Note: Descriptions are shown in the official language in which they were submitted.
LINEAR CROSS-HEAD BEARING FOR STIRLING ENGINE
10
20
TECHNICAL FIELD
The present invention relates to machines and more particularly, to a Stirling
cycle
machine and components thereof.
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BACKGROUND INFORMATION
Many machines, such as internal combustion engines, external combustion
engines,
compressors, and other reciprocating machines, employ an arrangement of
pistons and drive
mechanisms to convert the linear motion of a reciprocating piston to rotary
motion. In most
.. applications, the pistons are housed in a cylinder. A common problem
encountered with
such machines is that of friction generated by a sliding piston resulting from
misalignment
of the piston in the cylinder and lateral forces exerted on the piston by
linkage of the piston
to a rotating crankshaft. These increased side loads increase engine noise,
increase piston
wear, and decrease the efficiency and life of the engine. Additionally,
because of the side
loads, the drive requires more power to overcome these frictional forces, thus
reducing the
efficiency of the machine.
Improvements have been made on drive mechanisms in an attempt to reduce these
side loads, however, many of the improvements have resulted in heavier and
bulkier
machines.
Accordingly, there is a need for practical machines with minimal side loads on
pistons.
SUMMARY
In accordance with one aspect of the present invention, an external combustion
engine is disclosed. The external combustion engine containing a working fluid
and
includes a burner element for heating the working fluid of the engine, at
least one
heater head defining a working space containing the working fluid, at least
one piston
cylinder containing a piston for compressing the working fluid, a cooler for
cooling the
working fluid, a crankcase including a crankshaft for producing an engine
output, a rocking
beam rotating about a rocker pivot for driving the crankshaft, a piston rod
connected to the
piston, a rocking beam driven by the piston rod, and a connecting rod
connected at a first
end to the rocking beam and at a second end to a crankshaft to convert rotary
motion of the
rocking beam to rotary motion of the crankshaft. Also, the external combustion
engine
including a piston rod seal unit including a housing, a cylinder gland, and at
least one
floating rod seal assembly mounted in the cylinder gland, the floating rod
seal assembly
comprising at least one rod seal mounted onto the floating rod seal assembly.
Some embodiments of this aspect of the present invention include one or more
of the
following. Where the piston rod seal unit further includes a scraper ring.
Where the piston
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rod seal unit further includes a particle trap. Wherein the piston rod seal
unit further
includes a port.
Wherein the piston rod seal unit further includes a filter. Wherein the
floating rod seal
assembly further includes an outer ring, and at least one bushing. Wherein the
piston rod
seal unit further includes wherein the rod seal is a spring energized seal.
In accordance with one aspect of the present invention, a piston rod seal unit
is
disclosed. The piston rod seal unit includes a housing, a cylinder gland, and
at least one
floating rod seal assembly mounted in the cylinder gland, the floating rod
seal assembly
comprising at least one rod seal mounted onto the floating rod seal assembly.
Some embodiments of this aspect of the present invention include one or more
of the
following. Wherein the piston rod seal unit further includes a scraper ring.
Wherein the
piston rod seal unit further includes a particle trap. Wherein the piston rod
seal unit further
includes a port. Wherein the piston rod seal unit further includes a filter.
Wherein the
floating rod seal assembly further includes an outer ring, and at least one
bushing. Wherein
the piston rod seal unit further includes wherein the rod seal is a spring
energized seal.
In accordance with one aspect of the present invention, an external combustion
engine is disclosed. The external combustion engine containing a working fluid
and
including a piston rod seal unit including a housing, a cylinder gland, and at
least one
floating rod seal assembly mounted in the cylinder gland, the floating rod
seal assembly
including at least one rod seal mounted onto the floating rod seal assembly,
and an airlock
space separating a crankcase and a working space for maintaining a pressure
differential
between a crankcase housing and a working space housing.
Some embodiments of this aspect of the present invention include one or more
of the
following. Wherein the airlock pressure regulator is a bidirectional pressure
regulator for
maintaining a predetermined pressure differential between the crankcase and
one of the
airlock space and working space. Wherein the airlock pressure regulator
includes a filter, a
compressor, a pressure regulating spool valve, and a linear position sensor,
wherein the
linear position sensor produces a signal indicative of the regulating spool
valve position.
Wherein the airlock pressure regulator further includes a controller. Wherein
the linear
position sensor is an LVDT. Wherein the controller uses the linear position
sensor to
regulate a pump speed.
