US3238725A - Pump and turbine fluid drive unit - Google Patents

Pump and turbine fluid drive unit Download PDF

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US3238725A
US3238725A US221616A US22161662A US3238725A US 3238725 A US3238725 A US 3238725A US 221616 A US221616 A US 221616A US 22161662 A US22161662 A US 22161662A US 3238725 A US3238725 A US 3238725A
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rotor
fluid
pump
flow
turbine
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Ludin Ludwig
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    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16HGEARING
    • F16H41/00Rotary fluid gearing of the hydrokinetic type

Description

' March 8, 1966 L. L UDlN 3,238,725
PUMP AND TURBINE FLUID DRIVE UNIT Filed Sept. 5, 1962 3 Sheets-Sheet 1 INVENTOR Ludwig Ludin BY M, 7 ATTORNEYS March 8, 1966 LUDlN 3,238,725
PUMP AND TURBINE FLUID DRIVE UNIT Filed Sept. 5, 1962 3 Shuts-Sheet 2 44 #IFIGA.
INVENTOR BY Ludwig Ludin jmfx,
747% 7 ATTORNEYS March 8, 1966 UD|N 3,238,725
PUMP AND TURBINE FLUID DRIVE UNIT Filed Sept. 5, 1962 3 Sheets-Sheet 5 INVENTOR Ludwig Ludin jhuwD W 1170 w ATTORNEYS United States Patent 3,238,725 PUMP AND TINE FL DRIVE UNIT Ludwig Ludin, Niederwieierstr. 10, Wohlen, Switzerland Fiied Sept. 5, 1962, Ser. No. 221,616 3 Claims. (Ci. 60-54) This application is a continuation-in-part of pending application Ser. No. 671,114, filed July 5, 1957, now abandoned.
This invention relates to fluid drive units and more particularly to fluid drive units having at least one rotary pump which is adapted to be driven, at least one turbine adapted to drive other apparatus, and fluid connections joining the pump and turbine whereby power may be transmitted between the two.
The power transmitted between the pump and turbine of conventional fluid drive units is usually dependent on the amount of fluid circulated between the two, a lesser circulation resulting in smaller torque being transmitted to the turbine. In such units, however, there is usually an air space resulting in a boundary layer existing in the pump and turbine in which objectionable frothing will occur. Special fluids, for example, polyglycols, having low surface tensions have been utilized as the fluid circulated in the drive units to reduce frothing, however, these fluids are not entirely satisfactory because of their viscosity resulting in the units operating under low Reynolds number conditions where the efliciency of the units is low.
A further diificulty of these conventional units is that they require use of a separate fluid reservoir into which fluid may flow when a low amount of power is being transmitted between the pump and turbine and they further require separate servo pumps for pumping, the fluid either to or from the reservoir.
I intend to provide a fluid drive unit which eliminates the requirement of a separate reservoir and pumps and which will allow use of fluids having low viscosities in order that the drive units may operate under flow conditions having higher Reynolds number conditions, while at the same time providing a means whereby the power transmitted between the pump and turbine may be readily controlled. I further intend to provide for a unit having a closed fluid circuit filled with fluid to eliminate any air spaces to reduce the possibility of frothing of the fluid when the drive unit is operating.
A fluid drive unit constructed according to my invention comprises generally a pump having its inlet and outlet connected to the outlet and inlet respectively of a turbine to form a closed fluid circuit which in operation is maintained full of fluid and wherein the pump has a hollow cylindrical bladed rotor and guide means which guide the throughput of the pump through the rotor and twice through the path of the rotating blades in a direction transverse to the rotor axis. Adjustable control means may be provided to throttle the fluid circuit to any predetermined degree between unthrottled and minimum or no flow condition to vary the power transmitted between the pump and turbine.
This simple solution to the problem of controlling the power transmitted is not readily available in conventional fluid drive units which normally employ axial pumps. With such units the hydraulic efliciency decreases when the amount of fluid passing through the pump is reduced by throttling. The decrease of the power input is negligible, and it may even increase on throttling so that it reaches a maximum when the power transmitting circulation has been completely stopped. When the power transmitted is zero, an axial pump commonly absorbs about 130% of the power required to operate at full load.
