US2547392A - Continuous contact internal rotor for engines - Google Patents

Continuous contact internal rotor for engines Download PDF

Info

Publication number
US2547392A
US2547392A US659098A US65909846A US2547392A US 2547392 A US2547392 A US 2547392A US 659098 A US659098 A US 659098A US 65909846 A US65909846 A US 65909846A US 2547392 A US2547392 A US 2547392A
Authority
US
United States
Prior art keywords
teeth
tooth
rotor
ratio
rotors
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Expired - Lifetime
Application number
US659098A
Inventor
Myron F Hill
Francis A Hill
Original Assignee
Myron F Hill
Francis A Hill
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Myron F Hill, Francis A Hill filed Critical Myron F Hill
Priority to US659098A priority Critical patent/US2547392A/en
Application granted granted Critical
Publication of US2547392A publication Critical patent/US2547392A/en
Anticipated expiration legal-status Critical
Expired - Lifetime legal-status Critical Current

Links

Images

Classifications

    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01CROTARY-PISTON OR OSCILLATING-PISTON MACHINES OR ENGINES
    • F01C1/00Rotary-piston machines or engines
    • F01C1/08Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing
    • F01C1/10Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing of internal-axis type with the outer member having more teeth or tooth-equivalents, e.g. rollers, than the inner member
    • F01C1/103Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing of internal-axis type with the outer member having more teeth or tooth-equivalents, e.g. rollers, than the inner member the two members rotating simultaneously around their respective axes