In accordance with one aspect of the present invention, a rod seal assembly is
disclosed.
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The rod seal assembly includes a housing between two spaces configured to
receive a
reciprocating rod, the reciprocating rod disposed within a first space and a
second space, a
floating bushing configured to move axially and radially within the housing
and disposed
coaxially around the reciprocating rod, a rod seal configured to seal the
outside diameter of
the reciprocating rod relative to an inside surface of the floating bushing,
and at least one
stationary bushing fixed within the housing that may form a seal with the
floating bushing
to the axial flow of fluid in the presence of a pressure difference between
the two spaces.
Some embodiments of this aspect of the present invention include one or more
of the
following. Wherein the floating bushing is configured to move radially to
center on the
piston rod when the pressure difference between the first and second space is
small and
form the seal with the stationary bushing when the pressure difference is
larger. Wherein
the rod seal is a spring energized seal. Wherein the floating bushing further
comprises a
circumferential flange on the outside surface that is configured to extend
into the annular
space and form a seal with one of the stationary bushings. Wherein the rod
seal is formed
of a PTFE composite. Wherein the floating bushing and stationary bushing are
formed of a
wear resistance metal. Wherein the assembly further includes a scraper ring
disposed
coaxially around the piston rod and disposed within housing between the
floating seal and
the first space, and a passage connecting the first space to an annular gap
disposed around
the reciprocating rod between the scraper ring and the floating seal. Wherein
the assembly
further includes a magnetic particle trap disposed between the scraper ring
and floating seal.
In accordance with one aspect of the present invention, a rod seal assembly is
disclosed. The rod seal assembly includes a housing between two spaces
configured to
receive a reciprocating rod, the reciprocating rod disposed between a first
space and a
second space, a floating clearance bushing configured to move axially and
radially within
the housing and disposed coaxially around the reciprocating rod and forms a
clearance seal
with the reciprocating rod, and at least one stationary annular element fixed
within the
housing configured to form a face seal with the floating clearance bushing.
Some embodiments of this aspect of the present invention include one or more
of the
following. Wherein the floating clearance bushing is configured to move
radially to center
on the piston rod when the pressure difference between the first and second
space is small
and form the seal with the stationary annular element when the pressure
difference is larger.
Wherein the assembly further includes a spring energized face seal on at least
one end of the
floating clearance bushing. Wherein the assembly further includes a second
floating
clearance bushing disposed around the reciprocating rod, and two spring
energized lip seals
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disposed around the reciprocating piston rod and axially located within or
between the two
floating clearance bushings. Wherein the assembly further includes a scraper
ring disposed
coaxially around the piston rod and disposed within housing between the
floating seal and
the first space, and
5 a passage connecting the first space to an annular gap disposed around
the reciprocating rod
between the scraper ring and the floating seal. Wherein the assembly further
includes a
magnetic particle trap disposed between the scraper ring and floating seal.
In accordance with one aspect of the present invention, a floating rod seal is
disclosed. The floating seal includes a rod seal attached to a floating
bushing, wherein the
rod seal forms a leak tight joint with a floating bushing, two annular shaped
stationary
bushings that are located approximately coaxially with respect to the rod seal
and placed on
the inside diameter of an outer ring such that the two ends of the stationary
bushings form
an axial gap, and the floating bushing which includes an inner surface to seal
to the rod seal
and a circumferential rib on the outside surface, where the circumferential
rib is captured in
the axial gap and may move radially within the outer ring and may form a seal
with one
bushing.
In accordance with one aspect of the present invention, a floating rod seal is
disclosed. The floating seal includes a floating clearance bushing that forms
a clearance
seal with the piston rod and floats within cylinder-gland housing located
between a
workspace the air lock. The floating clearance bushing moves radially when the
pressure
difference between the workspace and the air lock are minimal and forms a seal
with a fixed
annular section of the housing when the pressure difference is large.
In accordance with one aspect of the present invention, an external combustion
engine is disclosed containing a working fluid and comprising a burner element
for heating
the working fluid of the engine, at least one heater head defining a working
space
containing the working fluid, at least one piston cylinder containing a piston
for
compressing the working fluid, at least one cooler for cooling the working
fluid and a
crankcase. The crankcase comprises a crankshaft for producing an engine
output, a piston
rod connected to the piston, a drive mechanism that converts the linear motion
of the piston
rod to rotary motion of the crankshaft, and a linear cross-head bearing
comprising a journal
and a guide. One end of the journal bearing is rigidly attached to the piston
rod and the
other end is attached to the drive mechanism. The guide is located outside the
working
space and the linear cross-head bearing solely constrains the motion of the
piston.