3,238,725 Patented Mar. 8, 1966 By contrast the pump of the drive unit according'to the invention requires less power as its throughput is throttled as will be pointed out more fully below, and when flow through the pump is stopped completely, the power consumption can be as little as 4% of power at full load.
The invention further provides for the use of fluids having a low kinematic viscosity resulting in a high efficiency of the unit because of the higher Reynolds numbers achieved. Examples of such fluids are water and mercury. Since once the fluid has been pumped into the casing it remains there, and also since no air space is present, Water can be used without its corrosive properties being detrimental to the unit. For the same reason mercury can now be used without the risk of mercury fumes escaping.
The casing of the unit is preferably built to withstand pressures so that the fluid inside can be pressurized to avoid cavitation. It is further advisable to provide for expansion of the fluid. It will be seen however that it is not necessary to provide a reservoir to receive fluid at low degrees of power transmission, nor a servo pump to pump fluid between the reservoir and the casing.
. In one preferred embodiment of the invention, the rotor and guide means of the pump are so designed that when the unit is operated under unthrottled conditions, the fluid flow through the pump sets up an approximately cylindrical vortex including a field region with a velocity profile approximating that of a Rankine vortex with a core region eccentric to the rotor axis to induce the stream tubes nearest the vortex core to change direction by at least in passing through the rotor from the suction to the pressure side thereof. Various forms of guide means for forming such a vortex are disclosed in application No. 671,114.
In another embodiment of the invention the pump guide means includes guide bodies located within the rotor to lead the fluid through it and through the path of the rotating blades. Various forms of such guide bodies are described in copending application 671,114.
The turbine, like the pump, conveniently comprises a cylindrical bladed rotor arranged for flow of fluid twice through the path of the rotating blades in a direction transverse to the turbine rotor axis and with such a turbine, a wide variety of arrangements of pump and turbine are possible, as will be shown, and the whole unit can if desired be made very compact.
Various embodiments of the invention are illustrated by way of example in the accompanying diagrammatic drawings, in which:
FIG. 1 is a transverse section of one form of fluid drive unit according to the invention;
FIG. 2 is a graph illustrative of the functioning of the FIG. 1 drive unit;
FIG. 3 is a transverse section of a second form of drive unit;
FIG. 4 is an axial section of the FIG. 3 unit taken along the line IV1V;
FIG. 5 is a transverse section of a further form of drive unit;
FIG. 6 is a transverse section of another embodiment of drive unit;
FIG. 7 is a transverse section of a further embodiment of drive unit;
FIG. 8 is an axial view of a different form of drive unit; and,
FIG. 9 is an axial section of a form of drive unit according to the invention.
Referring to the drawings, the drive unit of FIG. 1 comprises a pump designated generally P and a turbine designated generally T both mounted within a fluid-tight casing C which is entirely filled with liquid. A guide body 1 extends between the pump P and the turbine T and defines with the casing a pressure region 2 for flow between the pump and the turbine and a suction region 3 for return flow from the turbine back to the pump such that the pump and turbine are located in a closed hydraulic circuit.
