Patent Grant
6932562
This invention relates generally to a turbine fluid pump, and more particularly, to a multi-channel turbine fuel pump for use in a vehicle fuel delivery system.
Electric motor driven turbine pumps are customarily used in fuel systems of an automotive vehicle and the like. These pumps typically include an external sleeve which surrounds and holds together an internal housing adapted be submerged in a fuel supply tank with an inlet for drawing liquid fuel from the surrounding tank and an outlet for supplying fuel under pressure to an internal combustion engine of the vehicle. A shaft of the electric motor concentrically couples to and drives a pump impeller having an array of circumferentially spaced vanes disposed about the periphery of the impeller. An arcuate pumping channel carried by the housing substantially surrounds the impeller periphery and extends from an inlet port to an outlet port at opposite ends. Liquid fuel disposed in pockets defined between adjacent impeller vanes and the surrounding channel develops pressure through a vortex-like action induced by the three dimensional profile of the vanes and the rotation of the impeller.
Typically, impeller-type turbine fuel pumps have a stationary guide ring which strips fuel from the moving impeller vanes and diverts the fuel through an outlet port. The channel is located radially outward from the impeller vanes and radially inward from a substantial portion or trailing segment of the guide ring. In addition, the channel is located axially or laterally outward from both sides of the impeller at the circumferential array of vanes. In other words, the channel not only side-flanks or communicates axially with the impeller at the vane location from both sides, it also communicates with the vane pockets radially. A smaller portion, or striper segment of the guide ring, is disposed circumferentially between the inlet and outlet ports and is close to the impeller for striping the moving vanes of high pressure fuel, thereby, preventing the fuel at the outlet port from bypassing the fuel pump outlet and exiting back into the low pressure inlet port. Three examples of fuel pumps of this type are illustrated in U.S. Pat. No. 5,257,916 issued Nov. 2, 1993 to Tuckey, U.S. Pat. No. 6,068,456 issued May 30, 2000 to Tuckey et al. and U.S. Pat. No. 6,227,819 B1 issued May 8, 2001 to Gettel et al., each of which is assigned to the present assignee and is incorporated herein by reference.
A second type of turbine pump, such as that illustrated in U.S. Pat. No. 5,702,229 issued Dec. 30, 1997 to Moss et al. and incorporated herein by reference, has concentric dual circumferential arrays of vanes spaced radially apart by a mid-hoop or ring of the impeller, wherein both arrays communicate with a common channel. Similar to the first type of pump previously described, the outer array of vanes of this pump type project substantially radially outward from the periphery of the impeller toward a stationary guide ring. With this configuration, the fuel flows helically around the mid-hoop and through the channel. That is, the fuel flows about the mid-hoop as it is simultaneously circulating around the channel from an inlet to an outlet. Unfortunately, fuel flow cavitation within the pump, especially during hot fuel pumping conditions, continues to be a challenge.
A third type of turbine pump, as illustrated in U.S. Pat. No. 5,642,981 issued Jul. 1, 1997 to Kato et al. and incorporated herein by reference, is similar to the first example previously described, except that multiple pumps are arranged in series and powered by a common motor. Such pumps are better known as multi-stage pumps, or pumps having first and second stages, wherein the first stage (low pressure pump) feeds or flows fuel into a second stage (high pressure pump), thus being of a regenerative pump design. Unfortunately, multi-stage pump designs are expensive to manufacture and have an increased power consumption rate when compared to single stage designs.
Other types of turbine fuel pumps, such as that illustrated in U.S. Publication No. 2002/0021961 A1 published Feb. 21, 2002 to Pickelman et al. and U.S. Pat. No. 5,807,068 issued Sep. 15, 1998 to Dobler et al., both of which are incorporated herein by reference, do not utilize guide rings but instead have a peripheral hoop that is a unitary part of the impeller. The hoop engages the peripheral, radially outward distal ends of a circumferential array of impeller vanes. With this orientation, the impeller pockets only communicate with grooves of the channel in a lateral or axial direction. That is, communication between the impeller pockets and the channel is solely axial, or side-flanking. In contrast, the first and second types of turbine pumps have pockets that communicate with the channel in both an axial and a radial manner.
