the groove portion of the screw rotor and the tooth portions of the gate rotor can be made smaller, allowing the compression efficiency to be improved. Besides, wear of the seal portions of the gate rotor can be reduced, allowing an improvement in durability to be achieved. 
     Also, there is provided a compressor comprising: a disc-shaped screw rotor which rotates about a center axis and which has, in at least one end face thereof in a direction along the center axis, a plurality of spirally extending groove portions ( 10 ) radially outward from the center axis; and a gate rotor which rotates about a center axis and which has a plurality of tooth portions arrayed circumferentially on its outer circumference, the groove portions of the screw rotor and the tooth portions of the gate rotor being engaged with each other to form a compression chamber, wherein 
     with respect to a first plane containing the screw rotor center axis, a second plane which intersects orthogonally with the screw rotor center axis, and a third plane which intersects orthogonally with the first plane (S 1 ) and the second plane, 
     the gate rotor center axis is on the third plane, and 
     at least one of all the tooth portions of the gate rotor does not overlap with the first plane as viewed in a direction orthogonal to the third plane. 
     With such a compressor, the gate rotor center axis is on the third plane, and at least one of all the tooth portions of the gate rotor does not overlap with the first plane as viewed in a direction orthogonal to the third plane. Therefore, the side face of the groove portion of the screw rotor to be in contact with the tooth portions of the gate rotor can be set at approximately 90° against the rotational direction of the gate rotor (i.e. circumferential direction of the gate rotor) in its portion to be in contact with the side face of the groove portion of the screw rotor. Thus, the variation width of an angle formed by the side face of the groove portion of the screw rotor (hereinafter, referred to as screw rotor groove inclination angle) against a plane orthogonally intersecting with the rotational direction of the gate rotor (the circumferential direction of the gate rotor) can be made smaller. 
     Therefore, edge angles of the seal portions of the gate rotor to be engaged with side faces of the groove portion of the screw rotor can be made obtuse, so that the blow holes (leak clearances) present at engagement portions between the groove portion of the screw rotor and the tooth portions of the gate rotor can be made smaller, allowing the compression efficiency to be improved. Besides, wear of the seal portions of the gate rotor can be reduced, allowing an improvement in durability to be achieved. 
     In accordance with one aspect of the present invention, as viewed in the direction orthogonal to the third plane, a distance from an intersection point between a gate rotor plane formed by the first plane side end face of every tooth portion of the gate rotor and the gate rotor center axis ( 2   a ) to the first plane is 0.05 to 0.4 time as large as an outer diameter of the tooth portion of the gate rotor. 
     With such a compressor in accordance with this aspect of the present invention, as viewed in the direction orthogonal to the third plane, a distance from an intersection point between a gate rotor plane formed by the first plane side end face of every tooth portion of the gate rotor and the gate rotor center axis to the first plane is 0.05 to 0.4 time as large as an outer diameter of the tooth portion of the gate rotor. Therefore, the variation width of the screw rotor groove inclination angle can be made even smaller. 
     In accordance with one aspect of the present invention, as viewed in the direction orthogonal to the third plane, the gate rotor center axis is inclined by 5° to 30° against the second plane so that a tooth portion of the gate rotor closer to the screw rotor becomes closer to the screw rotor center axis than a tooth portion of the gate rotor farther from the screw rotor. 
     With such a compressor, as viewed in the direction orthogonal to the third plane, the gate rotor center axis is inclined by 5° to 30° against the second plane so that a tooth portion of the gate rotor closer to the screw rotor becomes closer to the screw rotor center axis than a tooth portion of the gate rotor farther from the screw rotor. Therefore, the variation width of the screw rotor groove inclination angle can be made even smaller. 
     In accordance with one aspect of the present invention, as viewed in a direction orthogonal to the first plane, a distance between the gate rotor center axis and the screw rotor center axis is 0.7 to 1.2 times as large as an outer diameter of the gate rotor. 
     With such a compressor, as viewed in a direction orthogonal to the first plane, a distance L between the gate rotor center axis and the screw rotor center axis is 0.7 to 1.2 times as large as an outer diameter D of the gate rotor. Therefore, the distance L can be made smaller, allowing a downsizing to be achieved. 
     In accordance with one aspect of the present invention, seal portions of the tooth portions of the gate rotor to be in contact with the groove portions of the screw rotor are formed into a curved-surface shape. 
