Fixed vane arrangement for a variable width diffuser

Abstract

 © A centrifugal compressor for a refrigeration or air conditioning system is provided having a variable width diffuser assembly (28) having a movable wall (34) and a plurality of vanes (82). The vanes (82) are angularly disposed relative to the impeller (16) within a critical range of angular values to provide high operating efficiencies and high pressure recoveries over a large operating range.

Description

Background of the Invention

[0001] The present invention relates to centrifugal vapor compressors, and more particularly to a fixed vane arrangement in a variable width diffuser for the compressor wherein the vanes are angularly disposed relative to the impeller within a range of angles for providing relatively high operating efficiencies and high pressure recoveries over relatively large operating ranges in, for example, refrigeration and air conditioning systems.

[0002] Flow stabilization through a centrifugal vapor compressor is a major problem when the compressor is used in situations where the load on the compressor varies over a wide operating range. The compressor inlet, impeller, and diffuser passage are designed to accommodate the maximum volumetric flow rate through the compressor. However, if the compressor inlet, impeller, and diffuser passage are designed to accommodate the maximum volumetric flow rate then flow through the compressor may be unstable when the volumetric flow rate is relatively low. As volumetric flow rate is decreased from a relatively high stable operating range, a range of slightly unstable flow is entered. In this slightly unstable range of flow there appears to be a partial reversal of flow in the diffuser passage which creates noise and lowers the efficiency of the compressor. Below this slightly unstable flow range, the compressor enters what is known as surge, wherein there are periodic complete flow reversals in the diffuser passage which lower the efficiency of the compressor and which may degrade the integrity of compressor components.

[0003] Numerous compressor modifications have been developed in attempting to improve flow stability through the compressor at low volumetric flow rates and to provide relatively high efficiencies throughout a large operating range. One such modification is the addition of guide vanes in the inlet passage to the compressor so as to vary the flow direction and quantity of the entering vapor.

[0004] Another modification is to vary the width of the diffuser passage in response to the load on the compressor. Normally, this is done by use of a movable diffuser wall which axially moves across the diffuser passage to throttle vapor flow therethrough.

[0005] Yet another modification includes utilizing vanes in the diffuser passage, usually in conjunction with a movable wall.

[0006] In spite of the above and other modifications, there still exists the need of a centrifugal compressor, and more particularly a diffuser assembly therefor, that will provide high operating efficiencies and high pressure recoveries over a relatively large operating range.

Summary of the Invention

[0007] Therefore, it is an object of the present invention to provide an improved centrifugal vapor compressor.

[0008] Another object of the present invention is to provide a centrifugal compressor having an improved diffuser assembly.

[0009] Yet another object of the present invention is to provide a unique diffuser assembly including a movable diffuser wall and fixed vanes for providing high operating efficiencies and high pressure recoveries over a relatively large operating range.

[0010] In one form of the present invention there is provided a centrifugal machine including a casing, an impeller rotatably mounted in the casing for moving a fluid therethrough, and a variable width diffuser assembly comprising a stationary wall member and a spaced-apart movable wall member generally radially disposed about the impeller to form therebetween a fluid passage leading from the impeller. The movable wall member is selectively movable relative to the stationary wall member, whereby the width of the fluid passage is selectively varied. A plurality of vanes are circumferentially disposed in the fluid passage and continuously span the distance between the wall members as the movable wall member moves relative to the stationary wall member. Each vane is angularly disposed in the fluid passage relative to the impeller such that each vane forms with an extended radius of the impeller passing through the vane an angle between about 73° to about 78°, whereby the impeller can be designed for any desired flow rate and yet maintain relatively high operating efficiencies and high pressure recoveries over a relatively large operating range.

Brief Description of the Drawings

[0011] The above-mentioned and other features and objects of the invention, and the manner of attaining them, will become more apparent and the invention itself will be better understood by reference to the following description of an embodiment of the invention taken in conjunction with the accompanying drawings, wherein:

Figure 1 is a fragmentary sectional side view of a centrifugal compressor incorporating a preferred embodiment of the present invention;

Figure 2 is a fragmentary sectional view of Figure 1 taken substantially along line II-II and viewed in the direction of the arrows;

Figure 3 is a graph of an optimum performance curve and plotted performance curves of a compressor with different diffuser setting vane angles;

Figure 4 is a graph of the same optimum performance curve and plotted performance curves of a second compressor with different diffuser setting vane angles; and

Figure 5 is a graph of the same optimum performance curve and plotted performance curves of a third compressor with different diffuser setting vane angles.

