Patent Publication Number: US-2016238019-A1

Title: Gas pipeline centrifugal compressor and gas pipeline

Description:
BACKGROUND OF THE INVENTION 
     1. Field of the Invention 
     The present invention relates to a gas pipeline centrifugal compressor having a centrifugal impeller and a gas pipeline, and more particularly, to a blade shape of the centrifugal impeller in a pipeline centrifugal compressor. 
     2. Description of the Related Art 
     Among industrial compressors, in a centrifugal compressor used as a booster for a gas pipeline, high efficiency and wide operating range are required. When the reserve of petroleum oil and natural gas pumped up from a well site of an oil field is reduced, the production is reduced due to depletion. Accordingly, flow rate control corresponding to the depletion is necessary. 
     As a flow rate control method for the centrifugal compressor, control of the number of units, valve control, rotation velocity control, inlet guide vane control and the like are known. When the flow rate is drastically reduced, the control of the number of units is effective. However, when the flow rate is changed (reduced) little by little, the control of the number of units is not available. When the flow rate is changed little by little, the rotation velocity control or the inlet guide vane control may be adopted, however, it is difficult to adopt these control methods from the points of cost, long-term reliability and maintainability. 
     Accordingly, as a gas pipeline centrifugal compressor, required is a compressor having a wide operating range corresponding to flow rate change to a certain degree without execution of controls as described above. 
     The operating range of the centrifugal compressor is generally determined based on surge on the low flow rate side while on choking on the high flow rate side, which much depends on design of the centrifugal impeller as a main element of the compressor. Accordingly, to realize a compressor having a wide operating range, the design of the impeller is important. 
     Note that as a designing method related to blades of the impeller of the centrifugal compressor, the methods described in the following patent literature 1 and 2 and non-patent literature 1 are known. 
     CITATION LIST 
     Patent Literature 
     [Patent Literature 1] Japanese Patent Laid-Open No. 2010-151126 
     [Patent Literature 2] Japanese Patent No. 3693121 
     Non-Patent Literature 
     [Non-Patent Literature 1] M. Zangeneh, A. Goto, and H. Harada: “On the Design Criteria for Suppression of Secondary Flows in Centrifugal and Mixed Flow Impellers”, ASME Journal of Turbomachinery, vol. 120, pp. 723-735, October 1998 
     SUMMARY OF THE INVENTION 
     Technical Problem 
     In the centrifugal compressor described in the above-described patent literature 1, to expand the operating range and improve the efficiency, and to increase the circumferential velocity of the impeller, the blade angle of the impeller blade is set as follows. 
     That is, the blade angle in a shroud-side blade angle curve of the blade takes a minimum value in the vicinity of a leading edge and is increased toward a trailing edge, and takes a maximum value between an intermediate point in the shroud-side blade angle curve and the trailing edge. On the other hand, the blade angle in a hub-side blade angle curve of the blade is increased from the leading edge toward the trailing edge, and takes a maximum value between an intermediate point in the hub-side blade angle curve and the leading edge. 
     In the centrifugal compressor described in the patent literature 1, on the shroud side of the impeller, the blade angle is minimum in the vicinity of the blade leading edge. In a status of the impeller viewed from the suction side (axial direction), the blade is closer to the circumferential direction in the vicinity of the shroud-side leading edge. Accordingly, a throat area as a minimum channel cross-sectional area between two adjacent blades is reduced especially on the shroud side. Accordingly, the flow velocity of the flow in the vicinity of the throat is increased, and choking easily occurs. When choking occurs, the operating range on the high flow rate side of the centrifugal compressor, i.e., the choke margin is narrowed. 
     On the other hand, as in the case of the patent literature 2, in the vicinity of the blade trailing edge of the impeller (in the vicinity of the impeller outlet), when the blade hub side is tilted such that it precedes the shroud side in the rotational direction of the impeller, the efficiency is improved as indicated in the non-patent literature 1, however, the operating range on the low flow rate side, i.e., the surge margin is narrowed. 
     The present invention has an object to obtain a gas pipeline centrifugal compressor in which the operating range on the low flow rate side can be expanded and the operating range on the high flow rate side can be maintained. 
     Another object of the present invention is to obtain a gas pipeline centrifugal compressor in which the operating range can be expanded and the efficiency can be improved while reduction of the efficiency can be suppressed. 
     Further object of the present invention is to obtain a gas pipeline to realize a compressor station provided with a high-efficient and low-price centrifugal compressor having a wide operating range. 
     Solution to Problem 
     To attain the above-described object, the present invention provides a gas pipeline centrifugal compressor used in a gas pipeline having gas piping to transfer gas and a plurality of compressors for gas pressurization provided on a route of the gas piping, wherein the centrifugal compressor has a centrifugal impeller fastened to a shaft, and the centrifugal impeller has a hub and a plurality of blades provided at intervals in a circumferential direction of the hub, and wherein blade angle distribution of the blade is configured such that, when a hub side camber line connecting a hub side leading edge as a suction side end and a hub side trailing edge as a discharge side end of the blade is indicated with a lateral axis, and a hub side blade angle of the blade is indicated with a vertical axis, a hub side blade angle is maximum on a side closer to the hub side leading edge than a central point of the hub side camber line, and from a part where the blade angle is maximum to the hub side leading edge, a hub side blade angle distribution curve indicating the hub side blade angle distribution is convex in a blade angle increasing direction, and configured such that, when a counter-hub side camber line connecting a counter-hub side leading edge as a suction side end on a counter-hub side and a counter-hub side trailing edge as a discharge side end of the blade is indicated with the lateral axis and a counter-hub side blade angle of the blade is indicated with the vertical axis, the counter-hub side blade angle is minimum at the counter-hub side leading edge of the counter-hub side camber line, or on a side closer to the counter-hub side leading edge than a central point of the counter-hub side camber line, further configured such that, in an arbitrary section including a part where the blade angle is minimum in a counter-hub side blade angle distribution curve indicating the counter-hub side blade angle distribution, the counter-hub side blade angle distribution curve is convex in a small blade angle direction, and from a downstream side of the section where the counter-hub side blade angle distribution curve is convex to the counter-hub side trailing edge, the counter-hub side blade angle distribution curve is convex in a large blade angle direction. 
