U.S. patent application number 14/518847 was filed with the patent office on 2015-04-23 for compressors.
The applicant listed for this patent is Steven Lee Wilhelm. Invention is credited to Steven Lee Wilhelm.
Application Number | 20150110647 14/518847 |
Document ID | / |
Family ID | 52826336 |
Filed Date | 2015-04-23 |
United States Patent
Application |
20150110647 |
Kind Code |
A1 |
Wilhelm; Steven Lee |
April 23, 2015 |
COMPRESSORS
Abstract
Disclosed compressor embodiments can prevent or substantially
reduce the inefficient upwardly spiraling heat buildup found in
typical compressors by significantly increasing the surface area of
mating piston and chamber walls, using the meshing and undulating
geometry of the piston/chamber walls to create increased air
turbulence in the chamber and generate more air flow against the
cooling walls, promoting turbulence even more by varying the axial
path of the piston head so that it is not simply linear, using
fluid to cool the non-working side of the piston, and/or providing
additional cooling features (such as fins) on the non-working side
of the piston to provide still further cooling means for the piston
head. Analogous principles can also be included in wobble-plate
compressors, scroll compressors, blowers, and other compressor
types.
Inventors: |
Wilhelm; Steven Lee;
(Corvallis, OR) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
Wilhelm; Steven Lee |
Corvallis |
OR |
US |
|
|
Family ID: |
52826336 |
Appl. No.: |
14/518847 |
Filed: |
October 20, 2014 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
|
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61961553 |
Oct 18, 2013 |
|
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|
61962523 |
Nov 12, 2013 |
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Current U.S.
Class: |
417/228 |
Current CPC
Class: |
F04C 18/12 20130101;
F04C 29/04 20130101; F04B 39/0005 20130101; F04B 53/08 20130101;
F04B 39/12 20130101; F04B 39/06 20130101; F04C 18/0207 20130101;
F04B 53/14 20130101 |
Class at
Publication: |
417/228 |
International
Class: |
F04B 53/08 20060101
F04B053/08; F04B 53/14 20060101 F04B053/14 |
Claims
1. A fluid compressor, which comprises: a housing; one or more
compression chambers that comprise both stationary and
non-stationary walls that define the compression chambers; wherein
the walls of the compression chambers are formed and structured to
have at least twice the surface area that the compression chamber
walls of a conventional compressor of the type has, and the
stationary and non-stationary walls of the compression chambers
partially or perfectly interleave or nest, or both, such that the
volume of the compression chambers can be reduced sufficiently to
effect a desired level of compression; wherein the walls of the
compression chambers are formed and structured such that a larger
portion of the fluid, compared to a conventional compressor of the
type, is close to one or more walls of the compression chambers for
most of the time during which the fluid is being compressed, such
that said portion of the fluid is in better thermal contact with
one or more walls of the compression chamber than in a conventional
compressor of the type; wherein at least some of the stationary
compression chamber walls are cooled by extraction of heat from
said walls by, or to, one or more external heat sinks, such that
said walls remain near or below the ambient temperature of the
fluid to be compressed prior to entering the compressor; wherein at
least some of the non-stationary compression chamber walls are
cooled by extraction of heat from said non-stationary walls by
fluid cooling, the extracted heat being conveyed by the cooling
fluid to one or more heat sinks, which heat sinks may be external
to the compressor, such that at least some of the non-stationary
walls of the compressor remain near or below the ambient
temperature of the fluid prior to entering the compressor; one or
more compressing features that reduce the volume of the fluid that
is to be compressed and thereby effect the compression of said
fluid, which compressing features may be the non-stationary walls
of the compression chambers, which non-stationary walls reduce the
volume of the compression chambers through their motion and thereby
effect the compression of the fluid wherein the one or more
compressing features are cooled by extraction of heat from said
compressing features by, or to, one or more heat sinks, such that
said compressing features, remain near or below the ambient
temperature of the fluid to be compressed prior to entering the
compressor.
2. The compressor of claim 1, wherein the manner or path in which
the non-stationary walls of the compression chambers move or are
moved is different from the manner or path of non-stationary walls
in a conventional compressor of the type; wherein said different
manner or path of movement of the non-stationary walls of the
compression chambers results directly or indirectly in increased
intra-chamber flow of the fluid to be compressed or results in
increased turbulence in the fluid to be compressed.
3. The compressor of claim 1, wherein the manner and/or path in
which the non-stationary walls of the compression chambers move or
are moved is that manner and path that are typical of a
wobble-piston compressor; wherein said wobble piston manner and
path of movement of the non-stationary walls of the compression
chambers results directly or indirectly in increased intra-chamber
flow of the fluid to be compressed and/or results in increased
turbulence in the fluid to be compressed.
4. The compressor of claim 2, wherein the compressor is a
reciprocating piston type compressor; wherein the stationary and
non-stationary walls of the compression chamber or chambers have a
geometry that is substantially rotationally symmetric, and the
different manner or path of movement of the non-stationary walls of
the compression chambers includes rotationally oscillatory motion,
and results in increased intra-chamber flow of the fluid being
compressed or results in increased turbulence in the fluid being
compressed.
5. The compressor of claim 1, wherein fluid cooling is not provided
for the non-stationary walls of the compression chambers.
6. A reciprocating piston-type compressor, comprising: a housing; a
piston that reciprocates within the housing along an axial
dimension of the compressor, the housing and the piston having
respective wall surfaces which together form a compression chamber
that varies in volume based on the axial position of the piston
relative to the housing; wherein the housing includes a fluid inlet
for receiving into the compression chamber a fluid to be compressed
and a fluid outlet for expelling out of the compression chamber the
fluid in a compressed state; wherein the housing comprises first
projections that project into the compression chamber; wherein the
piston comprises a piston head having a working side facing the
compression chamber and a non-working side facing away from the
compression chamber, wherein the working side of the piston head
includes second projections that project into the compression
chamber; wherein the second projections are oriented such that they
interleave between the first projections as the piston
reciprocates, and the first and second projections increase the
surface area of the wall surfaces and increase heat conduction away
from the fluid in the compression chamber when compared to a
cylindrical compression chamber; and wherein the compressor further
comprises a cooling fluid that is conducted into contact with the
non-working side of the piston head to conduct heat away from the
piston head via the cooling fluid.
7. The compressor of claim 6, wherein the cooling fluid is sprayed
against the non-working side of piston head.
8. The compressor of claim 6, wherein the non-working side of
piston head includes cooling fins that increase the surface area of
the non-working side of the piston and the cooling fluid is
conducted into contact with the cooling fins to conduct heat away
from the piston.
9. The compressor of claim 6, further comprising a cooling fluid
jacket positioned around the housing to conduct heat away from the
housing.
10. The compressor of claim 6, wherein the first and second
projections are arrayed circumferentially around a common central
axis and the first and second projections increase in
circumferential thickness as a function of increasing radial
distance from the central axis.
11. The compressor of claim 6, wherein the second projections
oscillate in a non-axial direction relative to the first
projections while the piston reciprocates axially within the
housing, thereby increasing turbulence within the fluid and further
increasing heat conduction away from the fluid.
12. The compressor of claim 6, wherein the first projections are
integral with side walls of the housing.
13. The compressor of claim 1, wherein the fluid enters and exits
the compression chamber via a common central port opposite from the
piston.
14. A reciprocating piston-type compressor, comprising: a housing;
a piston that reciprocates within the housing along an axial
dimension of the compressor, the housing and the piston having
respective wall surfaces which together form a compression chamber
that varies in volume based on the axial position of the piston
relative to the housing; wherein the housing includes a fluid inlet
for receiving into the compression chamber a fluid to be compressed
and a fluid outlet for expelling out of the compression chamber the
fluid in a compressed state; wherein the housing comprises first
projections that project into the compression chamber; wherein the
piston comprises a piston head having a working side facing the
compression chamber and a non-working side facing away from the
compression chamber, and the working side of the piston head
includes second projections that project into the compression
chamber opposite the first projections; wherein the second
projections are oriented such that they interleave between the
first projections as the piston reciprocates, and the first and
second projections increase the surface area of the wall surfaces
and increase heat conduction away from the fluid in the compression
chamber when compared to a cylindrical compression chamber; and
wherein the piston and the second projections oscillate in a
non-axial direction relative to the housing and the first
projections while the piston reciprocates axially within the
housing, thereby increasing turbulence within the fluid and further
increasing heat conduction away from the fluid.
15. The compressor of claim 14, wherein the piston moves through at
least one full oscillation during each axial compression stroke of
the piston.
16. The compressor of claim 14, wherein the oscillation of the
piston is caused by a mechanical interface between the piston and
the housing.
17. The compressor of claim 14, wherein the oscillation of the
piston and second projections comprises rotational motion of the
piston about a central axis of the compressor.
18. The compressor of claim 14, wherein the first and second
projections are arrayed circumferentially around a common central
axis and the first and second projections increase in
circumferential thickness as a function of increasing radial
distance from the central axis.
19. The compressor of claim 14, wherein each of the second
projections oscillates in a corresponding valley between two of the
first projections during the axial reciprocation of the piston.
20. The compressor of claim 14, wherein the first and second
projections have a generally trapezoidal cross-sectional shape when
viewed in a plane perpendicular to the axial direction.
Description
CROSS-REFERENCE TO RELATED APPLICATIONS
[0001] This application claims the benefit of U.S. Provisional
Patent Application No. 61/961,553, filed Oct. 18, 2013, and U.S.
Provisional Patent Application No. 61/962,523, filed Nov. 12, 2013,
both of which are incorporated by reference herein in their
entirety.
FIELD
[0002] This disclosure is related to air compressors, pneumatic
motors, and related devices, and particularly to near-isothermal
embodiments and/or those having a high degree of efficiency.
BACKGROUND
[0003] Compressors for compressing fluids, such as air, are
inefficient. The amount of energy consumed by such compressors
greatly exceeds the theoretical amount of work done merely to
compress the fluid. The remainder of the consumed energy ends up as
unwanted waste heat.
[0004] The true inefficiency of practical compressors appears not
to be widely known. Discussions of compressor inefficiency in the
literature tend to compare isothermal compression versus adiabatic
compression, with isothermal compression being theoretically the
most efficient compression. Adiabatic compression is sometimes
thought to be the worst case possible. The calculations are usually
done assuming reversible processes. A reversible adiabatic 10:1
volume reduction for air requires only 1.63 times as much as or 63%
more work than a reversible isothermal 10:1 volume reduction for
air.
[0005] However, actual compressors, working against their stated
maximum working pressure and fully warmed up to steady-state
operation, are far more inefficient than reversible adiabatic
theory suggests. A truer indication of how compressors tend to
operate appeared in a recent advertisement in an engineering
journal. Kaeser Compressors, Inc., a major manufacturer of
industrial compressors and blowers, noted, "It is a fact:
Compressed air is inherently inefficient. It takes 8 kW of
electricity to deliver 1 kW of power in compressed air--and almost
all of the remaining 7 kW is lost as heat." By this one measure,
the energy penalty in industrial compressors is about 700%. Some of
this loss, about 15% of the total 8 kW, is due to friction, and not
to the inefficiency of the compression process itself. That leaves
a penalty of about 5.8 kW, or 73% of the energy used in the best
compressors, being lost as waste heat. This is not how compressors
tend to be described and discussed in the literature.
[0006] The wastefulness of real-world compressors has long been
lamented, and many have opined that if only an isothermal
compressor were possible, much energy could be saved. But, an
isothermal compressor is widely believed to be impossible. So much
so that the concept of an isothermal compressor has been described
as something of a "Holy Grail." Conventional wisdom suggests that
development of an isothermal compressor is akin to trying to
develop a perpetual motion machine, or turn lead into gold.
[0007] Contrary to conventional wisdom, thermodynamic theory
indicates that an isothermal compressor is possible, but only if
the compressor is operated infinitely slowly, so that the amount of
heat that must flow out of the gas during isothermal compression
can flow out reversibly. If any compressor could operate infinitely
slowly, it would remain in thermal equilibrium with its
surroundings, and its operation would be truly isothermal. Thus,
almost any common piston-type compressor, for example, could be run
isothermally at near zero rotational speed (assuming perfect
sealing, so that the developed pressure is not lost). So, an
isothermal compressor is not impossible. It is just not possible in
any practical, useful sense. A compressor that requires an infinite
amount of time per compression cycle is not practical; in fact it
is of no value at all.
[0008] If means can be found to keep a compressor and the
compressing gas in near-thermal equilibrium with the surroundings,
but at operational speeds that are useful, something like
isothermal compression can be achieved. A problem which
conventional wisdom fails to recognize is that the heat developed
by the compression process must be drawn out of the gas quickly
enough, and during the compression if near-isothermal compression
is to be achieved. Thus, developing a practical near-isothermal
compressor is not just a thermodynamics problem; it is also
kinetics problem.
[0009] It is quite a kinetics problem. Reciprocating (piston-type)
compressors, for example, tend to run in the 800 to 1,000 rpm
range. At 900 rpm, the compression stroke lasts only 33
milliseconds (0.033 seconds). If the heat of compression is going
to be extracted from the gas during the compression, in emulation
of isothermal operation, the heat has to be extracted in a fraction
of this short time span. Perhaps something like 20 milliseconds is
available during each compression stroke to remove the heat, or
some large fraction of it.