The pump P has a hollow cylindrical rotor 4 the ends of which are closed and which is mounted for rotation about its axis. The rotor 4 has blades 5 which are concave facing the direction of rotation, indicated by the arrow 6, and have their outer edges leading their inner edges. One side of the rotor 4 lies close to the casing C, while the guide body 1 presents a wall 7 to the other side of the rotor which converges therewith in the direction of rotor rotation. By reason of the shape of the blades 5 and the character of the guide means provided by the wall 7, there is formed and stabilized, upon rotation of the rotor, a cylindrical vortex which is substantially a Rankine vortex. This vortex has a core indicated by the lines designated V and a field region wherein flow takes place as shown by the flow lines F, the flow tube with maximum velocity passing adjacent the coreand being designated MP. In passing through the pump P the whole throughput traverses the path of the rotating blades 5 twice and flows in a direction always perpendicular to the rotor axis. The direction of the maximum-velocity flow tube MP is changed approximately 180 in passing through the rotor from the suction region 3 to pressure region 2. The effect that the vortex has on flow fluid through the rotor is to increase the velocity of flow over portions of the rotor which in turn increases Reynolds number at these portions. Reynolds number is defined as R6=d'C/'y where d is the blade depth radially of the rotor, c is the peripheral speed of the rotor, and 'y is the kinematic viscosity of the fluid, the latter being equal to the quotient of the dynamic viscosity and density. A Reynolds number is considered herein to be low, as above defined, when it is less than 5x10 It is known that in flow machines which operate under low Reynolds number conditions, i.e., 50,000, where the flow machine is of the cross flow type having blades curved in the direction of rotation, that there is an initial acceleration and subsequent deceleration of the flow in boundary layers on the suction side of each blade as fluid passes over the blade. The higher the viscosity of the fluid in relation to its density or in relation to relative velocity between the fluid and the blade (i.e., the lower the Reynolds number),
the greater is the deceleration of the boundary layer in the deceleration zone of the blade. As the boundary layer is slowed down sufficiently, it separates from the blade and no longer follows the blade contour. The point at which separation occurs is known as the separation point. The separation point travels forward along the surface of the blade against the direction of flow in proportion to the increase in the effect of the viscosity relative to density or to the decrease in the relative velocity between the fluid and the blade.
The movement forward of the separation point along the blade because of low Reynolds number conditions produces a number of undesirable effects in a conventional cross flow machine. A vorticity zone in which the kinetic energy of the fluid is converted into thermal energy is produced after the separation point with the result that the efliciency of the flow machine drops. The degree of deflection of the fluid in passing through the rotor and through the path of the rotating blades decreases owing to the fact that the flow does not follow the full extent of blade profile but becomes nonlaminar beginning at the separation point. This results in less pressure gain in the machine since pressure gain is determined by the extent of the deflection of the stream tubes in the blade channel. In addition, the turbulent flow in back of the separation point effectively reduces a part of the cross section of the blade channel so that the throughput through the rotor of the machine also diminishes.
The vortex V shown in FIG. 1 is formed by action of the guide body 1 cooperating with the curved blades 5 of the rotor where a part of the output flow of the machine is directed back to the inlet. The vortex being of low pressure increases the velocity of flow of fluid immediately adjacent it. In addition, since the fluid passing through the,rotor of the FIG. 1 machine may be of low kinematic viscosity, the Reynolds number may be further increased thus still further increasing efliciency.
The turbine T has also a cylindrical rotor 10 closed at its ends and mounted for rotation about its axis which is parallel to that of the pump rotor. The turbine rotor has blades 11 which are convex facing the direction of rotation shown by the arrow 12 and have their inner edges leading their outer edge Within the rotor 16 are guide bodies 13-18 which guide fluid across the interior of the rotor so that it traverses the path of the rotating blades twice and flows always in a direction perpendicular to the rotor axis. The interior guide body 1'3 presents a wall closely adjacent the inner envelope of the rotor blades 11 opposite a wall presented to their outer envelope by the body 1 and these walls subtend an angle slightly over 45 at the rotor axis. The opposite side of the rotor lies closely adjacent the casing C over some 180 of its circumference, while opposite the casing, the interior guide body 14 presents a wall close to the inner envelope of the blades 11. Thus flow through the turbine T is substsntially confined to the central section of the rotor 10 wherein the other interior guide bodies 15-18, which are of aerofoil shape, are mounted.
The cross-section of the unit is constant over the interior of the casing and the various guide bodies extend the length of the rotor.
The pump is, in use of the unit, driven by some source of power while the turbine is made to drive some further apparatus. Power is transmitted between the pump and turbine by the liquid.