     Another characteristic feature of the present invention is a gas pipeline comprising: a gas piping to transfer gas from a gas source to a gas supply destination; a compressor station having a centrifugal compressor for gas pressurization set in a plurality of positions on a route of the gas piping; a pressure regulator and a flow rate measurement unit provided between the compressor stations provided in the plurality of positions; a valve system provided in the gas piping between a most upstream compressor station in the plurality of compressor stations and the gas source; and a controller that controls the valve system, the compressor stations, the pressure regulator and the flow rate measurement unit, wherein the centrifugal compressor for gas pressurization is the above-described gas pipeline centrifugal compressor. 
     Advantageous Effects of Invention 
     According to the present invention, it is possible to obtain a gas pipeline centrifugal compressor in which the operating range on the low flow rate side can be expanded and the operating range on the high flow rate side can be maintained. 
     Further, it is possible to obtain a gas pipeline centrifugal compressor in which the operating range can be expanded and the efficiency can be improved while reduction of the efficiency can be suppressed. 
     Further, it is possible to obtain a gas pipeline to realize a compressor station provided with a high-efficient and low-price centrifugal compressor having a wide operating range. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a line graph showing a blade angle distribution of a centrifugal impeller in an embodiment 1 of a gas pipeline centrifugal compressor according to the present invention; 
         FIG. 2  is an axial directional view of a blade of the centrifugal impeller having the blade angle distribution shown in  FIG. 1 ; 
         FIG. 3A  is an explanatory diagram of the definition of the shape of the centrifugal impeller; 
         FIG. 3B  is an explanatory diagram of a velocity triangle of the flow in the centrifugal impeller; 
         FIG. 4  is a line graph showing the blade angle distribution of the centrifugal impeller in an embodiment 2 of the gas pipeline centrifugal compressor according to the present invention; 
         FIG. 5  is an axial directional view of the blade of the centrifugal impeller having the blade angle distribution shown in  FIG. 4 ; 
         FIG. 6  is an explanatory diagram of the definition of the blade shape in the axial directional view of the centrifugal impeller;  FIG. 7  is an explanatory diagram of the blade shape of the centrifugal impeller in an embodiment 3 of the gas pipeline centrifugal compressor according to the present invention; 
         FIG. 8A  is an explanatory diagram of the flow between two adjacent blades in the centrifugal impeller; 
         FIG. 8B  is an explanatory diagram of the flow between the two adjacent blades in other centrifugal impeller than that in  FIG. 8A ; 
         FIG. 9  is a line graph showing the blade angle distribution of the centrifugal impeller in an embodiment 4 of the gas pipeline centrifugal compressor according to the present invention; 
         FIG. 10  is a meridional cross-sectional diagram showing an example of the gas pipeline centrifugal compressor according to the present invention; 
         FIG. 11  is an enlarged meridional cross-sectional diagram of a part of the centrifugal compressor shown in  FIG. 10 ; 
         FIG. 12  is a schematic diagram showing an example of the gas pipeline in the present invention; and 
         FIG. 13  is a line graph showing the relation between a flow rate and a head in the centrifugal compressor. 
     
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     Hereinbelow, particular embodiments of the present invention will be described based on the drawings. Note that in the respective drawings, elements having the same reference numerals indicate the same or corresponding elements. 
     Embodiment 1 
     First, the configuration of a gas pipeline centrifugal compressor and the configuration of a gas pipe line will be described in accordance with  FIGS. 10 to 13 .  FIG. 10  is a meridional cross-sectional diagram showing an example of the gas pipeline centrifugal compressor according to the present invention.  FIG. 11  is an enlarged meridional cross-sectional diagram showing a part (in the vicinity of a first stage impeller) of the centrifugal compressor shown in  FIG. 10 .  FIG. 12  is a schematic diagram showing an example of the gas pipeline according to the present invention.  FIG. 13  is a line graph showing the relation between a flow rate and a head in the centrifugal compressor. 
       FIG. 13  shows a characteristic curve of the centrifugal compressor. In  FIG. 13 , the lateral axis indicates a flow rate, and a vertical axis, a head. In the characteristic curve of a general centrifugal compressor, as indicated with a solid line in  FIG. 13 , an operating point at which the centrifugal compressor is actually activated is an intersection point between a duct resistance curve and the characteristic curve of the centrifugal compressor. 
     The system configuration of the gas pipeline will be described with the schematic diagram of  FIG. 12 . In the example shown in  FIG. 12 , a compressor station  2  ( 2   a,    2   b,    2   c ) is provided in three positions on the route of a gas piping  4  ( 4   a,    4   b,    4   c,    4   d,    4   e ) of a gas pipeline  1 . 