[0010] Another common misconception is made by those who assume
that an isothermal compressor, and nothing less, is what is needed.
It is not. Almost all of the value that would be obtained from a
fully isothermal compressor can be had from a compressor that is
only close to isothermal. In fact, it is not even necessary to get
very close. Any approximation of isothermal operation would reap
large benefits.
[0011] The reason for this is the compounding nature of the
temperature rise that occurs during adiabatic compression. The
initial part of the compression goes reasonably well, with little
increase in temperature. But, the temperature rise feeds on itself.
The higher the temperature, the faster the temperature rises. The
losses begin to compound. In the end, even a modest compression of
a fluid, say from 15 psi to 150 psi, results in a temperature rise
of hundreds of degrees.
[0012] As a specific example, compressing a quantity of a gas from
70.degree. F. and 15 psi to just over three times the pressure, 50
psi, results in a temperature rise of 335 degrees Fahrenheit, to
405.degree. F. But, triple the pressure again, to 150 psi, and the
temperature rises another 477 degrees Fahrenheit, to 882.degree. F.
Triple the pressure again, to 450 psi, and the temperature rises
another 739 degrees Fahrenheit, to 1,621.degree. F. An aluminum
compressor piston would melt. Yet, these are fairly modest
pressures.
[0013] One conventional method used to interfere with the
temperature rise during compression is water injection. Water
injection is currently used in some commercially available screw
compressors, and experiments have been done with water injection in
reciprocating (piston-type) compressors. Some energy savings have
been achieved. However, there are technical challenges with water
injection. One disadvantage of water injection is that it adds
significant water to the compressed gas. For most uses, the water
has to be removed, and dehumidification requires a
refrigerator-type drier that consumes significant energy (using,
incidentally, a compressor to achieve the cooling).
[0014] One conventional attempt to interrupt the temperature rise
is to use multi-stage compressor trains, with inter-stage cooling
of the gas. Depending on the final pressure required, two-stage, or
even three-stage compressors can be more economical than
single-stage compressors, in spite of the additional complexity,
friction, maintenance, and capital investment. (Pressures above 150
psi are generally not attempted with single-stage compressors, due
to excessively high operating temperatures.)
[0015] One reason why little effort (other than water injection)
has been made to interfere with the temperature rise in compressors
is that as the gas is compressed in the compression chamber (a
cylinder in piston-type compressors), it heats up essentially
homogeneously. Except for the gas that is immediately close to the
walls of the compression chamber (stator, rotor, scroll, cylinder
wall, piston, head, etc.), all of the gas heats throughout at the
same rate; and, there seems to be no means possible to stop the
heating (other than water injection).
[0016] The problem is that only the gas within a few mean free
paths of the walls of a compressor can lose some heat to the walls
and thus avoid some of the temperature rise that would otherwise
occur. The mean free path of air molecules, for example, at one
atmosphere pressure is on the order of 70 nM or about two and a
half millionths of an inch. At higher pressures, it is
proportionally less. Only the gas that is within a few microns of
the walls of a compressor is in thermal contact with the walls. The
remainder of the gas is thermally insulated from the walls by the
gas next to the walls. The great majority of the gas in a
compressor is well insulated from the (potentially cooler) walls.
Stagnant air, it turns out, is a good thermal insulator. This is
why fiberglass insulation works so well in homes. The fiberglass
itself provides almost no insulating value; it is the layer of
stagnant air it creates that is the actual insulator.
[0017] This helps explain why conventional attempts to cool
compressors, such as the use of water jackets, cooling fins and
cooling fans used to cool the outer wall of the compression
chamber, are sufficient to allow compressors to function
adequately, sometimes continuously, but do little to address the
fundamental efficiency problem of compressors.
[0018] More specifically, because most of the gas is well insulated
from the cooler chamber wall of a well cooled compressor, the gas
still begins to rise in temperature as the compression begins,
especially in the center of the gas. This causes excess work to be
done for further compression. Which causes the temperature to rise
further and faster, in a very quick upward spiral, as the gas is
suddenly compressed. The result is a highly inefficient
operation.
[0019] In conventional compressors, the compressed gas may lose
some of its heat to the cooler walls as it flows turbulently out of
the compression chamber.
[0020] However, post-compression cooling of the gas, once the
excess work is done and the game is lost, is not a good solution.
And such post-compression cooling results in a pressure drop which
is counter-productive. What is needed is a way to avoid the excess
work, to stop the upward temperature spiral as it is happening, or
better yet, to trip it up, so that it does not get started in the
first place. Contrary to conventional approaches, the solution is
to interfere with the compression heating of a gas in a compressor,
and thereby at least partially avoid it, rather than dealing with
its consequences after the fact.
[0021] Such a solution is not hopeless, as commonly believed,
because the amount of heat that must be removed from a gas as it is
isothermally compressed is quite small. Real-world compressors
generate so much heat, and that heat is such a monumental problem,
that it is easy to believe, incorrectly, that a large amount of
heat must be extracted from the gas as it is compressed to achieve
isothermal compression. This is far from true. The actual amount of
heat lost by a gas in a theoretical, infinitely slow isothermal
compression is actually quite small. All the other heat that we
associate with conventional compressors is excess heat that need
not be generated at all. The excess heat largely is a result of the
upward spiral of temperature rise, with each succeeding increment
of compression being harder than the last due to the rising
temperature and requiring even more work. It is a perfect
ratcheting upward of something that should be avoided altogether,
if possible.
[0022] As an example, it is easy to calculate for a very common,
rather small compressor that the theoretical, isothermal-equivalent
work done by the compressor is a fraction of the total amount of
shaft horsepower the compressor consumes from an electric motor in
actual operation. Assuming the following technical specifications:
Bore--2.375'', Stroke--2'', Rotational Speed 919 RPM, Maximum
Operating Pressure--125 psi, Flow Rate at Maximum Operating
Pressure--1.54 scfm (Factory claim=33% volumetric efficiency),
Power Requirement--1 Horsepower, the isothermal-equivalent work
being done is approximately 1/3 of a horsepower. And, this is the
full amount of heat that needs to be extracted from the gas, if the
compression can be kept isothermal throughout the stroke. All other
heat generated by this compressor is unnecessary. 1/3 of a
horsepower is only 250 Watts. It is quite easy to deal with 250
Watts of heat: that amount of heat put into one liter of water for
a full minute raises the temperature of the water by only 3.6
degrees Centigrade or 6.5 degrees Fahrenheit.
SUMMARY
[0023] One aspect of the disclosed technology is to significantly
increase the inner surface area of the piston head and compression
chamber to allow for much greater cooling of the air during the
compression cycle, thereby disrupting the rapidly compounding heat
buildup with each incremental increase in compression. In some
embodiments, the piston head and chamber are provided with opposing
interleaving or nesting projections that provide the piston head
and chamber walls with much greater surface area (compared to a
typical cylindrical piston design) to cool the air much more
effectively as it compresses. The moving piston can still move
axially within the compression chamber, allowing the mating piston
and chamber walls to compress the air as much as is needed and yet
also act as a more effective cooling means. As a another benefit,
the mating projections/protuberances on the piston head and chamber
wall surfaces also create greater air turbulence during the
compression stroke, thereby causing a larger volume of air to pass
closely by the wall surfaces and be cooled.
[0024] Another aspect of the disclosed technology is to laterally
or angularly shift or rotate slightly the piston during the
compression stroke such that the piston path is no longer perfectly
axial. This more complex movement differs from the traditional
linear axial stroke of conventional compressor pistons and creates
more air turbulence within the chamber, thereby promoting even more
cooling as a larger volume of air passes in close proximity to the
cooling wall surfaces of the piston and chamber.
[0025] Still another aspect of the disclosed technology is the use
of a fluid (oil or water, for example) to internally cool the
piston head and ultimately the compressed air in the chamber. Some
embodiments include a cooling oil spray on the internal non-working
side of the piston. Some embodiment include additional internal
(i.e., non-working side) cooling features, such as cooling fins,
projections, protuberances, etc. For example, the cooling fluid can
be sprayed against the internal cooling fins, projections, etc. on
the non-working side of the piston to provide even greater cooling
of the piston and ultimately the compressed air.
[0026] As a whole, disclosed compressor embodiments described
herein can prevent or substantially reduce the inefficient upwardly
spiraling heat buildup found in typical compressors by
significantly increasing the surface area of mating piston and
chamber walls, using the meshing and undulating geometry of the
piston/chamber walls to create increased air turbulence in the
chamber and generate more air flow against the cooling walls,
promoting turbulence even more by varying the axial path of the
piston head so that it is not simply linear, using fluid to cool
the internal non-working side of the piston, and/or providing
additional cooling features (such as fins) internal to the piston
to provide still further cooling means for the piston head.
[0027] The foregoing and other objects, features, and advantages of
the disclosed technology will become more apparent from the
following detailed description, which proceeds with reference to
the accompanying figures.
BRIEF DESCRIPTION OF THE DRAWINGS
[0028] FIG. 1 is a perspective view of an exemplary compression
chamber geometry.
[0029] FIG. 1A is a partially cross-sectional side elevation view
of an exemplary compressor.
[0030] FIG. 1B is an enlarged view of a portion of FIG. 1A.
[0031] FIG. 2A is a partially cross-sectional side elevation view
of another exemplary compressor.
[0032] FIG. 2B is an enlarged view of a portion of FIG. 2A.
[0033] FIG. 3 is a perspective view of a portion of an exemplary
compressor.
[0034] FIG. 4 is an exploded view of the compressor of FIG. 3.
[0035] FIG. 5 is a side view a piston and a housing head of the
compressor of FIG. 3, showing opposing projections nested
together.
[0036] FIG. 6 is a perspective view of the housing head of the
compressor of FIG. 3.
[0037] FIG. 7 is a side view of the housing head of the compressor
of FIG. 3.
[0038] FIG. 8 is a cross-section bottom view of the housing head
taken along section line 8-8 of FIG. 7.
[0039] FIG. 9 is a perspective view of the piston of the compressor
of FIG. 3.
[0040] FIG. 10 is a bottom view of the piston of the compressor of
FIG. 3.
[0041] FIGS. 11A-11C illustrate isolated non-axial relative motion
between opposing surfaces in a compression chamber and resultant
gas flow therebetween.
[0042] FIG. 12 illustrates a combination of axial compression
motion and non-axial oscillation motion between opposing
projections in a compression chamber.
[0043] FIG. 13 is a partially cross-sectional view of an exemplary
wobble plate-type compressor.
[0044] FIG. 14 is an enlarged view of cooling projections in the
compressor of FIG. 13.
[0045] FIG. 15A shows a port of the compressor of FIG. 13 in an
intake configuration.
[0046] FIG. 15B shows the port of the compressor of FIG. 13 in an
exhaust configuration.
[0047] FIG. 16 is a schematic illustration of an exemplary
scroll-type compressor.
[0048] FIGS. 17A and 17B are cross-sectional views of an exemplary
scroll-type compressor represented by FIG. 16, taken along section
lines 17A-17A and 17B-17B, respectively, of FIG. 16.
[0049] FIGS. 18A and 18B are cross-sectional views of another
exemplary scroll-type compressor represented by FIG. 16, taken
along section lines 18A-18A and 18B-18B, respectively, of FIG.
16.
[0050] FIG. 19 is a partially cross-sectional view of an exemplary
blower-type compressor.
[0051] FIG. 20 is a cross-sectional view of the blower-type
compressor of FIG. 19, taken along section lines 20-20 in FIG.
19.
[0052] FIG. 21 is an enlarged view of a portion of FIG. 19, showing
an interface between nesting projections of the housing and the
rotor.
DETAILED DESCRIPTION
[0053] For purposes of this description, certain aspects,
advantages, and novel features of the embodiments of this
disclosure are described herein. The disclosed methods,
apparatuses, and systems should not be construed as limiting in any
way. Instead, the present disclosure is directed toward all novel
and nonobvious features and aspects of the various disclosed
embodiments, alone and in various combinations and sub-combinations
with one another. The methods, apparatuses, and systems are not
limited to any specific aspect or feature or combination thereof,
nor do the disclosed embodiments require that any one or more
specific advantages be present or problems be solved.
[0054] Although the operations of some of the disclosed methods are
described in a particular, sequential order for convenient
presentation, it should be understood that this manner of
description encompasses rearrangement, unless a particular ordering
is required by specific language. For example, operations described
sequentially may in some cases be rearranged or performed
concurrently. Moreover, for the sake of simplicity, the attached
figures may not show the various ways in which the disclosed
methods and devices can be used in conjunction with other methods
and devices.
[0055] As used herein, the terms "a", "an" and "at least one"
encompass one or more of the specified element. That is, if two of
a particular element are present, one of these elements is also
present and thus "an" element is present. The terms "a plurality
of" and "plural" mean two or more of the specified element.
[0056] As used herein, the term "and/or" used between the last two
of a list of elements means any one or more of the listed elements.
For example, the phrase "A, B, and/or C" means "A," "B," "C," "A
and B," "A and C," "B and C" or "A, B and C."
[0057] As used herein, the term "coupled" generally means
mechanically, chemically, or otherwise physically coupled or linked
and does not exclude the presence of intermediate elements between
the coupled or associated items absent specific contrary
language.