     With such a compressor, since the seal portions of the tooth portions of the gate rotor to be in contact with the groove portion of the screw rotor are formed into a curved-surface shape, leakage of the compressed fluid from engagement portions between the tooth portions of the gate rotor and the groove portion of the screw rotor can be reduced, so that the compression efficiency can be improved. 
     Advantageous Effects of Invention 
     With a compressor in accordance with an embodiment of the present invention, the variation width of the inclination angle to which the side face of the groove portion of the screw rotor to be in contact with the tooth portions of the gate rotor is inclined against the circumferential direction of the gate rotor, the variation being over a range from radially outer side to inner side of the screw rotor, is made smaller than a variation width resulting when all the tooth portions of the gate rotor overlap with a plane containing the screw rotor center axis. Therefore, the blow holes can be made smaller, allowing the compression efficiency to be improved. 
     Also, with a compressor in accordance with an embodiment of the present invention, the gate rotor center axis is on the third plane, and at least one of all the tooth portions of the gate rotor does not overlap with the first plane as viewed in a direction orthogonal to the third plane. Therefore, the blow holes can be made smaller, allowing the compression efficiency to be improved. 
    
    
     
       BRIEF DESCRIPTION OF DRAWINGS 
         FIG. 1  is a simplified structural view showing an embodiment of the compressor of the invention; 
         FIG. 2  is a partial enlarged view of the compressor; 
         FIG. 3  is a simplified side view of the compressor; 
         FIG. 4  is a simplified plan view of the compressor; 
         FIG. 5  is a enlarged plan view of the compressor; 
         FIG. 6  is a graph showing a relationship between a gate rotor engagement angle γ and a screw rotor groove inclination angle β under the condition that a gate-rotor center axis inclination angle α is 12° and a positional-shift distance d is 0D; 
         FIG. 7  is a graph showing a relationship between a gate rotor engagement angle γ and a screw rotor groove inclination angle β under the condition that a gate-rotor center axis inclination angle α is 12° and a positional-shift distance d is 0.1D; 
         FIG. 8  is a graph showing a relationship between a gate rotor engagement angle γ and a screw rotor groove inclination angle β under the condition that a gate-rotor center axis inclination angle α is 12° and a positional-shift distance d is 0.2D; 
         FIG. 9  is a graph showing a relationship between a gate rotor engagement angle γ and a screw rotor groove inclination angle β under the condition that a gate-rotor center axis inclination angle α is 12° and a positional-shift distance d is 0.3D; 
         FIG. 10  is a graph showing a relationship between a gate rotor engagement angle γ and a screw rotor groove inclination angle β under the condition that a gate-rotor center axis inclination angle α is 0° and a positional-shift distance d is 0D; 
         FIG. 11  is a graph showing a relationship between a gate rotor engagement angle γ and a screw rotor groove inclination angle β under the condition that a gate-rotor center axis inclination angle α is 5° and a positional-shift distance d is 0D; 
         FIG. 12  is a graph showing a relationship between a gate rotor engagement angle γ and a screw rotor groove inclination angle β under the condition that a gate-rotor center axis inclination angle α is 12° and a positional-shift distance d is 0D; 
         FIG. 13  is a graph showing a relationship between a gate rotor engagement angle γ and a screw rotor groove inclination angle β under the condition that a gate-rotor center axis inclination angle α is 20° and a positional-shift distance d is 0D; 
         FIG. 14  is a graph showing a relationship between a gate rotor engagement angle γ and a screw rotor groove inclination angle β under the condition that a gate-rotor center axis inclination angle α is 0° and a positional-shift distance d is 0D; 
         FIG. 15  is a graph showing a relationship between a gate rotor engagement angle γ and a screw rotor groove inclination angle β under the condition that a gate-rotor center axis inclination angle α is 0° and a positional-shift distance d is 0.05D; 
         FIG. 16  is a graph showing a relationship between a gate rotor engagement angle γ and a screw rotor groove inclination angle β under the condition that a gate-rotor center axis inclination angle α is 0° and a positional-shift distance d is 0.1D; 
         FIG. 17  is a graph showing a relationship between a gate rotor engagement angle γ and a screw rotor groove inclination angle β under the condition that a gate-rotor center axis inclination angle α is 0° and a positional-shift distance d is 0.15D; 
         FIG. 18  is a graph showing a relationship between a gate rotor engagement angle γ and a screw rotor groove inclination angle β under the condition that a gate-rotor center axis inclination angle α is 0° and a positional-shift distance d is 0.2D; 
         FIG. 19  is a graph showing a relationship between a gate rotor engagement angle γ and a screw rotor groove inclination angle β under the condition that a gate-rotor center axis inclination angle α is 0° and a positional-shift distance d is 0.3D; 
         FIG. 20  is an enlarged sectional view of the compressor; 
         FIG. 21  is a graph showing a relationship between the positional-shift distance d and the degree of leakage effect with three screw rotor groove portions and twelve gate rotor tooth portions provided; 
         FIG. 22  is a graph showing a relationship between the positional-shift distance d and the degree of leakage effect with six screw rotor groove portions and twelve gate rotor tooth portions provided; 
     
    
    
     DESCRIPTION OF EMBODIMENTS 
     Hereinbelow, the present invention will be described in detail by way of embodiment thereof illustrated in the accompanying drawings. 