Description

April 3, 1951 I M. F. HILL ET AL 2,547,392
CONTINUOUS CONTACT INTERNAL ROTOR FOR ENGINES Filed April 2, 1946 3 Sheets-Sheet 1 VENTOR MYRO .H/LL, FRANCIS .4. H/LL 2nd ATTORNEY April 3, 1951 M. F. HILL; ETAL v CONTINUOUS CONTACT INTERNAL ROTOR FOR ENGINES Filed April 2, 1946 RIOR AR 3 Sheets-Sheet 2 INVENTOR MYRO/V HILL,
FRA/VG/SA. H/LL 2nd ATTORNF April 3, 1951 M. F. HILL ET AL commuous CONTACT INTERNAL ROTOR FOR ENGINES s Sheets-Sheet 5 Filed April 2, 1946 FIG. ZZZ /2/ INVENTOR MYfPO/V F. H/LL FRANCIS A. HILL 2nd uvg m/ 7. b/uu,
ATTORNEY Patented Apr. 3, 1951 I CONTINUOUS CONTACT INTERNAL ROTOR FOR ENGINES Myron F. Hill and Francis AT Hill, 2nd, Westport, Conn.
Application April 2, 1946, Serial No. 659,098
18 Claims. (Cl. 103-426) 1 This application is a continuation in part of our applications Nos. 227,954, filed September 1, 1938, now Patent 2,386,896; certain features of application No. 452,654, filed July 28, 1942, now
. abandoned; and 561,948, filed November 4, 1944; and contains new matter not therein shown or described; all constituting a step by step development over many years.
It relates generically to rotary fluid toothed displacement rotor mechanisms including liquid pumps and motors capableof high pressures, not omitting low pressure gas blowers and motors, and' speclfically having the inner displacement member or pinion rotor as the driving member connected to a drive shaft or other device; including also a difference of more than one tooth. It relates. also to ports for such rotors and driving relations, conforming to special contours of the rotor teeth.
These rotor contours are based on new prin- 1 ciples of geometry which have escaped the attention of designers of rotors and gears.
They are explained in other cases and herein also with diagrams. The invention itself however includes contours and cooperating forms of me'chanical displacement members producing new results in 2 applications to rotary displacement mechanisms J and gearing.
A;series of patents to M. F. Hill, particularly Re-issue Patent No. 21,316, described rotors having a diiierence of one in numbers of teeth 1 which have continuous contacts or fluidtight engagements forming rotor chambers which open andclose during rotation. As they rotate, acting as'a, pump for example, each tooth of one rotor enters and leaves all the tooth spaces of the other rotorf driving out their fluid contents and sucking in more. Each tooth slides or rolls over all the ,teth of the other rotor, providing a hunting relation (interchange of tooth engagements) between them all. such precision that this contact persiststhruout rotation. They have tight fits between all the teeth, and as pumps and hydraulic motors have been contributors to swift and precision operation of tanks, planes, ships and amphibians in war service. Also as lubricating pumps in autos and boats. Their high efficiency as lubricating pumps for engines and superchargers in planes provided flight at high altitudes. Across the open mesh region the tooth crowns of one rotor slide over the tooth crowns of the other rotor. With usual moderate oil pressures it is a pressure-less contact.
An open crescent space across open mesh between teeth out of contact, where notooth con- The rotors are made with I back lash betweenthe teeth, between teeth performing no useful pressure function. Contacts between teeth of opening chambers and teeth of closing chambers at the same time cannot take place with back lash. The difference of two (or more) in the numbers of teeth, such numbers having no common divisor, provides abasic fractional ratio having a dilierence of one which is essential for maintaining continuous travelling engagements between the teeth, and to generating the rotor curves accurately at a single setting.
Such rotor contours require specially designed ports. The difficult resettings in manufacturing rotors tacts occur and prevent contact having even numbers of teeth are eliminated, as
well as their hammering and uneven wear.
Back lash between our rotor teeth of continuous contact acquires a new function. Shaft bearings with enough looseness to run, let the teeth of one rotor ride on the other under high internal fluid pressures, tending to heat the rotors to the point of binding. Back lash and the crescent space at open mesh prevent it.
These traveling engagements and contacts, in
operation, lap themselves to a fit that is evident upon inspection of teeth in service. A polish occurs upon contacting surfaces. It may be observed that a mostly rolling contact occurs at iullmesh where the teeth of one rotor drive the teeth of the other, and between the tooth surfaces in contact along a port, where a rub occurs. If with manufacturing or mass production tolerances, some teeth fail to make actual contact, the rub elsewhere wears oiT the tooth surfaces until they all engage. Nevertheless, when first made they are to be in such proximity as to have driving and'pressure'holding engage merits capable of running themselves in, in
service. Departure from the Gerotor method of designing rotor contours makes possible the hunting relation between all the teeth of the rotors, larger displacement and improved pressure angles in the driving range between the teeth at full mesh.
I The crescent space, which requires no metal or other insert to seal the ports from each other as in other fluid mechanisms; also helps to prevent the pinion from contacting and. riding on the teeth of the" outer rotor from one end of the crescent to. the other. This crescent space is as the area at open mesh between two circles, one
passing thru the tips of the pinion teeth and the other thru the tips of the teeth of the outer rotor.
Back lash was old, but combined with the con,-
tinuous contacts, it prevents the teeth of the pinion from riding on the teeth of the outer rotor where no fluid pressure holding engagements are needed, between full mesh and one end of the crescent space leaving the continuous contacts and engagements between the other end of the crescent space and full mesh unafiected. A difference of two or more teeth between rotors or gears, old in themselves, performed unexpected results with continuous contact rotors, causing the rule of continuous contact to be modified, first to include multiples of the ratio having a difference of one; and then modified again to include the hunting relation which solved a manufacturing problem.
It also permitted rotor teeth to'be' built loose, with the contacts brought into action by driving, whether by a shaft or by fluid pressure. Thus rotors may lap themselves to a perfect fit in service to provide rotor chambers sealed at both ends, and to provide interchangeable and or reversible high and low pressure ports for pumps and motors for operation in opposite directions. The terms high and low merely differentiate between the ports. I
These step by step developments, each valuable in itself, produce new results in continuous contact rotors. I x
The first difference of twoteeth appeared when, inspecting Fig. IX of the reissue patent to M. F. Hill, No. 2l,3l6, the tooth spaces of the pinion appeared wide enough to hold another tooth, and the outer teeth wide enough for another tooth space. This doubled the number of teeth and they acquired a difference of two which provided the much needed crescent space to eliminatethe riding of one rotor on the teeth of the other at open mesh, without loss of the continuous contacts. See also Patent No. 2,386,- 896, Fig. XVI having even numbers of teeth. Useful in reduction gearing this idea was patented in patents, Nos. 2,091,317 and 2,209,202, to F. Hill. The disclosure in these patents was limited to multiples of rotor teeth having a. diiference of one for gears. The contours and pressure angles of the Reissue Patent 21,3'1'6 characterized the teeth. Furthermore, in multiplying the teeth, a part of the tooth height was sacrificed; and whatever was done to them thereafter, such as reducing the numbers of teeth again while increasing the eccentricity, the lost height still prevailed. The height of a tooth affects the size of rotor chambers between the teeth,- and so limits the displacement as compared with our later form. Incidentally the centers of the master curvesfor two sides of each tooth were doubled in number, and were evenly indexed with each other. The multip-led gear teeth had the same type of contours near the convex tops or crowns that the original rotors had.
The gear patents disclosed no pump casing or ports. V I
When teeth are multipled, even tho they had a hunting relation between them, the hunting relation between all the teeth is lost.
In our Patent #2,386,896 the off sides of adjoining teeth of the outer rotor are limited to a common center, and a similar relation exists between the near sides of alternate pinion teeth. Our present rotors have no such limitation.
Rotors having a difierence of two teeth, resultin from multiplying smaller ratios of integers differing by one, do not have the hunting relation so essential to easy manufacture and good service. To index two sets of teeth with relation to each other with the exactness needed for efiicient fluidtight high pressure relations, if not impossible is at least a great expense, and in mass production is difiiicult to maintain. In generation, one set of teeth is first generated on a blank. The blank must then be indexed exactly half way, and the second set generated. An
error in indexing results in hammering of teeth,
noise, lost efficiency and wear. The usual manufacturing tolerance would ruin rotors for efficient service. With a full hunting relation there is no second indexing, and the exact indexing char"- acteristic of generation of all the teeth is aocom plished. The hunting relation enables each tooth to engage in proper turn all the teeth of the other rotor. This evens up any possible wear due to diiferences in hardness or durability of teeth, or grit, and maintains a smooth operating mechanism. I
Realizing this, another tooth was added to each rotor, changing the ratio circles accordingly, and it was discovered to be possible. It was a departure from simple multipling, and eliminated the difiicult second indexing. Removing a tooth has the same effect. 8 x 10 rotors might be altered to 7 x 9 or g x 11 for example.
Then came another conception, a more scientific method of designing rotor contours for continuous cont-act, particularly useful for rotors having any difference in numbers of teeth greater than one. A diiference of two increased the height of a rotor tooth, and added displacement. It made possible a better location or inclination of the sides of rotor teeth to get better pressure angles. It made possible the location of driving surfaces across the ratio circles instead of mostly outside of them, thus reducing angular slip. It made possible the more intelligent layingout of rotor teeth, selecting the best radii of curvature, and determining the circroidal addition more accurately than with the generator machine. Mathematical calculations are possible; but are most intricate and take much time. But in a graph, comparison of diiferent curves is easy. It is regarded as the final step in adapting this new type of rotors for commercial development.
I In new geometrical relations rules or laws are sought to guide those who are to become skilled in the art. The first rule that appeared important, was that in gears, one within the other, there must be a difference of one tooth only, since a tooth of one had to travel continuously over all the teeth of the other and the speed ra= tio had to be the ratio of the teeth. This was accepted for many years by those developing and commercializing this art. The initial patents, to M. F. Hill contained claims limited to the differ ence of one tooth. I
Whenit was discovered that the teeth of a ratio differing by one could be doubled, to make a difference of two, or trebled to make a difierence of three, the rule had to be modnfied, and extended to include the multiples.
Then finally it was discovered that it was not the ratios of teeth that were controlling but the basic ratios themselves which must have a dinerence of one, even with ratios of fractional numbers.
and finally also it appeared that the hunting relation between all the teeth was possible only for teeth based either on integers having a diff'e'r'ence of one as in the original rotors orona fractional ratio, having a difference of one, the
t actual number of teeth being found by multiplying the ratio numbers by the denominator of their fractions. Merely multiplying teeth of rotors lost the hunting relation between all the teeth.
The lowest ratio having a good driving relation between the rotors may be 2% to 3 the rotors having five and seven teeth. They are not .multiples of any ratio having lesser numbers of teeth, and have greater displacement than four to five teeth of the same outside diameter.
A 1 /2 x 2 ratio, with 3 X 5 teeth is possible for low pressures tho the drive action is poor. A x 1 /2 ratio, having 1 x 3.teeth might require additional driving gears to keep them in registration. Multiples of very low ratios, 1 x 2 or2 x-3 doubled or trebled may fit some needs; all having the continuous tooth engagements.
With the pinion drive so called, a pinion is freely mounted on a drive shaft and key to find its own best running position, between the side walls while driving the outer rotor. Rotors as is well known, rotate relatively at speeds according to their tooth ratio. That is, speeds are determined by their relativenumbers of teeth. A 5 tooth pinion in driving a seven tooth outer rotor or displacement member, rotates faster than the seven tooth member. It rotates seven time to five turns of the outer rotor. Their rotor speeds therefore are, inversely proportional to their numbers of teeth. p
The five to seven tooth rotors have such radiallydeep rotor chambers'however that they leave little room for a shaft hole of a size needed for h igh pressure operation; so that then teeth of the pinion rotor are made integral with the shaft. Additional elements are required for endwise freedom; such as an Oldham clutch at the driving end'of the shaft. With seven to nine teeth or more (a ratio of 3 to l thereis ample room for a shaft hole for a high pressure shaftfwith the pinion rotor freely mounted and keyed to permit it to find its own best running position between side walls. 7
As these rotors rotate at ratio speeds wtih relation to each other, that is at speeds proportional to their numbers of teeth, directly or inversely, the new contours make continuous contacts and engagements with each other.
A difference in the numbers of teeth of a tooth ratio is not a matter of degree for another. reason." In rotors of the same size, a difierence of two gteeth increases displacement while a diiierence of three teeth reduces it, because of the longer; crescent space in the latter and fewer rotor chambers at any one instant to pump with. If tooth ratios have a common divisor, instead of beingindivisible by any common divisor, they require shifting the reiative settings of tool and rotor between generation of sets of tooth contours, introducing extra work, inaccuracies, noisy operation and more wear, so that the difierence is far reaching.