[0058] As used herein, the term "compressor" means a device that
use mechanical work to compresses a gas, such as air, and the term
excludes devices, such as internal combustion engines, that include
or operate using combustion or other chemical release of energy
within a chamber.
Compression Chamber Geometry
[0059] The following description recognized that, in order to
minimize the compression heating that occurs in a compressor, and
thereby increase the efficiency of the compressor, it is very
beneficial to extract heat from throughout the gas in the
compression chamber directly into surfaces having high thermal
conductivity. In a traditional cylinder shaped compression chamber,
a relatively low percentage of the gas in the chamber is adjacent
to the walls of the chamber, and the large percentage of the gas in
the middle of the chamber is only cooled via relatively inefficient
conduction of heat from the central volume of the gas to the outer
walls of the chamber. This leads to poor cooling of the gas in the
middle of the chamber. To more efficiently extract heat from the
entire gas volume in the chamber, additional cooling surfaces can
be placed inside the compression chamber such that the gas near the
center of the chamber is closer to the nearest surface and can
transfer heat to the closer surfaces faster.
[0060] A conventional cylindrical cylinder shape with flat upper
and lower walls makes sense from the perspective of maximizing the
total volume of the cylinder and keeping it free of complicating
features that take up space, space that can otherwise be filled
with more gas. However, the principles and designs disclosed herein
recognize that surface features, such as projections or
protuberances, that extend from the working face of the piston or
from the opposing face of the head or from the lateral walls of the
compression chamber do not necessarily result in less usable room.
In fact, in a piston-type compressor, the working volume of the
compression chamber is determined by the cross-sectional area of
the chamber and the stroke length of the piston (referred to as the
swept volume). The piston and the head can have any shapes, without
subtracting from or adding to the working volume or clearance
volume of the cylinder, as long as a sufficient cross-sectional
area of the chamber and a sufficient stroke length is provided. In
addition, if an irregular shape on the working face of the piston
nests matingly with a corresponding opposite irregular shape on the
face of the head, the volume in the chamber can reduce to near
zero, or whatever minimum volume is desired, just like with a
traditional cylinder with flat piston and head surfaces, such that
the same compression ratios can be accomplished with the same
stroke lengths.
[0061] As used herein, the terms "head" and "housing head" and
"cylinder head" are used to mean the stationary end portion of the
compression chamber (e.g., the top end in the orientation of FIG.
1A) that is opposite from the piston head, and which may or may not
be an integral portion of the overall stationary portion of the
compressor. The head and the lateral side walls of the stationary
portion of the compressor can, along with other stationary
components, be referred to collectively as the "housing" of the
compressor. The stationary lateral side surfaces of the compression
chamber may or may not have a circular cross-sectional shape. For
example, any cross-sectional shape (e.g., circular, elliptical,
polygonal, etc.) can be provided for the lateral side surfaces of
the compression chamber so long as that shape is consistent along
the axial path of the piston head so that a seal can be maintained
between the perimeter of the piston head and the lateral side
surfaces of the compression chamber.
[0062] At the top of the piston stroke (sometimes referred to as
"top dead center"), it can be desirable for the piston and head to
approach very closely, with little space between them, so that the
gas that has been compressed can be effectively cleared from the
compression chamber. It is conventionally believed that surface
features on the working side of the piston head and/or on the
opposing housing head will reduce the potential compression ratio
and clearance volume because they inhibit the piston and head from
coming into very close proximity at the top of the stroke. However,
as described herein, a near zero (or desirably small) compression
chamber volume can still be achieved even when the working side of
the piston head and/or on the opposing housing head include large
projections extending into the compression chamber.
[0063] Thus, a piston-type compressor can be designed with many
different shapes of piston and head, even shapes that are far
different from the nearly flat to completely flat shapes that are
conventional. And, the chamber side walls can be shapes other than
purely cylindrical. This freedom allows inclusion of increased
cooling surfaces that are disclosed herein. Piston and head
geometries are disclosed that extend cooling surfaces into the
inner, central regions of the compression chamber, and yet allow
for close approach of the piston and head at top dead center, which
makes it possible to provide piston-type compressors that compress
a gas, while at the same time extract significantly more heat from
the gas than in a conventional design, and thereby keep the gas
from rising significantly in temperature due to compression
heating.
[0064] In some of the embodiments disclosed herein, the lateral
walls of the housing that form lateral surfaces of the compression
chamber serve as enhanced heat extraction devices, shaped and
placed to maximize the extraction heat from the compressing and
exiting gas. The round, smooth side walls of a traditional
cylindrical compressor extract a minimal amount of heat from the
compressing and exiting gas due to their minimal surface area to
gas volume ratio. The minimal loss of heat from the gas to the
smooth side walls of a conventional cylindrical chamber is
incidental to the large rise in temperature in the gas. By
contrast, in disclosed compressor embodiments herein, the
compression chamber walls are much more effective at extracting
heat from the gas and thereby interfering with the temperature rise
in the gas. For example, in some embodiments, projections extending
axially from the piston head and the opposing housing head can also
be integral with and/or in direct contact with the lateral side
walls of the chamber, allowing them to efficiently conduct heat
from all parts of the chamber, including the central parts of the
chamber, both in the axial directions to the piston and head, but
also laterally (e.g. radially) to the side walls of the housing. In
some embodiments, for example, the projections can have major
dimensions in the axial and radial directions (the desired
directions of heat conduction), while being relatively thin in the
circumferential or angular directions to provide a maximum surface
area per volume ratio for the projections. The opposing piston and
head projections can also be tapered in the axial direction so that
they interleave and nest with each other to provide a minimal
chamber volume at the top of the piston stroke.
[0065] The concept of incorporating increased cooling surface area
in the compression chamber via projections that extend into the
inner regions of the compression chamber and extract heat from
throughout the compressing gas, without interfering with the
compression, applies equally well to all types of compressors that
involve compression chambers, such as screw compressors, wobble
plate-type compressors, and scroll-type compressors. But, for the
present discussion, the application of these concepts in a
reciprocating piston-type compressors will be used as a
representative, non-limiting example for ease of description.
[0066] Broadly described, exemplary devices for compressing gases
and other compressible fluids can include a compression chamber of
varying volume so that a compressible fluid can be drawn into the
chamber as the volume of the chamber increases and can be
compressed to a higher pressure and expelled from the chamber at a
higher pressure as the volume of the chamber decreases. They can
further include cooling surfaces that extend from the walls of the
compression chamber, whether stationary or not, that define the
compression chamber, into the inner regions of the compression
chamber and extract heat from the compressing fluid in the inner
regions of the compression chamber during the compression, without
significantly hindering the variability of the compression chamber
volume or substantially hindering the compression and expulsion of
the fluid. The cooling surfaces can also be coupled to effective
external cooling such that the heat extracted from the compressing
and exiting fluid is subsequently efficiently extracted from the
compression chamber, keeping the cooling surfaces from rising
significantly in temperature, so that a large portion of the heat
is extracted from the compressing fluid during each compression
cycle.
[0067] Many different geometries consistent with the herein
described improvements substantially improve heat extraction from
the gas, and all such geometries are included and encompassed by
this disclosure. Such geometries may include single cooling
surfaces, multiple cooling surfaces that are arranged in rows,
circles, other closed or annular shapes, spirals, are arranged
radially or in arrays, are arranged symmetrically within the
compression chamber, or are without significant symmetry or even
significant pattern. Various different geometries can offer a
combination of different advantages. Example advantages can include
the following:
[0068] (a) Some geometries offer greatly increased surface area,
such as additional surface expanses that the compressing and/or
exiting gas can come into contact with and lose heat to.
[0069] (b) Some geometries are more effective than others in
establishing effective flow paths for heat that is extracted from
the compressing or exiting gas and conveyed to the external cooling
sources.
[0070] (c) Some geometries, both symmetric and asymmetric, move
some of the gas from one or more regions of the compression chamber
to one or more other regions of the compression chamber during the
compression and/or expulsion of the gas, resulting in greater flow
of the gas within the compression chamber during compression and/or
expulsion of the gas.
[0071] (d) Some geometries create an advantageous pattern or
patterns to the flow of the gas as it enters the compression
chamber, or during the compression and/or expulsion of the gas,
patterns such as swirling or high-turbulence patterns of flow, that
result in increased thermal contact between the gas and the walls
of the chamber and/or the cooling surfaces.
[0072] (e) Some geometries create more turbulence in the flow of
the gas over the compression chamber walls and/or the cooling
surfaces during the compression and/or expulsion of the gas, and
thus improve the exchange of heat between the gas and the walls of
the compression chamber and/or the cooling surfaces; as examples,
some or all of the surfaces of the compression chamber walls and
cooling surfaces can be undulating, furrowed, dimpled, rough, or
polished.
[0073] There are other advantages, besides these few examples, that
can be achieved through various geometries of the walls of the
compression chamber and the cooling surfaces. All such geometries
and advantages are included in this disclosure and encompassed by
the claims of this application where applicable.
[0074] Continuing with the example of the application of this
technology to a piston-type compressor, FIG. 1 shows one of many
exemplary cooling surface geometries that can be incorporated in a
compression chamber. In FIG. 1, the upper geometry, including
arrayed projections 24, can be stationary and located in the
stationary upper end of the chamber and optionally integral with
the head and/or side walls of the housing. The lower geometry,
including arrayed projections 34, can be movable and optionally
integral with the piston head. For purposes of illustration, the
opposing cooling surfaces are shown separated more than their
maximum spacing when the piston is at the bottom of the stroke.
This exemplary geometry is circularly symmetric (a.k.a.,
rotationally symmetric) with twelve projections on each opposing
portion and a twelve-fold (C12) symmetry. In other embodiments, the
geometry can include 4, 6, 8, 10 or other numbers of projections
(including odd numbers) on each opposing section, and corresponding
different levels of symmetry. Also, it is not necessary that a
piston, chamber, and head have circular symmetry, or any symmetry
at all.
[0075] In any of the compressor embodiments disclosed herein, the
described geometry can significantly increase the surface area of
the compression chamber compared to a conventional cylindrical
chamber having flat top and bottom walls. In comparison to a
conventional cylindrical chamber having flat top and bottom walls
(taking for comparison a cylindrical chamber wherein the vertical
height of the cylinder is equal to the diameter of the cylinder), a
compression chamber having geometry as described herein and having
an equivalent diameter and height (vertical distance between
opposing top and bottom surfaces) can have a surface area that is
at least 200%, at least 300%, at least 400%, and/or at least 500%
greater. The particular example geometry shown in FIG. 1 has more
than five and a half times the surface area of an equivalent
cylinder with a flat piston and a flat head. This increase in
surface area, optionally along with any combination of other
cooling features disclosed herein, results in an increase in the
head extraction from the compressing gas and a corresponding
increase in efficiency of the compressor. For example, for the
disclosed compressors, the energy efficiency of the compressor can
be greater than 33%, greater than 50%, greater than 66%, and/or
greater than 75%.
[0076] In this example geometry, the cooling surfaces also have two
optional features: (1) the opposing surfaces of the housing head
and of the piston head are identical; and (2) they also "nest"
perfectly. That is, if brought together all the way to the point of
contact, there is no space between the opposing surfaces. The
ability to nest perfectly makes higher compression ratios (lower
clearance volumes) possible; and, efficient chamber designs may
tend in this direction. But, in some cases, perfect nesting may
well be sacrificed to some extent, in order to capture other
advantages of geometries that do not evince perfect nesting.
[0077] Another advantage of the exemplary geometry of FIG. 1, and
many others, is that the cooling surfaces of the head can be
continuous with (i.e., part of or integral with), or contiguous
with and in intimate thermal contact with, the side walls of the
housing, which can itself be cooled externally (e.g., with cooling
fins, a water jacket, fluid spray, and/or other cooling
mechanisms). Whether the projections in the compression chamber are
continuous/integral with or contiguous to the housing walls or not,
the projections can convey heat to and lose heat to external
cooling that is applied alongside of the lateral housing walls that
form the compression chamber. The heat extraction, therefore, is
not limited to extraction through the head and piston only. This
greatly increases the efficiency of removing heat from the gas as
it compresses and as it is expelled from the compression
chamber.
[0078] A useful effect of this exemplary geometry, and many other
possible geometries, is exchange of heat between the stationary and
the non-stationary walls of the compression chamber. Due to
differences in the ability to cool the stationary and
non-stationary walls of the compression chamber, it can be more
difficult to keep the temperature of the non-stationary wall(s) of
the compression chamber (e.g., the piston head in a piston-type
compressor) as low as desired. With the exemplary geometry of FIG.
1, and many others, the (possibly turbulent) flow of gas into the
compression chamber during the inlet phase of the cycle can extract
heat from higher temperature wall portions of the compression
chamber, and can release heat to cooler temperature wall portions
of the chamber as the gas enters the chamber. Heat is thus
transferred between the different walls of the compression chamber,
tending to even out the temperatures of the compression chamber
walls. This effect can prove especially useful in keeping the
non-stationary wall(s) of the compression chamber cooler, as they
can, in some geometries, prove to be the more difficult walls to
cool. To enhance this effect, it is beneficial for the compression
chamber geometry to increase the flow of gas within the chamber
(intra-chamber flow) and increase the turbulence of the gas as it
moves in the compression chamber, which thereby increases heat
transfer between the walls and the gas.
