EP0009915A1 - Verdrängungsmaschinen mit rotierenden Kolben - Google Patents

Verdrängungsmaschinen mit rotierenden Kolben Download PDF

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Publication number
EP0009915A1
EP0009915A1 EP79301948A EP79301948A EP0009915A1 EP 0009915 A1 EP0009915 A1 EP 0009915A1 EP 79301948 A EP79301948 A EP 79301948A EP 79301948 A EP79301948 A EP 79301948A EP 0009915 A1 EP0009915 A1 EP 0009915A1
Authority
EP
European Patent Office
Prior art keywords
rotor
hub
rotors
lobe
angle
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.)
Withdrawn
Application number
EP79301948A
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English (en)
French (fr)
Inventor
Arthur E. Brown
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Individual
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Individual
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Filing date
Publication date
Application filed by Individual filed Critical Individual
Publication of EP0009915A1 publication Critical patent/EP0009915A1/de
Withdrawn legal-status Critical Current

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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01C—ROTARY-PISTON OR OSCILLATING-PISTON MACHINES OR ENGINES
    • F01C1/00—Rotary-piston machines or engines
    • F01C1/08—Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing
    • F01C1/12—Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing of other than internal-axis type
    • F01C1/123—Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing of other than internal-axis type with tooth-like elements, extending generally radially from the rotor body cooperating with recesses in the other rotor, e.g. one tooth