' With difierent numbers of teeth, critical intersections of certain normals to certain curves have different relative positions in the geometry of the Hill Theorem to be discussed later.
Rotors have bearings for the pinion shaft in side walls inside of ports which are often located in sidewalls, providing larger port access to ro-- tor chambers than radial ports. Seven to nine teeth allow ample room even for antifrictionbearings inside of the ports. These and many other factors difierentiate the various types and ratios of teeth.
The next stumbling block. encountered in the development of rotors with crescent spaces and back lash was port relations. As already noted rotors having a difiference of one tooth, and having no crescent space, had continuous contacts between all the teeth. Ports areseparated by abutm'ents, that is, areas in the enclosing walls between the ports to shut off escape, of liquid from one port to the other. It is the practice to make the abutments of such length that one rotor chamber is not disconnected froma port until it has connected to the other, directly or thru the crescent space, to avoid liquid look. It is called the overlap. But the ports are separated substantially by a tooth engagement that prevents dissipation of pressure. Now the crescent space shortens the'intake port of a liquid pump. The intake volume is increased by-the greater eccentricity (that of 3 inch 6 x '7 rotors is while that of 5 x 'Z, same'outs'ide diameter, rotors is the volume increasing with the ra tio and its eccentricity). 'Eadially deeper ports compensate, lying outside and inside of the path of tooth. engagement.
If in doubling the numbers of teeth in a liquid pump two fluidtight engagements occur between closing chambers disconnected from crescent or port areas, liquid lock ensues .and stoppage" of rotation. Increasing the difierence in the numbers of teeth therefore requires modification of port relations relative to the tooth contours.
With back lash between the teeth of a pump only one end of the crestcent space may be separated from a tooth chamber containing high liquid pressure by a fluidtight tooth engagement. The separation may take place at one end of the crescent only and that end depends on the kind of rotor drive applied, whether fluid or mechanical, and the direction of rotation.
Enough back lash to prevent the teeth on both sides of one .rotor from riding on the teeth of the other opens a crevice between the teeth which in the sizes shown, would'dissipate pressure, un'-' less otherwise prevented. The end of the path of contact occurs when the crevice starts to open. The location of a fiuidtight engagement overa range from one tooth to the next is afiected by the ending of the path of continuous engagement. In our present invention a fiuidtight contact or tooth engagement occurs in the full mesh region and prevents the escape of fluid between the teeth from one port to the other.
Paths of continuous engagement vary in different types of continuous engagement rotors. There are hosts of types all based upon the geometrical principle of the circroidal addition. There are three types of tooth ratios and variations of them. In the well known Gerotor" type of the prior art, having a difference of one tooth;
all of the teeth are tightly fitted to each other. The outer rotor has a circular curve on each tooth extending over the top and sides with a common center, all the tooth centers being evenly indexed. The second type is that shown in the gear Patents Nos. 2,091,317 etc. The outer teeth have the same side arcs of circles, but no completelypconvex crowns, and the arc centers are evenly indeXed. The third type, in this case, has teeth differing by two or more with a full hunting relation, preferably with a center of the arc on one side of each outer rotor tooth unevenly spaced between all the centers of the arcs on the other sides of the teeth, affording larger radii better pressure angles and greater displacement.
The paths of tooth engagements difier in these types. In the first or Gerotor form, the path is continuous around the teeth, with a :loop at full mesh, '(Fig. .XXI) of prior ,art. Without back lash in the second and third'type the path starts from the end of a crescent space and stops at full mesh. Another path .is the :reverse on the other side of the rotors. The third form is complicated 'further'by-the odd'ratio based on fractional numbers having a difference of one,xwhich has the hunting relation. The location of the fluid tight tooth engagement, or its range from one tooth to the next, is :affected'by these dinerences. With back lash between the teeth, one of the paths of contact in the second and third types, is removed.
Pressure angles between the teeth in the driving range have an important efiect upon the internal resistance of a fluid mechanism and the power required to drive it. The pressure angles in the six and seven tooth rotors of the prior art, having a difference of one tooth, used as'motors, are prevented from rotating by heavy fiuidpressure thrusting the outer rotor away from its axis, and thrusting the teeth at open mesh against each other. In the gear patents the tooth contours (and pressure-angles at full mesh) were the same as in Gerotors. Pressure angles between our 'Rotoid teeth making continuous contacts are not fixed as in gearing of the-present day. The pressure angles vary as one tooth rolls against the other across full mesh. The angles of the six to seven tooth Gerotors were of the order of 30 to 50, while in Rotcids having five and seven teeth the angles are nearer 9 to 17. The radii of curvature may be increased, and the driving curves more steeply inclined. ,Also they may now be located wherever desired, inside, across or outside of the ratio circles, aways obeying the 116- quirements of the circroidal addition. Low pressure angles and fluid'pressure balances ,favor operation as a motor. The constructionprovides bearings favoring starting as a motor, and favoring metering. The low pressure angles also provideleast friction between the teeth,.so that fluid pressures needed for operation are low. :As a meterthe difierence between inflow and outflow pressures is slight, favoringgreater accuracy. As a motor, particularly for hydraulic operatiomeficiencyishigh.
.One of the factors of lowinternal resistance is the character of the tooth engagementin the driving range at full mesh. .In .X '7 ,toothrotors it is :within a few .per cent of va pure ;roll of a curvex curve on a concave curve having slightly larger radii-of curvature.
Continuous .contact rotors are now usually manufactured by breaching the outer rotor and regenerating the pinion. The broach has circular cutting edges to cut thecircular arcs on the outer rotor teeth. The bottoms of the tooth spaces .do not have to ihave aigenera-ted form as long as they clear'the tops of the pinion teeth during rotation. They-are usedmainly for liquid mechanisms. For air or other gas the bottoms of the itooth'spaccsmay be closer to the-tops of the piston teeth and even-have the same form, making allowance in the bottoms of the tooth spaces-for the escape of liquid in the tooth spaces acrossjfull mesh. For partof the distance across full mesh this liquidmay assist or act as a substitute for tooth contact,-and thus partake of the character of afluid pressure holding engagement.
The pinion is generated by means of a tool, be it a miliing cutter or: grinding wheel, .whiclrhas the rshape of ;a tooth f the outerrotor, that is, 1
the circular arc. The larger the curve of this tooth is, the larger the tool and'thc faster itsoperation. The relative size shown in (Figs. I- IIJ was selected as a mean between extreme sizes each of which in that form has a disadvantage, in reducing displacement.
Fluid displacement mechanisms are apt to be taken 'for granted. In gear fluid mechanisms havingno continuous contacts, liquid locking and jamming are not critical factors. But in fluid mechanisms having continuous tooth engage,v
ments, the exact location of ports as to tooth contours, .or critical portions of them, are so important that wrong locations of them result either in liquid looking or loss of efficiency thru leakage. In Gerotor mechanisms having a difference of one tooth, the rotors when manufactured areso tight that pressure is used toput them together. They are rotated slowly at;first, and as the lap themselves to a running fit during rotation, they are enabled to run faster and faster. The power for rotating them, high at first, steadily reduces until when completely lapped to each other, they may be run at high speeds without generating heat, provided .fiuid pressures are not too high.
In thenew rotors, having the crescent and back lash, the pinion can drop thru the outer rotor withouttouching it.
With crescents and back lash there are inhibitions to avoid port losses or liquid locks, involving an end of a path of tooth engagement. The overlap of the ends of ports, where a rotor chamber connects with one port before it leaves the other, is not critical at open mesh inGerotors. In the new rotors with back lash it is atan end of the crescent. It has to beat the properend, otherwise the port losses injure the utility of the pumpor other fluid mechanism. That end is determined by the drive direction, also the direction of the drive. It is the drive thatbrings the teeth together to form a fluid pressure holding engagement. Without it' fluid can flow thru chambers having no fluid pressure holding engagements.
:Our rotors and ports have-difierent uses.
One may be with the pinion driving the outer rotor, clockwise. One abutment at full mesh and one at the right hand end of the crescent space arein action aided by passingtooth contacts to bar escape of pressure fluids back into the suctionport. The abutment at the left hand end of the crescent space area does not bar pressure from the discharge port irom being communicated to the crescent area thru the crevices between the teeth due to back lash, so thatthe crescent area has the same fluid pressure asthe ischarge port.
Another use be with the pinion driving the outer rotor anti-clockwise which reverses the operation of abutments and ports. At full mesh the contacts maintained tight by the driving action are passing the abutment in the opposite direction to that above noted. The left hand crescent abutment with tooth engagements kept tight by the driving action passing over it, bars the escape of fluid pressure from the crescent areaback into the intake port along which the rotor chambers are opening.
When two abutments are used, one at each end of the crescent area, one is out of action due to back lash and the other has the overlap. This is the reversible construction. Back lash and overlap are, modified for successful useas. motors.
The different operations as motor and pump,
and their directions of travel involve tight tooth continuous contacts on the one hand and'back lash with free leakage on the other hand. If the shaft drives the pinion in a pump, the contacts between the teeth to keep the chambers sealed from each other occur where the chambers are opening, and the chambers that are closing are connected thru back lash. to the crescent space, which is sealed off from the opening chambers and intake portby an abutment area, and the tight tooth contacts. a a As a hydraulic motoigthe reverse is true, the closing chambers being sealed from each other and from the crescent space. The latter is then connected tothe opening chambers thru back lash. v
" In thedrawings:
Fig. I is a sectional elevation of the liquid pump or-motor construction on line- I -I, Fig. II, the rotors being indicated in broken lines, full lines being shown in Fig. IV. iFi'g. II is asection of Fig. I on line 11-11.
6 Fig. IIIis an elevation of the left side of Fig. II omitting the boss, seal'members, shaft and thrust plate. l r
- Fig. IV is to indicate tooth-relations between the rotorsJ 4 Fig. V shows rotor members and ports tocooperate'with them for a one-direction pump or motor.
Fig. VI shows the rotors in another position indicating port modifications at full mesh.
Fig. VII shows the parts of an Oldham clutch for driving the shaft avoiding end thrust on the rotor teeth.
Fig. VIII showsrotors of the prior art having a difference of one tooth, with the teeth doubled in number, resulting in 8 and 10 teeth, and their paths of contact.
"*FigrIX shows rotors having the same tooth forms and the same eccentric relations, with the teeth reduced in number to 5 and 7.
Fig. X shows a modification of the teeth of the rotors in Fig. IX.
' Fig. XI shows'rotors having 9 and 11 teeth built from the same type of curves.
Fig. XII shows a diagram of a more scientific method of designing rotor contours.
Fig. XIII shows a diagram of ratio circles and circroids upon which Fig. XII is based.
Fig. XIV shows a detail of the balancing duct system.
Fig. XV shows another detail in a different position.
Fig. XVI shows the ducts upon the side of the casing with the seal cap removed.
Fig. XVII is a vertical section of a modified form of fluid mechanism on the line XVII-XVII in Fig. XVIII.
Fig. XVIII is a right hand elevation of Fig. XVII with the cover removed. Fig. XIX shows a spring to keep the seal inFig.
XVII tight.
I Fig. XX is an enlargementof the seal in Fig.
XVII.
g-Fig. XXI illustrates a form of rotors of the prior art now on the market. I The figures in the drawings illustrate the adap versible liquid pumps and motors, the useful in other fluid mechanisms. Low pressure gas may be pumped without much loss of power in fluid mechanisms adapted to handle liquids.
Rotors are shown in a series of positions to" illustrate the port relations necessary for such rotors having contours providing continuous fluid pressure holding engagements between the teeth. I .7
Two typical units are shown, onerequiring that the teeth of the pinion be integral'with a driving shaft,'and others having the pinion keyed to the driving shaft. The former has the small est size with relation'to volume of displacement. while the latter has greater pressure holding capacity. 1 1
Ports and continuous contacts leakage of fluids between the teeth, are the types of contours used. and the port forms necessary for providing inlets and outlets for fluidwithout looking or jamming on the one hand or copious leakage on the other. Copious leakage of the prior art requires larger pumps and more horsepowerto operate them, and are at a disadvantage in competition;
i In Fig. I the casing member 12 is shown in section, with ports l3 and I 4, in this case, either one of which may be the inlet and the other the outlet. f -In {other words this is a reversible mechanism.='v p Fluid displacement (rotor) members are indicated in broken lines 15 and as, and in various positions in Figs. IV, V, and VI. Inorder to design ports, firstthe path of contact i'i between the teeth has to be determined Fig. IV shows this path of contact 6! and path of tooth contact or engagement Ila on the other sides of the centerline CL (which passes thru the two rotor axes, ac and p0).
Diiferentrotor contours have correspondingly difierent paths of tooth engagements. The con tour of a pinion is determined by generating it, using as a generating tool a cutter or grinding wheelhaving the form and size of a dominating convex master curve of a tooth Q8 of the'outer rotormember '2, this tool being caused to cut or otherwise shape .the teeth of the pinion, while the pinion blank. rotates on its own axis 1/. and