As mentioned heretofore any liquid can be used although, preferably, a liquid having a low viscosity is utilized since such a liquid effectively increases Reynolds number conditions of fluid flow as compared with a liquid having a high viscosity. Further the liquid utilized can be chosen for its flow characteristics without reference to its frothing properties because, since there is no surface between the liquid and air, no possibility of frothing exists. As previously explained, it is desirable to provide for expansion of the liquid and to pressurize the casing. The region of the vortex core is a low pressure region and there would be a tendency to cavitation if the liquid was not under static pressure. It should be noted, that kinetic energy in the fluid which is not absorbed by the turbine is returned to the pump and is not lost. Thus the efiiciency of the unit is greater than the efliciencies of a pump and a turbine considered in isolation from one another. The remarks of this paragraph apply not only to the embodiment illustrated in FIG. 1, but also to the other embodiments described hereafter.
FIG. 2 is a graph showing the effect of throttling a pump constructed according to the invention. In this graph, the ordinates represent power input and the abscissae the flow rate. The full line represents the performance of the pump and the dotted line that of a comparable axial flow pump. The point a corresponds to maximum throughput (or zero throttling) and the po1nt b corresponds to zero throughput. At zero throughput, the pump consumes about 4% of the power at point a, while the axial pump consumes some of that power. Studies of the operation of the pump indicate that as the throughput is throttled at constant speed, the fluid vortex moves gradually towards the axis of the rotor, and when flow is shut off entirely, the vortex is virtually concentric with the rotor so that the very small retarding forces remaining are mainly frictional.
FIG. 3 shows a fluid drive unit where the turbine rotor 32 coaxially surrounds the pump rotor 31, the blades of the two rotors being spaced radially only by a working clearance. Apart from this difference, the pump and turbine rotors which are enclosed in a fluid filled casing C are generally similar to those of the unit of FIG. 1. Within the rotor 31 are situated a pair of guide bodies 33 and 34, and also a series of vanes extending across the rotor with opposite ends spaced only slightly from the inner blade envelope. The bodies 33 and- 34 co-operate respectively with an exterior guide body 36 and the casing C in the same general manner as the bodies 13 and 14 in FIG. 1 co-operate with the body 1 and casing of that figure. The vanes 35 perform the same general function of guiding flow of hydraulic fluid across the interior of the rotor 31.
The casing C defines with the turbine rotor and guide body 36 a crescent-shaped passage 37 for flow of fluid emerging from the turbine rotor over the arc thereof 38 back into the turbine over the are 39. Because of this shape, the passage acts as a difluser followed by a nozzle and thereby minimizes energy losses in turning the flow through the desired angle. Between the guide body 36 and the casing are located three balanced butterfly throttling valves 40 movable in unison between a position in which they block or throttle flow completely to an unthrottled position in which they present negligible resistance to flow.
The cross-section of the unit is the same throughout the length of the pump and turbine blades so that flow is always perpendicularly to the axis. The pump and turbine are located in a closed hydraulic circuit with the result that substantially the whole throughput twice traverses the path of the rotating blades of both rotors. The degree of coupling or power transmission between the rotors can be varied evenly and with negligible expenditure of energy by control of throttling valves 4t). The relation between throttling and power taken by the pump rotor of the FIG. 3 unit will be substantially as shown in FIG. 2.
FIG. 4 illustrates further constructional details of the FIG. 3 drive unit. The stationary guide bodies 33 and 34 and vanes 35 form an assembly shown diagrammatically at 42 which is carried on a fixed cylindrical post 43 projecting horizontally from a fixed support 44. Both rotors 31 and 32 have rigid end walls 31a and 32a. The rotor 31 is carried on the post 43 by means of a coaxial sleeve 31b freely rotatable thereon and extending from one end wall 31a. A spindle 310 extends from the other end wall 31a of the rotor 31 and is coaxial therewith. The spindle 31c carries a pulley 45 whereby drive is transmitted to the unit and its inner extremity 31d projects within a bore 42a of the assembly 42 so as to be steadied thereby. The turbine rotor 32 is carried on a pair of coaxial sleeves 32b, and 320 each extending from an end wall 32a with the sleeve 32b being freely rotatable on the spindle 31c and the sleeve 320 being freely rotatable on the sleeve 31b of the rotor 31. A pulley 45' fastened with the sleeve 32c transmits drive from the unit. The casing C is not shown in FIG. 4.