       FIG. 1  shows a simplified structural view which is an embodiment of the compressor of the invention.  FIG. 2  shows a partial enlarged view of the compressor. As shown in  FIGS. 1 and 2 , the compressor includes: a disc-shaped screw rotor  1  which rotates about a center axis  1   a  and which has, in its end face in a direction along the center axis  1   a , a plurality of spirally extending groove portions  10  radially outward from the center axis  1   a ; and a disc-shaped gate rotor  2  which rotates about a center axis  2   a  and which has a plurality of tooth portions  20  arrayed circumferentially on its outer circumference, the groove portions  10  of the screw rotor  1  and the tooth portions  20  of the gate rotor  2  being engaged with each other to form a compression chamber  30 . 
     That is, this compressor is a so-called PP-type single screw compressor. The term ‘PP-type’ means that the screw rotor  1  is formed into a plate-like shape while the gate rotor  2  is formed into a plate-like shape. This compressor is to be used in, for example, air conditioners, refrigerators and the like. 
     The groove portions  10  are formed in each of two end faces of the screw rotor  1 . The gate rotor  2  is provided two in number on each end face of the screw rotor  1 . Then, as the screw rotor  1  rotates about the screw rotor center axis  1   a  along a direction indicated by an arrow, each gate rotor  2  subordinately rotates about the gate rotor center axis  2   a  along an arrow direction by mutual engagement of the groove portions  10  and the tooth portions  20 . 
     On an end face of the screw rotor  1  are provided a plurality of thread ridges  12  spirally extending radially outward from the screw rotor center axis  1   a , where the groove portions  10  are formed between neighboring ones of the thread ridges  12 ,  12 . With one of the tooth portions  20  engaged with one of the groove portions  10 , side faces (i.e. seal portions) of the tooth portion  20  come into contact with side faces  11  of the groove portion  10  to seal the compression chamber  30 , while the tooth portion  20  is rotated by the side faces  11  of the groove portion  10 . 
     On an end face of the screw rotor  1  is attached a casing (not shown) which has grooves that allow the gate rotors  2  to rotate. A space closed by the groove portion  10 , the tooth portion  20  and the casing serves as the compression chamber  30 . 
     In the casing is provided a suction port (not shown) communicating with the groove portions  10  on the outer peripheral side of the screw rotor  1 . In the casing is also provided a discharge port (not shown) communicating with the groove portions  10  on the center side of the screw rotor  1 . 
     Referring to action of the compressor, a fluid such as refrigerant gas introduced to the groove portion  10  through the suction port is compressed in the compression chamber  30  as the capacity of the compression chamber  30  is reduced by rotation of the screw rotor  1  and the gate rotor  2 . Then, the compressed fluid is discharged through the discharge port. 
     As shown in the simplified front view of  FIG. 3  and the simplified plan view of  FIG. 4 , there are defined a first plane S 1  containing the screw rotor center axis  1   a , a second plane S 2  orthogonally intersecting with the screw rotor center axis  1   a , and a third plane S 3  orthogonally intersecting with the two planes of the first plane S 1  and the second plane S 2 . The second plane S 2  is coincident with the axial end face of the screw rotor  1 .  FIG. 3  is a view taken along an arrow A direction of  FIG. 2 , and  FIG. 4  is a view taken along an arrow B direction of  FIG. 2 . 