     Gas is sent from a natural gas well site (gas source)  3  such as an oil field or a gas field, via a gas piping  4   a,  first to a gas processing facility  5 , in which the gas is subjected to processing such as gas gathering or gas treatment, then is sent via a valve system (including a valve)  6  and a gas piping  4   b,  to a first compressor station  2   a.  The compressor station  2   a  has a centrifugal compressor (gas pipeline centrifugal compressor)  200  for gas pressurization, a bypass piping system  201  and the like. Next, the gas pressurized with the first compressor station  2   a  is sent via a gas piping  4   c  to a second compressor station  2   b,  and further, sent via a gas piping  4   d  to a third compressor station  2   c.  These second and third compressor stations  2   b  and  2   c  also have the same configuration as that of the first compressor station  2   a.    
     The gas pressurized with the third compressor station  2   c  is sent through a gas piping  4   e  to each of various plants (gas supply destinations)  7  such as an LNG plant. The gas piping  4   c  on the downstream side of the first compressor station  2   a  is provided with a pressure regulator  8 , a flow rate measurement unit  9  and the like. Reference numeral  10  denotes a controller to control the respective compressor stations  2   a,    2   b  and  2   c,  the valve system  6 , the pressure regulator  8 , the flow rate measurement unit  9  and the like, via a control signal transmitter (control line)  11 . 
     The compressor station ( 2   a,    2   b,    2   c ) shown in  FIG. 12  is provided by, e.g., several 10 km. Accordingly, the resistance curve of the pipeline centrifugal compressor  200  depends on the duct resistance (loss) of this long gas piping. The specification of the gas pipeline centrifugal compressor  200  is determined based on the prediction of the duct resistance. When the predictive accuracy regarding the duct resistance of the duct having the long gas piping is insufficient, the resistance curve in  FIG. 13 , i.e., the flow rate at the operating point of the centrifugal compressor varies. However, when the operating range of the centrifugal compressor is sufficiently wide, it is possible to perform operation corresponding to the variation. 
     Further, when the amount of gas gathered from the well site  3  is reduced, the flow rate also changes. In such case, in the conventional centrifugal compressor with a narrow operating range, it is impossible to continue the operation in some cases. Accordingly, in this case, a part of the compressed gas is returned with the bypass piping system  201  to the suction side of the centrifugal compressor  200 , thus a circulation channel is formed. With this arrangement, operation on the high flow rate side is possible in the centrifugal compressor  200 . However, when this operation is performed, as a part of the compressed gas is returned with the bypass piping system  201  to the suction side, operation to send a low rate gas to the downstream side is performed in the first compressor station  2 . At this time, as high flow rate operation is performed in the centrifugal compressor  200 , the motive power is wasted. 
     Then, when the centrifugal compressor  200  having a wide operating range is realized, it is possible to perform long-term operation of the gas pipeline centrifugal compressor  200  without bypass operation with the bypass piping system  201 . In the gas pipeline centrifugal compressor  200  in the present embodiment, as described later, the operating range can be wide. Accordingly, it is possible to perform long-term operation without the bypass piping system  201  even when the amount of gas at the well site  3  is reduced, by adopting the gas pipeline centrifugal compressor  200  in the present embodiment as a centrifugal compressor for gas pressurization in the first compressor station  2 . Accordingly, it is possible to obtain an efficient gas pipeline where waste of power consumption is suppressed. 
     Next, using  FIG. 10  and  FIG. 11 , an example of the centrifugal compressor  200  adopted in a gas pipeline shown in  FIG. 12  will be described. 
       FIG. 10  shows the entire configuration of the gas pipeline centrifugal compressor  200 . The centrifugal compressor  200  is a uniaxial multistage centrifugal compressor in which a single shaft  108  is provided with a multistage (two stages in this example) centrifugal impeller (hereinbelow, it may be simply referred to as an “impeller”)  100  ( 100 A and  100 B). 
     The centrifugal impeller  100  ( 100 A and  100 B) rotates integrally with the shaft  108 , to apply the rotational energy to fluid. 
     The shaft  108  is rotatably supported with radial bearings  109  provided at both ends of the shaft  108 . Further, a thrust bearing  110  to support the shaft  108  in an axial direction is provided at one end of the shaft  108 . Further, a seal  114  is respectively provided inside of the radial bearings  109  at both ends of the shaft  108 . 
     A diffuser  104  ( 104 A,  104 B) to convert the dynamic pressure of the fluid made to flow from the centrifugal impeller  100  to static pressure is provided outside of the centrifugal impeller  100 A,  100 B in the radial direction. A return channel  105  to lead the fluid to a downstream channel  107  is provided downstream of the diffuser  104 A. The gas is led from the downstream channel  107  to the subsequent stage centrifugal impeller  100 B. 
     The impellers  100 A and  100 B, the diffusers  104 A and  104 B and the return channel  105  are accommodated in a casing  111 . Further, a suction casing  112  is provided on the suction side of the casing  111 . A discharge casing  115  is provided on the discharge side of the casing  111 . 
     The gas (fluid) sucked from the suction casing  112  as indicated with an arrow  116  is sucked from a suction port of the initial stage impeller  100 A, then it is pressurized while it is made to pass through the impeller  100 A, the diffuser  104 A and the return channel  105 , and sent to the subsequent stage impeller  100 B. Further, the gas made to flow from the subsequent stage centrifugal impeller  100 B is made to pass through the diffuser  104 B, then is made to pass through a scroll  113 , then finally it is pressurized to have predetermined pressure and discharged to the outside from the discharge casing  115  as indicated with an arrow  117 . 
       FIG. 11  shows an enlarged view around the initial stage (first stage) impeller  100 A in  FIG. 10 . The configuration of the initial stage impeller  100 A will be described using  FIG. 11 . 