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Some embodiments of this aspect of the present invention include one or more
of the
following: the ratio of the linear cross-head bearing length over diameter is
greater than 2.0;
the linear cross-head bearing is a hydrodynamic bearing supplied with
pressurized oil from
an annulus on the guide; a piston rod seal unit including a housing; a
floating clearance
bushing configured to move axially and radially within the housing and
disposed coaxially
around the piston rod and at least one stationary annular element fixed within
the housing
configured to form a face seal with the floating clearance bushing; a piston
rod seal unit
including a housing, a floating rod seal assembly with at least one rod seal
mounted onto the
floating rod seal assembly. a piston rod seal unit including a housing, a
cylinder gland, and
at least one floating rod seal assembly mounted in the cylinder gland where a
rod seal is
mounted onto the floating rod seal assembly; the rod seal is an spring
energized seal; the
piston includes a clearance seal where the clearance seal may include radial
grooves on the
outside diameter and 0-rings on the inside diameter and the clearance seal may
include at
least one of the following materials; graphite, PTFE. UHMWPE, antimony.
In accordance with one aspect of the present invention, an external combustion
engine is disclosed containing a working fluid and comprising a burner element
for heating
the working fluid of the engine, at least one heater head defining a working
space
containing the working fluid, at least one piston cylinder containing a piston
for
compressing the working fluid, at least one cooler for cooling the working
fluid and a
crankcase. The crankcase comprises a crankshaft for producing an engine
output, a piston
rod connected to the piston, a rocking beam rotating about a rocker pivot for
driving the
crankshaft, a rocking beam driven by the piston rod, and a connecting rod
connected at a
first end to the rocking beam and at a second end to a crankshaft to convert
rotary motion of
the rocking beam to rotary motion of the crankshaft and a linear cross-head
bearing
comprising a journal and a guide. One end of the journal bearing is rigidly
attached to the
piston rod and the other end is attached to the drive mechanism. The guide is
located
outside the working space and the linear cross-head bearing solely constrains
the motion of
the piston.
Some embodiments of this aspect of the present invention include one or more
of the
following: the ratio of the linear cross-head bearing length over diameter is
greater than 2.0;
the linear cross-head bearing is a hydrodynamic bearing supplied with
pressurized oil from
an annulus on the guide; a piston rod seal unit including a housing; a
floating clearance
bushing configured to move axially and radially within the housing and
disposed coaxially
around the piston rod and at least one stationary annular element fixed within
the housing
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configured to form a face seal with the floating clearance bushing; a piston
rod seal unit
including a housing, a floating rod seal assembly with at least one rod seal
mounted onto the
floating rod seal assembly. a piston rod seal unit including a housing, a
cylinder gland, and
at least one floating rod seal assembly mounted in the cylinder gland where a
rod seal is
mounted onto the floating rod seal assembly; the rod seal is an spring
energized seal; the
piston includes a clearance seal where the clearance seal may include radial
grooves on the
outside diameter and 0-rings on the inside diameter and the clearance seal may
include at
least one of the following materials; graphite, PTFE, UHMWPE, antimony.
In accordance with one aspect of the present invention, an external combustion
engine is disclosed containing a working fluid and comprising a burner element
for heating
the working fluid of the engine, at least one heater head defining a working
space
containing the working fluid, at least one piston cylinder containing a piston
for
compressing the working fluid, at least one cooler for cooling the working
fluid and a
crankcase. The crankcase comprises a crankshaft for producing an engine
output, a piston
rod connected to the piston, a drive mechanism that converts the linear motion
of the piston
rod to rotary motion of the crankshaft, and a piston comprising a piston base
with a seal, an
outer shell that mounts on the base and an inner shell that is shorter and
narrower than the
outer shell and defines with the piston base a volume above the inner sutface
of the piston
base.