FIG. 5 shows a heavy duty drive unit where the turbine rotor is concentric with and surrounds the pump rotor 51 but where, however, a considerable annular space separates the rotors. The rotors 50 and 51 are similar to those of FIG. 1 and will not be further described. A sealed casing C completely encloses both pump and turblue and is completely full of fluid. A pair of stationary guide bodies 53 and 54 are located in the space between the rotors 5t) and 51 and define therewith a suction region S and a pressure region P. The casing C together with a guide body 55 outside the turbine rotor 50 defines a crescent-shaped passage 56 for flow of fluid between the arcs of the turbine rotor opposite the pressure and suction regions respectively, the shape of the passage minimizing energy losses in bending of the fluid. Opposite the passage 56 the casing C and guide body 54 present arcuate walls closely conforming to the inner and outer envelopes of the turbine rotor and the guide bodies 53 and 55 also present walls conforming to those envelopes but at the side of the rotor adjacent the passage 56. By this means substantially all the fluid is forced to flow in the closed circuit from the suction region through the pump rotor to the pressure region and thence to the passage 56 and back to the suction region. The whole throughput apart from negligible leakages will thus pass twice through the path of the rotating blades of both rotors 50 and 51 in a direction perpendicular to their common axis.
A portion 53a of the guide body 53 presents to the rotor 51 a concave Wall which considered as a whole converges with the rotor in the direction of rotation. An auxiliary guide body 53b is located in the Space so formed. Opposite the guide body portion 53a, the guide body 54 presents a wall 54a starting from a point near the rotor and extending away therefrom with gradually increasing curvature. In operation the rotor 51 and guide body 53 set up and stabilize a vortex approximating a Rankine vortex, having a core region indicated diagrammatically by the flow line shown in chain-lines at V, and a field region where flow takes place as shown by the flow lines F. The flow conditions in the region of the rotor 51 are generally similar to those in the region of the pump rotor of FIG. 1 and will not be further described. The wall 54a in combination with the wall 530 of the guide body portion 53a defines the pressure region P and these walls diverge to form a diffuser section. Sets of vanes 57, 58, 59, are provided as shown to guide fluid into the blades of the rotors 50, 51.
FIG. 6 shows a fluid drive unit which makes use of the appreciation that a pump and turbine of the type described, that is having cylindrical bladed rotors the path of the rotating blades of which are traversed twice by the whole fluid throughput in a direction transverse to the rotor axis, can operate eflectively with only a portion of the cross-section of the rotor available for flow. Certain flow machines of the type just mentioned and utilizing this appreciation are described in copending application 671,114.
In the FIG. 6 unit, a turbine rotor 69) surrounds a pump rotor 61 in closely spaced coaxial relation. The rotor arrangement itself resembles that of FIG. 3 and will not be further described. Both rotors are enclosed in a casing C of elliptical cross-section which is completely filled with fluid. At opposed portions 62 on the minor axis of the ellipse, which intersects the axis of the rotor, the casing approaches the rotor so as to be spaced therefrom by only a working clearance. The casing C accord ingly defines with the rotor 60 crescent-shaped circulation spaces 63a and 63b which are symmetrical upon the major axis of the ellipse. A stationary body 64 within the interior of the pump rotor 61 presents similar concave walls 64a and 64b thereto which are symmetrical about both axes of the ellipse and which define with the rotor 61 two chambers 65a and 65b. The body 64 presents convex Walls 64c opposite the casing portions 62 and spaced from the rotor 61 by only a working clearance. Thus the chamber 65a with the circulation space 63a is completely separated from the chamber 65b and the circulation space 63b apart from negligible leakages. In each chamber and circulation space, flow takes place as shown by the dotted lines F, and a fluid vortex V approximating a Rankine vortex is formed. Each chamber and circulation space can thus be regarded as a separate closed circuit with its own pump and turbine, the pumps and turbines of the two circuits being, however, mechanically connected. In passing around each circuit, the fluid throughput twice traverses the path of the rotating blades of both pump and rotor.