Cooling Moving Walls with Cooling Fluid
[0079] The stationary walls of a compression chamber are relatively
easy to provide with sufficient external cooling to keep the
stationary surfaces of the chamber at low enough temperatures to
continuously extract heat from the gas over many compression
cycles. However, a non-stationary wall of a compression chamber,
such as the piston head in the present example of a piston-type
reciprocating compressor, can be more difficult to provide with
effective cooling. In piston-type compressors described herein, the
piston head is used not only as a compressing surface, but also as
an active member in extracting heat from the gas. The heat that the
piston head extracts (or accepts) from the gas must be carried away
and delivered to an effective heat sink, in order to keep the
piston from eventually heating up and becoming ineffective as a
cooling surface.
[0080] Generally, a non-stationary wall of a compression chamber
cannot be cooled with a stationary water jacket or with stationary
external cooling fins, as in the case of a stationary walls,
because there is no pathway to efficiently convey heat from the
moving compression chamber wall to a heat sink such as a water
jacket or a stationary cooling fin. However, in some embodiments
disclosed herein, a moving heat conveyance pathway can be provided
to carry the heat extracted from the compressing gas by the
non-stationary compression chamber wall to one or more, possibly
stationary, external heat sinks. In some embodiments, an
intermediary substance or object is provided that can both extract
(or accept) heat from a non-stationary compression chamber wall and
convey it to and release it to a, possibly stationary, heat sink.
Some such embodiments utilize a fluid that can serve as the
intermediary heat transfer substance. In some embodiments of a
piston-type compressor, oil cooling of the non-working side of the
piston can convey the heat extracted by the piston from the
compressing and exiting gas to an external heat sink, such as, for
example, a radiator. Cooling fluids other than oil can
alternatively be used. In some embodiments, the cooling fluid can
also act as a lubricant, such as using the same fluid that is used
to cool the non-stationary compression chamber walls. As used
herein, the term "fluid" includes any combination of gas, liquid,
vapor, suspension, or other substance that flows under applied
shear stress.
[0081] The use of a cooling fluid (e.g., oil) conducted into
contact with the non-working side of the piston head, or other
moving compression chamber wall, serves the purpose of extracting
heat from the compressing gas, through the piston head, to a remote
heat sink. This is fundamentally different than the use of oil to
cool a piston in an internal combustion engine, where the purpose
of the cooling is to keep the piston itself from getting extremely
hot and melting or otherwise failing as extreme heat is generated
in the chamber from fuel explosions. In that context, any
rudimentary cooling means can be sufficient due to the extreme
temperature gradient between the exploding fuel and the ambient
surroundings. By contrast, in the disclosed compressors, it is
desired to keep the low temperature gas in the chamber from heating
up in the first place, and therefore the cooling fluid is used to
quickly remove any heat that the piston head extracts from the
compressing gas so the piston does not heat up and remains at or
near ambient temperature or otherwise low temperature. In fact,
cooling of a piston in an internal combustion engine to near
ambient temperature is very undesirable, because the piston would
extract too much heat from the chamber (high chamber temperature is
good in the engine context) and would reduce the efficiency of the
engine and constitute over-cooling of the piston.
[0082] Broadly described, exemplary devices for compressing gases
and other compressible fluids can include a non-stationary wall of
a compression chamber that in moving varies the volume of the
compression chamber, causes or corresponds to gas being drawn into
the chamber, causes the gas to be compressed to a smaller volume
and higher pressure, and causes the gas to be expelled from the
chamber at or near the higher pressure, and the non-stationary wall
of the chamber is effectively cooled by occasional, continual, or
continuous contact with a cooling fluid of lower temperature to
which it can lose heat and thereby remain as close as possible to
the ambient temperature, or to be cooled to a temperature below
ambient.
[0083] Fluid cooling of non-stationary compressor chamber walls
applies to all types of compressors, not just piston-type
compressors. Further examples include: (1) The non-stationary
scroll in a scroll-type compressor can be effectively cooled
through contact with oil or another fluid on its non-working side;
(2) The counter-rotating rotors of screw-type compressors and
roots-type blowers can be effectively cooled by pumping water or
another fluid through them; and (3) The diaphragm of a
diaphragm-type compressor can be cooled by contact with a cooled
fluid on its non-working side.
Increased Surface Area on Non-Working Side of Moving Wall
[0084] It can be difficult to extract sufficient heat from a piston
or other moving chamber wall due to the relatively low surface area
on the non-working side of the piston head or other moving chamber
wall. For example, the non-working side of a conventional air
compressor piston is lacking in any cooling fins or other irregular
features that would increase surface area and therefore heat loss
to oil or other cooling fluid that they might be in contact with,
either occasionally, continually or continuously.
[0085] In order to serve a significant role in extracting heat from
the gas during compression, as described herein, the non-stationary
wall(s) of the compression chamber can be maintained at a low
enough temperature (e.g., near ambient) to be effective in
extracting heat from the compressing and exiting gas. In order for
the non-stationary wall(s) to be maintained at a low enough
temperature, in some embodiments, the non-working side of the
non-stationary wall(s) can be structured to provide increased
surface area and to effectively transfer heat to the cooling fluid
that contacts it. For example, the non-working side of a piston
head can include fins, ribs, ridges, grooves, projections, bumps,
recesses, and/or other non-flat surface shapes that increase the
surface area, such as to a surface area substantially greater than
the cross-sectional area of the bore in which the piston head
reciprocates. Such a surface structure can increase the area that
contacts the cooling fluid to increase the amount and rate of heat
transfer from the piston head to the cooling fluid. Examples of
such piston head geometry are shown in the exemplary embodiments
included herein.
[0086] Broadly described, exemplary devices for compressing gases
and other compressible fluids can include a non-stationary (moving)
wall of a compression chamber that in moving varies the volume of
the compression chamber and causes a gas to be drawn into the
chamber and causes the gas to be compressed to a smaller volume and
a higher pressure and to be expelled from the chamber at a higher
pressure, the non-stationary wall of the chamber being designed and
structured to incorporate one or more surfaces (on the non-working
side of the non-stationary wall) that are not in contact with the
compressible fluid being compressed, said surface(s) increasing or
maximizing the area of the non-working side of the non-stationary
(moving) wall, the effect of which is to increase the ability of
the non-stationary wall to lose heat from its non-working side to a
cooling fluid, said surface(s) being in occasional, continual, or
continuous contact with a cooling fluid of lower temperature to
which they can lose heat.
Non-Axial Motion During Axial Compression
[0087] With the example geometry in FIG. 1, the gas flows over the
cooling surfaces to some extent during compression, and much more
during expulsion of the gas from the compression chamber. This is
true of many of the other possible geometries of the chamber, head,
and piston, to a greater or lesser degree. However, with the
example geometry of FIG. 1 and many of the other geometries of the
chamber, head, and piston, during the majority of the compression
of the gas, flow of the gas over the surfaces of the compression
chamber may be limited. Given the insulating properties of stagnant
or slow moving gas, the gas may rise more in temperature than would
be desirable, at least in some areas of the compression chamber. In
this sense, the compression can be partially, locationally,
adiabatic, and this increases the work required to compress the
gas. In the example geometry of FIG. 1, and many of the other
possible geometries, the excess heat generated from this partially
adiabatic compression subsequently is at least partly lost to the
cooling surfaces as the gas exits the compression chamber, because
during that part of the cycle (expulsion) the gas develops
relatively higher velocities and greater turbulence. And, this
later loss of excess heat to the cooling surfaces is helpful in
reducing the total amount of work required to compress the gas and
force it out of the compression chamber. As the gas loses heat to
the cooling surfaces, the pressure of the gas relaxes considerably,
easing the final stages of the compression. Still, even though the
excess heat generated by the partially adiabatic compression is
subsequently partially removed before and during the expulsion of
the gas from the compression chamber, and even though the gas exits
the compression chamber at a temperature much closer to the
temperature of the gas as it entered the compression chamber, the
temperature did rise during compression, and excess work was
therefore done, which decreases the overall efficiency of the
compressor. To minimize adiabatic compression of the gas, greater
flow, and especially more turbulent flow, of the gas over the
cooling surfaces during the compression is advantageous and
desirable.
[0088] Augmented flow, and especially turbulent flow, of the gas
during compression can be achieved, in some embodiments, by
shifting a portion of the gas from one or more areas of the
compression chamber to one or more other areas (or between areas)
of the compression chamber during the compression. That is, flow in
and about the compression chamber that does not involve expulsion
of the gas can increase the transference of heat from the gas to
the cooling surfaces during the compression. In the example of a
piston-type compressor, this can be achieved by manipulating the
head, side walls, and/or the piston to provide a stroke that is
other than the direct, straight-line axial stroke typical of many
piston-type compressors. For the example geometry of FIG. 1, and
many of the other possible geometries, one way to shift the gas in
and about the compression chamber (intra-chamber flow) is to
rotate, shift, or oscillate the head and/or the piston slightly, in
a non-axial direction, during the axial compression motion of the
piston. Such non-axial motion can, for example, be oscillatory
(back and forth) or in a single direction. There are many other
non-axial motions or manipulations of the head and/or piston that
can be advantageous for increasing flow and turbulence in the
chamber, including in other exemplary compressors, such as
wobble-piston compressors and swash plate-driven compressors, both
of which can involve a tipping-back-and-forth motion of the piston
(among other non-axial motions) as it reciprocates axially in the
cylinder(s) of such compressors.
[0089] In some embodiments, the housing head and/or side walls can
be oscillated or otherwise moved during the straight line axial
stroke of the piston, instead of oscillating the piston itself. Or,
both the piston and the housing can be moved, such as in opposite
directions, to create a relative oscillatory motion between the
piston and the housing.
[0090] FIG. 12 shows a rotationally oscillating path of approach
for the example geometry similar to that shown in FIG. 1, in which
one or more of the housing head, side walls, and the piston are
manipulated to create a rotational oscillation about a central
axis. FIGS. 11A-11C shows the non-axial rotation motion in
isolation from the axial, straight-line reciprocating approach of
the head and piston that is typical of many piston-type
compressors. The side-to-side relative motion reduces some volumes
and increases or creates other volumes between the interleaving
projections, causing the gas to flow along the surfaces. The flow
is largely across or parallel to the faces of the cooling surfaces,
which enhances heat transfer into the projections. Turbulence
during such flows can be increased by detailed sculpting of the
surfaces. Undulating surfaces, for example, can be included in some
embodiments to increase turbulence. Such detailed sculpting is
omitted from FIGS. 11 and 12 for clarity. As a cooling surface of
the piston (or head) makes a near approach to its adjacent cooling
surface on the head (or piston), higher velocity flow (sometimes
referred to as "squish") is developed, creating significant and
useful additional turbulence in the gas, some at points distant
from the near-contacting surfaces. Flow rates of up to a few
hundred miles per hour can be achieved, for example, and turbulence
of any practically significant level is readily attainable. And, as
the opposing surfaces then move apart again, a considerable
low-pressure point is created between the separating surfaces,
causing what might be termed "back-squish". This also can be highly
turbulent flow, further enhancing heat transfer to or from the
gas.
[0091] In addition to the exemplary oscillatory path shown in FIG.
12, various alternative paths and orientations can be achieved in
other embodiments, and many would be as effective or more effective
in actual practice.
[0092] During both the inlet phase and during the
compression/exhaust phase, the increased flow and turbulence caused
by the non-axial motion of piston and/or head can help maintain
more even temperatures in the gas and in the walls of the
compression chamber.
[0093] Broadly described, exemplary devices for compressing gases
and other compressible fluids can include one or more compression
chamber walls that move (e.g., oscillate) in any motion that is
different from the normal motion that is characteristic of that
type of compressor (such as moving non-axially during an axial
compression stroke of reciprocating piston-type compressor, or
non-orbital motion during the normal orbiting motion for a
scroll-type compressor, and equivalent non-normal motions for a
blower type compressor, screw-type compressor, wobble-piston
compressor, or a swash plate-driven compressor), an effect of which
non-normal movement is to create or increase intra-chamber flow of
the compressing gas and/or increase turbulence in the gas.
Exemplary Embodiments
[0094] FIGS. 1A and 1B show an exemplary reciprocating piston-type
compressor 10 having many of the features disclosed herein. The
compressor 10 generally includes a stationary upper housing 12, a
reciprocating piston 20 that forms a sealed compression chamber
along with the housing, and a lower casing 13 that can comprise a
crank case for driving the piston and a cooling fluid reservoir.
This disclosure encompasses compressor embodiments having any
number of such compression chambers.
[0095] The housing 12 includes an upper housing head 14, side walls
18, and other stationary components. A piston 20 reciprocates
within the housing side walls 18 using annular seals 22 (e.g.,
O-rings) to create a sealed internal compression chamber 16. The
side walls 18 and piston 20 can have any cross-sectional bore
profile, such as cylindrical or otherwise. The material of the
housing 12 and the piston 20 can comprise any sufficiently strong,
thermally conductive material (e.g., any of various metals) that
facilitate heat conduction away from the gas in the compression
chamber. The stationary components, including the head 14 and side
walls 18, can be one-piece/integral with each other or can be
comprised of components that that are contiguous and in
sufficiently close contact with each other to transmit heat across
the interfaces without substantial thermal insulation therebetween.