     The impeller  100 A has a disk-shaped hub  102  fastened to the shaft  108 , a shroud (side plate)  101  provided oppositely to the hub  102 , and plural blades  103 , positioned between the hub  102  and the shroud  101 , provided at intervals in a circumferential direction. Note that the subsequent stage (second stage) impeller  100 B (see  FIG. 10 ) has the same configuration as that of the initial stage impeller  100 A. Further, the impeller  100  shown in  FIG. 11  has the shroud  101 , however, it maybe a so-called half shroud type impeller which does not have the shroud  101 . 
     Further, in the present embodiment, as the diffuser  104 A, a vaned diffuser having plural vanes in the circumferential direction is adopted. The subsequent stage diffuser  104 B (see  FIG. 10 ) has the same configuration. Note that a vaneless diffuser which does not have any vane may be used. 
     Note that numeral  106  denotes the above-described suction port of the initial stage impeller  100 A; and  107 , the above-described downstream channel. 
     In the centrifugal compressor  200 , especially in a centrifugal compressor to handle gaseous matter, a phenomenon that the flow is stalled in the centrifugal impeller  100  and the diffuser  104  in accordance with reduction of flow rate, and even when the flow rate is reduced by using a flow rate regulating valve or the like, the pressure is not raised from that level, and a large pressure variation and flow rate variation are caused occurs. This phenomenon is surge (or surging), which indicates a limiting point on the low flow rate side of the centrifugal compressor  200 . 
     On the other hand, when the flow rate regulating valve or the like is opened so as to increase the flow rate from the surge-occurred limiting flow rate, a phenomenon that the discharge pressure is lowered and the flow rate is not increased from that level occurs. This phenomenon is called choking, which indicates a limiting point on the high flow rate side of the centrifugal compressor  200 . The section between these two limiting points, surge and choking, is called an operating range of the centrifugal compressor. It is required that the operating range is expanded without lowering the efficiency of the centrifugal compressor. 
     Hereinbelow, the centrifugal compressor  200  in which the operating range can be expanded without lowering the efficiency will be described. 
     Using  FIGS. 1 to 3 , an embodiment 1 of the gas pipeline centrifugal compressor  200  used in the gas pipeline according to the present invention will be described. Note that in the following description, a centrifugal impeller having a shroud will be described, however, a half-shroud type centrifugal impeller without shroud is also applicable. In the case of the half-shroud type centrifugal impeller, the “shroud side” in the following description is a “counter-hub side”. Further, in the case of the centrifugal impeller having a shroud, the “counter-hub side” means the “shroud side”. 
       FIG. 1  is a line graph showing the blade angle distribution of one blade  20  (see  FIG. 2 ) among the blades  103  in the centrifugal impeller  100  of the gas pipeline centrifugal compressor  200 . In  FIG. 1 , the lateral axis indicates a non-dimensional blade center line (camber line) S plotted by connecting points, where the distances from pressure surface and suction surface of the blade  20  are equal, with regard to hub side end and shroud side (counter-hub side) end. Further, the vertical axis in  FIG. 1  indicates a blade angle β (°). 
     Numeral  12  denotes a hub side blade angle distribution curve showing the blade angle distribution on the hub side; and  13 , a shroud side (counter-hub side) blade angle distribution curve showing the blade angle distribution on the shroud side (counter-hub side). In the lateral axis in  FIG. 1 , the total camber line length from the leading edge side to the trailing edge side in the respective curves  12  and  13  is 1, i.e., the leading edge side of the blade  20  is expressed as “S=0” while the trailing edge side of the blade  20 , “S=1”. In the figure, S m  denotes an intermediate point (S=0.5). 
     The distribution of the blade angle β at the hub side end of the blade  20  is as shown with the hub-side blade angle distribution curve  12  as a broken line. Further, the distribution of the blade angle β at the shroud side end of the blade  20  is as shown with the shroud-side blade angle distribution curve  13  as a solid line. 
       FIG. 2  is an axial directional view of one blade  20  among the blades  103  in the centrifugal impeller  100 . The hub side end of the blade  20  is indicated with a curve  23 , while the shroud side end of the blade  20 , with a curve  24 . Note that in the description of  FIG. 2  and the subsequent description, the camber line is used as a representative curve of the blade  20 . A leading edge  21  as a suction side end and a trailing edge  22  as a discharge side end of the blade  20  in the centrifugal impeller  100  are respectively linear shaped. 
     The blade angle β is expressed as inclination from the circumferential direction. For example, the blade angle β s  in the position of the radius R on the shroud side is expressed as a ratio between a circumferential minute length R·dθ and a distance dm on a meridian plane. The distance dm on the meridian plane is a distance between points obtained by, assuming that the shroud side end  24  has changed from a point s 1  to a point s 2 , projecting the points s 1  and s 2  on a meridian plane of the impeller  100  (R-Z plane) (R: radial coordinate, Z: axial coordinate) in the circumferential minute length R·dθ on the blade  20 . Accordingly, the blade angle β on the camber line between the points s 1  and s 2  is indicated with the following expression (1). Note that in  FIG. 2 , N denotes a rotational direction; and O, an origin. 
         B =tan −1 ( dm /( R·d θ))   (1)
 
       FIG. 3A  is an axial perspective diagram of two adjacent blades A and B in an arbitrary radial positions. A broken line indicates a case where the blade angle β is large and β=β G  holds, and solid line, a case where the blade angle β is small and β=β s  holds. The suction surfaces of the blades A and B are denoted by numerals  31 A and  31 B, and the pressure surfaces, by numerals  32 A and  32 B. A perpendicular line drawn from the blade A of the two adjacent blades A and B onto the suction surface of the other blade B is a blade passage width L. The blade passage width L is L G  when blade angle β=β G  holds, and is L s  when blade angle β=β s  holds. 