Definitions

  • This invention relates to rotary positive displacement machines.
  • Figs. 1 to III herein are prior art drawings, the purpose of which are t) illustrate certain problems with the prior art.
  • Figs. 1 to III are similar to Fig. XIX of U.S. Patent 3,472,445.
  • the diameter B of the first rotor hub is larger than diameter C of the second rotor hub and thus the radial dimension A (between hub and casing) is relatively small. Therefore, the flow of air or gas is restricted at dimension A and there is a throttling loss (energy loss) there.
  • the dwell period can be seen in Figs. IV and V of Patent 3,472,445 and lasts about 90 degrees. During the dwell period, no gas is drawn into the inlet port and the flow through same is completely stopped once per rotation. Thus, the flow of gas into the inlet port has a start - stop - start - stop action which would have a detrimental effect on efficiency and noise.
  • Fig. XIX of Patent 3,472,445 is non workable as shown because the right rotor is not balanced and sever vibration would take place. It would be difficult to counterbalance same in a practical manner because the hub is so small in diameter; and thus there is little room for counter-weighting inside the small hub.
  • a rotary positive displacement machine adapted to handle a working fluid, comprising a casing structure having two intersecting bores, a first rotor mounted for rotation in one said bore, a second rotor mounted for rotation in the other said bore, each rotor having a hub and at least two lobes, each lobe attached to a respective hub and projecting radially outward to the outer radius of the rotor, each hub having at least two grooves therein, each groove being located adjacent a respective lobe, timing gear means constraining said two rotors to rotate in opposite directions of rotation, said two rotors interengaging as they rotate, said two hubs being profiled so as to rotate in sealing relation to each other during a portion of each rotation, said casing structure having a lower pressure port for the passage of the working fluid at lower pressure, said casing structure also having a higher pressure port for the passage of the working fluid at higher pressure, said higher pressure port being located in an end wall of the bore containing said first
  • a first rotor 1 and a second rotor 2 are rotatably mounted in the intersecting bores 3 and 4 in the casing structure 5.
  • the first rotor has a large diameter hub 6 and two small angle teeth or lobes ? projecting radially outward from the hub.
  • the second rotor has a smaller diameter hub 8 and two larger angle teeth or lobes 9 projecting radially outward from the hub.
  • the first rotor hub 6 has two large angle grooves 10 located therein and each groove 10 is located adjacent a respective lobe 7.
  • the second rotor hub 8 has two smaller angle grooves 11 located therein and each groove 11 is located adjacent a respective lobe 9.
  • Timing gears 12 and 13 (Fig.
  • X (mounted on the rotor shafts) constrain the two rotors to rotate in timed interengaging relation.
  • a source of power is applied to the second rotor so as to rotate the second rotor in the direction shown.
  • a portion of the total power is then transmitted through the timing gears so as to rotate the first rotor in the opposite direction as shown.
  • the working fluid or gas to be compressed enters the lower pressure port 14, is compressed internally within the machine, and is then delivered through two higher pressure ports 15 which are located one in each flat end wall of the casing structure.
  • the ports 15 are alternately covered and uncovered by the first rotor 1 so as to control the flow of the working fluid through the ports.
  • the compressed gas is then conducted from the two ports 15 to a common outlet (not shown).
  • Fig. IV shows the rotors at the start of compression in chamber 16.
  • the smaller chamber 17 is near the end of delivery and is being closed out.
  • all the gas in chamber 17 is to be delivered through the ports 15 so as to avoid wasting any compressed gas.
  • the following requirements are needed: (a) the trailing edge of port 15 should be a circular arc projected from or by the outer radius of the second rotor, (b) the convex face of the second rotor lobe should be tangent to the outer radius of the same lobe, and (c) the circumferential width of said convex face (at the pitch circle) should be at least as large as the radial height of said convex face from the pitch circle outward.
  • Fig. V shows how chamber 16 and 18 equalize pressure after about 20 degrees rotation from Fig. IV.
  • the "precompression" in chamber. 18 is negligible due to the small angle lobes 7.
  • the volume of chamber 16 is larger than with equal diameter hubs and thus the effect of precompression is even further reduced because the tip of lobe 7 projects briefly into an even larger volume 16 (prior to release) and thus the precompression pressure is released at even lower pressure.
  • Fig. VII shows the rotor positions where the ports 15 are still covered by the first rotor. but will start to be uncovered in the next few degrees of rotor rotation.
  • the rotor and port profiles shown in Figs. IV to IX are calculated and drawn to scale for a 3 to 1 pressure ratio.
  • the discharge pressure of the second stage would be 3 x 3 x 14.7 + 132.3 PSIA - 117 PSIG.
  • the ports 15 start to be uncovered 20 degrees ahead of the theoretical pressure ratio of 3 location.
  • Such backflow represents a small energy loss which is more than compensated for in increased port area so that the net loss due to throttling through the ports 15 is less.
  • the early opening of port 15 might be compared (in a very general way) to advancing the spark in an internal combustion engine.
  • Embodiments of this invention teach the use of two lobes per rotor and no more. If (for instance) the machine instead had three or four lobes per rotor, then each lobe would have less angular distance to travel during the compression phase, and thus the discharge port 15 would have to be smaller in angle to secure the same built-in pressure ratio - a serious disadvantage.
  • the ports 15 are referred to as the higher pressure ports and the port 14 is referred to as the lower pressure port since this designation is applicable for operation as both a compressor and an expansion engine. Most of the discussion herein pertains to operation as a compressor, however, those improvements described for a compression cycle would also benefit the operation as an expansion engine.
  • Fig. VI shows the rotors at the formation of dump pockets 19.