The stationary components can, for example, be glued, soldered,
welder, brazed, or bonded together.
[0096] The stationary head 14 includes projections or protuberances
24 that extend into the compression chamber 16 and the piston 20
includes corresponding opposing projections or protuberances 34
that extend into the compression chamber. The projections 24 and 34
partially divide the compression chamber 16 into smaller spaces,
such that a substantial fraction of the gas in the chamber is at
all times relatively close, compared to an open compression
chamber, to one or more thermally conductive surfaces to which the
compressing gas can give up heat. The projections 24 and 34 can
have various geometries, with one example shown in FIG. 1. A
further example of such geometry is shown in FIGS. 4-9.
[0097] As illustrated in FIG. 1, the opposing projections 24 and 34
can be shaped to interleave or mesh with each other to avoid
interfering with each other at the top of the compression stroke,
as the piston pushes the lower projections 34 up into and between
the upper projections 24. At the extremity of the compression
stroke (top dead center), the lower projections 34 can approach the
upper projections 34 so as to define a minimum volume in the
compression chamber that is a small fraction of the working volume
(swept volume) of the compressor, thus resulting in the compressor
having a relatively small clearance volume and high compression
ratio. The lower projections 34 may or may not actually contact the
upper projections 24 at the top of the stroke.
[0098] The projections 24 can be integral with the rest of the head
14 and the lateral/radial surfaces of the projections 24 can be
integral with or in thermally conductive contiguous contact with
the side walls 18, since the projections 24, the rest of the head
14, and the side walls 18 can all be (not necessarily) stationary.
This allows heat from the projections 24 to be conducted both
axially up through the head 14 (as illustrated by arrows 26) and
laterally/radially through the side walls 18 (as illustrated by
arrows 28). The lower axial ends of the projections 34 can be
integral with the rest of the piston 20 while the lateral/radial
surfaces of the projections 34 can be in sliding contact or
near-contact with the stationary side walls 18, such as with a
lubricant therebetween. This allows heat from the projections 34 to
be conducted both axially down through the piston 20 (as
illustrated by arrows 36) and laterally/radially through the side
walls 18 and sealing rings 22.
[0099] As shown, the projections 24 and 34 can be gradually tapered
and can have a truncated flat distal end, such as with slight
rounded edges, or can have smoothly rounded distal end. The
projections can be, though not necessarily, arrayed angularly
around a central vertical axis, with each projection extending
radially away from the central axis in a generally wedge or
trapezoid shape. The radially outward facing surfaces can be
rounded to match the side walls 18 of the housing, thereby
facilitating heat flow in the radial direction through the
projections and into the surrounding side walls. The projections
can have broad, flat angularly-facing sides that form valleys
between them, the valleys having the same or similar size and shape
as the projections, but axially flipped, such that the valleys can
fittingly receive the opposing interleaving projections at the top
of the compression stroke. The projections 24 of the housing head
14 can angularly surround an open central bore that couples the
intake 50 and exhaust 52 with the compression chamber 16. Each of
the open valleys between the projections 24 can extend radially
from the central bore. The lower projections 34 of the piston head
can also include/surround a corresponding central open area. The
geometry of the projections 24 and 34 ensures that the gas in the
compression chamber 16 is always close to at least one surface that
is ready to accept heat from the gas as it is compressed.
[0100] The compressor 10 includes a gas intake 50 and exhaust 52
that are fluid coupled to the compressor chamber via at least one
valve 48. The intake and exhaust can flow through a common upper
central passage where the valve 48 is located. The valve can be
mounted in the housing head 14 and/or in another component of the
housing 12.
[0101] The stationary housing 12 can include a cooling jacket 30,
external fins, and/or other features that accept heat from the head
14, side walls 18, projections, and other components and conduct
the heat away from the compressor. In the illustrated embodiment,
the cooling jacket 30 (which can utilize any cooling fluid) extends
around a substantial portion of the side walls 18 and above the
head 14. The cooling jacket 30 includes a lower outlet 62 coupled
to a heat sink 32 via conduit 54, and an upper inlet 64 coupled to
the heat sink via conduit 56. The cooling fluid cycles through the
cooling jacket 30 and the heat sink 32 to continuously remove heat
from the compressor. The heat sink 32 can comprise any
configuration sufficient to cool the cooling fluid below the
temperature of the housing 12 so that the fluid can effectively
extract heat from the housing. Fins, fans, and/or other means can
also be included with the housing 12 to help remove heat from the
outer surfaces of the housing walls.
[0102] The motion of the reciprocating piston 20 can be driven by a
crank shaft located in the casing 13 and coupled to the piston head
with a piston shaft 60. The piston shaft 60 is pivotably joined to
the head of the piston 20 such that the head of the piston moves up
and down axially while the shaft 60 moves in a more complex
motion.
[0103] The piston 20 has a working side that includes the
projections 34 and which forms surfaces of the compression chamber,
and a non-working side facing away from the compression chamber.
Heat generally flows from the working side through the piston head
to the non-working side, and the heat is then conducted away from
the non-working side to a heat sink. The non-working side of the
head of the piston 20 includes features that increase its surface
area, such as fins 40, to enhance heat conduction away from the
piston to a cooling fluid, such as air and/or a cooling fluid. The
non-working side can also or alternatively include other surface
features, such as ribs, grooves, projections, bumps, etc., that
increase the surface area and/or facilitate the flow of a cooling
fluid over the surface to effect heat extraction from the
piston.
[0104] The compressor 10 can include a piston cooling system that
utilizes a cooling fluid that is cyclically conducted into contact
with the non-working side of the piston 20 and then conducted to a
heat sink. The piston cooling system can include a cooling fluid
reservoir 42, such as located within the casing 13, a pump 58 to
conduct the cooling fluid through a conduit and cause the cooling
fluid to be sprayed from an nozzle 44 in the direction of the
non-working side of the piston 20 and/or the piston shaft 60, and
an external heat sink 46 coupled to the reservoir 42 and pump 58.
Alternatively, the cooling fluid can be circulated through a
portion of the head and/or lateral walls of the compressor and
thereby lose heat to the cooling mechanism(s) provided for the head
and/or lateral walls. This can eliminate the need for a separate
heat sink for the piston cooling fluid. The dashed lines in FIGS.
1A, 1B, 2A, and 2B emanating from the nozzles/sprayers 44 and 144
illustrate exemplary spray paths for the cooling fluid, however
many other pray patterns can be achieved, causing the fluid to
contact any combination of the piston surfaces. The cooling fluid
can be sprayed in such a way to effect substantial contact of the
fluid with the large surface area of the non-working side of the
piston. For example, if fins 40 are present, the cooling fluid can
be sprayed onto the fins and into the valleys between the fins to
cover as much of the surface area of the fins as possible. The
cooling fluid can be allowed to run via gravity along the surfaces
of the piston and/or fall down toward the reservoir, taking with it
heat from the piston. The relatively hotter cooling fluid is then
pumped through an external heat sink 46, or through a portion of
the head and/or walls, to be cooled down and then pumped back
through the cycle to be sprayed at the piston again to help
maintain the piston at or near ambient temperature. In some
embodiments, more than one spray nozzle or other spraying device
can be used to maximize the volume and coverage of the fluid in
contact with the piston at any given time.
[0105] In some embodiments, one or more of the heat sinks used to
remove heat from cooling fluids (e.g., heat sink 32 and/or heat
sink 46) can comprise a well or other subterranean cavity
containing water that is below the ambient air temperature, or can
comprise a fluid cooled to below ambient air temperature of the air
surrounding the compressor by colder outdoor air during colder
times of the year.
[0106] FIGS. 2A and 2B show another exemplary reciprocating
piston-type compressor 100 that is similar to the compressor 10,
but also includes means for causing the reciprocating piston to
oscillate non-axially (e.g., rotationally about its central axis)
while reciprocating along the central axis as usual. The compressor
100 generally includes a stationary upper housing 112, a
reciprocating piston 120 that forms a sealed compression chamber
along with the housing, and a lower casing 113 that can comprise a
crank case for driving the piston and optionally a cooling fluid
reservoir.
[0107] The housing 112 includes an upper housing head 114, side
walls 118, and other stationary components. A piston 120
reciprocates within the housing side walls 118 using annular seals
122 (e.g., O-rings) to create a sealed internal compression chamber
116. The side walls 118 and piston 120 have a cylindrical
cross-sectional bore profile to allow for rotational oscillations
of the piston, though in alternative embodiments the oscillatory
motion of the piston can be other that rotational and thus the
cross-sectional bore profile can be other than cylindrical.
[0108] The stationary head 114 includes projections or
protuberances 124 that extend axially into the compression chamber
116 and the piston 120 includes corresponding projections or
protuberances 134 that extend axially into the compression chamber.
The projections 124 and 134 partially divide the compression
chamber 116 into smaller spaces, such that a substantial fraction
of the gas in the chamber is at all times relatively close,
compared to an open compression chamber, to one or more thermally
conductive surfaces to which the compressing gas can give up heat.
The projections 124 and 134 can have various geometries, with one
example shown in FIG. 1 described elsewhere herein.
[0109] The compressor 110 includes a gas intake 150 and exhaust 152
that are fluid coupled to the compressor chamber via at least one
valve 148. The intake and exhaust can flow through a common upper
central passage where the valve 148 is located. The valve can be
mounted in the housing head 114 and/or in another component of the
housing 112.
[0110] The stationary housing 112 can include a cooling jacket 130,
external fins, and/or other features that accept heat from the head
114, side walls 118, projections, and other components and conduct
the heat away from the compressor. The cooling jacket 130 includes
a lower outlet 162 coupled to a heat sink 132 via conduit 154, and
an upper inlet 164 coupled to the heat sink via conduit 156. The
cooling fluid cycles through the cooling jacket 130 and the heat
sink 132 to continuously remove heat from the compressor. The heat
sink 132 can comprise any configuration sufficient to cool the
cooling fluid below the temperature of the housing 112 so that the
fluid can effectively extract heat from the housing. Fins, fans,
and/or other means can also be included with the housing 112 to
help remove heat from the outer surfaces of the housing walls.
[0111] The motion of the reciprocating piston 120 can be driven by
a crank shaft located in the lower casing 113 and coupled to the
piston head with a piston shaft 160. The motion of the piston head
is more complex than simple axial reciprocation along the central
axis of the cylindrical housing bore. For example, the piston head
can move rotationally, or angularly, within the bore during the
axial compression stroke. In some embodiments, the piston can tilt
or wobble, such as about an axis perpendicular to and/or
intersecting the reciprocation axis) during reciprocation. The
effect of such complex motion is to increase flow and turbulence of
the gas in the compression chamber, especially during compression
when heat is being generated, as is described in more detail
elsewhere herein.
[0112] In some embodiments, the piston can complete at least one
full oscillation during each compression stroke. A full oscillation
includes motion in one rotational direction followed by motion in
the opposite rotational direction such that the piston returns to
its original rotational position. The piston may complete more than
one oscillation during the stroke in some embodiments, such as at
least two oscillations. In some embodiments, the piston may
complete a half oscillation during the compression stroke, and
complete the rest of the oscillation during the down stroke, such
that one full oscillation is completed in one full
reciprocation.
[0113] An exemplary axial and rotational path 170 of the distal tip
of one of the lower projections 134 is shown in dashed lines in
FIGS. 2A and 2B. In this example, one full oscillation is completed
per reciprocation cycle of the compressor. This exemplary path
includes a portion between point 172 at the bottom of the stroke
and point 174 part way through the compression stroke, wherein the
projections 124 and 134 approach each other and create
intra-chamber flow of the gas from a shrinking space on one side of
the projections to a space on the other side of the projections,
thus creating increased flow velocity and turbulence. As the
opposing surfaces approach at point 174 without touching, "squish"
occurs and causes rapid flow of the gas being squeezed out from
between the approaching surfaces, creating higher velocity and
increased turbulence. The path can then include a portion between
points 174 and 176 where the opposing projection surfaces, having
approached one another closely at 174, now separate to a greater
distance, resulting in "back-squish," wherein the gas rapidly flows
back into the low pressure void growing between the surfaces. This
again generates increased gas flow and turbulence. The path can
further include a portion between points 176 and 178 where the
projections 134 are aligned in the valleys evenly between the
projections 124 such that the piston moves in a substantially
straight line motion (e.g., only axially) to allow tight
intermeshing of the opposing projections without contact between
them. This motion is reversed during the down stroke.
[0114] An alternative path 300 is illustrated in FIG. 12, wherein a
piston completes more than one complete oscillation during the
compression stroke. The path 300 shown in FIG. 12 is not
necessarily due to rotational motion of the piston, as other types
of non-axial motion (e.g., in non-cylindrical bore shapes) can
result in the path shown in FIG. 12 as well. As shown in FIG. 12,
the lower projection 302 moves right, then left, then right and
finally back to a central aligned position at top dead center. This
motion may or may not be reversed during the down stroke.
[0115] FIGS. 11A-11C illustrate the induced flow of the gas caused
by the non-axial relative motion between the opposing projections
of the piston and the head. The illustrations of FIG. 11A-11C
ignore the actual axial motion occurring simultaneously between the
opposing projections in order to more clearly describe the effects
of the non-axial motion. In FIG. 11A, the upper projections 300 are
shifted to the right relative to the lower projections 302, causing
opposing surfaces 304 and 306 to come into close proximity and
causing gaps 312 between opposing surfaces 308 and 310. In FIG.