       FIG. 3B  is a vector diagram showing a velocity triangle of the flow in the impeller  100 . When the circumferential velocity of the impeller  100  is U and the blade angle β is β G , a relative velocity of the flow in the impeller  100  is W, and an absolute velocity of the flow in the impeller  100  is C. When the blade angle β is β s , the relative velocity of the flow in the impeller  100  is W′, and the absolute velocity of the flow in the impeller  100  is C′. C m  is a meridional component of the absolute velocity and is a velocity component related to the flow rate. 
     Returning to  FIG. 1 , the shroud-side blade angle distribution curve  13  showing the distribution of the shroud side blade angle β s  of the blade  20  takes a minimum value β s   _   min  at a blade leading edge S L   _   s , and is increased toward the downstream side. The shroud-side blade angle distribution curve  13  is downwardly convex within the range of the camber line length S A  from the blade leading edge S L   _   s , and is upwardly convex within the range of the camber line length S B  from the point of the camber line length S A  to the blade trailing edge S T   _   s . Note that the camber line length S A  is smaller than the flow-directional intermediate point S m  (non-dimensional camber line length S=0.5). 
     That is, in an arbitrary section including a part (β s   _   min ) where the blade angle β s  in the shroud-side blade angle distribution curve  13  is minimum, the shroud-side blade angle distribution curve  13  is convex in a small blade angle direction, and in a section from the downstream side of the section S A  to the shroud side trailing edge, the shroud-side blade angle distribution curve  13  is convex in a large blade angle direction. 
     On the other hand, the hub-side blade angle distribution curve  12  showing the distribution of the hub side blade angle β h  forms maximum blade angle β h   _   max  between a blade leading edge S L   _   h  and the flow-directional intermediate point S m  (non-dimensional camber line length S=0.5). From the maximum blade angle part (β h   _   max ) to the hub side leading edge, the hub side blade angle distribution curve showing the distribution of the hub side blade angle is convex in the blade angle increasing direction. Between the blade leading edge S L   _   h  and the blade angle β h   _   max , the distribution curve  12  showing the hub-side blade angle β h  has no inflection point. 
     The ground of the setting of the shape of the blade  20  in this manner is as follows. 
     In  FIG. 3A , the difference between the blade angle β G  and β s  appears as a difference in the shape of the velocity triangle in  FIG. 3B . When the meridional components C m  of the absolute velocities C, C′ in  FIG. 3B  are approximately the same in the same radial position, the relative velocity vector W′ in the case of β s  when the blade angle β is small is larger than the relative velocity vector W in the case of β G  when the blade angle β is large. 
     In the general centrifugal impeller  100 , the deceleration of the shroud-side relative flow velocity is higher than that of the hub-side relative flow velocity. Accordingly, it is possible to improve the impeller efficiency and the impeller stall characteristic determined based on the values of wall friction loss, deceleration loss (loss due to increase in thickness of wall boundary layer toward the downstream side in the flow direction by deceleration of the relative flow velocity) and the like by appropriately setting the deceleration of the relative flow velocity on the shroud side. 
     Accordingly, in the present embodiment, the distribution is set such that the shroud side blade angle β s  is minimum at the blade leading edge, and in the section of the camber line length S A , the blade angle distribution curve  13  is downwardly convex. With this arrangement, it is possible to suppress increase of the blade angle β s  in the first half on the shroud side where the deceleration of the relative flow velocity is large and the blade  20  is easily stalled, and to reduce the deceleration of the relative flow velocity. Accordingly, it is possible to suppress the stall of the impeller to the further low flow rate side. 
     Further, when it is arranged such that the relative flow velocity is not decelerated on the shroud-side leading edge side (in the camber line length S A ) of the blade  20 , a high relative flow-velocity region is expanded from the blade leading edge  21  toward the flow direction downstream side. In the high relative flow-velocity region, the wall friction loss is large, and the increase of the high relative flow-velocity region causes reduction of the impeller efficiency. According to the present embodiment, in the distribution on the shroud-side blade trailing edge  22  side (within the camber line length S B ), the blade angle β s  is upwardly convex, to decelerate the relative flow velocity so as to prevent increase of the wall friction loss. 
     That is, in the shroud-side blade leading edge side (within the camber line length S A ), the increase of the blade angle β s  in the vicinity of the leading edge  21  is suppressed, and thereafter, the blade angle β s  is radically increased so as to increase the deceleration of the relative flow velocity. That is, in the region where the increase of the blade angle β s  is suppressed, the relative flow velocity becomes high as shown in  FIG. 3B , and this high relative flow-velocity region is expanded to the downstream side. As a result, the impeller stall on the low flow rate side due to relative flow velocity reduction is suppressed, and it is possible to improve the impeller efficiency. 
     In the impeller in the present embodiment, since the increase of the blade angle β s  on the shroud-side leading edge side (within the range of the camber line length S A ) is suppressed, the blade passage width L is narrowed as shown in  FIG. 3A  on the shroud-side leading edge side (within the range of the camber line length S A ). Regarding the camber line length S direction, the blade passage width L is minimum at the blade leading edge  21 , and further, is smaller on the shroud  23  side than that on the hub  24  side. 
     In the blade passage formed with the two adjacent blades A and B, regarding the direction of the camber line length S, a part where the channel cross sectional area is minimum is called a “throat”. In this throat, when the Mach number of the relative flow velocity exceeds 1, choking occurs and it is impossible to increase the flow rate. Accordingly, in high flow rate operation in the centrifugal compressor where the relative flow velocity is increased, the operating range is narrowed. 