Some embodiments of this aspect of the present invention include one or more
of the
following: a small orifice in the piston base, an orifice in the piston base
with a diameter
between 0.002 and 0.008 inches, and port in the inner shell. In accordance
with one aspect
of the present invention, an external combustion engine is disclosed. The
external
combustion engine containing a working fluid and including a burner element
for heating
the working fluid of the engine, at least one heater head defining a working
space
containing the working fluid, at least one piston cylinder containing a piston
for
compressing the working fluid, a cooler for cooling the working fluid, a
crankcase. The
crankcase includes a crankshaft for producing an engine output, a rocking beam
rotating
about a rocker pivot for driving the crankshaft, a piston rod connected to the
piston, a
rocking beam driven by the piston rod, and a connecting rod connected at a
first end to the
rocking beam and at a second end to a crankshaft to convert rotary motion of
the rocking
beam to rotary motion of the crankshaft. The external combustion engine also
includes an
airlock space separating the crankcase and the working space for maintaining a
pressure
.. differential between the crankcase housing and the working space housing
and an airlock
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pressure regulator connected between the crankcase and one of the airlock
space and
working space.
Some embodiments of this aspect of the present invention include one or more
of the
following. A first seal for sealing the crankcase from the airlock space,
wherein the seal is a
rolling diaphragm. A seal for sealing the workspace from the airlock space,
wherein the
seal is a pair of oppositely disposed rolling diaphragms. A second seal for
sealing the
workspace from the the airlock space, wherein the seal is a high pressure
seal. Wherein the
airlock pressure regulator is a bidirectional pressure regulator for
maintaining a
predetermined pressure differential between the crankcase and one of the
airlock space and
working space. Wherein the airlock pressure regulator includes a filter, a
compressor, a
pressure regulating spool valve, and a linear position sensor, wherein the
linear position
sensor produces a signal indicative of the regulating spool valve position.
Wherein the
airlock pressure regulator further includes a controller. Wherein the linear
position sensor
is an LVDT. Wherein the controller uses the linear position sensor to regulate
a pump
speed. Wherein the airlock pressure regulator further includes a drain port
and a fill port,
wherein the drain port and the fill port are selectively connected. Wherein
the controller
sends a stop command to the drive based at least in part on the position of
the linear position
sensor. Wherein the external combustion engine further includes a burner
controller for a
multi-burner including a master controller and an individual combustion
control circuit.
Wherein the master controller controls a variable resistance element connected
to the fuel
line in each burner in the multi-burner. Wherein the external combustion
engine further
includes an over-temperature circuit, wherein the over-temperature circuit
monitors a
temperature on each of the heater head temperatures and may disable the fuel
valve
supplying a burner heating a heater head. Wherein the external combustion
engine further
includes a flow-detection circuit wherein the flow-detection circuit disables
all the fuel
valves when the flow-detection circuit detects a flow that is below a
predetermined
threshold.
In accordance with one aspect of the present invention, a rocking beam drive
mechanism for a machine is disclosed. The drive mechanism includes a rocking
beam
having a rocker pivot, at least one cylinder and at least one piston. The
piston is housed
within a respective cylinder. The piston is capable of substantially linearly
reciprocating
within the respective cylinder. Also, the drive mechanism includes at least
one coupling
assembly having a proximal end and a distal end. The proximal end is connected
to the
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piston and the distal end is connected to the rocking beam by an end pivot.
The linear
motion of the piston is converted to rotary motion of the rocking beam.
Some embodiments of this aspect of the present invention include one or more
of
the following: where the rocking beam is coupled to a crankshaft by way of a
connecting
rod. In this embodiment, the rotary motion of the rocking beam is transferred
to the
crankshaft. Also, where the cylinder may further include a closed end and an
open end.
The open end further includes a linear bearing connected to the cylinder. The
linear bearing
includes an opening to accommodate the coupling assembly. Also, where the
coupling
assembly further includes a piston rod and a link rod. The piston rod and link
rod are
coupled together by a coupling means. The coupling means is located beneath
the linear
bearing. Also, where the drive mechanism also includes a seal, where the seal
is sealably
connected to the piston rod. Also, where the seal is a rolling diaphragm.
Also, in some
embodiments, the coupling means is a flexible joint. In some embodiments, the
coupling
means is a roller bearing. In some embodiments, the coupling means is a hinge.
In some
embodiments, the coupling means is a flexure. In some embodiments, the
coupling means is
a journal bearing joint.