FIG. 7 shows a further refinement of the FIG. 6 unit. In this construction pump and turbine rotors and 71 similar to those of FIG. 6 are enclosed in a casing C completely full of fluid. The casing C provides ten circulation spaces '72. A body 73 positioned within the pump rotor 70 provides ten chambers 74 each radially opposite one of the spaces 72. Intermediate the spaces 72 and chambers 74, the casing C and body 73 are spaced by only a working clearance from the respective rotor blade envelope so that each circulation space with the corresponding chamber forms an independent circuit in which flow takes place as shown in FIG. 6.
FIGS. 8 and 9 show two arrangements of a fluid drive unit such as shown in FIG. 1 where the pump and turbine rotors are not disposed one inside the other. In these figures the pump and turbine rotors S and 81 and 90 and 91 respectively are shown in elevation and the connecting fluid passages 82 and 92 respectively in section. In each case a casing C completely filled with fluid encloses both rotors and provides the connecting passages. In FIG. 8 the input and output shafts 83 and 84 carrying the pump and turbine rotors 80 and 81 respectively lie at an angle. Although the shafts 83 and 84 are shown with their axes coplanar and intersecting, they can if desired be made skew. In FIG. 9 the corresponding shafts 93 and 94 are coaxial but revolve in opposite directions. Vanes to guide flow in the bends in the passages 82 and 92 are shown at 85' and 95. Guide means of the types disclosed associated with the rotors in FIGS. 1, 3 or are associated with the rotors in order to achieve the desired flow within the rotor, however, for clarity, they have not been illustrated.
I claim:
1. A fluid drive unit comprising a cross-flow type pump having a hollow bladed rotor, auxiliary guide means associated with said rotor whereby, on rotation thereof, fluid is caused to flow from a suction side of said rotor through the path of the blades into said rotor and thence out through the path of the blades to a pressure side, a bladed turbine wheel concentric with said rotor, a casing surrounding said rotor and wheel, guide bodies for directing flow from said rotor towards a portion of said wheel whereby fluid flows centrifugally through said wheel, and a discharge guide body positioned between said casing and radially outwardly and adjacent to said wheel to form with said casing a recirculating conduit for guiding flow from said wheel around a portion of the circumference thereof to an inlet area where the fluid flows radially inwardly through said wheel to the suction side of said rotor.
2. A fluid drive unit according to claim 1 having in addition guide vanes positioned on the pressure side of said rotor and adjacent said turbine wheel for directing flow against the blades of said wheel.
3. A fluid drive unit according to claim 1 wherein a portion of the flow from the outlet of said rotor is recirculated back to the rotor inlet without passing through the blades of said wheel.
References Cited by the Examiner UNITED STATES PATENTS 1,141,812 6/1915 Michell et al 54 1,291,871 1/1919 Hein 6054 1,428,586 9/1922 Garrison 60-54 1,920,952 8/1933 Anderson 230- 2,032,398 3/1936 Brady 6054 X 2,658,700 11/1953 Howell 24415 2,942,773 6/1960 Eck 230125 2,965,284 12/1960 Coester 230125 X FOREIGN PATENTS 291,007 8/1928 Great Britain.
JULIUS E. WEST, Primary Examiner.