ing the fifth; then skipping the first, forming the second, skipping the third and forming the fourth when the .rotor is completed. This unusual generating process results from unusual tooth ratios and their ratio circles. j. The; five and seven tooth rotors have a ratio which-dif" fers by one, this difference being a necessity for continuous contact rotor teeth contours. The ratiois not between integers but between fractional numbers, in this case, 2 and But discharge of fluid, is indicated at 2i.
since rotors cant run with a half tooth, these numbers are multiplied by the denominator of the fraction, that is by 2, providing the numbers of teeth of and 7. If the fractional ratio was 2 and 3 /2,, the tooth numbers would be 9 and 13, the ratio numbers being multiplied the denominator 4. These fractional ratios have what is called a hunting relation by which each toothof one rotor in some order, makes engagement, a travelling engagement, with every tooth of the other rotor. That makes it possible for a tool representing one tooth of one rotor to out every tooth of the other rotor. When the tooth forms of one rotor are selected or form-ed arbitrarily, the bottoms of narrow tooth spaces may be of minor importance, particularly on the outer rotor, and as long as they keep out of contact with the tops of the teethf the other rotor. The depth or form of the bottom of the is unimportant, so far as continuous engagement is concerned. For use with liquids deepened tooth spaces in. the places described, work well, and in fact may provide easier inlet and outlet of liquid from the chambers between. the rotor teeth nearing full mesh. Reference is made to the reissue patent to Myron Hill, 21,316, for further description of the generating process and. to the patent to Hugo Bilgram and M. Hill, 'No. 1,798,659 describing a suitable gen.- erator. For these contours the mill or grinding wheel travels across the rotor blank, in. a direction parallel to the rotor axes; accomplished by mounting the Bilgram Hill machine upon a shaper, which may provide the cutting tool, or which may carry a grinding head.
Other ratios may be provided, such as 3 and 4 4 and 5 /2; etc, or 3 /4 and 4 /4; or any other fractional ratio differing by one. Only these fractional ratios have the. desired hunting relation. Other ratios based on integers, such as 4L and 6, 6 and 8, and the like, for many purposes may be useful. and novel combinations with ports and novel features of the rotors and mechanisms, aside from the hunting relation, lie within the scope oi our invention.
The path of contact between rotor contours varies with the type of contours used; with the ratio, whether of even or odd fractional numbers; and with the presence or absence of back lash. Fig. XXI shows a type of rotors now on the market, and the path of tooth engagement 24. This path 24 is endless and has a loop 24s at full mesh. Similar rotors with teeth doubled in number are shown in Fig. VIII. One path of contact is shown from 59 to its and another from 26a to 29. If an abutment should be shifted somewhat, angularly (to right or left), it would have little effect.
In Fig. I the abutment at full mesh performs a similar service, being located between the ends 4'! and 48 of the two ports i3 and M- In a one direction pump as in Fig. VI the ends of the ports 35 and Ma may be extended into this abutment area as indicated at 32 and 4 3 for very close fits of rotors between their side walls, or for large pumps or motors, a motor being more or less a reverse of a pump. If back lash is applied to these rotors, one of the paths is eliminated and the ports have to be accommodated to the single path of tooth engagements. Deepening the tooth spaces of the outer rotor slightly, provides easier Removing three teeth'from each' rotor produces a ratio of 5 X '7 shown in Fig. IX, the basic ratio being 2 x 3% which has the difference of one needed for continuous tooth contacts. These long tooth divisions complicate the port problem, particularly inv View of the crescent space at open mesh, The" radial increase in the heights of teeth in Fig. X increases displacement and provides for radially large port areas. In Fig. X the. sides of the pinion teeth. shown in Fig. IX are brought closer together.
The ports for'the different types of rotor operation vary, the ports for one type often not working with another type, sometimesbeing' inoperative and. sometimes being subject to the evil of all liquid pumps, cavitation, with its attendant hammering, noise and vibration. While this application is primarily concerned with liquid pumps and motors, it is also concerned in a generic way with any kind of. fluid mechanism where the invention can be of service.
A drive. shaft may be. either the shaft; that drives a. pump rotor or that. is driven by the rotors acting as a hydraulic motor;
In Fig. I ports. were shown. for a reversible operation and they differ. from. the; ports in Figs. V and VI because the latter two are not designed to be reversible. They are alternative: as to Fig. IV. Any type with. its ports. may be used in Figs. I and II.. The port conditions at full mesh and at the two ends. of. the crescent are. the. factors that have to be dealt with. Another factor that has to be dealt with is the path of continuous tooth contact and whether, with back lash between the teeth, it is a single'path on one side of the center line (with its book at f-ull mesh), or, without back lash, paths on both sides of the center line. In those forms of rotors proposed heretofore, such as the patent to Feuerheerd in which the contours made accurately are inoperative. In. commercial pumps the'teeth have loose relations, and port relations are not exacting, since fairly free leakage between the teeth elimihates the. possibility of jamming of; liquid and liquid look, so that. it is merely a question of. lost displacement and of higher power requirements. Such mechanisms are incapable of efficient service.v Viking and Tuthill pumps have no suchcontinuous tooth contacts and their ports are separated at open mesh by crescent inserts, over which the teeth of both rotors travel out of contact.
Hydraulic motor, pinion drive Operating anticlockwise, and illustrating with Fig. I, high pressure (as distinguished from low pressure) enters thru the port !3. This pressure fluid should not escape into the outlet port 14 except thru opening chambers as they pass the abutment 34a. The ends of the abutment if modified as at 40, let a slight amount of liquid escape but pressure liquid opens the rotor chambers causing rotation, minimizing the escape of liquid. A rotor chamber in the position of 46 in Fig. VI has not passed the critical point 29 (inboth figures) and so contains low pressure liquid. As the chamber leaves the low pressure port M (Ma in Fig. VI) it connects with the high pressure port I3 (35 in Fig. VI). If it connects before reaching full mesh, and is disconnected from the port l4, it would have to try to compress liquid, and in the effort become liquid locked. So that the chamberconnects with the high pressure port I3 after it has stopped closing at full mesh. The port extension 32 in Fig. VI freely connecting with the chamber across full mesh favors this relation, but is not useful in reversible motors.
In Fig. I the ports inside of the paths-of contact i! and Ha may extend closer to the full mesh point than those portions outside of the paths of contact. The extension 44 extends to the chamber at full mesh and its limit 48 corresponds to the limit of the extension 44 in Fig.
VI. If these ends do not quite connect with the and is still closing, it must have another outlet 1 a to avoid liquid lock. Therefore the point 28 is located by the length of the chamber at full mesh, since that is where closing has ceased. The left end of the chamber theoretically has reached the point 29. In practice an overlap is needed t9 preventlmomentary jamming in the chamber, be
iore it disconnects from the port 4 la which must have made connection with the port 35 containing the high pressure liquid that operates the motor.
to open and thereafter is oadsed to continue to open until its left or forward end has passed the abutment area near the crescent space. A corresponding overlap at this abutment remedies cavitation noises.
All that an overlap needs is that the theoretical boundary of a port be cut away slightly. It makes the abutment smaller in effective area than the chamber. high pressure port 35 to the low pressure port 41a. With the narrow spaces between the rotor teeth andthe inertia and momentum of oil flying by at high speed, the loss of volume is negligible. But it softens a tendenc to knock. Any extension of the high pressure port 35 towards the chamber at full mesh lies radially inside of the path of tooth contact li. The end oi the port Ma may extend to the contact closing point 23 (or slightly beyond for the overlap). The ports [4 and iii in Fig. I are for reversible operation,
extending to 47 and 48 respectively, corresponding to theextension M in Fig. VI and the end of the port Ala without the extension 32, the points 28 and 29 in Fig. I being the same points in Fig.
In Fig. V the chamber between the points 33 and 26 is about to connect with the crescent space between 26 and 260. where low pressure prevails, so it has to disconnect (except for the overlap) from the high pressure port 36. long low pressure port in Fig. V, including the crescent space favors high speed of operation. The abutment for these five to seven tooth rotors is shown in this figure, its outer boundary extendin from the end 3! of the port 36, along the lower side of the outer rotor tooth in this rotor position, to the path of contact ll, up along the path to the contact 33, thence along the lower edge of the pinion tooth to the inner boundary 29 of the port 36. The outer boundary may be circular around the outer tips of the tooth spaces.
If the chamber has reached its limit of: closing it has reached a point where it is ready It provides a leak from the 1 The These outlines may be modified as already described. They are adapted to high pressure operation.
Owing to the need of precision here and there, the wide lines of drawings are representative only.
Liquid pumps In our pinion drive liquid pump, the drive shaft is connected to drive the pinion. The low presat 13c and Nil).
sure intake ports in Figs. I, V and, VI are I3, 36, and 35 respectively when operated anticlockwise (see arrows in Figs. V and VI). Chambers along the other high pressure discharge ports are subject to an openingforce opposing the driving force on the shaft, which causes the teeth to make driving contacts and pressure holding engagements between opening chambers. Overlap of ports is useful to minimize cavitation noise. It may be accomplished by difierent modifications of ports, among which are shown the chopped-off portions of the abutment 34 one or both, Fig. V, and 34a and 5| in Fig. I, at the two ends 40. The mass and momentum of the liquid operate to prevent any substantial leakage when the overlap is correctly proportioned. The overlaps at 4% may be varied to accommodate the speed at which the mechanism is to run,a very small overlap doing for low speed and a much larger overlap for high speed. This overlap also may be accomplished by shifting the two sides of the abutment a little closer to each other or similar shifting of other contours of the ports. This overlap is not a necessity in all cases in the hydraulic motor, pinion drive, described above.
Reversibility rated from a neutral port area d9 by abutments at 340. and 5|, similar to the abutment 34 in Fig. V. If the pinion in a pump with back lash drives the outer rotor, shown in broken lines, clockwise, the drive engages the teeth of the opening chambers, so that only the teeth forming the opening chambers are kept tight b the driving relation. The teeth, with their back lash, between the closing chambers are. therefore not tight, and fluid pressure in the outlet port is transmitted between the teeth to the neutral port area 49. -The abutment 5! is the one that prevents pressure from the port l3 from flooding into the intake port 14.
That the neutral areais not required to be separate in all units is indicated in Fig. VI. In
low pressure pumps and low speed pumps, or other fluid mechanisms, the neutral port is of lesser consequence.
The ports have another feature of importance. It may be noted that with rotors having a difference of two teeth, the ports are shorter than with rotors having a difference ofone tooth, but they are wider radially. Rotor chambers have greater displacement volume. More port area may be added on the opposite side of'the rotors as indicated in Fig. II, where port is is shown This doubles the already enlarged port area, allowing free intake of low pressure liquid, drawn in by suction (popularly speaking). These ports communicate with piping passages 52 and 53. Having the same fluid But the port extensions 32 and 44 are no longer permisare equal.
pressures on both sides, rotors do not of themselves tend to press against either front or rear .wall and so create wear.
Bearings In the ordinary pump art it is customary to allow up to seven thousandths of an inch clearance in the journal per inch of shaft diameter. It is evident that if rotor contacts are as loose as that, they could not exert the continuous "liquid pressure holding engagements intended.
and prevent binding.
Radial pressures on the bearings may be opposed by fluid pressures in the bearings having the same pressure as in the rotor chambers counteracting the load pressure upon the bearings. This mechanism is adapted to be cast, die cast or plastic molded, so that the liquid pressure balancing system is designed to permit withdrawal of forming members of molds from casing members containing the bearings. Liquid bal ancing is accomplished preferably by grooves parallel to the rotor axes, around the outer rotor and around the shaft carrying the pinion. These groo es are grouped in zones 59 and 50; opposing pressures, high or low, in the ports 45, i3 and Hi respectively, so that whatever the fluid pressure in any port area may be it is instantly communicated to the zone of grooves that will oppose the pressure in the port androtor chambers connected to it. The grooves are connected together, at their open ends by circular recesses so that the pressure in the grooves of any zone These recesses are capable of being molded, and with a press member capable oi being withdrawn after molding. Circular ducts impressed upon the end face of the casing member, indicated in Fig. III furnish the means of making these various interconnections for the grooves around the shaft bearing in the casing member l2. The circular duct 82 is connected to the zone and its recess, 58, as indicated and by a hole 63 thru the casing 2 to the port 49. The duct E l, connected by the hole 85 to port i3, is connected to the zone 59 by holes 66 drilled as indicated in Fig. XIV, until they meet. The duct 61, connected by the hole 58, to the port M, is connected to the zone 60 by holes similar to 5 8 in Fig. mV, as shown at 59 in Fig. XV. Thus each port, intake, outlet and neutral has its zone of opposing pressures, regardless of what the pressure may be under the conditions of operation heretofore described. The various ducts are closed by the gasket 10 clamped to the surface by means of the screw seal cap H. The zones or recesses at the ends of the bearing grooves are closed by the plate 12 pressed against the casing member 'H by the spring 18. The seal cap may contain the Beach seal comprising a sponge rubber or neoprene sleeve 13 clinging tightly to the driving shaft is journalled within the zones 58, 59 and 0d, and additional zones at its other end to be described later. Upon both ends of this sleeve 13 are mounted caps 75 and '16 free to move as on a universal joint and thus to always lie flat against an end seal plate H,