11B, the top projections have shifted partially to the left
relative to the lower projections 302, causing the gaps 312 to be
reduced and opening gaps 320 between the surfaces 304 and 306. This
causes the gas to flow from the gaps 312, around the valleys 314
and 316, into the gaps 320. In FIG. 11C, the gaps 312 are
substantially closed and the gaps 320 are increased as the gas
continues to flow around the valleys 314, 316 and into the gaps
320. This flow of gas can be at high speed due to the "squish"
effect and therefore create desirable and substantial turbulence.
The flow is also primarily directed parallel to and close to the
surfaces of the projections, which maximizes heat transfer between
the walls and the gas.
[0116] Referring again to FIG. 2A, an exemplary mechanical means
for causing the illustrated path 170 is shown. The inner walls of
the housing 12 can include one or more grooves 180, and the outer
wall of the piston 20 can include one or more corresponding pins
182 that extend radially into the grooves and move along the
grooves as the piston moves axially. Or, this can be reversed, with
the piston having the pins and the cylinder walls having the
grooves. The shape of the grooves 180 defines the non-axial motion
of the piston 20 and the corresponding path of the lower
projections 134. In the example of the compressor 100, each groove
180 has a shape that matches the shape of the path 170.
[0117] The piston shaft or connecting rod 160 can include a thrust
bearing 184 that allows the piston head freedom to move in
oscillatory motions in addition to its normal reciprocation motion
while the lower portion of the piston shaft 160 moves in its normal
elliptical/reciprocating motion.
[0118] The non-working side of the head of the piston 120 can
include features that increase its surface area, such as fins 140,
to enhance heat conduction away from the piston to a cooling fluid,
such as air and/or a cooling liquid. The non-working side can also
or alternatively include other surface features, such as ribs,
grooves, projections, bumps, etc., that increase the surface area
and/or facilitate the flow of a cooling fluid over the surface to
effect heat extraction from the piston.
[0119] Similar to the compressor 10, the compressor 100 can also
include a piston cooling system that utilizes a cooling fluid that
is cyclically conducted into contact with the non-working side of
the piston 120 and then conducted to a heat sink. The piston
cooling system can include a cooling fluid reservoir 142, such as
located within the lower casing 113, a pump 158, one or more
nozzles 144 to spray the fluid in the direction of the non-working
side of the piston 120 and/or the piston shaft 160, and an external
heat sink 146. In some embodiments, more than one spray nozzle or
other spraying device can be used.
[0120] FIGS. 3-10 show various views of an exemplary compressor 200
that includes many of the features disclosed herein. FIG. 3 shows
an assembled view and FIG. 4 shows an exploded view. The compressor
200 includes a base 202, main body 204, outer sheath 206, inner
liner 208, piston 210, head 212, valve housing 214, and/or various
other components. All of the illustrated components form a
stationary housing, except for the piston 210, which reciprocates
within the housing and can be driven by any suitable power source
(not shown). The piston 210 and head 212 include opposing,
interleaving projections similar to those described with
compressors 10 and 100.
[0121] The compressor 200 can include an intake/exhaust port 216 at
the center top of the head 212 and valve housing 214, which can
include a valve to selectively allow gas intake or gas exhaust.
[0122] The compressor 200 can also include a drive mechanism (not
shown) for causing the piston reciprocation, and can include a
lower casing that houses the drive mechanism and/or includes a
fluid cooling system (e.g., a fluid reservoir and spraying
mechanism for cooling the non-working side of the piston, as
described elsewhere herein).
[0123] The housing can include a fluid cooling jacket in a space
formed in the housing between the outer surface of main body 204
and the outer sheath 206, as previously described in other
embodiments, and the cooling jacket can extend over the top of the
compression chamber in a space within the head 212 or in a space
between the head and the valve housing 214. The cooling jacket can
be fluidly coupled to an external heat sink through one or more
inlets and outlets, such as via port 218 (FIG. 3) in the base 202.
The main body 204 can be secured to the base 202 such that the
stepped lower end 218 of the main body seals against the lower
surface 220 of the base, while the outer sheath 206 secures against
the top of the base, creating a sealed internal void for cooling
fluid to flow from the cooling jacket to the port 218.
[0124] The entire structure shown in FIG. 3 can be mounted to a
crank case or other stationary lower structure that contains a
power means for driving the piston 210 and/or a fluid cooling means
for cooling the non-working side of the piston 210, such as is
described herein with reference to compressors 10 and 100.
[0125] The inner liner 208 is optional and can be mounted inside
the main body 204 to provide a cylindrical surface across which the
piston slides, such as via one or more O-rings in a sealed manner.
The inner liner 208 can comprise a different material than the main
body 204, and/or can be replaceable when worn.
[0126] FIG. 5 shows the piston 210 and the head 212 fully mated
together, illustrating the corresponding but opposite geometry of
the opposing projections 228 and 230. Each projection can fit very
tightly in the valley between two of the opposing projections, such
that the volume of the compression chamber formed therebetween is
minimal at top dead center and the compression ratio and swept
volume are increased. Flow and turbulence of the compressed gas is
also increased with such geometry, especially at the later stages
of compression.
[0127] FIGS. 6-8 show the head 212 in various views, and FIGS. 9-10
show the piston in various views. The head includes tapered
projections 230 arrayed angularly around a central opening 236,
with evenly sized valleys 231 formed between the projections.
Similarly, the working side of the piston 210 includes projections
228 arrayed around a central opening 252 and sized to fit in the
valleys 231 of the head, and valleys 229 sized to receive the
projections 230 of the head. The projections of the piston and the
head can optionally include notched regions (for example as shown
at 254 in FIG. 9) at their radially inner ends near the
intake/exhaust port 216, such as to provide increased space for
intake and exhaust flow.
[0128] The head 212 can also include an upper body that includes
two concentric rims 232 and 234. The inner rim 232 can be secured
to and contiguous with the top end of the main body 204 while the
outer rim 234 can be secured to the top end of the outer sheath
206, allowing the cooling jacket to extend into the upper body of
the head 212 and or between the head and the valve housing 214.
[0129] As shown in FIGS. 5 and 9, the piston can include one or
more recesses 226 in its outer surface to accommodate O-rings or
other sealing members that provide for sealed sliding against the
inner liner 208. The piston 210 can also include openings 250,
reinforced wall structures 264, and/or other features that
facilitate the jointed attachment of a piston rod or other
mechanical linkage.
[0130] As shown in FIG. 10, the lower, non-working side of the
piston 210 can include a recessed volume, generally referred to as
260, that includes one or more fins 262 that project downward from
the otherwise flatter surface 266 on the non-working side of the
piston. The illustrated embodiment includes five fins 262 having
straight and curved shapes, and formed valleys between them, which
substantially increase the surface area on the non-working side of
the piston and enhance heat conduction through and away from the
piston. As discussed elsewhere herein, a fluid cooling system can
be used to spray or otherwise conduct cooling fluid into contact
with the non-working side of the piston 200 to assist in removing
heat from the piston. The increased surface area on the non-working
side of the piston and/or such a fluid cooling system for the
piston, can help keep the piston at or near ambient air
temperature, so that the piston can actively conduct heat away from
the compressed gas before the temperature builds up in the gas.
[0131] FIGS. 13-15 illustrate another exemplary compressor 400 that
embodies many of the features disclosed herein. The compressor 400
is a wobble plate compressor that includes a plurality of
reciprocating piston-based compression chambers, such as chambers
402 and 404 (only two are shown for clarity), which embody many of
the principles that are described in connection with other
reciprocating piston-type compressors disclosed herein, along with
other features.
[0132] FIG. 13 shows a partially cross-sectional view of the wobble
plate compressor 400, with certain features not shown in full
cross-section. For example, the opposing projections 420 and 422 in
the compression chambers are shown in a view that illustrates the
interleaving structure and orientation of the opposed projections,
similar to the illustrations in FIGS. 1A, 1B, 2A, and 2B, all of
which can represent various geometries of projections, including
the geometries shown in FIG. 1 and in FIGS. 4-9.
[0133] The disclosed technology can also be applied in swash-plate
compressor embodiments in similar manner. In a wobble plate
compressor, as shown in FIG. 13, a tilted drive plate 406 (as
opposed to a crank shaft in a reciprocating compressor) is rotated
by the shaft 408 of the compressor. In this manner, the drive plate
processes about its center line 410, and in this manner forces the
pistons 412, through the connecting rods 414, to reciprocate in the
cylinders 416. Thus, a wobble-plate compressor (as well as a
swash-plate compressor) is a species of the reciprocating
compressor genus, and can include any combination of the features
disclosed herein that are associated with a reciprocating
compressor.
[0134] In a conventional wobble-plate compressor, the pistons and
the heads are flat, and the compression chambers are simple
cylinders, thus minimizing the surface areas of the compression
chamber walls, and reducing and perhaps minimizing turbulence in
the cylinders. For clarity, only two pistons are shown. However,
the disclosed wobble-plate compressors can have more pistons,
including odd numbers of pistons. Three, five, and seven piston
wobble plate compressors are exemplary configurations that can
include the disclosed technologies. The disclosed wobble-plate
compressors can be used, for example, in automotive air
conditioning systems, refrigerators, and air conditioning systems
for buildings.
[0135] In FIG. 13, the compressor 400 includes contoured surfaces
that project into the compression chamber, including projections
420 that protrude from the piston and opposing offset projections
422 that protrude from the stationary end of the chamber. As with
other reciprocating compressors disclosed herein, many other
geometries are possible, and many of those other geometries may be
more or less efficient or otherwise advantageous or disadvantageous
over the geometry shown in FIG. 1 and FIG. 13. All possible
geometries that increase the surface area of the stationary and/or
non-stationary walls of the compression chambers are envisioned and
encompassed by this disclosure.
[0136] While not shown in FIG. 13, in some embodiments, a wobble
plate compressor can further include a mechanism that causes the
pistons to move or oscillate non-axially as the piston reciprocates
axially, as is disclosed in reference to the compressor 100 and in
other places herein. Any of the non-axial piston motions, and
related mechanisms for causing such motions, described herein can
also be incorporated in the modified versions of the compressor 400
or other wobble plate compressors.
[0137] As in other types of compressors, there are other ways
(besides manipulating the stationary and/or non-stationary walls of
the compression chambers) to increase the turbulence in the
compression chambers of a wobble-plate compressor, in addition to
increasing surface area within the compression chamber. Sculpting
of the cooling surfaces in the chambers, for example, can increase
turbulence locally. The opposing compression chamber surfaces can,
for examples, be undulating, furrowed, dimpled, rough, polished,
and/or have other characteristics.
[0138] As with other reciprocating compressors, any and all of the
walls of the compression chambers of a wobble-plate compressor
(either or both stationary and non-stationary) can serve as heat
sinks to absorb and draw away the heat of compression. It is not
necessary to use all three surface regions (the head, the cylinder
side walls, and the piston) in this way, however. Significant
improvement in efficiency can be gained by cooling only the
stationary walls (the head and the side walls of the cylinder). In
addition, cooling of both the stationary and non-stationary walls
is also possible, and both can increase the efficiency of a
wobble-plate compressor.
[0139] The head, the piston, and the cylinders side walls can be
provided with significant externally applied cooling, to draw away
the heat of the compression and to keep the compression chamber
walls from heating up. The non-moving compression chamber walls
(the head and cylinder side walls) can be air-cooled by cooling
fins and optionally a fan on its exterior, as shown in FIG. 13.
They can also be cooled by a cooling fluid, such as a water
jacket.
[0140] In the example embodiment depicted in FIG. 13, the cooling
of the compressor head and cylinders (and, indirectly, the cooling
of the pistons) is accomplished through fan-cooled external cooling
fins, as opposed to incorporation of a water jacket(s) or other
means. The example external cooling fins are shown at the points
labeled 424. The fan 426 draws in ambient air, or air that has been
conducted from a cooler location, and blows it first across the
radial external cooling fins that occupy the center portion of the
head at the point labeled 428. The air then courses over a manifold
including several inlet ports 430 and a similar manifold including
several exhaust (outlet) ports 432. Such manifolding is not
depicted in FIG. 13. The air then blows across the radial cooling
fins on the outer portion of the head, at the points labeled 434,
before turning approximately 90 degrees to travel in the path
labeled 436 between the radially arrayed external cooling fins
along case 438 (the sides) of the compressor.
[0141] Internally, in this example embodiment, internal cooling
fins 440 are shown on the inside surface(s) of the case of the
compressor. These, or otherwise-shaped, cooling surfaces afford
greater surface area for absorbing heat from the oil (or other
fluid), helping to keep the oil as close to ambient temperature as
feasible. The heat absorbed from the oil is conducted to the outer
surface of the compressor case and is partially conducted to the
cooling surfaces (fins in the case of this example embodiment)
thereon. FIG. 14 shows a possible arrangement of the external
cooling fins 462 on the outside surface of the case 460 of the
compressor and the internal cooling fins 464 on the inside surface
of the case of the compressor. This exemplary arrangement is
optional in the functioning of the compressor, but can increase the
efficiency of removing heat from the case and the oil and component
parts therein over other arrangements which do not align the
internal and external cooling fins for improved heat
conduction.