     The gate rotor center axis  2   a  is on the third plane S 3 . None of the tooth portions  20  of the gate rotor  2  overlaps with the first plane S 1  as viewed in a direction orthogonal to the third plane S 3 . 
     As viewed in the direction orthogonal to the third plane S 3 , a distance d from an intersection point between a gate rotor plane SG formed by an first plane S 1  side end face of every tooth portion  20  of the gate rotor  2  and the gate rotor center axis  2   a  to the first plane S 1  (hereinafter, referred to as positional-shift distance d) is 0.05 to 0.4 time as large as an outer diameter D of the tooth portion  20  of the gate rotor  2  (0.05D≦d≦0.4D). 
     As viewed in the direction orthogonal to the third plane S 3 , the gate rotor center axis  2   a  is inclined against the second plane S 2  so that a tooth portion  20  of the gate rotor  2  closer to the screw rotor  1  becomes closer to the screw rotor center axis  1   a  than a tooth portion  20  of the gate rotor  2  farther from the screw rotor  1 . An inclination angle α of the gate rotor center axis  2   a  is 5°-30°. In this case, an engagement depth of the tooth portions  20  with the groove portions  10  is 0.2 time as large as an outer diameter D of the gate rotor  2 . 
     As viewed in a direction orthogonal to the first plane S 1 , a distance L between the gate rotor center axis  2   a  and the screw rotor center axis  1   a  (hereinafter, referred to as axis-to-axis distance L) is 0.7 to 1.2 time as large as the outer diameter D of the gate rotor  2  (0.7D≦L≦1.2D). 
     In the gate rotor plane SG, an angle that a center line of the tooth portion  20  engaged with the groove portion  10  forms against a reference line parallel to the axial end face (second plane S 2 ) of the screw rotor  1  is referred to as a gate rotor engagement angle γ, and the angle of the center line (an intermediate line between leading side and unleading side) of the tooth portion  20  is measured from the reference line on a side of engagement starting. 
     The enlarged plan view of  FIG. 5  shows, in a tooth portion  20  of the gate rotor  2 , a minimum diameter, an intermediate diameter and a maximum diameter of engagement of the gate rotor  2 , the engagement being done with the groove portions  10  of the screw rotor  1 . Also in the tooth portion  20 , a side face on the downstream side of the rotational direction of the gate rotor  2  is assumed as a leading-side side face  20   a  while a side face on the upstream side of the rotational direction of the gate rotor  2  is assumed as an unleading-side side face  20   b.    
     Next,  FIGS. 6 to 9  show relationships between the gate rotor engagement angle γ (see  FIG. 4 ) and the screw rotor groove inclination angle β when the positional-shift distance d of the gate rotor center axis  2   a  (see  FIG. 3 ) is changed as 0D, 0.1D, 0.2D and 0.3D with the inclination angle α of the gate rotor center axis  2   a  (see  FIG. 3 ) set at 12°. In the figures are plotted engagement maximum diameters and intermediate diameters (see  FIG. 5 ) of the gate rotor  2  with respect to the leading-side side face  20   a  and the unleading-side side face  20   b  (see  FIG. 5 ), respectively. The number of groove portions  10  of the screw rotor  1  is three, and the number of tooth portions  20  of the gate rotor  2  is twelve. 
     It is to be noted here that the screw rotor groove inclination angle β, as shown in  FIG. 20 , refers to an angle β formed by the side face  11  of a groove portion  10  of the screw rotor  1  against a plane St which orthogonally intersects with the rotational direction (indicated by an arrow RG) of the gate rotor  2  (i.e. a circumferential direction of the gate rotor  2 ) at a contact portion of the side face  11  of the groove portion  10  and the tooth portion  20  of the gate rotor  2 . In addition, with the plane St taken as a reference, the screw rotor groove inclination angle β is expressed in positive values (+ direction) on the gate rotor rotational direction (arrow RG direction) side, and in negative values (− direction) on the side opposite to the gate rotor rotational direction (arrow RG direction). 
       FIG. 6  shows a chart when the positional-shift distance d is 0D, where variation widths of the screw rotor groove inclination angle β become larger with respect to engagement maximum diameters and intermediate diameters of the gate rotor  2  in the leading-side side face  20   a  and the unleading-side side face  20   b , respectively. 