and the thrust plate '72. The rubber may-exert presture thru the seal caps against these end plates; assisted, for higher pressure or vacuum service, by a helical spring 18. The rubber sleeve, caps and spring rotate with the shaft. The caps may be of a special plastic or Bakelite, and the thrust plates 12 and ll of bronze. The-special plastic is a thermosetting compound of resin, fibre and graphite. The space around the 'rotating seal members is connected to a low pressure port in each pump for best service, as by a duct '19, Figs. II and III.
The right end of the shaft, 73a, is similarly mounted in a bearing in the cover member of the casing, the bearing having similar grooves, recesses and zones as the other end of the shaft. The ZOIlfS and ports are interconnected by ducts in the cover member. They are also illustrated in Fig. XVI. Around the outside edge of the cover is the duct 8| connected to the Zone 5911 by the hole 32 and by the slot 83, Figs. II and XVI, to the port area {3b. On the outer face of the cover is the duct 8% connected to zone 60a by the hole 35, and to port M by a hole 35. Another duct 8'! connected to the zone 580; by the hole" 88, is connected to the neutral port by a hole '89. Thus connected these zones reinforce the zones at the left end of the shaft and together in area should balance as near as may be, the pressures in the rotor chambers and the ports connected to them. In 5 and 7 tooth rotors there'is a variation of pressure of some 40%. But rotating parts have momentum and inertia, and an average balance for usual motor speeds accomplishes excellent results. The balancing pressures counteract the shaft pressures on solid bearing walls, reducing both Wear and friction.
The space at the right end of the shaft, .at .90, Fig. II, is connected to the low pressure area in each type of mechanism as by a slot at 9!.
The open ends of the grooves and recesses 58a, 53a and 50a are sealed by the ring 92, adrive fit in the cover member and freely fitted to the shaft. The cover member 89 is fitted to the casing i2 and against a shoulder 93 to provide correctrunning clearance for the rotors 42 and 96. A gasket 9 covers the ducts 8t and Bi, and both cover and gasket held tightly against the casing shoulder 93 by the hollow nut 85. The threadsof this screw joint are loose enough to let the screw adjust itself to the cover and shoulder. To enable the teeth 9*5 of the pinion rotor, integral with the shaft M. to find their easy running position between the cover 80 and the back wall in the casing 12, the shaft has sufficient freedom endwise. To avoid interference from outside thrusts an Oldham clutch is employed, illustrated in Figs. II, VII, and XVII. Lugs 97 and 98, at right angles to each other, on the shaft ends, engage recesses in a middle disc member 99, Fig. VII, with allowance for radial and endwise freedom to accommodate axes not always in perfect alignment. In case the shaft M is molded plastic, the lugs 98 are on a metal cap piece shown in Fig. II having a rough interior to grip the plastic, during molding at 375 F. and under 5,000 lbs. p. s. 1. pressure.
Ratios Our invention is applicable as to some of its features to many con-tant contact ratios. The 5 x 7 ratio is unique in that it has the fewest permissible teeth of the basic fractional ratios for general use, its displacement for high pressures requiring a driving shaft so large that the teeth pinion tooth at full mesh in Fig. VIII with enough of the middle of the tooth, as between broken lines IZI, removed to add another tooth to each rotor, the outer teeth being correspondingly narrowed. In Fig. XI the complete tooth curve I! is generative of the dominating or master curve lllla which may or may not be circular. A part of this tooth curve may be employed to form the tooth 102, its other side being the same curve in reverse, or some other curve adapted to a corresponding other curve on the teeth of the outer rotor. These tooth forms, and their pressure angles may be improved by redesigning in accordance with the system of designing rotor con"- tours illustrated in Figs. XII and XIII.
Our earliest efiorts in rotors, indicated in Patents 1,682,563-4-5, were based upon the hypo system of generation, where a pinion tooth form, preferably circular, was used as the master form by means of which the rotor contours were designed by generation. The shortcomings of such rotors were expensive. They were remedied by a reversal of the method, so that a master form was the tooth form of the outer rotor, still preferably circular and the outer rotor tooth spaces were narrow. The main advantage of the cicular form is the comparative ease of making tools, since other forms require special and intricate mechanical equipment. The narrow tooth spaces in the outer rotor help to eliminate liquid jamming at full mesh.
New geometry Figs. XII and XIII illustrate the method of designing comparative rotor contours. The ratio of the rotors having five to seven teeth is 'used to design diiferent circroids and corresponding rotor contours, to estimate their relative values. They comprise an outer ratio circle Bor arcs of it having a radius 3 times the eccentricity (see pc to ac, Fig. IV), a pinion ratio circle A having a radius 2% times the eccentricity (pc to ac, Fig. IV), a radicroid R, that is, a radius of the outer ratio circle, extended to a circroid at a chosen distance which may be varied for varying the circroid, and positions along a cycloid which assistin locating the radicroid in successive rolling positions, that is, rolling of an outer ratio circle upon the pinion ratio circle as its outer tip traces a circroid. Fig. XII utilizes the pinion ratio circle and the circroid to explain the critical factor that determines the inner limits of the pinion contour, generated by a selected form fixed to the end of the radicroid R. A circular form is usually employed, but other forms may be used.
. In Fig. IGII the pinion ratio circle A is divided into an equal number of arcs. Ten are convenient for a five tooth pinion, starting from the radicroid position R, some of which are marked 00, 11, 22, 33, 44, 55, 66, and 77. An extra division between 22 and 33 assists accuracy for reasons to follow. It is marked 25. An eccentricity. circle E is described from the center Y of the circle A, with a radius equal to the eccen-. tricity for the proposed rotor curve. This circle is also divided into ten parts, 0, l, 2, 3, 4, 5, 6, and 7, etc. The other points are not used. The starting position of the radicroid is from the point 0 in E thru the point 00in A, and on out to the point 000 selected as an experiment. The
- ratio circle of the outer rotor is not drawn in full to save unnecessary confusion, and is indicated by the arc B in the starting position. As the ratio circle B rolls on A it assumes successive positions indicated in broken lines Bi, B2, B3, B4, B5, B6, and BI. As the ratio circle rolls, and assumes the various positions BI, B2, etc., a point 00 travels along the cycloid thru the various points 10, 20, 25, 30, 40, 50, 60 and 70, sincea point in one circle rolling on another travels along a cycloid. Meanwhile the cente of this ratio circle, point 0 in the circle of eccentricity E, travels around this circle E thru the various points 1, 2, etc. to swing the radicroid. As the radicroid R coincides withxand includes the radius of B, and moves with it, its tip, out beyond the ratio circle, traces a curve, which runs thru the points 000, 100, 200, 250, 300, 400, 500, 600, 700, etc., or as many others in between as may be needed for accuracy. This is the circroid wanted for trying out rotor curves. Another might be C I.
length of the radius of this master generating circular arc Ml. If it is too long, a critical portion of the envelope T will be broken into parts crossing some normals at different angles. There must be a critical point outside of which the perfect rotor tooth contour must lie. If it is to be parallel to the circroid it must have all its normalsi. e. normals to its tangents-also normal to the envelope, and for an equidistance envelope all such normals must be of equal length. If the curve Ml should be changed in curvature to a non-circular curve, the envelope corresponds to its irregularity. But in that direction lie complex tooth forms with circroidal additions to correspond. I
Just as a radius of a circle is normal to it, so circroids for circular curves have radii normal to them. They illustrate the principle involved to better advantage. Such radii are the instant lines from given points of the circroid to corresponding points on the ratio circle A to which the ratio circle B is tangent during its rotation. The end of the radicroid, while tracing the circroid, is swingingor turning on a travelling point the point of tangency between the circle A and B. These instant radii are indicated, or a few of them are indicated in this Fig. XII, one from 200 to 22, and another from 300 to 33. They converge more than any of the other lines from other points. Another line from 250 to 25 also converges toward 300 to 33, even more. That point of intersection between these various instant radii nearest to the circroid, is the critical point that We are after, since any envelope beyond it is broken up by arcs lying at interfering angles. This point of intersection is indicated at N. An envelope between N and the circroid 0 described by any curve Ml will provide a tooth curve having a continuous contact relation with an outer rotor tooth having that curve.
The distance of the tip oi the radicroid to its ratio circle is termed the circroidal addition," and with a given ratio, this circroidal addition determines the distance of N from it. Circular master forms have been described, but any tooth curve designed for rotors or gears having radii of curvature greater than zero where it crosses the ratio circle, must observe the requirement of the circroidal addition in order to maintain continuous fluid pressure holding contacts at steady speed. (The same is true as to such tooth curves cut off at the ratio circle as in certain gears.)
The nearer the tip of the radicroid is to its ratio circle, the lesser the radius M. By such reductions the instant, radii of the circroid are reduced. If this reduction is carried to its limit zeroethe circroid is merged into a cycloid as L in Fig. XIII, and the radius of MI is correspondingly reduced to zero. To put it another way, there is no envelope possible within a cycloid and equi-distant from it. It is the failure of gear designers generally tounderstand this fact that is responsible for noisy gears and limited durability.
One would naturally suppose that in order to generate a tooth of a rotor or gear, a blank would be located on the inner ratio circle axis, and a tool to generate with, located onthe outer ratio circle radius and that generation would produce continuous contact tooth curves When this process failed, as it always must, to produce the smooth acting curve, it was a puzzle. The generating tool, one would argue, certainly could not be carried on a greater radius than that of the ratio circle, because the speed ratio would be changed. That is where efforts to solve this rotary problem undoubtedly stopped. It is the. il-- logical idea that solved the enigma. While a master form is mounted on a radicroid of greater length than the radius of a ratio circle, nevertheless the resulting generated tooth contour may lie across the inner ratio circle as gear teeth should, and thus travel at thespeed ratio. It also might lie outside of the ratio circle with continuous tooth contactsv but that location introduces angular slip and poorer pressure angles, Circular type rotor contours may have more than one circroidal addition if merged into each other. If a master form is a composite of different radii of curvature, the point N is to be determined from the several segments which lie nearest to the nearest circroid. Contours once determined may be modified where not needed for engagements.
In order to locate as accurately as possible, the intersection N in Fig. XII a large chart was used, and instruments of accuracy located the various points involved. The relative location shown is approximate. A number of normals were drawn from points on the circroid between 200 and 300 to corresponding points upon the ratio circle A before the point N was finally located. Varying the ratio or circroidal addition shifts the point N. It must always be theintersection that is nearest to the circroid. For the relations shown in Fig. XII the intersecting normals or instant radii of the circroid are between the 200 and 300 positions. For other relations it is nearer the point 000 or further away from it. Some of the normals diverge and requirenoconsideration. The intersection might be considered to be an apex of a triangle whose base is between two points on the circroid at an infinitely small distance from each other, and sides converging at an infinitely small angle. While the diagram method is a shorter and easier one, differential 20 equations may be used to. find the mathematically correct N. The idea that a generating tool should be centered upon a rolling circle larger than a ratio circle appeared impossible for producingteeth of the correct ratio. In effect our invention altered the speed of this larger-than-a-ratiocircle to compensate for its disproportionate size.
Practically, the tooth curve is drawn as an envelope outlined by arcs Ml, M2 etc. having a, radius of M and centered at successive points all along the circroid. The radiusv M must locate, the envelope between the point N and the circroid. The nearer this envelope isv to the circroid, the more it partakes of its curvatures. The nearer it is to the point N the sharper the curvature around N. If carried to its limit, a corner around N is arrived at, too sharp for use. By having the curve located a slight distance from, N the best results are attained. Also it allows. forminor drafting errors.
Pressure angles are involved in the radius. of the curve MI. The less the radius, the greater the variationin the pressure angles inthe driving range. The best average pressure angles for a given ratio are derived from the largest useful radius of curvature. The inclination of thecurve is also better with a larger radius, due to the curve centers being so much farther. away around the ratio circle, as indicatedin Fig,.,XII where 000 is far to the right of the vertical axis, while, the. toothcurve Mi is far to the left- The pressure angle is equal to that between atangentto the driving curve at any point, and a radius of av ratiov circle to that point. During a driving relation, over a driving range of one tooth division, the tangent point is travelling, hence its angl'evaries. A fixed pressure angle bars continuous tooth contact in the driving range.