Description of a Preferred Embodiment

[0012] Referring primarily to Figure 1, there is illustrated a centrifugal compressor 10 including main casing 12 having an inlet 14 that directs the refrigerant into a rotating impeller 16 through a series of adjustable inlet guide vanes 18. Impeller 16 is secured to drive shaft 20 by any suitable means to align impeller 16 along the axis of compressor 10. Impeller 16 includes central hub 22 supporting a plurality of blades 24. Blades 24 are arranged to create passages therebetween that turn the incoming axial flow of refrigerant fluid in a radial direction and discharge the compressed refrigerant fluid from respective blade tips 26 into diffuser section 28. Diffuser section 28 is generally circumferentially disposed about impeller 16 and functions to direct the compressed refrigerant fluid into a toroidal-shaped volute 30, which directs the compressed fluid to the compressor outlet (not shown).

[0013] Diffuser section 28 includes a radially disposed stationary wall 32 and radially disposed movable wall 34 which is spaced-apart from stationary wall 32. Movable wall 34 is arranged to move axially towards and away from stationary wall 32 to vary the width of diffuser passage 36 formed therebetween, thereby altering the operating characteristics of compressor 10 in regard to varying load demands or flow rates.

[0014] Movable wall 34 is secured to carriage 38 by screws 40 received through aligned openings (not shown) in movable wall 34 and carriage 38. Screws 40 draw movable wall 34 tightly against the front of carriage 38. Carriage 38 is movably mounted in compressor 10 between shroud 42 and main casing 12. Movable wall 34 is accurately located by means of dowel pins (not shown) received in aligned holes (not shown) in movable wall 34 and carriage 38.

[0015] Carriage 38 is illustrated as being fully retracted against stop surface 44 of main casing 12 to open diffuser passage 36 to a maximum flow handling position. Carriage 38 is securely fixed by screws 46 to a double-acting piston 48. Although the piston may be driven by either gas or liquid, it shall be assumed for explanatory purposes that it is liquid actuated. By introducing fluid under pressure to either side of piston 48, its axial position and thus that of carriage 38 and wall 34 can be controlled. Piston 48 is slidably mounted between shroud 42 and main casing 12 so that it can move movable wall 34 by means of carriage 38 between the previously noted maximum flow position against stop surface 44 and a minimum flow position wherein the piston is brought against shroud wall 50.

[0016] A first expandable chamber 52 is provided between piston front wall 54 and casing wall surface 56. Delivering fluid under pressure into chamber 52 drives piston 48 toward stationary wall 32. A second expandable chamber 58 is similarly located between piston back wall 60 and shroud wall 50. Directing fluid under pressure to chamber 58 causes piston 48 to be driven forward to increase the width of diffuser passage 36.

[0017] Fluid is delivered into chambers 52, 58 from a supply reservoir (not shown) by means of a pair of flow circuits. The first flow circuit leading to chamber 52 includes channels 62, 64. The second circuit includes channels 66, 68, 70 and 72 which act to deliver the drive fluid into chamber 58. Channels 62-72 are formed by drilling communicating holes into the machine elements and plugging the holes where appropriate. Channels 62, 66 are drilled one behind the other and thus appear as a single channel in Figure 1. Both channels 62, 66 are connected to supply lines 74 in any suitable manner.

[0018] A suitable control system 76 containing electrically actuated valves regulates the flow of the fluid into and out of expandable chambers 52, 58 to either move piston 48 towards or away from stationary wall 32. A series of 0-ring seals 78 encircle piston 48 and prevent fluid from passing between chambers 52, 58. Control system 76 controls the position of carriage 38 and thus movable wall 34 to vary the width of diffuser passage 36. Although described in terms of control system 76, the present invention includes or contemplates other types of systems or methods for moving wall 34.

[0019] Referring now to Figures 1 and 2, a plurality of fixed vanes 82 are disposed in passage 36 in any suitable manner, for example, by securement to stationary wall 32. Vanes 82 may be of any suitable contour, such as NACA airfoils, and are equally spaced in passage 36 so as to be slidably received in complementary-shaped slots 84 in movable wall 34. Vanes 82 continuously span diffuser passage 36 regardless of its width as determined by the position of movable wall 34.

[0020] Critical to the present invention is the angular relationship between vanes 82 and extended radii of impeller 16. Compressors generally have an operating range at which the compressor optimally operates. However, once outside this range, the compressor's efficiency decreases to a point that it is desirable, and possibly a necessity, to design and construct another compressor having better efficiencies in these other ranges. Thus, over a relatively large operating range, several compressors having different efficiencies may be required to provide high operating efficiencies throughout this relatively large operating range.

[0021] This problem is addressed by the present invention and its solution was arrived at by theoretical and empirical analyses of fluid flow through impeller and diffuser section assemblies having movable walls and fixed vanes. As a result of the analyses, it has been shown that the diffuser setting vane angle is substantially independent of design pressure ratios or flows created by impellers. The term "diffuser setting vane angle" is that angle formed between a chord line 88 (Fig. 2) of a vane 82 and an extended radius 90 of impeller 16 passing through the leading edge point 92 of a respective vane 82. By "chord line" is meant a straight line drawn between the leading edge point 92 and the trailing edge point 94 of a vane 82.