     In the present embodiment, to avoid this inconvenience, it is arranged such that the hub side blade angle β h  is maximum (β h   _   max ) from the blade leading edge (non-dimensional camber line length S=0) to the point where the non-dimensional camber line length S=S m =0.5 holds. Further, from the part where the blade angle is maximum to the hub side leading edge, the curve indicating the hub-side blade angle distribution (hub-side blade angle distribution curve  12 ) is convex in the blade angle increasing direction. Further, in the section from the blade leading edge  21  to the point where the blade angle β h  is maximum (the section where the hub-side blade angle distribution curve  12  is convex in the blade angle increasing direction), the distribution curve  12  of the hub-side blade angle β h  has no inflection point. 
     With this arrangement, the hub side blade angle β h  is increased smoothly and radically between the throat, often formed until the non-dimensional camber line length S=0.5 holds, and the blade leading edge  21  (non-dimensional camber line length S=0). As a result, the hub side blade angle β h   _   throat  in the throat is increased, and in the throat, a blade passage width L h  is increased in the vicinity of the hub side. Accordingly, even when a blade passage width L s  is narrowed on the shroud side, as the blade passage width L h  is increased in the vicinity of the hub side, the area of the throat can be maintained. Since the hub side blade angle distribution has no inflection point and is upwardly convex, the increase of the hub-side blade passage width L h  is realized. As a result, it is possible to expand the flow rate region where the Mach number of the relative flow velocity exceeds 1 to the further high flow-rate side, to suppress the occurrence of choking in the impeller  100 , and to ensure the high flow-rate side operating range in the centrifugal compressor. 
     Note that to increase the hub side blade angle β h  in the throat, the hub-side blade angle maximum value β h   _   max  is brought closer to 90° as much as possible within a range where separation of the hub side surface of the blade  20  does not occur. In this manner, when the hub-side blade angle maximum value β h   _   max  is brought closer to 90°, the hub-side blade angle maximum value β h   _   max  is often greater than a hub-side outlet blade angle β h   _   T . Accordingly, it is desirable that the blade angle β h  distribution from the point where the hub side blade angle is the maximum value β h   _   max  to the hub side outlet is smoothly reduced. 
     Embodiment 2 
     An embodiment 2 of the centrifugal compressor  200  of the present invention will be described using  FIGS. 4 to 6 . In the present embodiment, the difference from the centrifugal compressor shown in the above-described embodiment 1 is that the position of the minimum value in the shroud-side blade angle distribution of the blade of the centrifugal impeller  100  is changed. 
       FIG. 4  shows an example of the blade angle distribution of the centrifugal impeller  100  according to the present embodiment. A hub-side blade angle distribution curve  40  is similar to that in the embodiment 1. On the other hand, a shroud-side blade angle distribution curve  41  as the counter-hub side is once reduced from the blade leading edge S L   _   s  toward the flow direction downstream side, then takes a minimum value β s   _   min  in a position closer to the shroud side leading edge than the intermediate point S m  (camber line length S=0.5), then is increased thereafter. Further, the blade angle between the shroud-side blade leading edge S L   _   s  and the blade trailing edge S T   _   s  is downwardly convex initially, and then upwardly convex around the end, along the camber line in the downstream direction. 
     In  FIG. 4 , the blade angle distribution curve  41  is downwardly convex in a section S c  on the upstream side from the intermediate point S m  and is upwardly convex in a section S D  following the section S c . In the section S c , in which the blade angle is downwardly convex may exceed the intermediate point S m . 
     In the centrifugal impeller  100  having the above arrangement shown in the embodiment 2, it is possible to further reduce the deceleration of the relative flow velocity in the vicinity of the shroud side leading edge of the impeller  100  in comparison with the centrifugal impeller  100  shown in the above-described embodiment 1. With this arrangement, it is possible to obtain a centrifugal impeller in which the operating range on the low flow rate side is further expanded. 
     Note that in the embodiment 2, the blade passage width L is further smaller on the shroud side of the throat in comparison with the impeller shown in the above-described embodiment 1. Accordingly, in the present embodiment, to ensure the operating range of the centrifugal impeller  100  on the high flow-rate side, the hub-side maximum blade angle β h   _   max  is equal to or greater than that in the embodiment 1. Further, as the hub-side maximum blade angle B h   _   max  is often wider than the hub-side outlet blade angle β T   _   h , the distribution is set such that the blade angle is smoothly reduced from the position of the hub-side maximum blade angle β h   _   max  to the hub-side outlet S T   _   h . 
       FIG. 5  is an axial directional view of one blade  50  of the centrifugal impeller having the blade angle distribution shown in  FIG. 4 . A shroud side camber line  54  of the blade  50  has an approximately S shape having a part A 5 A the blade leading edge  51  side of which is radial outwardly convex (outer diameter side). On the other hand, a hub side camber line  53  of the blade  50  has an approximately S shape having a part A 5 B the blade leading edge  51  side of which is radial inwardly convex (inner diameter side). The grounds will be described also using  FIG. 6 . 
       FIG. 6  is a coordinate system and an axial directional view regarding the centrifugal impeller  100 .  FIG. 6  is a diagram viewed from the suction side. The centrifugal impeller  100  rotates about a shaft O in a rotational direction N. To assist explanation of the operation of the blade of the centrifugal impeller  100 , a blade  60  having a linear blade camber line will be described. 
     The figure shows that, assuming that the blade angle at a blade leading edge  61  is β L , the blade angle β is in a position  62  on the downstream side from the blade leading edge  61 . The position  62  is away from the blade leading edge  61  by Δθ in the circumferential direction. The blade angle β in the position  62  is represented from geometrical relation as β=β L +Δθ. 