As the circroidal addition is reduced as derscribed, the critical normal from thecircroid; to, the ratio circle is ever shifting along the circroid towards 000. With changes in the numbers of teeth, and with variations of the other factors mentioned, the location of the normal also changes in one direction or another, and in designingdifferent rotors, with different relativeradii of curvature; these various changes arestudied to select the forms most suitable for the rotors desired, compromising upon numbers of teeth, size of master curves, circroidaladditions, for strength, low-pressureangles, and displace ment. By making graphs of the effect of the changes of each of the factors, one is enabled to select more intelligently the form best suited to the problem in hand.
In Fig. X is illustrated a comparison between the displacement of 6 and '7 tooth rotors at I50; and l6a'and5' and 7 tooth rotors at I51) and I6b;. the former being in broken lines. The great difference in displacements is evident.
The rotors in Fig. IX, tho having a difierent ratio have the teeth of thee and 5 tooth form (hence of greater displacement than- 6 x 7 rotors) indicated in Fig. VIII, omitting half of the teeth andtooth spaces. After doubling these numbers 4 and- 5, the ratio circles-were reduced by subtracting three from both numbers to the 5 and? ratio, and the teeth disposed accordingly. The design in Fig. X, in full lines resulted from the geometrical studies in Figs. XII'and- XIII. The pinion shaft diameter is the same in both figures. In Fig. X the rotor chambersreach-the shaft almost, while in Fig. IXthe chambers-are shallower. The pinion teeth in Fig, IX have thesame tops as those in Fig.'VIII, losing that'extra depth 21 of rotor chambers gained partly by'bring'ing the sides of a pinion tooth closer together. In our gear Patent 2,091,317 the type of tooth corre- 'sponds to Fig. IX, but of a different ratio, not with odd numbers of teeth but with even numbers of teeth (which lose the hunting relation) between all the teeth. The teeth of the rotors in Fig. XI, 9 and 11 in ratio, have deeper rotor chambers and the full hunting relation. p
In order to design these 9 and 11 tooth rotors the ratio circle A of Fig. XIII has a radius 4 times the eccentricity, and the ratio circle B has a radius of 5 /2 times the eccentricity. Circle A has a number of divisions laid off on it of equal length and the points of the cycloid L are located to accord with them. The rest of the procedure is similar to that for the 5 and '7 tooth rotors, except that the radius M and the circroidal addition have to be experimented with to get the best form of tooth curve, the best pressure angles, sufficient tooth size and greatest displacement for the different ratio. sists of varying the circroidal addition by varying the extension of the radicroid beyond'its ratio circle, then finding the point N for such circroid as may be described by the radicroid, and then with a radius a little short of the point N, outlining a rotor tooth curve.
After describing a curve that appears satisfactory for the tooth ratio, its driving relation, its displacement, its pressure angles, etc., the next step is to select the portion of such a tooth curve desired for a rotor tooth for a five tooth pinion. The curve T may be one side of a tooth. Obviously half a tooth is limited to one fifth of 360 divided by 2, which is 36. So next it is desired to find out what part of the curve T may be utilized for one half of the tooth. The broken lines GY and HY are drawn from the ratio circle center Y at such an angle of 36. By swinging them around the center, back and forth, keeping their angularity to each other fixed they include different portions of the curve T. The partincluded in this figure is from G to 1-1. If swung to the right, the end H is nearer the center Y, and there is little change of diameter at G. This might be desirable as it increases displacement, but for many uses the shaft usable with this contour would be too small. Also the teeth might lack strength as being too thin. The teeth shown in Figs. IV, V and VI show the final compromise between these factors. They would have to be integral with the shaft for purposes of strength, a unique case. Smaller ,ratios usually need outside gears to drive them. The next larger .ratio having the desirable hunting relation is the 7 and 9 ratio. The next, the Q and ll ratio. The first three ratios more than cover the ground of the Gerotors now in use having ratios of 4 and 5, 6 and 7, 8 and 9, and 10 and 11; so that the new rotors cut the range of manufacturing equipment down to 75% of that required for the older form. Furthermore, the 5 and 7 tooth ratio supplants both the popular 4 and 5 tooth, and the 6 and 7 tooth rotors-for many uses.
Another popular size of rotors having eight and nine teeth is 1% inches in diameter, the same size as the rotors shown in Fig. XVIII, at
Such experimentation con-.
Md and 32d. These rotors have 11 and 13 teeth, and have a greater displacement than the 8x 9 rotors, thereby reducing friction and increasing pressure capacities, as well as delivering a steadier' flow of fluid. Rotors of this size are :useful in oil burner pumps, and in supercharger, scavenging and lubricating pumps (double) of superior service for high altitude flying on account of their high suction characteristics. Ports are designed in accordance with the conditions and limitations described with relation to Figs. I, V, and VI, etc.', with port areas inside and outside of the path of tooth engagement. The pipe connections to the ports are threaded as shown at |03.- The pinion is separate from the shaft if desired, in which case it is keyed as at I04. The section in Fig. XVII shows the same balancing of fluid pressures, the ducts, etc. as shownin Fig. II, the connections following the method shown in Figs. III, XIV, XV and XVI; with the exception that the nut Hi5 replaces the seal cap ll in Fig. II. The screw cap Hi6 holds the seal thrust plate it? in position. Between the thrust plates 161 and H38 in'the casing, is located the seal. It'may be combined with the clutch disc I09 shown in larger size in Fig. XX. As shown in Fig. XVII there is endwise freedom between the shafts I H and H2 so that thrust moments on the shaft H2, as when it is a motor shaft, will not put a corresponding thrust on the pinion lill against the casing wall.
It hasbeen the custom, as shown in Fig. II, to put a sponge rubber sleeve 53 on the shaft M, which is supposed to rotate with the shaft. If neoprene is substituted in order to pump oil (rubber being softened to impotence in oil) the neoprene slips on the shaft and does not rotate withit. Soon leakage develops. Therefore we prefer to make a union between the sleeve and the rotating memberthe Oldham clutch disc such as by vulcanization, there being a fixed ring or band I 13 between the disc and the sponge rubber or neoprene ring I I4, vulcanized or cemented to both. The object of this sponge rubber, neoprene, or other resilient substance is to prevent idiosyncracies of the disc from being communicated to the seal caps I I5 and I I6 mounted and vulcanized to the rubber, etc. These maybe of stainless steel or bronze, spaced apart and pressed against the thrust plates 59'! andlOB by a spring Ill, which for convenience may be a spring washer of hardened stainless spring steel, crimped into a wavy form indicated in Fig. XIX, hich shows a section of it. The strength of the 3 spring is adjusted to resist separation of the bronze shoes from the thrust plates. Without this rubber, a disc corresponding to E09, once tried in the 8 x 9 toothed rotor mechanisms,
leaked. Therecesses in the disc indicated by broken lines, fit the lugs I I3 and H9 allowing radial freedom, and have the same shape as in Fig. VII exceptthat the diameter is larger as vindicated in Fig. XIX, and adapted to fit the ring H3.
The form of casing members in Figs. XVII and II permits them to be molded from the special plastic composed of graphite with strength fibres thru it with a thermoset resin. This powder does not easily flow in transfer molds and is made of biscuits, preformed cold from powder with mild pressure; which in the molds under some 5,000 lbs. p. s'. i. and at a temperature of around 375 F. becomes a strong structure with very slight coefiicients of expansion; and impervious, to water, oil, and most acids. It may be cooled in the molds before removing pressure. Theaccuracy possible makes machining unnecessary. The casing members are each molded in a mold to form their exteriors, with rains to form the interiors including port areas. The casing members 1 instead of being boltedl together are, united by hollow nuts which bin'dthe covers (Fig. II) "and various needs.
1 {9 (Fig. XVII) against-gasketsfl and H1 (Fig. II) and IZO'andI-2i (Fig. XVII) which are thus bound tight against shoulders in the casingmembers respectively. The gasketsmay beof rubber and fabric with slight compressibility to'equalize of plastic, it is molded into a driving cap having the lugs 98. The inside ofthe cap is knurled or roughened for a secure grip onzthe plastic material. If its coeflicient of expansion is greater than that of the plastic, as it cools it grips the plastic ever tighter.
There .is no limit'theoreti'cally to the highest numbers of teeth possible with-.our ratios having a difference of two or. more teeth. The least number of teeth is determined by the manner of driving relation that keeps them running .;at the steady ratio speeds necessary for continuous fluidtight engagements. Aratio of.5.x 7 teeth provides an excellent driving relation between the teeth, far better than with teeth having commercial forms now generally used outside of the Gerotor art. A ratio of3 x 5 is possible, particular-ly for an oil pump due to the long. driving con- "tactacross full mesh of a rolling character. Howeverit has to be assisted in part of the driving range by a very considerable radial rubbing action between the teeth. With the plastic or other durable materialfor the tooth surfaces even this rubbing action is not prohibitive. -With rotors having 1 x 3 teeth however outside gearing may be used to keep 'them in notvery :accurate-regis 'tration, but enough so to *do some pumping or blowing of air or gas, as well as liquid. Other low ratios doubled, trebled or multipled may fit All these modifications lie within the scope of our invention.
The importance of continuous engagement at steady'ratio speed is-perhaps realized in connection'with such low ratio rotors-where continuous contact means one contact, not two-or three, and its continuity means nothing if tooth curves are irregular, since the teeth might have unsteady speeds and still maintain some continuouscontact. 'Only correctcontours based on the circroidal addition make possible steady ratiospeeds.
Other master generating. contours 'suchas oycloids, ellipses, oval curves or a series of one or more of them and mated contours of them, are useless unless subject to our correct circroidal addition, hence, made operable by "the light thrown on the problem by our invention.
In operation as a pump, a shaft 14 '(Fig; '11) driven by an outside source of power'applied tln'u an Oldham clutch 9T, 93, v99 turnsithe pinion or displacement member 96 which in turn drives the teeth of the outer rotor displacement "member 42 where the rotor chambers are opening. The
' driving action of the shaft and the resistance of the outer rotor due to high pressurein-the closing chambers (which vtends to ;open;them by-reverse rotation) exerts an opposing force ;that keeps the contacts and pressure holdingengagements between the driving teethtight. "The closing chambers are connected'to' eachpther and to the crescent. space through backlash between the teetn but p essu e is un l to pass the tee h c n ac ndabutment-a e endjo t e at s space nearest the intake port.
When .operating .as a hydraulic motor, ,in the reverse direction, pressure .fluids enter thru the same high -pressure port, expand the chambers and cause rotationin ,a reversedirection to that of -,a;pump. 'Pressureliquids impel the rotors in the reverse direction, but the pinion driving shaft, carrying a load resists so that the outer rotor runs ahead of the pinion to the extent of the backlash, andis then stopped byengaging closing teeth of the: pinion, thusassisting in maintaining rotation. These pressure holding toothengagements prevent liquid from forcing its way thru .the device exceptaspermitted by rotation.
.The foregoing description explains the various features of our invention and difierentiates it .fromthe-prior art. The novel features and functionslie within the suspect our invention.
What we. claim is:
1. In va rotary fluid mechanism, a casing, toothed displacement rotor members in said casingyone rotor member within and eccentric to the other, one ,rotor member having inwardly projecting teeth and the other rotor member having outwardly projecting teeth meshing therewith, said teeth having contours providing a crescent space between the teeth where no tooth engagement occurs, saidtooth contours providing for continuous, drive contacts and fluid pressure holding engagements, along a path of contact between full n1esh and saidcrescent space where needed for the performance of fluid pres sure functions,..while traveling at steady ratio speeds, drive. means for said rotary displacement members providing-for saidvdrive contacts and .said pressure holding engagements along a path ofcontact, said tooth. contours providing rotor chambers between said teeth which open and ..close during. rotation; high .,and low pressure ports in :saidcasing located along said path, a
port communicating.withopening chambers and a .port communicatingwith closing chambers, abutment areas located. ona said .path between said ports to check leakage from'a high pressure port to a low pressure port, said-abutment areas located to provide an interval between them for fluid displacement insaid rotor. chambers, said tooth contacts and engagements cooperatingwith said abutment areas in sealing said ports from each other, said tooth contours being located around ratiocircles or curves and including centers of curvature traveling far enough outside of said ratio circles or curves to provide said-con- "tinuouscontacts and engagements, the numbers of-teeth of said two displacement rotor members difiering by two or-more and having a basic fractional ratio differing by one.
2.-The combination claimed in claim -1 having the contours on one side of the teeth shifted angularly to providebacklash in reversible rotors. 1 3. The combination claimed in claim 1 having mainly convex curves on the crown portions of the teeth of the outer rotor rolling upon the mainly concave tooth curves of the inner rotor. L'The combinationclaimed in claim 1 having circular arcs for convex curves on the teeth of one rotor and the other tooth contours of both rotors determined by mutual generation.