Despite the variety of turbine-type pumps and significant improvements in the design and construction of turbine fuel pumps on the market today, they are still somewhat inefficient. The efficiencies are generally between about 35%-45%, and when combined with a typical electric motor having an efficiency of about 45%-50%, the fuel pumps have an overall efficiency of between 16%-22%, in general. Higher flow and pressure requirements in the fuel pumping industry are exceeding the capabilities of conventional 36-39 mm diameter regenerative turbine fuel pumps. To increase fuel output and pressure, pumps must operate at higher speeds which aggravates cavitation concerns. Higher speed results in armature viscous drag (lost efficiency), noise and commutator wear. Maximum flow output under hot conditions is around 150 liters per hour for a conventional, single stage, turbine pump. Conventional alternatives to improve hot fuel flow are adding multi-pressure stages to the turbine pump, or oversizing the first stage of a two stage pump to accommodate a 30%-40% flow loss typical for regenerative pumps. However, such alternatives are costly and have an increase in power consumption, thus, which in turn decreases pumping efficiency.
The above-noted shortcomings of prior art fluid pumps are overcome by the turbine fluid pump assembly of the present invention, which, according to one embodiment, generally includes a lower casing, an upper casing, an impeller cavity, an electric motor and an impeller. The lower casing has a fluid inlet passage and first and second lower annular grooves; similarly, the upper casing has a fluid outlet passage and first and second upper annular grooves. The impeller has a first vane array that communicates with the first lower and upper annular grooves, and a second vane array that communicates with the second lower and upper annular grooves, such that rotation of the impeller causes a portion of the incoming fluid to enter the first lower annular groove and a portion to enter the second lower annular groove.
Objects, features and advantages of this invention include providing a turbine fluid pump assembly that has an improved pump efficiency, an increased displacement or output without loss of pumping efficiency or adding of additional components, improved hot fuel performance at high flow rates over a wide pressure range, that does not require adding additional components as with conventional multi-stage designs, has a higher efficiency then conventional single stage and dual stage designs, is easier to manufacture than multi-stage pumps, has a flat performance curve through various pressures and voltages, and where multiple stages can be added without significant cost or complexity, to name but a few. Furthermore, the design is relatively simple and economical to manufacture, and has a significantly increased useful life in service.
These and other objects, features and advantages of the present invention will be apparent from the following detailed description of the preferred embodiments and best mode, appended claims and accompanying drawings, in which:
The pumping section 32 includes an upper casing 42 and a lower casing 44, which are held together externally and generally encircled by the outer housing 38. An impeller cavity 46 is defined between, as well as being disposed substantially concentric to, the upper and lower casings 42, 44, and carries an impeller 48 of the present invention which rotates about the axis 34. A rotor (not shown), an integral shaft 35 of the motor, and impeller 48 all co-rotate about the axis of rotation 34. The shaft 35 projects downward through the upper casing 42, is fixedly coupled to and projects through the impeller 48, and bears against a bearing 49 that is located in a blind bore 51 in the lower casing.
A fuel inlet passage 50 communicates through the lower casing 44 in a substantially axial direction, through which low pressure fuel flows upward from a fluid reservoir or surrounding fuel tank (not shown) to the impeller cavity 46. Similarly, the upper casing 42 carries a fuel outlet passage 52 (shown in phantom), which provides a passage for pressurized fuel to flow in an axially upward direction out of the cavity 46. Inner and outer circumferential vane arrays 56A, 56B of impeller 48 respectively propel the fuel through circumferentially extending inner and outer pumping chambers 54A, 54B, which are primarily deposed between upper and lower casings 42, 44. The inner and outer vane arrays 56A, 56B are radially aligned with inner and outer pumping chambers, respectively, which generally extend for an angular extent of about 300-350°, or in any case, less than 360°. The pumping chambers 54A and 54B extend about the rotational axis 34 from the inlet passage 50 to the outlet passage 52. There is generally no, or only a limited amount, of cross fluid communication between the inner and outer pumping chambers 54A, 54B. Very limited cross fluid communication between the pumping chambers may be desirable where fuel is needed to act as a lubricant or a fluid bearing between the moving surfaces.