Claims (1)

1. A FLUID DRIVE UNIT COMPRISING A CROSS-TYPE PUMP HAVING A HOLLOW BLADED ROTOR, AUXILIARY GUIDE MEANS ASSOCIATED WITH SAID ROTOR WHEREBY, ON ROTATION THEREOF, FLUID IS CARRIED TO FLOW FROM A SUCTION SIDE OF SAID ROTOR THROUG THE PATH OF THE BLADES INTO SAID ROTOR AND THENCE OUT THROUGH THE PATH OF THE BLADES TO A PRESSURE SIDE, A BLADED TURBINE WHEEL CONCENTRIC WITH SAID ROTOR, A CASING SURROUNDING SAID ROTOR AND WHEEL, GUIDE BODIES FOR DIRECTING FLOW FROM SAID ROTOR TOWARDS A PORTION OF SAID WHEEL WHEREBY FLUID FLOWS CENTRIFUGALLY THROUGH SAID WHEEL, AND A DISCHARGE GUIDE BODY POSITIONED BETWEEN SAID CASING AND RADIALLY OUTWADLY AND ADJACENT TO SAID WHEEL TO FORM WITH SAID CASING A RECIRCULATING CONDUIT FOR GUIDING FLOW FROM SAID WHEEL AROUND A PORTION OF THE CIRCUMFERENCE THEREOF TO AN INLET AREA WHERE THE FLUID FLOWS RADIALLY INWARDLY THROUGH SAID WHEEL TO THE SUCTION OF SAID ROTOR.
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Cited By (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3481530A (en) * 1968-01-17 1969-12-02 Anatoly Grigorievich Korovkin Diametral fan
US4042355A (en) * 1974-10-15 1977-08-16 Pearson Paul W Pollution control device
US6047765A (en) * 1996-08-20 2000-04-11 Zhan; Xiao Cross flow cooling device for semiconductor components
US20080004090A1 (en) * 2006-06-30 2008-01-03 Ricketts Jon E Rotating inlet for cross flow fan

Citations (9)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1141812A (en) * 1912-12-05 1915-06-01 Anthony George Maldon Michell Hydraulic mechanism for transmission of power.
US1291871A (en) * 1918-06-17 1919-01-21 William Hein Transmission mechanism.
US1428586A (en) * 1919-02-04 1922-09-12 Garrison Benjamin Flexible power transmission
GB291007A (en) * 1927-05-23 1928-08-02 Harald Dalin Improvements in rotary fans
US1920952A (en) * 1931-01-02 1933-08-08 American Blower Corp Line flow fan
US2032398A (en) * 1934-03-30 1936-03-03 Rca Corp Film drive mechanism
US2658700A (en) * 1943-07-28 1953-11-10 Power Jets Res & Dev Ltd Turbocompressor power plant for aircraft
US2942773A (en) * 1953-07-17 1960-06-28 Paul Pollrich & Comp Fans
US2965284A (en) * 1955-11-24 1960-12-20 Benninger Ag Maschf Turbo-machine

Patent Citations (9)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1141812A (en) * 1912-12-05 1915-06-01 Anthony George Maldon Michell Hydraulic mechanism for transmission of power.
US1291871A (en) * 1918-06-17 1919-01-21 William Hein Transmission mechanism.
US1428586A (en) * 1919-02-04 1922-09-12 Garrison Benjamin Flexible power transmission
GB291007A (en) * 1927-05-23 1928-08-02 Harald Dalin Improvements in rotary fans
US1920952A (en) * 1931-01-02 1933-08-08 American Blower Corp Line flow fan
US2032398A (en) * 1934-03-30 1936-03-03 Rca Corp Film drive mechanism
US2658700A (en) * 1943-07-28 1953-11-10 Power Jets Res & Dev Ltd Turbocompressor power plant for aircraft
US2942773A (en) * 1953-07-17 1960-06-28 Paul Pollrich & Comp Fans
US2965284A (en) * 1955-11-24 1960-12-20 Benninger Ag Maschf Turbo-machine

Cited By (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3481530A (en) * 1968-01-17 1969-12-02 Anatoly Grigorievich Korovkin Diametral fan
US4042355A (en) * 1974-10-15 1977-08-16 Pearson Paul W Pollution control device
US6047765A (en) * 1996-08-20 2000-04-11 Zhan; Xiao Cross flow cooling device for semiconductor components
US20080004090A1 (en) * 2006-06-30 2008-01-03 Ricketts Jon E Rotating inlet for cross flow fan
US7731577B2 (en) 2006-06-30 2010-06-08 Cnh America Llc Rotating inlet for cross flow fan

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