       FIG. 7  shows a chart when the positional-shift distance d is 0.1D, where variation widths of the screw rotor groove inclination angle β are smaller than those of the screw rotor groove inclination angle β shown in  FIG. 6 . 
       FIG. 8  shows a chart when the positional-shift distance d is 0.2D, where variation widths of the screw rotor groove inclination angle β are smaller than those of the screw rotor groove inclination angle β shown in  FIG. 7 . 
       FIG. 9  shows a chart when the positional-shift distance d is 0.3D, where variation widths of the screw rotor groove inclination angle β are smaller than those of the screw rotor groove inclination angle β shown in  FIG. 6 . 
     Also,  FIGS. 10 to 13  show relationships between the gate rotor engagement angle γ and the screw rotor groove inclination angle β when the inclination angle α of the gate rotor center axis  2   a  is changed as 0°, 5°, 12° and 20° with the positional-shift distance d set at 0D. The rest of the conditions are similar to those of  FIGS. 6 to 9 . 
       FIG. 10  shows a chart when the inclination angle α of the gate rotor center axis  2   a  is 0°,  FIG. 11  shows a chart when the inclination angle α of the gate rotor center axis  2   a  is 5°,  FIG. 12  shows a chart when the inclination angle α of the gate rotor center axis  2   a  is 12°, and  FIG. 13  shows a chart when the inclination angle α of the gate rotor center axis  2   a  is 20°, where the variation width of the screw rotor groove inclination angle β becomes smaller as the inclination angle α of the gate rotor center axis  2   a  becomes larger. 
     That is, in  FIGS. 11 to 13 , since at least one of all the tooth portions  20  of the gate rotor  2  does not overlap with the first plane S 1 , the variation width of the screw rotor groove inclination angle β can be made smaller as compared with the case where all the tooth portions  20  of the gate rotor  2  shown in  FIG. 10  overlap with the first plane S 1 . 
     Also,  FIGS. 14 to 19  show relationships between the gate rotor engagement angle γ and the screw rotor groove inclination angle β when the positional-shift distance d is changed as 0D, 0.05D, 0.1D, 0.15D, 0.2D and 0.3D with the inclination angle α of the gate rotor center axis  2   a  set at 0°. The rest of the conditions are similar to those of  FIGS. 6 to 9 . 
       FIG. 14  shows a chart when the positional-shift distance d is 0D,  FIG. 15  shows a chart when the positional-shift distance d is 0.05D,  FIG. 16  shows a chart when the positional-shift distance d is 0.1D,  FIG. 17  shows a chart when the positional-shift distance d is 0.15D,  FIG. 18  shows a chart when the positional-shift distance d is 0.2D, and  FIG. 19  shows a chart when the positional-shift distance d is 0.3D, where the variation width of the screw rotor groove inclination angle β is smaller when the positional-shift distance d is larger than 0D. 
     That is, in  FIGS. 15 to 19 , since none of the tooth portions  20  of the gate rotor  2  overlaps with the first plane S 1 , the variation width of the screw rotor groove inclination angle β can be made smaller as compared with the case where all the tooth portions  20  of the gate rotor  2  shown in  FIG. 14  overlap with the first plane S 1 . 
     As shown in the enlarged sectional view of  FIG. 20 , seal portions  21   a ,  21   b  of the tooth portions  20  of the gate rotor  2  to be in contact with the groove portions  10  of the screw rotor  1  are formed into a curved-surface shape. 
     That is, a leading-side seal portion  21   a  is formed at the leading-side side face  20   a  of the tooth portion  20 , while an unleading-side seal portion  21   b  is formed at the unleading-side side face  20   b  of the tooth portion  20 . 
     The screw rotor  1  moves along a downward-pointed arrow RS direction, while the gate rotor  2  moves along a leftward-pointed arrow RG direction. 
     At engagement portions between the groove portion  10  of the screw rotor  1  and the tooth portion  20  of the gate rotor  2 , blow holes (leak clearances)  40 ,  50  shown by hatching are present. 
     More specifically, a leading-side blow hole  40  (shown by hatching) is present on an upstream side (compression chamber  30  side shown by hatching) of the leading-side seal portion  21   a  in the moving direction of the screw rotor  1 , while an unleading-side blow hole  50  (shown by hatching) is present on an upstream side (the compression chamber  30  side) of the unleading-side seal portion  21   b  in the moving direction of the screw rotor  1 . 