     In the blade where the blade camber line is linear shaped, the blade angle β is linearly increased with respect to a circumferential angle θ from the blade leading edge  61  toward the downstream side. 
     An example where the blade angle β is not linearly changed with respect to the circumferential angle θ of the blade camber line and the increase of the blade angle β is gradually reduced from the position  62  toward the downstream side, and another example where the increase of the blade angle β is increased, will be described. When the increase of the blade angle β is reduced with respect to the circumferential angle θ of the camber line from the position  62 , the shape of the camber line is as indicated with a curve  63  in  FIG. 6 . That is, it is in contact with the linear camber line passing through the position  62 , and is convex radial outwardly. On the other hand, when the increase of the blade angle β is radically increased with respect to the circumferential angle θ of the camber line, the shape of the camber line is as indicated with a curve  64  in  FIG. 6 . That is, it is in contact with the linear camber line passing through the position  62  and is convex radial inwardly. 
     In the centrifugal impeller  100  having the blade angle distribution shown in  FIG. 4 , the shroud-side blade angle distribution is once reduced from the blade leading edge toward the downstream side, to minimum, and is increased thereafter. Accordingly, as shown in  FIG. 5 , the shroud-side camber line shows an approximately S shape where the blade leading edge  51  side is convex radial outwardly. Further, as the hub-side blade angle distribution is maximum without inflection point from the leading edge  51  to the flow direction intermediate point, and is smoothly reduced on the downstream side from the position of the maximum value, the hub side camber line has an approximately S shape where the blade leading edge  51  side is convex radial inwardly. In this manner, the blade angle distribution shown in  FIG. 4  has the above-described approximately S shape in appearance. 
     Embodiment 3 
     An embodiment 3 of the gas pipeline centrifugal compressor of the present invention will be described with reference to  FIGS. 7 and 8 . In the embodiment 3, the difference from the centrifugal compressor  200  shown in the above-described embodiments 1 and 2 is that, in the embodiment 3, in addition to the arrangement of the embodiments 1 and 2, the inclination direction at the blade trailing edge in the centrifugal impeller  100  is tilted backward with respect to the rotational direction. With this arrangement, as shown in  FIG. 7 , when the centrifugal impeller  100  is viewed from the axial direction, a hub side camber line  73  and a shroud side camber line  74  of a blade  70  intersect each other. 
     That is,  FIG. 7  is an axial directional view of the one blade  70  among the blades  103  (see  FIG. 11 ) in the centrifugal impeller  100 . On the blade trailing edge  72  side of the blade  70 , the trailing edge of the shroud side camber line  74  is positioned on the rear side than the trailing edge of the hub side camber line  73  with respect to the rotational direction (N direction in the figure). Note that the blade angle distribution of the hub side camber line  73  and that of the shroud side camber line  74  are similar to that in the above-described embodiment 1 or the embodiment 2. 
     The operation of the centrifugal impeller  100  in the embodiment 3 having the above-described arrangement will be described below using  FIGS. 8A and 8B . In these figures, the blade of the centrifugal impeller  100  is denoted by numeral  80 . 
       FIG. 8A  is a diagram of the impeller  100  in which the camber line on the shroud side  83  of the blade  80  is tilted frontward from the camber line on the hub side  84  on the trailing edge  86  side of the blade  80  (hereinbelow, also referred to as a “forward tilted impeller”), and a diagram of two adjacent blades  80  forming the blade passage. As shown in  FIG. 8A , at the trailing edge  86  of the blade  80 , when the shroud side  83  of the blade  80  is tilted forward from the hub side  84  with respect to the rotational direction, it is possible to reduce the centrifugal force acting on the blade  80 . 
     On the other hand, regarding the inner flow, a blade force F acting from each blade  80  to the fluid acts in a vertical direction with respect to the blade pressure surface  81 , in other words, the direction of the hub side  84  of the blade suction surface  82 . As the static pressure is raised in the direction where the blade force F acts, the static pressure is raised on the hub side  84  of the blade suction surface  82 . On the other hand, the static pressure is lowered on the shroud side  83  of the blade suction surface  82 . 
     In the blade passage of the centrifugal impeller  100 , a wall velocity boundary layer where the flow velocity is lower than the main flow velocity and the energy is low occurs in the vicinity of the wall surface. The fluid in the wall velocity boundary layer cannot overcome the gradient of the static pressure in the blade passage cross section, and it drifts from a high static pressure region to a low static pressure region. Note that the blade passage cross section is a cross section obtained by cutting the blade passage in a radius r=predetermined cylindrical surface from the center of the shaft. The drifting flow forms a secondary flow having a flow velocity component in the vertical direction with respect to the main flow in the blade passage cross section. 
     As described above, the secondary flow from the blade pressure surface  81  having high static pressure toward the blade suction surface  82  having low static pressure occurs in the vicinity of the wall velocity boundary layer in the blade passage cross section of the centrifugal impeller  100 . Further, in the forward-tilted impeller, a secondary flow from the hub side  84  to the shroud side  83  also occurs in the vicinity of the wall velocity boundary layer of the blade suction surface  82 . Accordingly, the low energy fluid is accumulated on the shroud side  83  of the blade suction surface  82 , and the pressure loss is increased. In addition, the uniformity of the flow in the blade passage cross section is degraded, and the loss in the diffuser and the return channel on the downstream side from the impeller  100  is increased. 
     Note that in  FIG. 8A , numeral  85  denotes the blade  80  leading edge. 