- 5. The combinationclaimed in claim 1 having one or-more of saidports located radially inside and-outside ofsaid paths of contact.
" 6. The combination claimed; in claim 1 having I 25 a ratio of seven, to five teeth on said rotor members, the five teeth on the inner member being integrally united to a supporting and drive shaft, said drive shaft having two sections one within said mechanism and one without with an endwise loose connection between them. 7
7. The combination claimed in claim 1 having a two part drive shaft, one inside and one outside of said mechanism, loosely united end to end through a seal member acting as a seal around said shaft.
8. The combination claimed in claim 1, including a drive shaft, and a seven tooth inner rotor member removably mounted on said shaft.
9. The combination claimed in claim 1 including a drive shaft, and a nine tooth inner rotor member removably mounted on said shaft.
10. In a rotary fluid mechanism, a casing, toothed displacement rotor members in said casing, one rotor member within and eccentric to the other, one rotor member having inwardly projecting teeth and the other rotor member havin outwardly projecting teeth meshing therewith, said teeth having contours providing a crescent space between the teeth where no tooth engagement occurs, and providin back lash between them, said tooth contours providing for continuous drive contacts and fluid pressure holding engagements where needed for the performance of fluid pressure functions, along paths between full mesh and said crescent space while traveling at steady ratio speeds, drive means for said rotary displacement members providing for said drive contacts and said pressure holding engagements, said tooth contours providing rotor chambers between said teeth opening and closing during rotation, high and low pressure ports in said casing located along said paths, a port communicating with opening chambers and a port communicating with closing chambers, abutment areas between said ports located on a said path to check leakage from a high pressure port to a low pressure port, said abutment areas providing an interval between them for fluid displacement in said rotor chambers, said tooth contacts and engagements cooperating with said abutment areas in sealing said ports from each other, said tooth contours being located around ratio circles or curves and including centers of curvature traveling far enough outside of said ratio circles or curves to provide said continuous contacts and engagements, the numbers of teeth of said two displacement rotor members diifering by two or more and having a basic fractional ratio differing :by one.
11. The combination in claim 10 having said crescent space communicating with a high or low pressure port, thru back lash. I
12. In a rotary fluid reversible mechanismja casing, toothed displacement rotor members in said casing, one rotor member within and eccentric to the other, one rotor member having inwardly projecting teeth and the other rotor member having outwardly projecting teeth meshing therewith said teeth having contours providing a crescent space between the teeth where no tooth engagement occurs, said tooth contours providin for continuous drive contacts and fluid pressure holding engagements along paths between full mesh and each end of said crescent space While travelling at steady ratiospeeds', and
. where needed for the performance-,of fluid pressure functions {drive mean for said rotary dis- Y placement members providingfor said drive cons tacts and said pressure holding engagement along either of said paths, said contours providing rotor chambers between said teeth opening and closing during rotation; high and low pressure ports in said casing located along said paths, a port communicating with opening chambers and a port communicating with closing chambers, abutment areas in said casing along said paths between said ports to check leakage from one port to the other, located between full mesh and each end of a said crescent space, said tooth contacts and engagements cooperating with said abutment areas in sealin said ports from each other, said abutment areas located to provide an interval between them along either of said paths for fluid displacement in said rotor chambers, while rotating in one direction or the other, said tooth contours being located around ratio circles or curves and including centers of curvature travelling far enough outside of said ratio circles or curves to provide said continuous contacts and engagements, the numbers of teeth of said two displacement rotor member differing by two or more and having a basic fractional ratio diiiering by one.
13. In a rotary fluid mechanism, a casing, toothed displacement rotors in said casing, one rotor member within and eccentric to the other, one rotor member having inwardly projecting teeth and the other rotor member having outwardly projecting teeth meshing therewith, said teeth having contours providing a crescent space between the teeth where no tooth engagement occurs, said tooth contours providing for continuous drive contacts and fluid pressure holding engagements along paths between full mesh and said crescent space while travelling at steady Y ratio speeds, and where needed for the performance of the fluid pressure functions, drive means for said rotary displacement members maintaining said drive contacts and pressure holding engagements, said contours providin rotor chamalong a path of contact between said ports, said abutment areas adjusted in area with relation to said rotor chambers to provide over-lap of said chambers between said ports, said tooth contacts and engagements cooperating with said abutment areas in otherwise sealing said ports fromeach other, said tooth contours being 10- cated around ratio circles or curves and includ- '-'ing* centers of curvature travelling far enough outside of said ratio circles or curves to provide ments occur, said contours providing for continuous drive contacts and fluid pressure holding engagements along paths'between the full'mesh region and-said crescent space where needed for the performance of fluid pressure functions and While travelling at steady ratio speeds, said contours providing backv lash between said teeth, drive means connected to said inner rotor members maintaining said drive contacts and pressure holding engagements along a path of contact, said tooth contours providing rotor chambers between said teeth which open and close during rotation; high and low pressure ports in said casing located along said paths, a port communicating with opening chambers and a port communicating with closing chambers, abutment areas in said casing along a said path between said ports to check leakage from a high pressure port to a low pressure port, said abutment areas located to provide an interval between them for fluid displacement in said rotor chambers, said tooth contacts and engagements cooperating with said abutments in sealing said ports from each other, said tooth contours being located around ratio circles or curves and including centers of curvature travelling far enough outside of said ratio circles or curves to provide said continuous contacts and engagements, the numbers of teeth of said rotor members differing by two or more, and having a basic fractional ratio differing by one.
15. In a rotary fluid mechanism, a casing, toothed displacement rotor members in said casing, one within, and eccentric to the other, one rotor member having inwardly projecting teeth and the other rotor member having outwardly projecting teeth meshing therewith, said teeth having contours providing a crescent space at open mesh where no tooth contacts or engagements occur, said contours providing for continuous drive contacts and fluid pressure holding engagements along paths of contact between full mesh and said crescent space where needed for the performance of fluid pressure functions while travelling at steady ratio speeds, a shaft drive means for said rotor members maintaining said drive contacts and pressure holding engage- :ments along paths of contact between f-ull mesh and said crescent space, said tooth contours providing rotor chambers between said teeth which open and close during rotation, high and low pressure ports in said casing, a port communicating with opening chambers and a port communicating with closing chambers, abutment areas in said casing between said ports along a said path of contact, said abutment areas providing an interval between them for fluid displacement in said chambers, said tooth contacts and engagements cooperating with said abutments in sealing said ports from each other, said tooth contours being located around ratio circles or curves and including centers of curvature traveling .far enough outside of said ratio circles or curves to provide said continuous contacts and engagements, the numbers of teeth of said rotor members differing by two 'or more and having a basic fractional ratio differing by one, zones of fluid pressure grooves in said casing-surrounding said shaft balancing radial fluid pressures in said rotor chambers on said shaft, a zone for high pressure connected to said high pressure port, and a zone for low pressure connected to said low pressure port.
16. In a rotary fluid mechanism, a casing, toothed displacement rotor members in said casing, one within, and eccentric to the other, one rotor member having inwardly projecting teeth and :the other rotor member having outwardly projecting teeth meshing therewith, said teeth having contours providing a crescent space at open mesh where no tooth contacts or engage:
ments occur, said contours providing for continuous drive contacts and fluid pressure holding engagements along paths of contact between full mesh and said crescent space where needed for the performance of fluid pressure functions, while traveling at steady ratio speeds, said contours providing back lash between said teeth, drive means for said rotor members maintaining said drive contacts and pressure holding engagements, said tooth contours providing rotor chambers between said teeth which open and close during rotation; high and low pressure ports in said casing located along said paths, a port communieating with opening chambers, and a port communicating with closing chambers, abutment areas in said casing along said paths between said ports, said abutment areas providing an interval between them for fluid displacement in said rotor chambers, said tooth-contacts and engagements cooperating with said abutment areas in sealing said ports from each other, said tooth contours being located around ratio circles or curves and including centers of curvature traveling far enough outside of said ratio circles or curves to provide said continuous contacts and engagements, the numbers of teeth of said rotor members differing by two or more and having a basic fractional ratio differing by one, a shaft for said driving means, fluid pressure zones along said shaft opposite to said ports balancing port fluid pressures, a zone connected to said high pressure port, a zone connected to said low pressure port; and a third zone connected to said crescent space providing automatic fluid balancing pressures in said zone for different fluid pressures in said crescent space.
17. In a rotary fluid mechanism, a casing, toothed displacement rotor members in said casing, one rotor member within and eccentric to the other, one rotor member having inwardly projecting teeth and the other rotor member having outwardly projecting teeth meshing therewith, said teeth having contours providing a crescent space between the teeth where no tooth engagement occurs, said contours providing for continuous drive contacts and fluid pressure holding en-' gagements along paths of contact between full mesh and said crescent space while travelling at steady ratio speeds, drive means for said rotary displacement members maintaining said drive contacts and said pressure holding engagements,
said tooth contours providing rotor chambers between said teeth which-open and close during rotation; high and low pressure ports in said casing, a port communicating with opening chambers and a port communicating with closing chambers, abutment areas in saidcasing located along a said path of contact between said ports, said areas providing an interval between them for fluid displacement in said chambers, said tooth contacts and engagements cooperating with said abutment areas in sealing said ports from each other, said tooth contours being located around ratio circles or curves and including centers of curvature travelling far enough outside of said ratio circles or curves to provide said continuous contacts and engagements, the numbers of teeth of said two displacement rotor members differing by two or more and having a basic fractional ratio differing by one, one side of the teeth of a said displacement member being relieved or cut away to providezback lash.
18. In a reversible rotary fluid mechanism, a
casing, toothed displacement rotor members in said casing, one rotor member within and eccen- 29 tric to the other, one rotor member having inwardly projecting teeth and the other rotor memher having outwardly projecting teeth meshing therewith, said teeth having contours providing a. crescent space between the teeth where no tooth engagement occurs, said contours providing for continuous drive contacts and fiuid pressure holding engagements along paths of contact between full mesh and the two ends of, said crescent space where needed for the performance of fluid pressure functions, while travellingat steady ratio speeds, drive means for said rotary displacement members providing said drive contacts and said pressure holding engagements along said paths of contact; said contours providing rotor chambers between said teeth which-open and close during rotation; high and low pressure ports in said casing, a port communicating with opening chambers and a port communicating with closing chambers, abutment areas in said casing located along said paths of contact at full mesh and at each end of said crescent space, and providing intervals between them for fluid displacement in said rotor chambers, said tooth contacts and engagements cooperating with said abutment areas in sealing said ports from each other, said tooth contours being located around'ratio circles or curves and including centers'of curvature travelling far enough outside of said ratio circles or REFERENCES CITED The following references are of record in the file of this patent:
UNITED STATES PATENTS Number Name Date Re. 21,316 Hill Aug. 24, 1928 1,538,328 Holdener May 19, 1925 1,646,615 Furness Oct. 25, 1927 1,682,563 Hill Aug. 28, 1928 1,682,564 Hill Aug. 28, 1928 1,682,565 Hill Aug. 28, 1928 1,798,059 Bilgram et al Mar. 24, 1931 1,927,799 Mann Sept. 19, 1933 1,972,565 Kempton Sept. 4, 1934 2,031,888 Hill Feb. 25, 1936 2,091,317 Hill Aug. 31, 1937 2,209,201 Hill July 23, 1940 2,209,202 Hill July 23, 1940 2,336,479 Graef Dec. 14, 1943 2,386,896 Hill et al Oct. 16, 1945 2,389,728 Hill Nov. 27, 1945
US659098A 1946-04-02 1946-04-02 Continuous contact internal rotor for engines Expired - Lifetime US2547392A (en)