  • the gas contained in the pockets 19 is only slightly pressurized and in about the next five degrees of rotor rotation this low pressure gas is dumped back to inlet pressure.
  • the calculated power loss due to dump pockets 19 is only 0.08% (less than one tenth of one per cent) of the adiabatic work of compression.
  • Double lobe rotors have 18% more displacement than single lobe rotors (even after deducting for the dump pockets previously described). This is because single lobe rotors have a dwell period during which no displacement occurs as can be seen in Figs. IV and V of Patent No. 3,472,445. More displacement per rotation (for a given rotor diameter and a given rotor width) is a very desirable feature since it increases capacity and reduces per cent leakage. Thus all embodiments herein use double lobe rotors instead of single lobe rotors.
  • a good rotor width for the first stage rotors is four to five inches.
  • Increased rotor width permits the machine to be more compact for a given capacity as follows: If (for example) the width of a rotor is doubled, then the capacity is more than doubled (less per cent leakage) but the overall size and weight of the machine is not doubled due to the space requirements of bearings, seals, timing gears, casing, etc. Conclusion: Wider rotors (made possible by increased port area) permits the machine to be more compact for a given capacity.
  • the hub diameter B' (Fig. VI) is made larger so as to increase the area of the ports 15.
  • a conflicting feature is that the radial dimension A' becomes relatively small and thus the possibility of a throttling loss (pressure drop) at this point is of concern.
  • a particular feature of this embodiment is that the throttling loss due to flow of the working fluid between the first rotor hub and the casing bore is much less (Figs. IV to X) than with the prior art (Figs. I to III) and the reasons therefore are explairin paragraphs (a), (b), (c). and (d) as follows: (a) Figs. I, II, and III are prior art drawings which illustrate single lobe retors in successive rotative positions. The porting shown is for 3 to 1 pressure ratio.
  • Figs. IV to IX illustrate anbodiments of the invention wiht to is for 3 to 1 pressure ratio.
  • the restriction A' (Fig. VI) is no longer present in Figs. VIII and IX because the dimension D' and E' (at the groove) are much larger than dimension A' (at the hub) hence less throttling (pressure drop).
  • the narrow restriction A' is in effect a much smaller percentage of the time.
  • the grooves 10 and 10 in the embodiments are larger in angle 4 i (Fig. X) compared to the small angle groove in Fig. II - the prior art.
  • the large angle grooves 10 result in less restriction more of the time.
  • the restriction A' is no longer in effect because dimension D' (at the groove) is much larger than dimension A' (at the hub), hence less throttling at D'.
  • Fig. IX illustrates how the large angle grooves 10 continue to present a wide open flow path E' for the gas so as to reduce the problem described for Fig. I. (the prior art) which has only one small angle grooved.
  • the volume contained in the two large angle grooves 10 of embodiments of this invention is about 3 or 4 times the volume contained in the single small angle groove (Fig. III of prior art). This means therefore that more volume is contained in the first rotor and less volume of gas needs to flow from right to left past the restriction A' (these embodiments) compared to the volume which must flow past restriction A in Fig. III (the prior art).
  • the torque (due to fluid pressure) is less on the first rotor 1 than on the second rotor 2; and this is because the hub diameter B' exceeds hubs diameter C'.
  • the following is a calculated example.
  • Two 12 inch diameter rotors compress air from atmospheric pressure to 44.1 PSIA.
  • the diameter of the first rotor hub is 10.5 inches (267 mm) and the diameter of the second rotor hub is 7.5 inches (190 mm).
  • the drive shaft is attached to the second rotor.
  • the per cent of power transmitted through the timing gears throughout a 360 degree rotation equals 38% of the total rotary shaft power input.
  • the drive shaft should be attached to the second rotor so as to reduce timing gear power. If the rotors are designed to run at a faster RPM than the drive motor, then a larger diameter drive gear may mesh with the timing gear on the second rotor shaft so as to drive same first and thereby reduce timing gear power.
  • the rotors shown in Fig, X are the same as the Fig. IV to IX rotors but to a larger scale.
  • the rotor axes are at X and Y.
  • the arcs T-N, U-P, 0-R, are circular with centres at X and Y.
  • the arcs U-P and I-J are at the hub diameters.
  • Profiles U-O are generated by tips G.
  • Profiles G-K are relieved in of the path generated by the tips N so as to permit easierfilling on the low pressure (suction) side of one machine. All remaining profiles are the running clearance of the two rotors as they
  • the so called knees P-Q are rounded so as to reduce the throttling loss between profile P-Q and the housing as shown as dimension D' in Fig. VIII.
  • each pitch circle has a diameter equal to the distance between the axes of rotation of the two rotors.
  • Each pitch circle has its centre on the axis of rotation of its respective rotor.
  • the rotor hubs are shown to be truly circular,although other forms are possible such as two slight spirals - as long as the two hubs rotate in sealing proximity to each other.
  • the two rotors are shown to have the same cuterdiameter and the two casing bores are shown to be equal, although it would be possible to make the bigger in diameter than the first two rotors rotate at equal rotative
  • the first rotor may be provided with small notches at location 0 in the sides of the rotor so as to provide more port area (particularly at the last portion of each delivery phase).
  • the ports 15 would be modified radially inward slightly so as to register with the notches.
  • the two rotors are shown to have straight profiles which are non-helical, and unlike screw machines, internal compression of the working fluid is made possible without the need of making the rotors helical.
  • helical rotors of low wrap angle could be used.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Applications Or Details Of Rotary Compressors (AREA)
  • Rotary-Type Compressors (AREA)
EP79301948A 1978-09-28 1979-09-19 Verdrängungsmaschinen mit rotierenden Kolben Withdrawn EP0009915A1 (de)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
US94659178A 1978-09-28 1978-09-28
US946591 1978-09-28