     The fluid compressed in the compression chamber  30  passes through the blow holes  40 ,  50  to leak outside the casing  3  (shown by imaginary line). 
       FIGS. 21 and 22  show a relationship between the positional-shift distance d (see  FIG. 3 ) and the degree of leakage effect. In this case, only the positional-shift distance d is changed within a range of 0D to 0.4D without any inclination of the gate rotor center axis  2   a  (α=0°). A degree of leakage effect of the leading-side blow hole  40  (see  FIG. 20 ), a degree of leakage effect of the unleading-side blow hole  50  (see  FIG. 20 ), and a total degree of leakage effect of the leading-side blow hole  40  and the unleading-side blow hole  50  are shown. It is noted here that the term, “degree of leakage effect,” refers to a degree obtained by converting areas of the leading-side blow hole  40  and the unleading-side blow hole  50  into corresponding leak amounts, respectively, wherein a degree of 100 corresponds to a leak amounts when the positional-shift distance d is 0D (as in the conventional case). 
       FIG. 21  shows degrees of leakage effect when the number of groove portions  10  of the screw rotor  1  is three and the number of tooth portions  20  of the gate rotor  2  is twelve. As the positional-shift distance d becomes larger, the degree of leakage effect becomes smaller, so that the compression efficiency is improved. 
       FIG. 22  shows degrees of leakage effect when the number of groove portions  10  of the screw rotor  1  is six and the number of tooth portions  20  of the gate rotor  2  is twelve. As the positional-shift distance d becomes larger, the degree of leakage effect becomes smaller, so that the compression efficiency is improved. 
     According to the compressor of the above-described constitution, since the gate rotor center axis  2   a  is present on the third plane S 3  and since at least one of all the tooth portions  20  of the gate rotor  2  does not overlap with the first plane S 1  as viewed in a direction orthogonal to the third plane S 3 , side faces  11  of a groove portion  10  of the screw rotor  1  to be in contact with the tooth portion  20  of the gate rotor  2  can be set at approximately 90° against the rotational direction (indicated by arrow RG) of the tooth portion  20  of the gate rotor  2  to be in contact with the side faces  11  of the groove portion  10  of the screw rotor  1  (i.e. against the circumferential direction of the gate rotor  2 ) as shown in  FIG. 20 . Thus, the variation width of the screw rotor groove inclination angle β can be reduced. 
     More specifically, in cases where the positional shift or inclination of the gate rotor  2  as in the present invention is not used (prior art), the changing width of the screw rotor groove inclination angle β during the course from suction to discharge becomes 16.0° at the leading-side side face  20   a  and 15.6° at the unleading-side side face  20   b . In contrast to this, in a case where the positional shift or inclination of the gate rotor  2  of the invention is applied to a compressor whose configuration (gate rotor tooth number, screw rotor groove number, gate rotor diameter, axis-to-axis distance, gate rotor tooth width, and suction cut angle) is similar to that of the prior art, the results are 6.5° at that the leading-side side face  20   a  and 13.8° at the unleading-side side face  20   b.    
     In other words, the variation width of the inclination angle of the side faces  11  of the groove portion  10  of the screw rotor  1  to be in contact with the tooth portion  20  of the gate rotor  2 , the inclination being against the circumferential direction of the gate rotor  2  and the variation width measuring from a radially outer side of the screw rotor  1  to its inner side, is made smaller, as compared with the variation width resulting when all the tooth portions of the gate rotor  2  overlap with the first plane S 1  containing the screw rotor center axis  1   a . In addition, the term, “circumferential direction of the gate rotor  2 ,” can be reworded as the rotational direction of the tooth portion  20  of the gate rotor  2  to be in contact with the side faces  11  of the groove portion  10  of the screw rotor  1 . Also, the term, “variation width of the screw rotor  1  from a radially outer side of the screw rotor  1  to its inner side,” refers to a variation width of the inclination angles of all the groove portions  10  from radially outer side to inner side of the screw rotor  1  to be concurrently in contact with the tooth portions  20  of the gate rotor  2 . 