With specific reference now to
The upper and lower grooves 58A, 58B and 62A, 62B are concentric, arcuate grooves that each circumferentially extend around a surface of the upper and lower casings, respectively, such that they open into the impeller cavity 46. Each of these grooves preferably has an oval or elliptical cross-sectional shape, as opposed to a semi-circular cross sectional shape, as commonly seen on prior art pumps. For purposes of clarity, the following description of the shape of the grooves will be provided with specific reference to one of the grooves, but equally applies to the remaining grooves as well. The oval cross-sectional shape of the grooves is comprised of a first radial section 63, a linear or flat section 64, and a second radial section 65, and can increase the efficiency of the pump by reducing the effect of dead or stagnate zones in the pumping chambers where fuel stalls and does not adequately flow. This phenomenon sometimes occurs in semi-circular cross sectional grooves where the groove is too deep, which causes fuel to collect and sit at the bottom of the groove instead of circulating with the rest of the fuel flowing through the pumping chamber. The two radial sections 63, 65 are semi-circular portions of the groove, and may have radii (designating r1 and r2) of a common length or they may have radii with differing lengths. Likewise, the length of the flat section may be uniform amongst the different grooves, or its length may vary with respect to the length of the individual radial sections. In a preferred embodiment, the flat section 64 has a length of between 0.25 mm-1.00 mm. Due to the intervening flat section 64, center points C1 and C2, which correspond to radii r1 and r2, are separated by a certain distance. This distance may vary to suit the particular performance needs of the pump, and can be a function of one of the other dimensions of the grooves. For instance, either the length of flat section 64 or the distance separating the center points may be defined as a function of the length of r1 and/or r2. The upper and lower grooves 58A, 58B and 62A, 62B, which are stationary during operation as they are formed in the upper and lower casings 42, 44, interact with the circulating vane arrays, which will now be described in greater detail.
The vane pockets 60A and 60B are part of the impeller 48 and are formed between adjacent vanes in the inner and outer vane arrays 56A and 56B, respectively. Both the inner and outer vane pockets are open on both their upper and lower axial ends, such that they are adjacent surfaces 59, 69 and are in fluid communication with the upper and lower grooves. Furthermore, the inner vane pocket includes a surface 66A and the outer vane pocket includes a surface 66B, each of which is located on a radially inward side of the vane pocket and includes a circumferential ridge or rib 92A, 92B, respectively. Each of the vane pockets also includes a surface 67A, 67B that is located on the radially outward side of the vane pocket and is flat or extends in an axially straight line. Surfaces 66A and 66B are each partially partitioned by the ridges 92A, 92B such that curved surfaces 73A, 73B are formed on the upper axial halves of surfaces 66A and 66B, and curved surfaces 75A, 75B are formed on the lower axial halves of surfaces 66A and 66B. It follows, that the inner pumping chamber 54A includes a vane pocket 60A having a radially inward surface 66A with a ridge 92A. That ridge partitions surface 66A such that upper and lower curved surfaces 73A and 75A are formed. These curved surfaces may be semi-circular in shape and preferably have a radius equal to that of the first radial section 63 of the corresponding groove. Accordingly, each curved surface 73A, 75A extends away from the ridge 92A in an axial direction towards the upper and lower grooves, respectively, and continues across the small gap separating the grooves from the vane pocket. This continuation causes the curved surfaces 73A and 75A to effectively join with the first radial sections 63 of the grooves 58A and 62A, respectively, thus forming a larger, combined semi-circle or arcuate surface that extends from the ridge to the flat section 64. Of course, other pumping chamber arrangements could also be used, such as where the ridge culminates in a rounded, flat or blunt end, as opposed to the pointed end shown in the drawings. Furthermore, the grooves could be longer in the radial dimension than are the corresponding vane pockets, etc.
Turning now to
With reference now to
Each vane 78A of the inner vane array 56A and each vane 78B of the outer vane array 56B radially extends within the impeller 48 in a non-linear fashion, such that it increases the pumping efficiency of the impeller. The vanes will now be described in connection with several Figures, each of which shows the vanes from a different perspective and highlights different attributes of the vanes and/or the impeller.