Publications (1)

Publication Number Publication Date
EP0009915A1 true EP0009915A1 (de) 1980-04-16

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ID=25484701

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EP79301948A Withdrawn EP0009915A1 (de) 1978-09-28 1979-09-19 Verdrängungsmaschinen mit rotierenden Kolben

Country Status (5)

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EP (1) EP0009915A1 (de)
JP (1) JPS5591702A (de)
AU (1) AU5101479A (de)
CA (1) CA1119139A (de)
ZA (1) ZA794573B (de)

Cited By (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
FR2520451A1 (fr) * 1982-01-25 1983-07-29 Ingersoll Rand Co Machine volumetrique rotative, notamment compresseur bietage a deux paires de rotors a lobes imbriques et rotor pour une telle machine
EP0133629A1 (de) * 1981-01-02 1985-03-06 Ingersoll-Rand Company Rotationskolben-Verdrängungsmaschine
US6776594B1 (en) * 2003-06-02 2004-08-17 Liung Feng Industrial Co., Ltd. Rotor mechanism
DE102008003077A1 (de) * 2008-01-03 2009-09-17 Grobe, Gerhard Explosionsmotor
CN1904365B (zh) * 2005-07-29 2010-06-16 良峰塑胶机械股份有限公司 爪式转子设计方法
CN103696961A (zh) * 2013-12-30 2014-04-02 南京航空航天大学 一种双转子相向旋转压缩机
CN103982425A (zh) * 2014-05-20 2014-08-13 上海齐耀螺杆机械有限公司 一种干式双螺杆压缩机转子的齿型
US9702361B2 (en) 2011-09-30 2017-07-11 Anest Iwata Corporation Claw pump with relief space

Families Citing this family (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPS60138202A (ja) * 1983-09-02 1985-07-22 インガ−ソル・ランド・カンパニ− 回転容積式機械
CA2814396A1 (en) 2010-10-22 2012-04-26 Peter South Rotary positive displacement machine

Citations (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3472445A (en) * 1968-04-08 1969-10-14 Arthur E Brown Rotary positive displacement machines
US3535060A (en) * 1969-03-21 1970-10-20 Arthur E Brown Rotary displacement machines

Family Cites Families (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPS5110362A (ja) * 1974-07-15 1976-01-27 Matsushita Electric Works Ltd Tasoinsatsuhaisenbanno seizohoho

Patent Citations (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3472445A (en) * 1968-04-08 1969-10-14 Arthur E Brown Rotary positive displacement machines
US3535060A (en) * 1969-03-21 1970-10-20 Arthur E Brown Rotary displacement machines

Cited By (9)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP0133629A1 (de) * 1981-01-02 1985-03-06 Ingersoll-Rand Company Rotationskolben-Verdrängungsmaschine
FR2520451A1 (fr) * 1982-01-25 1983-07-29 Ingersoll Rand Co Machine volumetrique rotative, notamment compresseur bietage a deux paires de rotors a lobes imbriques et rotor pour une telle machine
DE3248225A1 (de) * 1982-01-25 1983-08-04 Ingersoll-Rand Co., 07675 Woodcliff Lake, N.J. Rotationsmaschine mit zwangsverdraengung
US6776594B1 (en) * 2003-06-02 2004-08-17 Liung Feng Industrial Co., Ltd. Rotor mechanism
CN1904365B (zh) * 2005-07-29 2010-06-16 良峰塑胶机械股份有限公司 爪式转子设计方法
DE102008003077A1 (de) * 2008-01-03 2009-09-17 Grobe, Gerhard Explosionsmotor
US9702361B2 (en) 2011-09-30 2017-07-11 Anest Iwata Corporation Claw pump with relief space
CN103696961A (zh) * 2013-12-30 2014-04-02 南京航空航天大学 一种双转子相向旋转压缩机
CN103982425A (zh) * 2014-05-20 2014-08-13 上海齐耀螺杆机械有限公司 一种干式双螺杆压缩机转子的齿型

Also Published As

Publication number Publication date
ZA794573B (en) 1980-08-27
CA1119139A (en) 1982-03-02
AU5101479A (en) 1980-04-03
JPS5591702A (en) 1980-07-11

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