       FIG. 8B  is a diagram of the impeller  100  in which the camber line on the shroud side  83  is tilted further backward than the camber line on the hub side  84 , on the blade trailing edge  86  side (hereinbelow, also referred to as a “backward-tilted impeller”), and a graph showing the two adjacent blades  80  forming the blade passage. In the backward-tilted impeller, the blade force F acts in the direction of the shroud side  83  of the blade suction surface  82 . Accordingly, on the hub side  84  of the blade suction surface  82 , the static pressure is lowered, while on the shroud side  83  of the blade suction surface  82 , the static pressure is raised. With this arrangement, it is possible to suppress the secondary flow toward the shroud side  83  of the blade suction surface  82 . The uniformity of the flow in blade passage cross section is improved, and the efficiency of the centrifugal impeller  100  is improved. That is, it is possible to realize an impeller with higher efficiency and wide operating range by combining the backward-tilted impeller and the blade angle distribution according to the embodiment 1 or 2. 
     Further, it is possible to obtain a gas pipeline centrifugal compressor with higher efficiency and a wider operating range in comparison with conventional devices by applying the above impeller to a gas pipeline centrifugal compressor. 
     Embodiment 4 
     Using  FIG. 9  and the above-described  FIG. 10 , an embodiment 4 of the gas pipeline centrifugal compressor according to the present invention will be described. The embodiment 4 is advantageous when the present invention is applied to a uniaxial multistage centrifugal compressor as shown in  FIG. 10  (two stage device in  FIG. 10 ).  FIG. 9  corresponds to  FIG. 1  in the above-described embodiment 1. As in the case of  FIG. 1 ,  FIG. 9  illustrates the hub-side blade angle distribution curve  12  and the shroud-side blade angle distribution curve  13 . The hub-side blade angle distribution curve  12  shown in  FIG. 9  is similar to the hub-side blade angle distribution curve  12  in  FIG. 1 . 
     In the present embodiment, as shown in  FIG. 9 , as the shroud side (counter-hub side) blade angle distribution curve  13 , two types of distribution curves, i.e., a shroud-side (counter-hub side) blade angle distribution curve  13 A of an upstream stage impeller indicated with a solid line, and a shroud-side (counter-hub side) blade angle distribution curve  13 B of a downstream stage impeller indicated with an alternate long and short dash line, are shown. 
     The shroud-side blade angle distribution curve  13 A of the upstream stage impeller indicated with the solid line corresponds to the blade angle distribution in the initial stage (first stage) centrifugal impeller  100 A of the two stage centrifugal compressor shown in  FIG. 10 . The shroud-side blade angle distribution curve  13 B indicated with the alternate long and short dash line corresponds to the blade angle distribution in the subsequent stage (second stage) centrifugal impeller  100 B shown in  FIG. 10 . 
     The shroud-side blade angle distribution curve  13 B of the subsequent stage centrifugal impeller  100 B indicated with the alternate long and short dash line is set such that the blade angle of the downstream centrifugal impeller  100 B is smaller than that of the upstream stage centrifugal impeller  100 A. At least in a part of the shroud-side blade angle distribution curve which is convex in the small blade angle direction, the blade angle of the downstream centrifugal impeller  100 B is smaller than that of the upstream stage centrifugal impeller  100 A. 
     That is, the blade angle distribution in the vicinity of the blade leading edge (inlet) of the subsequent stage centrifugal impeller  100 B is smaller than that of the initial stage centrifugal impeller  100 A. With this arrangement, the blade load in the vicinity of the inlet (in the vicinity of blade leading edge) of the subsequent stage centrifugal impeller  100 B is relatively small, and the surge margin is wider in the subsequent stage impeller  100 B. 
     Generally, the surge in the uniaxial multistage centrifugal compressor such as a two stage centrifugal compressor is determined based on downstream-stage surge margin rather than upstream-stage surge margin. Accordingly, it is possible to further expand the surge margin of the entire multistage centrifugal compressor by changing the blade angle distribution in correspondence with each stage of the multistage centrifugal impeller  100  as described in the present embodiment. Especially, in a pipeline centrifugal compressor requiring a wide operating range, it is possible to obtain a gas pipeline centrifugal compressor with high efficiency and wide operating range by changing the blade angle distribution from the upstream-stage side centrifugal impeller toward the downstream-stage side centrifugal impeller as described above. 
     As described above, as the gas pipeline centrifugal compressor according to the present embodiment has the blade angle distribution as described above, on the low flow rate side, the blade load is small on the shroud side in the vicinity of the impeller inlet. Thus it is possible to suppress occurrence of stall and to obtain wide surge margin. Further, as the blade angle is large immediately rear of the impeller inlet on the hub side, the throat area is large. Thus it is possible to ensure the throat area in the entire impeller. Accordingly, it is also possible to suppress the reduction of choke flow rate. Further, on the blade trailing edge side of the shroud side, as the blade angle distribution curve is upwardly convex, the relative flow velocity is decelerated, and the increase of the wall friction loss is suppressed. With this arrangement, it is possible to design an impeller with high efficiency and wide operating range, and it is possible to obtain a gas pipeline centrifugal compressor with high efficiency and wide operating range. 
     Further, it is possible to obtain a gas pipeline to realize a compressor station having a low-price centrifugal compressor with wide operating range and high efficiency by adopting the above-described gas pipeline centrifugal compressor of the present embodiment as a centrifugal compressor for gas pressurization in a gas pipeline compressor station of a gas pipeline. That is, even when the flow rate in the gas pipeline is changed little by little, as it is possible to expand the operating range of the centrifugal compressor, it is not necessary to perform rotation velocity control, inlet guide vane control or the like, and it is possible to realize a low price compressor station.