Turning now to
The advance in circumferential travel of the tip segment 90 is generally not as great as the retard in circumferential travel of the root segment 88. Therefore, the overall radial projection of the vanes between the outer hub surface 66A and the inner mid hoop surface 67A, is slightly retarded when considered in the direction of impeller rotation 102. In other words, the radially innermost point 114 on the leading surface of the vane is advanced when compared to the radially outermost point 142 on the leading surface the vane, when considered in the direction of rotation 102. This retarded or trailing alignment is demonstrated as angle β, which represents the angular separation between the impeller radius 144 and straight line 146, which connects points 114 and 142. It follows, that during rotation of the impeller, point 114 reaches a particular angular position before point 142. Angle β is in the range of 0°-10°, is desirably between 0°-5°, and is preferably about 2°.
For the purposes of clarity and simplicity, the following paragraphs will only describe vanes of the inner vane array with the understanding that the vanes of the outer vane array are substantially identical unless otherwise stated. Referring now to
During manufacturing of the impeller 48, the impeller must be released from the mold via a rotational motion. Therefore, the root segment 88 of the vane has an incline angle α(R) which is equal to, or preferably slightly less than, an incline angle α(T) of the tip segment 90. The incline angles α(R) and α(T) can be measured from either the leading or the trailing sides of the vane, as they are preferably parallel. Preferably, the incline angle α of the inner vane array gradually increases from the root segment 88 through the tip segment 90, and is in the range of 10°-50°, is desirably in the range of 20°-40°, and is preferably about 25° at the radially innermost point of the root segment and is preferably 35° at the radially outermost point of the tip segment. An equivalent relationship exists for the vanes of the outer array, however, their incline angle is in the range of 15°-55°, is desirably between 20°-45°, and is preferably about 30° at the radially innermost point of the root segment and 40° at the radially outermost point of the tip segment. Accordingly, the following relationship between the incline angle at the root versus that angle at the tip holds true for both the inner and outer vane array: 10°≦α(R)≦α(T)≦55°. The incline angle α(R) of the root segment is measured in degrees between a reference line 113, which is parallel to the rotating axis 34, and an incline line 116 which lies along a leading surface of vane 78A at the root segment 88. As previously stated, each of the vane upper and lower halves 100, 104 have leading and trailing surfaces 108, 112 that are parallel; that is, the vane has a uniform vane thickness in the circumferential direction. Thus, incline line 116 could alternatively be located along the trialing vane surface as well. Reference line 113 and incline line 116 preferably intersect each other at a point that lies on the leading face of the vane. Separately, the radially innermost ends of the leading intersection line 106 and the trailing intersection line 110 are contiguous to the ridge 92A, as best shown in
The incline angle α(T) of the tip is measured in degrees between reference line 122, which is parallel to both the rotating axis 34 and the reference line 113, and an incline line 124, which preferably lies along the leading surface 108 of the vane in the region of the tip segment 90. As previously explained, incline line 124 could lie along the trailing vane surface 112 as well.
Also, the incline angles α(R) and α(T) of the vanes of the inner vane array 56A are respectively less than those of the vanes of the outer vane array 56B. Amongst other benefits, this difference in angles allows the impeller to be rotated out of a single rotational mold during manufacturing. This incline angle arrangement does not sacrifice pump performance, since the vanes of the inner vane array 56A operate with a higher pressure coefficient and thus require a smaller incline angle α for optimum performance than do the vanes of the outer vane array 56B.
As previously discussed, the root segment 88 radially extends outward from the outer hub surface 66A in a retarded or trailing manner, with respect to the radius of the impeller 144. It follows, that the leading intersection line 106, which separates the upper and lower halves 100, 104 of the vane, includes a radially inward portion that also extends in a retarded or trailing manner, with respect to radius 144 when considered in direction 102. This radially inward portion of the leading intersection line 106 is the portion that linearly extends from the ridge 92A to the radially outer terminus of the root segment. Leading intersection line 106 also includes a radially outward portion that extends in an advanced, curvilinear direction, just like the tip segment 90. This radially outward portion is the portion of the leading intersection line 106 that begins where the radially inward portion left off, and extends outward to the inner mid hoop surface 67A. Stated differently, the leading intersection line 106 includes a radially inward portion that is part of the root segment 88 and thus extends in a retarded, linear direction, and a radially outward portion that is part of the tip segment 90 and thus extends in an advanced, curved direction. As previously indicated this pocket forming or cupped vane configuration, when considered in both the radial and the axial directions, enhances pumping efficiency.