     Therefore, edge angles δ 1 , δ 2  (see  FIG. 20 ) of the seal portions of the gate rotor  2  to be engaged with the side faces of the groove portions  10  of the screw rotor  1  can be made obtuse, so that the blow holes (leak clearances) present at engagement portions between the groove portions  10  of the screw rotor  1  and the tooth portions  20  of the gate rotor  2  can be made smaller. Thus, the compression efficiency can be improved. Besides, wear of the seal portions of the gate rotor  2  can be reduced, allowing an improvement in durability to be achieved. 
     In consequence, in the present invention, it has been found that in the PP-type single screw compressor, the angle of side faces of the groove portions  10  of the screw rotor  1  to be in contact with the tooth portions  20  of the gate rotor  2  is varied by shifting the position of the gate rotor  2  relative to the screw rotor  1 . 
     Also, since the positional-shift distance d is 0.05 to 0.4 time as large as the outer diameter D of the tooth portion  20  of the gate rotor as viewed in the direction orthogonal to the third plane S 3 , the variation width of the screw rotor groove inclination angle β can be made even smaller. 
     Also, as viewed in the direction orthogonal to the third plane S 3 , the gate rotor center axis  2   a  is inclined by 5° to 30° against the second plane S 2  so that a tooth portion  20  of the gate rotor  2  closer to the screw rotor  1  becomes closer to the screw rotor center axis la than a tooth portion  20  of the gate rotor  2  farther from the screw rotor  1 . Therefore, the variation width of the screw rotor groove inclination angle β can be made even smaller. 
     That is, in the PP-type single screw compressor, the velocity of the screw rotor  1  engaged with the gate rotor  2  has large differences between outer peripheral portions and central portion. In particular, at the central portion of the screw rotor  1 , the rotational speed of the gate rotor  2  becomes larger relative to the rotational speed of the screw rotor  1 , so that the screw rotor groove inclination angle β is varied to a large extent. 
     As a solution to this, it can be conceived to increase the axis-to-axis distance L between the screw rotor  1  and the gate rotor  2  so that velocity changes of the screw rotor  1  between outer peripheral portions and central portion of the screw rotor  1  becomes small. However, this incurs a problem that the outer diameter of the screw rotor  1  is increased, leading to an increased maximum diameter of the compressor. 
     Accordingly, by making the gate rotor center axis  2   a  inclined by 5° to 30° against a plane orthogonal to the screw rotor center axis  1   a , the variation width of the screw rotor groove inclination angle β can be made smaller without increasing the outer diameter of the screw rotor  1 . 
     Also, as viewed in the direction orthogonal to the first plane S 1 , the distance L between the gate rotor center axis  2   a  and the screw rotor center axis  1   a  is 0.7 to 1.2 times as large as the outer diameter D of the gate rotor  2 . Therefore, the distance L can be made smaller, allowing a downsizing to be achieved. 
     In other words, since the changing width of the screw rotor groove inclination angle β can be made small, the variation width of the contact angle between the gate rotor  2  and the screw rotor  1  can be suppressed even if the distance L is reduced. Thus, the downsizing can be achieved while the compression efficiency is maintained. 
     Also, since the seal portions  21   a ,  21   b  of the tooth portions  20  of the gate rotor  2  to be in contact with the groove portions  10  of the screw rotor  1  are formed into a curved-surface shape, leaks of the compressed fluid from engagement portions between the tooth portions  20  of the gate rotor  2  and the groove portions  10  of the screw rotor  1  can be reduced, so that the compression efficiency can be improved. 
     In other words, since the variation width of the screw rotor groove inclination angle β can be made small, the seal portions  21   a ,  21   b  of the gate rotor  2  can be formed into a curved-surface shape. More specifically, without increasing the thickness of the gate rotor  2 , maximum and minimum values of the inclination angle of the seal portions  21   a ,  21   b  can be fulfilled by machining the groove portions  10  of the screw rotor  1  with an end mill and by forming the seal portions  21   a ,  21   b  of the tooth portions  20  of the gate rotor  2  into a curved-surface shape with an end mill. 
     The present invention is not limited to the above-described embodiment. For example, the groove portion  10  may be provided only in one of the end faces of the screw rotor  1 . Also, the number of the gate rotors  2  may be freely increased or decreased. Further, the seal portions  21   a ,  21   b  of the tooth portions  20  of the gate rotor  2  to be in contact with the groove portions  10  of the screw rotor  1  may also be formed into an acute-angle shape. Besides, the screw rotor  1  and the gate rotor  2  may be rotated in opposite directions.