[0022] Thus, it has been found that vanes 82 can be angularly disposed relative to respective impeller radii 90 within a critical range of angles independently of impeller design. This critical range of angles is between about 73° to about 78°, as measured between a chord line 88 and an impeller radius 90. Further testing revealed that an angle of 76.6° between a chord line 88 and a radius 90 is the optimum value within the just-mentioned critical range of angles. Once the vanes have been established within this critical range of angles, the impeller tip width can be appropriately designed and used for any desired flow rate.

[0023] Referring now to Figures 3-5, each graph is a plot of the head factor versus the flow factor of the optimum performance curve 96 of a designed prototype compressor. The head factor Y is defined as 1,000 times the ratio of the polytropic head to the square of the sonic velocity at compressor inlet conditions, and flow factor X is defined as the ratio of the flow rate at compressor inlet conditions to the sonic velocity at compressor inlet conditions. The Figures 3, 4, and 5 further include plotted performance curves of three different compressors, which differ physically in the design of their respective impellers. Furthermore, the nominal pressure ratio and lift capability of the compressor of Figure 3 is less than the nominal pressure ratio and lift capability of the compressor of Figure 4; and the nominal pressure ratio and lift capability of the compressor of Figure 4 is less than the nominal pressure ratio and lift capability of the compressor of Figure 5.

[0024] Preferably, for a given decrease in flow rate, it is desired to maximize the increase in head, as indicated by optimum performance curve 96. The graphs of Figures 3-5 illustrate that with vanes 82 angularly disposed relative to respective impeller radii 90 within the critical range of about 73° to about 78° yields the best trade off for decreasing flow rate and increasing head.

[0025] Referring to Figure 3, optimum performance curve 96 is plotted with head factor versus flow factor and illustrates the desired or optimum performance of a designed compressor. Plotted against optimum performance curve 96 are plotted performance curves A, B, C and D of a first compressor. In testing the first compressor, the diffuser setting vane angle was varied so as to determine which angle resulted in a plotted performance curve most closely paralleling optimum performance curve 96. In Figure 3, plotted performance curve D was obtained by setting the diffuser setting vane angle to about 66.2°. Similarly, plotted performance curves C and B were obtained by setting the diffuser setting vane angle to about 70.2° and about 73.2°, respectively. When the diffuser setting vane angle was set to about 76.6°, plotted performance curve A resulted therefrom. As can be seen, plotted performance curves B, C, and D do not closely parallel optimum performance curve 96, whereas plotted performance curve A does closely parallel curve 96. Further testing of the first compressor by setting the diffuser setting vane angle at lower and higher angles resulted in even less desirable performance curves relative to optimum performance curve 96.

[0026] Once plotted performance curve A is obtained by a selected diffuser setting vane angle, it is next desired to modify the first compressor so that plotted performance curve A overlies optimum performance curve 96 as closely as possible. This can be accomplished by appropriately designing the impeller tip width for the first compressor.

[0027] Turning now to Figure 4, the same optimum performance curve 96 is illustrated along with three plotted performance curves A', B', and E of a second compressor having greater nominal pressure ratios and lift capabilities than the first compressor of Figure 3. In Figure 4, plotted performance curve A' was obtained from a diffuser setting vane angle of about 76.6°, plotted performance curve B' from a diffuser setting vane angle of about 73.2°, and plotted performance curve E from a diffuser setting vane angle of about 78.0°. As was determined with the compressor of Figure 3, the plotted performance curve of the second compressor of Figure 4 most closely parallels optimum performance curve 96 with a diffuser setting vane angle of about 76.6°. The second compressor of Figure 4 is then optimally designed to approach optimum performance curve 96 by appropriately designing the impeller tip width so that plotted performance curve A' overlies curve 96 as closely as possible.

[0028] Referring now to Figure 5, the same optimum performance curve 96 is illustrated along with three plotted performance curves of a third compressor having nominal pressure ratios and lift capabilities greater than the compressors of Figures 3 and 4. In Figure 5, plotted performance curve A" was obtained from a diffuser setting vane angle of about 76.6°, plotted performance curve B" from a diffuser setting vane angle of about 73.2°, and plotted performance curve F from a diffuser setting vane angle also of about 73.2°. The difference between the plots of performance curves B" and F, which were both based on the same diffuser setting vane angle, is due to a minor difference in design of the third compressor. As with the above two compressors of Figures 3 and 4, it can be seen again that the plotted performance curve that most nearly parallels optimum performance curve 96 is that plotted performance curve resulting from a diffuser setting vane angle of about 76.6°. Again, diffuser setting vane angles less than about 73° and greater than about 78° resulted in even more undesirable plotted performance curves. Plotted performance curve A" can now be made to overlie as closely as possible optimum performance curve 96 by appropriately designing the impeller tip width.