[0142] The non-stationary compression chamber walls (e.g., the
pistons 412) can be cooled by being in occasional, continual, or
continuous contact with a cooling fluid (e.g., liquid, vapor, or
gas) of lower temperature, to which they can lose heat. The
non-working sides of the pistons can be in contact with oil or
other fluid that provides lubrication, in some embodiments. Such
oil, and/or other fluids, can also be utilized as a heat sink, to
extract significant heat from the pistons so that they can in turn
serve in their role as heat sinks, to absorb and draw away some of
the heat of compression that is produced in this, and all other
types of gas compressors. FIG. 13 shows one possible approach as an
example: spraying a cooling fluid 442 at the non-working sides of
the pistons. In this example embodiment, use of spray nozzles or
oil jets 444 to spray oil at the non-working sides of the pistons
is shown. This use of fluid cooling of the non-stationary
compression chamber walls of a wobble-plate compressor is another
example of the general concept of cooling the non-working side of a
moving wall of the compression chamber in order to use that wall to
more efficiently extract heat from the compressing gas. FIG. 13
does not show either oil pump(s) or conducting tubes or other means
to convey oil to the oil jets 444, though this can be accomplished
in various ways.
[0143] As with the other reciprocating-type compressors discussed
herein, the non-working sides of the non-stationary walls of the
compression chambers (e.g., the pistons) can be shaped to increase
the transfer of heat to the cooling fluid (e.g., liquid, vapor, or
gas) to help to keep the pistons as close to ambient temperature as
feasible. This shaping can, for example, include cooling fins, as
shown in FIG. 13 at the points labeled 446. Many other geometries
are possible for the non-working sides of the non-stationary walls
of the compression chambers. Shaping of the non-working sides of
the non-stationary walls of the compression chambers of a
wobble-plate compressor to increase the surface area, or to
otherwise increase the ability of the pistons to transfer heat to
the cooling fluid (liquid, vapor, or gas), is another example of
the concept of including increased surface area on the non-working
side of a non-stationary wall of the compression chamber to
increase heat conduction through and away from the non-stationary
wall.
[0144] As shown in FIGS. 15A and 15B, gas can enter and exit the
compression chambers through individual valve bodies 470 that
function as both inlet and exhaust (outlet) valves 450 and 452. The
valve 470 at configuration 450 is in the exhaust position, and the
valve at 452 is in the intake position. The valves 470 can include
spring mechanisms 472 that cause them the open and close in the
proper functioning manner based on pressure differential on either
side of the valve. The dimensions, including the relative
proportions of the valve dimensions, as shown in FIG. 13, are
exemplary for illustration and not necessarily optimized for
efficient operation. The schematic representations of the valves
shown in FIG. 13 depict the concept of concentric intake and
exhaust (outlet) that can occupy one location as just one
illustrative example.
[0145] FIGS. 16-18 illustrate embodiments of a scroll-type
compressor 500 that include many of the features disclosed herein.
The scroll-type compressor 500 can include two interlaid spiral
plates (e.g., Archimedes spirals), with one of the spiral plates
502 serving as a stator and the other 504 serving as a rotor. Each
spiral plate includes a flat disk portion with a spiral wall
extending perpendicularly from one side of the disk portion, such
that the spiral walls project towards each other and interleave, as
shown in FIG. 16. The two disk portions are substantially parallel
and form top and bottom walls of the compression chamber. The top
and bottom disk portions are not shown in FIG. 16, only the spiral
walls. The rotor 504 orbits about the center axis of the stationary
stator 502. During operation, the two scrolls define a progression
of compression chambers that form between the spirals of the stator
and rotor, spiral inward, toward the centers of both the stator and
the rotor, with decreasing volume, thus compressing the gas between
them, and disappear at reaching the center of the compressor, at
which point the compressed gas is expelled through an exhaust
valve.
[0146] In FIGS. 17A and 17B, the basic functioning of a
conventional scroll compressor is further illustrated. As the rotor
504 "orbits" the vertical center axis of the stator 502, without
itself rotating, the vertical walls 507 of the rotor 504 oscillate
back and forth between the vertical walls 506 of the stator 502.
The moving vertical wall 507 can approach very close to both of the
stationary walls 506 (e.g., within a few thousandths of an inch)
but not actually touch the stationary walls 506. As in some other
types of compressors (e.g., blowers, screw compressors, etc.), the
non-stationary (i.e., moving) walls of the compression chamber need
not actually contact the stationary walls of the compression
chamber, though such contact is optional in some embodiments. This
clearance between the stator 502 and rotor 504 is beneficial in a
scroll-type compressor because, as the walls of the stator and
rotor come into near proximity to one another, there is a sideways
motion of the two relative to one another that can result in a
scraping action if they touch.
[0147] However, in embodiments wherein the vertical walls 506, 507
of the stator and rotor do not touch, leakage can occur between
them. Leakage between the vertical walls of the stator and rotor,
at the interface labelled 508 (and other analogous interface in
18B) results in significant losses of efficiency. There is also
leakage of the compressed and compressing gas at the interfaces
labeled 510, where the tops of the rotor walls 507 interface with
the bottom of the stator disk, and the bottoms of the stator walls
506 interface with the tops of the rotor disk, which interfaces
extend over the entire spiral scroll length. At all of these
interfaces, the compressed and compressing gas can leak out of the
compression chambers toward lower pressure regions of the
compressor.
[0148] The straight, vertical walls 506, 507 of the rotor and
stator shown in FIGS. 17A and 17B are analogous to the flat piston
and flat head of the conventional reciprocating piston-type
compressor discussed herein. The flat vertical walls 506, 507 offer
the minimum surface area to absorb heat from the compressed and
compressing gas and avoid putting anything in the space reserved
for the gas that is to be compressed.
[0149] As with reciprocating compressor embodiments disclosed
herein, the walls of the compression chambers of a scroll-type
compressor (either or both stationary and non-stationary) can also
serve as heat sinks to absorb and draw away the heat of
compression, heat that is produced in all compression of any
compressible gas, even purely isothermal compression. It is not
necessary to use both the stator and the rotor in this way.
Significant improvement in efficiency can be gained by cooling
those compression chamber walls that are easiest to cool, the
stationary walls of the stator. But, cooling of both the stationary
and non-stationary walls can further increase the efficiency of a
scroll-type compressor.
[0150] FIGS. 18A and 18B shows an exemplary geometry for the scroll
compressor 500, in which the purely vertical walls 506, 507 of the
rotor and stator shown in FIGS. 17A and 17B are replaced with walls
512, 513 having much greater surface areas. As in the reciprocating
compressors discussed herein, the vertical walls 512 of the stator
502 include projections extending perpendicularly from both sides
of the wall, and the vertical walls 513 of the rotor 504 also
includes projections extending perpendicular from both sides of the
wall. Each opposing set of projections can be offset and
correspondingly shaped such that the opposing projections can
closely interleave with each other when the moving wall 513
approaches the stationary wall 512, similar to the reciprocating
compressor embodiments and other compressor embodiments described
herein. The illustrated projections of the walls 512, 513 can
increase the surface area of these walls more that 400% compared to
the flat faces of walls 506, 507. In alternative similar
geometries, the surface can be increased by at least 200%, at least
300%, at least 400%, and/or at least 500%, compared to the flat
faces of walls 506, 507. Many other geometries are possible at the
interfaces between the moving and stationary vertical walls that
increase the surface area of the walls of the compression chambers
and thereby increase heat conduction between the gas in the
chambers to the walls, and all such possible geometries are
encompassed by this disclosure. In addition, the increased width of
the vertical walls 512, 513 compared to the narrow width of the
conventional straight vertical walls 506, 507, as seen at the
points labeled 514, can substantially improve the ability of the
vertical walls 512, 513 to transfer heat vertically to the disk
portions of the stator and rotor that form the upper and lower
horizontal walls of the compression chambers.
[0151] Many other geometries can alternatively be utilized to
achieve the increased heat transfer away from the compressing gas.
For example, some geometries can permit actual contact between the
stator and rotor. Such contact between the rotor and stator offers
the possibility of higher efficiency, by (1) reducing the losses of
the compressed and compressing gas at the (near-contact) points of
nearest approach, and (2) increasing turbulence within the
compression chambers by improving the "squish" that happens at the
points of near-contact or contact. Geometries that involve actual
contact between the rotor and stator walls can include a mechanism
that allows the rotor to rotate back-and-forth, or oscillate
freely, a few degrees on its own axis, while orbiting the other
scroll (the stator) about its axis. This motion can allow for the
sideways (sliding) action at the points of near-contact (or of
actual contact) that can be included in the orbiting motion of the
one scroll (rotor) in the other scroll (stator). This manipulation
of the rotor and/or stator, in ways not customary in the operation
of a conventional scroll-type compressor, to allow actual contact
between the two compression chamber walls, is another example of
the concept described herein in various other examples, wherein the
path of at least one compression chamber wall (rotor, stator, or
both) has been altered from the conventional path to achieve
increased gas flow and turbulence in the compression chamber, and
thereby increased heat transfer within the gas and between the gas
and the chamber walls.
[0152] The increased surface area of the example geometry shown in
FIGS. 18A and 18B (and in many other alternative geometries for the
cooling surfaces) and the shearing action that is created between
these faces by the irregular sideways motion described above,
creates significant turbulence of the gas that is caught between
the rotor and stator and compressed as it moves toward the center
of the spiral (scroll). As in other types of compressors described
herein, there are additional and/or alternative ways to increase
the turbulence in the compression chambers of a scroll-type
compressor. Sculpting of the cooling surfaces, for example, can
increase turbulence locally. Any of the surfaces of the compression
chambers surfaces can, for examples, be undulating, furrowed,
dimpled, rough, or polished.
[0153] In some embodiments, the stator or the rotor or both can be
provided with substantial externally applied cooling, to draw away
the heat of the compression and to keep the stator and/or rotor
from heating up. As with a reciprocating compressor, the non-moving
compression chamber walls (the stator) can be fluid-cooled, such as
air-cooled by cooling fins and/or a fan on its exterior, or by a
circulating cooling fluid, as in a water jacket. FIGS. 18A and 18B
show an exemplary water jacket 516 located in the horizontal disk
portion of the stator 502 to extract heat from the stationary walls
of a scroll-type compressor.
[0154] As in other compressor embodiments disclosed herein, the
non-stationary compression chamber walls (e.g., the rotor) can be
cooled by being in occasional, continual, or continuous contact
with a cooling fluid (liquid, vapor, or gas) of lower temperature,
to which they can lose heat. Analogous to the configuration of many
piston-type compressors, the non-working sides of rotors in many
scroll-type compressors are in contact with oil that provides
lubrication. However, in the disclosed scroll compressors,
lubrication oil and/or other fluids serve as a substantial heat
sink, to extract significant heat from the non-working side of the
rotor so that it can in turn serve in its role as a heat sink, to
absorb and draw away the heat of compression that is produced in
the compression chambers. FIG. 18B shows one possible approach as
an example: spraying a cooling fluid at the non-working side of the
rotor from a nozzle 522. In some embodiments, such cooling fluid
can be circulated through an external heat sink and returned cooled
to be sprayed at the rotor again in a cyclical manner.
[0155] As with reciprocating-type compressors disclosed herein, the
non-working side of the non-stationary walls of the compression
chambers of a scroll-type compressor (e.g., the rotor) can be
shaped to increase surface area and thereby increase the transfer
of heat to the cooling fluid. This shaping can, for example,
include cooling fins 518, as shown in FIGS. 18A and 18B, and/or
other geometries. Shaping of the non-working side of the
non-stationary walls of the compression chambers of a scroll-type
compressor to increase the surface area, or to otherwise increase
the ability of the rotor to transfer heat to the cooling fluid
(liquid, vapor, or gas), is another example of the concept that can
be applied analogously across many different species of compressors
disclosed herein.
[0156] FIGS. 19-21 illustrate an exemplary low-pressure, high-flow,
rotary blower-type compressor 600 (referred to herein simply as a
blower), which incorporates many of the principles disclosed
herein. Application of the disclosed principles in screw-type
compressor embodiments (not illustrated) is analogous and
encompassed by this disclosure. In the blower 600, two rotors 610
are counter-rotated by the shaft 612 (FIG. 19) of the compressor
and two gears 614 (FIG. 19). In this manner, the rotors 610 capture
a quantity of air at the inlet 616 (FIG. 20), and carry it along a
path toward the outlet 618 (FIG. 20) of the blower, without
substantial compression.
[0157] A blower is conventionally thought of as simply taking in
air and conveying it to the outlet of the blower, and the
compression that is achieved is conceptually thought of as
happening outside of the compressor; the process can be described
as involving "external compression," as opposed to the "internal
compression" that is characteristic of many other compressors. Even
similar screw-type compressors can achieve compression within the
compressor, and can be described as an "internal compression" type
of compressor.
[0158] However, this conventional concept of a blower is not
correct: the compression that is achieved by a blower occurs,
though rather indirectly, within the blower. The compression occurs
in what would otherwise be described as a compression chamber, even
though the chamber does not compress the gas through operation of a
decreasing compression chamber volume, as do many other types of
compressors.