As shown in FIG. 11 and as previously mentioned, each half 100, 104 of each vane 78A also has a back angle γ which is preferably equal to the opposite front incline angles α(R) and α(T). This results in a uniform vane thickness when considered in a circumferential direction, and eases the manufacturing process by allowing for the release of the impeller following the molding process. It is possible, however, for the back angle γ to be greater than the corresponding front incline angle (“corresponding” means the portion of the front surface 108 that is at the same radial position on the vane), which would result in vanes having front and rear surfaces that converge together as they approach the axial side walls or ends of the vane. Consequently, because the minimum value of α(R) is 10° and because α(T) is equal to or greater than α(R), then the minimum value of γ, along the entire radial extent of the vane, is also 10°.
Each vane also includes two radii 120, 130 formed along edges located between the trailing vane surface 112 and adjacent upper and lower side walls 121, 131. Sidewall 131, best seen in
Of course, the previous explanation of impeller components, particularly the linear root segment, curved tip segment, circumferential ridge, vane pockets, upper vane half, lower vane half, leading intersection line, trailing intersection line, and radius, as well as all angles, reference lines, imaginary planes, etc. pertaining thereto, apply equally to the outer vane array 56B, unless stated otherwise. Moreover, the previous discussion is not specifically limited to a dual vane array impeller, as it could equally apply to other multi vane array impellers having three, four, or any other number of vane arrays that may practicably be utilized by the impeller.
Turning now to
With reference to
The upper casing 42 is quite similar to the lower casing just described, and generally includes a lower surface 59 having upper inner and outer grooves 58A, 58B formed thereon, an outlet passage 52, and a circumferentially extending lip or flange 160. The upper inner and outer grooves 58A, 58B each includes an axially tapered section, third section 162, but does not include two axially tapered sections as with the lower grooves. Third section 162 is tapered in an opposite or complimentary manner to that of first section 152; that is, while first section 152 of the lower grooves is decreasing in cross-sectional area, third section 162 of the upper grooves is increasing in cross-sectional area over the same angular extent. Complimentarily shaped tapers such as these promote adequate fuel distribution into both the upper and lower grooves, as opposed to a disproportionate amount remaining in the lower grooves because they are in direct communication with the inlet passage 50. Lip 160 circumferentially extends around the outer periphery of the upper casing 42 and provides a surface for the lower casing 44 to rest upon. By resting upon the lip 160, as opposed to surface 59 itself, the lower casing 44 and upper casing 42 create impeller cavity 46 which is located there between. The height and other attributes of the lip can vary, as they are dependent upon the thickness of the impeller 48 as well as other design considerations.
In operation, rotation of impeller 48 causes fuel to flow into the pumping section 32 via the fuel inlet passage 50, which directly communicates with independent, lower inner and outer grooves 62A, 62B. During its propulsion through the first section 152, fuel is forced into the upper inner and outer grooves 58A, 58B, such that an appropriate distribution of fuel is achieved between the upper and lower grooves. This produces a somewhat uniform fuel distribution between the upper and lower parts of the inner and outer pumping chambers 54A and 54B, such that approximately equal forces reside on both axial sides of the impeller. As best seen in
According to the alternative embodiment shown in
It will thus be apparent that there has been provided in accordance with the present invention a turbine fluid pump assembly which achieves the aims and advantages specified herein. It will, of course, be understood that the foregoing description is of preferred exemplary embodiments of the invention and that the invention is not limited to the specific embodiments shown. Various changes and modifications will become apparent to those skilled in the art and all such changes and modifications are intended to be within the scope and spirit of the present invention as defined in the following claims.
Applicant claims the benefit of U.S. Provisional Application No. 60/389,676, filed Jun. 18, 2002.
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| 60389676 | Jun 2002 | US |