[0029] From the above, it can be seen that the diffuser setting vane angle is substantially independent of design pressure ratios or flows created by impellers. Further, when the diffuser setting vane angle of vanes 82 are within the critical range of angles of about 73° to about 78°, the impeller tip width can be appropriately designed to maximize the increase in head for a given decrease in flow rate. The optimum performance of a compressor design in accordance with the principles of the present invention is obtained when the diffuser setting vane angle is about 76.6°.

[0030] Thus, a centrifugal compressor 10 designed as described above having the angular relationship between vanes 82 and impeller 16 will have high efficiencies and high pressure recoveries in a relatively large operating range.

[0031] While this invention has been described as having a preferred embodiment, it will be understood that it is capable of further modifications. This application is therefore intended to cover any variations, uses, or adaptations of the invention following the general principles thereof, and including such departures from the present disclosure as come within known or customary practice in the art to which this invention pertains and fall within the limits of the appended claims.

Claims

1. In a centrifugal compressor including a casing, an impeller rotatably mounted in said casing for moving a fluid therethrough, and a chamber generally circumferentially disposed about said impeller for receiving the fluid therefrom; a variable width diffuser assembly for said centrifugal compressor, comprising:

a fluid passage means being generally disposed about said impeller and including generally oppositely disposed wall members forming therebetween a passage extending between said impeller and said chamber for delivering the fluid therebetween, one of said wall members being fixed and the other said wall member being movable relative to said fixed wall member,

means operatively connected to said movable wall member for selectively moving said movable wall member relative to said fixed wall member, and

a plurality of fixed vanes generally circumferentially disposed in said passage,

each said fixed vane being angularly disposed in said passage such that a chord line of said each vane forms with an extended radius of said impeller passing through a leading edge of said each vane an angle between about 73° to about 78°.

r 2. The compressor of claim 1 wherein said angle is preferably between about 76° to about 77°.

3. The compressor of claim 1 wherein said angle, optimally, is substantially 76.6°.

4. The compressor of claim 1 wherein said vanes are slidably received in a respective plurality of complementary-shaped openings in said movable wall member.

5. In a centrifugal machine including a casing and an impeller rotatably mounted in said casing for moving a fluid therethrough, a variable width diffuser assembly, comprising:

a stationary wall member being generally radially disposed about said impeller,

a movable wall member being generally radially disposed about said impeller and spaced-apart from said stationary wall member to form therewith a fluid passage leading from said impeller,

means operatively connected to said movable wall member for selectively moving said movable wall member relative to said stationary wall member, whereby the width of said fluid passage can be selectively varied, and

a plurality of vanes circumferentially disposed in said fluid passage and being slidably disposed in a respective plurality of complementary-shaped openings in said movable wall member, said vanes continuously spanning the distance between said wall members as said movable wall member moves relative to said stationary wall member,

each said vane being angularly disposed in said fluid passage such that it forms with an extended radius of said impeller passing through said each vane an angle between about 73° to about 78°, whereby said impeller can be designed for any desired flow rate and yet maintain high operating efficiencies and high pressure recoveries over a relatively large operating range.

6. The machine of claim 5 wherein said angle is formed by a respective chord line extending between respective leading and trailing edges of each said vane and said extended impeller radius passing through said leading edge.

7. The machine of claim 6 wherein said angle is between about 76° to about 77°.

8. The machine of claim 7 wherein said angle is about 76.6°.

9. The machine of claim 5 wherein said angle is formed by a respective chord line of each said vane and an extended radius of said impeller passing through a leading edge of a respective said vane.

10. The machine of claim 9 wherein said angle is about 76.6°.

11. In a refrigeration system comprising a condenser, an evaporator, and a centrifugal compressor including therein a rotatable impeller and a chamber generally disposed about said impeller; a variable width diffuser and fixed vane assembly for said centrifugal compressor, comprising:

a stationary wall member being generally radially disposed about said impeller,

a movable wall member being generally radially disposed about said impeller and spaced-apart from said stationary wall member to form therewith a fluid passage leading from said impeller,

means operatively connected to said movable wall member for selectively moving said movable wall member relative to said stationary wall member, whereby the width of said fluid passage can be selectively varied, and

a plurality of vanes circumferentially disposed in said fluid passage and being slidably disposed in a respective plurality of complementary-shaped openings in said movable wall member, said vanes continuously spanning the distance between said wall members as said movable wall member moves relative to said stationary wall member,

each said vane being angularly disposed in said fluid passage such that a chord line of said each vane forms an angle between about 73° to about 78° with an extended radius of said impeller passing through a leading edge of said each vane.

12. The refrigeration system of claim 11 wherein said angle is between about 76° to about 77°.

Drawing