[0159] The chamber that captures a quantity of the gas to be
compressed (and "blown") can be seen at 620 in FIG. 20. In the
chambers 620, a quantity of gas is carried along an angular path
around the shaft 612, without compression. For example, in FIG. 20,
the left-hand rotor 610 rotates counter-clockwise and the
right-hand rotor rotates clockwise, such that the chambers 620
moves outward and down around the peripheral sides of the blower
toward the outlet 618. At the point that a chamber 620 opens to the
outlet, as shown at 624, it becomes a compression chamber, as it is
connected to the outlet chamber and to the closing chamber of the
opposing rotor, both of which are filled with higher pressure gas.
The higher pressure is back-pressure, caused by a resistance
downstream, that is, the load against which the blower is working.
The higher pressure gas rushes into the "compression chamber" 620
when it joins at 624. In doing so, it does work against the gas
that has been conveyed, in theory at essentially ambient pressure,
to this point. The new gas is compressed as it rises to the
pressure of the outlet; the outlet pressure falls momentarily, as
previously compressed gas flows backward, into the "compression
chamber" that has just opened. Thus, the compression actually
happens inside the blower; it is not actually an "external
compression" compressor.
[0160] Unlike other compressors that achieve compression of the gas
by trapping it in a chamber of varying volume and compress the gas
by reducing the volume of the compression chamber, in a blower, the
equivalent of the non-stationary walls in other compressors, is a
moving, higher-pressure gas, which effects the compression. Though
it is not an accurate literal description, in a blower the
"non-stationary walls" can be thought of as a pressure wave that
compresses the newly admitted gas to a fraction of its original
volume at 620. In reality, the newly admitted gas and the
higher-pressure gas mix freely under the considerable turbulence
that occurs when the chamber opens to the outlet. Like other
conventional compressors, the compression-effecting feature in a
conventional blower tends to be hotter than the new gas was when it
was admitted to the inlet of the compressor at 616. Being hotter,
the already compressed gas heats the newly admitted gas directly,
while it is also being heated by the work of compression. This
unfortunate feature of a conventional blower, as in other types of
compressors, results in much of the inefficiency of a conventional
blower.
[0161] The gas is also heated during the entire trip from the inlet
616 to the moment of opening of the chamber 620 to the outlet 618.
This results from the general high temperature of the conventional
blower, which heats up during operation, such as to temperatures
higher than an internal combustion engine is allowed to reach.
Thus, in a conventional blower, the gas is heated at essentially
every moment between inlet and full compression.
[0162] If any of the direct heating of the gas in a blower can be
avoided, that can improve the efficiency of the blower. If, in
addition, the temperature rise that results from compressing the
gas can be reduced by extraction of heat from the gas during the
compression, efficiency can be further improved.
[0163] FIGS. 19-21 illustrate the application of the principles
disclosed herein to a rotary-lobe blower. The blower shown has a
straight-lobe design. The principles disclosed herein also apply
equally to helical-lobe blowers.
[0164] The case of the blower can be fluid-cooled using a cooling
jacket 626 in FIG. 20 (e.g., a water jacket). The end plates of the
compressor can also be fluid-cooled, as shown at 630 in FIG. 19,
such as by a portion of the cooling jacket that extends around the
end plates. As in any application of the disclosed principles to
any type of compressor, it can be valuable to cool any or all of
the walls of the compressor, both the stationary and the
non-stationary walls. In the blower 600, the non-stationary walls
are the rotary lobes of rotors 610. In this embodiment, the rotary
lobes can be hollow and can be cooled by flowing a cooling fluid
through them, as shown at 632 in FIG. 19. The connection of the
spinning shafts 612 to the coolant flow is accomplished through
rotary unions 634. Within the lobes, the flow path of the coolant
can be directed to be more general and to more effectively cool the
surfaces of the lobes. In this embodiment, internal ducting of the
coolant is shown at 640; while other means to direct the flow
within the lobes can be included in other embodiments. A port 642
can be included to admit any gathered gas (air or other) within the
lobes: a gas, being lower density than the cooling fluid, can
migrate to the center and can escape into the shaft, from where it
can exit the blower.
[0165] In the thoroughly cooled blower 600, rather than heating the
gas during both its trip through the blower and during compression,
the walls of the blower's chambers extract heat from the gas. The
effectiveness of this cooling can be improved by addition of
cooling fins, as shown in example at 644, or by other surface
features that increase the surface areas of the lobes and the
blower case. The example geometry shown in FIG. 21, which include
close intermeshing between projections of the rotating lobes 660
and offset opposing projections of the stationary case 662, can
increase the surface area of the lobes and the case by more than
700%. In other exemplary geometries, the increase in surface area
at the lobe-case interface can be greater than 200%, greater than
300%, greater than 400%, greater than 500%, and/or greater than
600%. A limiting factor in adding surface area to the lobes and
walls is that it may increase leakage of higher pressure gas
backwards through the gap that is required between the lobes and
the case. However, a sufficient seal between the lobes and the
blower case can be maintained by configuring the intermeshing
cooling surfaces to fit closely, as shown in FIG. 21.
[0166] As noted herein, it can be valuable to cool all of the
"walls" of the blower, so that the gas is not heated by the blower,
or such heating is minimized. As also noted herein, the
compression-effecting feature of a blower is the high-pressure gas
in the outlet rushing into the chamber on opening. As in other
types of compressors, it is valuable to effectively cool the
compressing feature of a blower. This is analogous to cooling the
piston in a reciprocating compressor, or cooling the scrolls and
plates of a scroll-type compressor. To effect the sufficient
cooling of the higher-pressure gas in a blower, a heat sink 650
(FIG. 20) can be positioned in the outlet 618 of the blower. A
large rectangular or similarly shaped outlet port can be included
here, rather than a circular or otherwise smaller outlet port used
in some conventional blowers; this can allow the heat sink 650 to
serve its cooling function along the entire length, or most of the
length, of the rotors.
[0167] The heat sink 650 may be in the form of a radiator, as
shown, and/or in the form of plates that are integral with the
blower case or thermally connected to the blower case, which can
then be cooled by an external heat sink, such as a water jacket or
cooling fins. In some embodiments, the heat sink, whether integral
or not, can be cooled by flowing the fluid in the cooling jacket
through the heat sink. A bar and plate type radiator shown at 650
may instead be a tube and fin type radiator. The plates or tubes of
a radiator, if any, can also be placed transversely, instead of
longitudinally, as shown; this can also be efficient in cooling the
higher pressure gas as it enters the alternately opening chambers,
in allowing the left-right component of the flow as the chambers of
the two rotors open alternately, and in facilitating use of the
cooling jacket fluid to cool the heat sink. In any configuration,
the location of the heat sink in the outlet can be valuable in
removing heat from the gas and increasing efficiency.
[0168] It can be desirable for the heat sink be located close to
the faces of the spinning lobes; that is, the heat sink may be more
effective when closely coupled to the lobes. Placement of
radiators, sometimes known as after-coolers, downstream of a blower
cannot serve the function described herein in cooling the gas that
is about to back up into the opening chambers. Because the volume
of gas that flows backward into the opening chamber is less than or
not much greater than the volume of the chamber, the cooling that
is required must be accomplished very close to the opening point of
the chamber, so that all or nearly all of the gas that enters the
chamber has been cooled. The higher pressure gas in the outlet that
is going to compress the gas in the chamber, on opening, is
desirably cooled before it enters the chamber, as is the case in
the blower 600.
[0169] Many of the features and principles disclosed herein can be
applied to and included in various other embodiments that those to
which they refer in this disclosure, and can be so applied or
included in any practical combination, all of which embodiments are
expressly included and encompassed by this disclosure. Such
features and principles include, in addition to others disclosed
elsewhere herein, the following:
[0170] (a) Concentric valves, such as those shown in the Figures
depicting reciprocating compressors and wobble-plate
compressors.
[0171] (b) Operating compressors at faster rates, such as above
1000 rpm or 2000 rpm. Conventional compressors run fairly slowly,
at least compared to gasoline and diesel engines, such as about 900
rpm, just above an idle for most engines. The main reason is that
the temperature rise of the compressor would be even worse, and
most compressors could not be run continuously. This is not a
problem solely for reciprocating compressors. Scroll compressors
can also include excessive temperature rise as a particular problem
that keeps them from being run faster. Without the cooling of the
non-stationary surfaces, compressor cannot be operated very fast,
even if the stationary surfaces are fluid-cooled. Thoroughly cooled
as they are, the disclosed compressors can be run faster than
conventional compressors.
[0172] (c) Operating compressors at higher pressures than is
possible with conventional compressors, due to the ability to
maintain the compressor at that temperature without the temperature
increasing further.
[0173] (d) Compressors that take advantage of additional, even what
might be called excessive, cooling. If there is a source of
coldness available, such as water from a well, or water cooled in a
well, or, in wintertime a radiator placed outside, a compressor can
be cooled to below ambient air temperature, taking advantage of the
low temperature source. Without the features and principles
disclosed herein, even the extra coldness of the outdoor air cannot
be used to its full advantage. With the disclosed technology, it is
possible to cool a compressor down, such as below the ambient air
temperature, and convey that lower temperature into the compression
chamber and this to the compressing gas, to the point that the work
of compression falls to exceptionally low levels.
[0174] (e) Cooling the scrolls of a scroll compressor. For
refrigerators, at least, conventional compressors are sealed up in
a can and get no cooling whatever. Cooling the scrolls provides a
huge efficiency advantage.
[0175] (f) Cooling fins on the non-working sides of the
non-stationary scroll walls.
[0176] (g) For blowers: positioning a radiator directly in the
outlet port very close to the lobes, as noted elsewhere herein,
instead of farther down the outlet conduit like conventional
after-coolers.
[0177] (h) For blowers and other rotary compressors (e.g., screw
compressors): fluid-cooling of the rotors and the axial ends of the
blower cases.
[0178] (i) Continuously cooling the non-stationary walls of any
compressor.
[0179] Conventional compressors have not been able to take
significant advantage of the cooling that has been applied to them,
due to:
[0180] (a) lack of sufficient surface area in the compression
chamber walls to which the compressing fluid can lose heat during
the compression;
[0181] (b) the walls of the compression chamber, especially the
non-stationary walls of the compression chamber, not being cooled
sufficiently to serve in the role of extracting heat from the fluid
that is being compressed, primarily during the compression; and
[0182] (c) lack of sufficient turbulence in the compressing fluid
to bring it into effective thermal contact with the walls of the
compression chamber, such that the fluid can lose heat to the walls
of the compression chamber, during the compression, at the rates
required to effect near-isothermal compression of the fluid.
[0183] The technology disclosed herein solves all of these problems
(though not every disclosed embodiment necessarily solves all of
these problems) and, for the first time, results in compressors
that can take fuller advantage of external cooling, specifically to
change the compression process over the short time scales involved,
and thereby effect compression that is significantly closer to
isothermal compression and significantly more efficient in the use
of energy. For the first time, compressors can make effective use
of external cooling, even cooling below ambient temperature when
reasonably possible, to change the nature and progression of the
compression process itself, moving it to near-isothermal
compression, and possibly, with sufficient cooling, even beyond the
theoretical efficiency of isothermal compression.
[0184] Additional exemplary claims include:
[0185] A. A reciprocating piston-type compressor, comprising:
[0186] a housing;
[0187] a piston that reciprocates within the housing along an axial
dimension of the compressor, the housing and the piston having
respective wall surfaces which together form a compression chamber
that varies in volume based on the axial position of the piston
relative to the housing;
[0188] wherein the housing includes a fluid inlet for receiving
into the compression chamber a fluid to be compressed and a fluid
outlet for expelling out of the compression chamber the fluid in a
compressed state;
[0189] wherein the housing comprises first projections that project
into the compression chamber;
[0190] wherein the piston comprises a piston head having a working
side facing the compression chamber and a non-working side facing
away from the compression chamber, wherein the working side of the
piston head includes second projections that project into the
compression chamber opposite the first projections;
[0191] wherein the second projections are oriented such that they
interleave between the first projections as the piston
reciprocates, whereby the first and second projections increase the
surface area of the wall surfaces and increase heat conduction away
from the fluid in the compression chamber when compared to a
cylindrical compression chamber; and
[0192] wherein the first projections and the second projections are
each arrayed circumferentially around a common central axis, the
first and second projections extend both axially and radially away
from the central axis, and the first and second projections
increase in circumferential thickness with increasing radial
distance from the central axis.
[0193] B. The compressor of claim A, wherein the first and second
projections have a generally trapezoidal-shaped cross-sectional
shape when viewed in a plane perpendicular to the axial
direction.
[0194] C. The compressor of claim A, wherein the second projections
have the same shape and size as the first projections, the second
projections fit flushly in corresponding valleys between the first
projections, and the second projections are offset
circumferentially relative to the first projections about one half
the angular width of one of the first projections.
[0195] D. The compressor of claim A, wherein the housing has a
cylindrical side wall and the first projections are integral with
the cylindrical side wall such that heat is conducted radially
through the first projections into the sidewall.
[0196] In view of the many possible embodiments to which the
principles of the disclosed technology may be applied, it should be
recognized that the illustrated embodiments are only examples and
should not be taken as limiting the scope of the disclosure.
Rather, the scope of the disclosure is at least as broad as the
following claims. I therefore claim all that comes within the scope
of these claims.
* * * * *