Priority Applications (1)

Application Number Priority Date Filing Date Title
US659098A US2547392A (en) 1946-04-02 1946-04-02 Continuous contact internal rotor for engines

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
US659098A US2547392A (en) 1946-04-02 1946-04-02 Continuous contact internal rotor for engines

Publications (1)

Publication Number Publication Date
US2547392A true US2547392A (en) 1951-04-03

Family

ID=24644009

Family Applications (1)

Application Number Title Priority Date Filing Date
US659098A Expired - Lifetime US2547392A (en) 1946-04-02 1946-04-02 Continuous contact internal rotor for engines

Country Status (1)

Country Link
US (1) US2547392A (en)

Cited By (22)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US2601397A (en) * 1950-04-11 1952-06-24 Hill Myron Francis Rotary fluid displacement device
US2693313A (en) * 1952-05-09 1954-11-02 Wetmore Hodges Motor pump or compressor package
US2728300A (en) * 1951-08-27 1955-12-27 Aero Supply Mfg Co Inc Gear pump
US2739538A (en) * 1951-12-14 1956-03-27 Eaton Mfg Co Pumping unit with multiple intake ports
US2760348A (en) * 1952-08-05 1956-08-28 Wetmore Hodges Motor-compressor in plural temperature refrigerating system
US2853023A (en) * 1955-08-12 1958-09-23 American Brake Shoe Co Fluid energy translating apparatuses
US2872872A (en) * 1954-11-23 1959-02-10 Gerotor May Corp Of Maryland Hydraulic pump or motor
US2965039A (en) * 1957-03-31 1960-12-20 Morita Yoshinori Gear pump
US2990724A (en) * 1956-04-06 1961-07-04 Borg Warner Internal-external gears
US3015282A (en) * 1959-02-16 1962-01-02 Viking Pump Company Pump
US3026809A (en) * 1956-04-06 1962-03-27 Borg Warner Internal-external gear pump
US3157350A (en) * 1963-06-11 1964-11-17 Ingersoll Rand Co Rotary fluid machine
US3250459A (en) * 1964-06-15 1966-05-10 Ingersoll Rand Co Gear-rotor motor-compressor
US3275225A (en) * 1964-04-06 1966-09-27 Midland Ross Corp Fluid compressor
DE1284156B (en) * 1963-08-13 1968-11-28 Borg Warner Gear pump for supplying internal combustion engines with fuel
US3424095A (en) * 1965-03-04 1969-01-28 Danfoss As Gear pumps and gear power units
US3513727A (en) * 1966-09-19 1970-05-26 Toyo Kogyo Co Transmission controls
DE2024339A1 (en) * 1969-10-27 1971-05-13
US3619093A (en) * 1968-11-18 1971-11-09 Hohenzollern Huettenverwalt Gear-type hydraulic machine
WO2002063151A1 (en) * 2001-02-08 2002-08-15 James Brent Klassen Two-dimensional positive rotary displacement engine
US6676394B2 (en) * 2000-07-21 2004-01-13 Robert Bosch Gmbh Internal-gear pump having a pinion with radial play
US20080017437A1 (en) * 2006-07-19 2008-01-24 Hitachi, Ltd. Internal gear pump and power steering device

Citations (16)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1538328A (en) * 1924-03-13 1925-05-19 Holdener Simon Tooth formation of internal gears
US1646615A (en) * 1924-10-02 1927-10-25 Cellocilk Company Pump
US1682564A (en) * 1923-02-15 1928-08-28 Myron F Hill Compressor
US1682563A (en) * 1921-11-05 1928-08-28 Myron F Hill Internal rotor
US1682565A (en) * 1921-11-05 1928-08-28 Myron F Hill Rotary compressor
US1798059A (en) * 1922-07-08 1931-03-24 Hill Engineering Co Inc Machine for making rotors
US1927799A (en) * 1932-03-07 1933-09-19 Goulds Pumps Rotary pump
US1972565A (en) * 1928-11-14 1934-09-04 Tuthill Pump Co Rotary engine
US2031888A (en) * 1928-08-24 1936-02-25 Hill Engineering Company Inc Tooth curve for rotors and gears
US2091317A (en) * 1934-10-13 1937-08-31 Myron F Hill Gear tooth curve
USRE21316E (en) * 1940-01-09 Tooth curve fob rotors and gears
US2209202A (en) * 1937-12-31 1940-07-23 William D Horne Method of consolidating sugar refining with the manufacture of milk chocolate, condensed milk, and other products
US2209201A (en) * 1937-08-28 1940-07-23 Myron F Hill Change speed gear
US2336479A (en) * 1939-05-08 1943-12-14 Tokheim Oil Tank & Pump Co Pump construction
US2386896A (en) * 1938-09-01 1945-10-16 Myron F Hill Balanced compressor
US2389728A (en) * 1943-10-14 1945-11-27 Myron F Hill Elliptical contour for rotor teeth

Patent Citations (16)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
USRE21316E (en) * 1940-01-09 Tooth curve fob rotors and gears
US1682563A (en) * 1921-11-05 1928-08-28 Myron F Hill Internal rotor
US1682565A (en) * 1921-11-05 1928-08-28 Myron F Hill Rotary compressor
US1798059A (en) * 1922-07-08 1931-03-24 Hill Engineering Co Inc Machine for making rotors
US1682564A (en) * 1923-02-15 1928-08-28 Myron F Hill Compressor
US1538328A (en) * 1924-03-13 1925-05-19 Holdener Simon Tooth formation of internal gears
US1646615A (en) * 1924-10-02 1927-10-25 Cellocilk Company Pump
US2031888A (en) * 1928-08-24 1936-02-25 Hill Engineering Company Inc Tooth curve for rotors and gears
US1972565A (en) * 1928-11-14 1934-09-04 Tuthill Pump Co Rotary engine
US1927799A (en) * 1932-03-07 1933-09-19 Goulds Pumps Rotary pump
US2091317A (en) * 1934-10-13 1937-08-31 Myron F Hill Gear tooth curve
US2209201A (en) * 1937-08-28 1940-07-23 Myron F Hill Change speed gear
US2209202A (en) * 1937-12-31 1940-07-23 William D Horne Method of consolidating sugar refining with the manufacture of milk chocolate, condensed milk, and other products
US2386896A (en) * 1938-09-01 1945-10-16 Myron F Hill Balanced compressor
US2336479A (en) * 1939-05-08 1943-12-14 Tokheim Oil Tank & Pump Co Pump construction
US2389728A (en) * 1943-10-14 1945-11-27 Myron F Hill Elliptical contour for rotor teeth

Cited By (25)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US2601397A (en) * 1950-04-11 1952-06-24 Hill Myron Francis Rotary fluid displacement device
US2728300A (en) * 1951-08-27 1955-12-27 Aero Supply Mfg Co Inc Gear pump
US2739538A (en) * 1951-12-14 1956-03-27 Eaton Mfg Co Pumping unit with multiple intake ports
US2693313A (en) * 1952-05-09 1954-11-02 Wetmore Hodges Motor pump or compressor package
US2760348A (en) * 1952-08-05 1956-08-28 Wetmore Hodges Motor-compressor in plural temperature refrigerating system
US2872872A (en) * 1954-11-23 1959-02-10 Gerotor May Corp Of Maryland Hydraulic pump or motor
US2853023A (en) * 1955-08-12 1958-09-23 American Brake Shoe Co Fluid energy translating apparatuses
US2990724A (en) * 1956-04-06 1961-07-04 Borg Warner Internal-external gears
US3026809A (en) * 1956-04-06 1962-03-27 Borg Warner Internal-external gear pump
US2965039A (en) * 1957-03-31 1960-12-20 Morita Yoshinori Gear pump
US3015282A (en) * 1959-02-16 1962-01-02 Viking Pump Company Pump
US3157350A (en) * 1963-06-11 1964-11-17 Ingersoll Rand Co Rotary fluid machine
DE1284156B (en) * 1963-08-13 1968-11-28 Borg Warner Gear pump for supplying internal combustion engines with fuel
US3275225A (en) * 1964-04-06 1966-09-27 Midland Ross Corp Fluid compressor
US3250459A (en) * 1964-06-15 1966-05-10 Ingersoll Rand Co Gear-rotor motor-compressor
US3424095A (en) * 1965-03-04 1969-01-28 Danfoss As Gear pumps and gear power units
US3513727A (en) * 1966-09-19 1970-05-26 Toyo Kogyo Co Transmission controls
US3619093A (en) * 1968-11-18 1971-11-09 Hohenzollern Huettenverwalt Gear-type hydraulic machine
DE2024339A1 (en) * 1969-10-27 1971-05-13
US6676394B2 (en) * 2000-07-21 2004-01-13 Robert Bosch Gmbh Internal-gear pump having a pinion with radial play
WO2002063151A1 (en) * 2001-02-08 2002-08-15 James Brent Klassen Two-dimensional positive rotary displacement engine
US20030209221A1 (en) * 2001-02-08 2003-11-13 Klassen James B. Rotary positive displacement device
US7111606B2 (en) 2001-02-08 2006-09-26 Klassen James B Rotary positive displacement device
US20080017437A1 (en) * 2006-07-19 2008-01-24 Hitachi, Ltd. Internal gear pump and power steering device
US7857092B2 (en) * 2006-07-19 2010-12-28 Hitachi, Ltd. Internal gear pump and power steering device

Similar Documents

Publication Publication Date Title
US2312891A (en) Hydrodynamic machine
CA2419068C (en) Ring gear machine clearance
US2622787A (en) Helical rotary engine
Mabie et al. Mechanisms and dynamics of machinery
Colbourne Gear shape and theoretical flow rate in internal gear pumps
JP2818723B2 (en) Gear type machine
US2354992A (en) Gear pump
US4558998A (en) Variable capacity type vane pump with balancing groove in the cam ring
US3267763A (en) Variable-ratio toothed gearing mechanism
US2640428A (en) Drive for fluid handling devices of the rotary, positive displacement type
US2988008A (en) Rotary piston machines
US3226013A (en) Rotary machine
US3787154A (en) Rotor profiles for helical screw rotor machines
US3371552A (en) Rolling contact gear
US3869231A (en) Vane type fluid energy translating device
WO1995000761A1 (en) Rotary positive displacement device
US2165963A (en) Constant flow nonpulsating pump
US4224015A (en) Positive displacement flow meter with helical-toothed rotors
US2572334A (en) Gearing utilizing detachable gear teeth
EP0332450B1 (en) Double-toothed gear, system for generating its tooth profiles and differential speed reduction apparatus using it
US3247736A (en) Involute gear combinations
GB1197432A (en) Improvements in and relating to Rotary Positive Displacement Machines of the Intermeshing Screw Type and Rotors therefor
GB1255799A (en) Rotary positive fluid displacement apparatus
GB2189569A (en) Mounting of a planetary gear assembly in a casing
JP5465366B1 (en) Hydraulic device