In accordance with another aspect of the present invention, a Stirling cycle
machine is disclosed. The machine includes at least one rocking drive
mechanism where
the rocking drive mechanism includes: a rocking beam having a rocker pivot, at
least one
cylinder and at least one piston. The piston is housed within a respective
cylinder. The
piston is capable of substantially linearly reciprocating within the
respective cylinder. Also,
the drive mechanism includes at least one coupling assembly having a proximal
end and a
distal end. The proximal end is connected to the piston and the distal end is
connected to
the rocking beam by an end pivot. The linear motion of the piston is converted
to rotary
motion of the rocking beam. Also, a crankcase housing the rocking beam and
housing a
first portion of the coupling assembly is included. A crankshaft coupled to
the rocking
beam by way of a connecting rod is also included. The rotary motion of the
rocking beam is
transferred to the crankshaft. The machine also includes a working space
housing the at
least one cylinder, the at least one piston and a second portion of the
coupling assembly. A
seal is included for sealing the workspace from the crankcase.
Some embodiments of this aspect of the present invention include one or more
of
the following: where the seal is a rolling diaphragm. Also, the cylinder may
further include
a closed end and an open end. The open end further includes a linear bearing
connected to
the cylinder. The linear bearing includes an opening to accommodate the
coupling
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assembly. Also, where the coupling assembly further includes a piston rod and
a link rod.
The piston rod and link rod are coupled together by a coupling means. The
coupling means
may be located beneath the linear bearing. Also, the machine may also include
a lubricating
fluid pump in the crankcase. In some embodiments, the lubricating fluid pump
is a
5 mechanical lubricating fluid pump driven by a pump drive assembly, the
pump drive
assembly being connected to and driven by the crankshaft. In some embodiments,
the
lubricating fluid pump is an electric lubricating fluid pump. The machine may
also include
a motor connected to the crankshaft. The machine may also include a generator
connected
to the crankshaft.
10 In accordance with another aspect of the present invention, a Stirling
cycle
machine is disclosed. The machine includes at least two rocking drive
mechanisms. The
rocking drive mechanisms each include a rocking beam having a rocker pivot,
two
cylinders, and two pistons. The pistons each housed within a respective
cylinder. The
pistons are capable of substantially linearly reciprocating within the
respective cylinder.
Also, the drive mechanisms include two coupling assemblies having a proximal
end and a
distal end, the proximal end being connected to the piston and the distal end
being
connected to the rocking beam by an end pivot. The linear motion of the piston
is converted
to rotary motion of the rocking beam. The machine also includes a crankcase
housing the
rocking beam and housing a first portion of the coupling assemblies. Also, a
crankshaft
coupled to the rocking beam by way of a connecting rod. The rotary motion of
the rocking
beam is transferred to the crankshaft. The machine also includes a lubricating
fluid pump in
the crankcase for pumping lubricating fluid to lubricate the crankshaft and
the rocking beam
and the first portion of the coupling assemblies. Also, a working space
housing the
cylinders, the pistons and the second portion of the coupling assemblies. A
rolling
diaphragm for sealing the workspace from the crankcase is also included.
Some embodiments of this aspect of the present invention include one or more
of
the following: where the cylinder may further include a closed end and an open
end. The
open end further includes a linear bearing connected to the cylinder. The
linear bearing
includes an opening to accommodate the coupling assembly. Also, where the
coupling
assembly further includes a piston rod and a link rod. The piston rod and link
rod are
coupled together by a coupling means. The coupling means may be located
beneath the
linear bearing. Also, where the coupling means is a flexible joint. In some
embodiments,
also disclosed is where the coupling means is a roller bearing.
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Other embodiments of this aspect of the present invention relate to one or
more of
an external combustion engine containing a working fluid comprising a burner
element for
heating the working fluid of the engine, at least one heater head defining a
working space
containing the working fluid, at least one piston cylinder containing a piston
for
compressing the working fluid, a cooler for cooling the working fluid, a
crankcase
comprising a crankshaft for producing an engine output, a rocking beam
rotating about a
rocker pivot for driving the crankshaft, a piston rod connected to the piston,
a rocking beam
driven by the piston rod, and a connecting rod connected at a first end to the
rocking beam
and at a second end to the crankshaft to convert rotary motion of the rocking
beam to rotary
motion of the crankshaft wherein the piston reciprocates along a substantially
linear piston
axis in the crankcase and the crankshaft is arranged below a limit of the
piston axis in the
crankcase.
A still further embodiment of the invention relate to one or more embodiments
of
an external combustion engine containing a working fluid comprising a burner
element for
heating the working fluid of the engine, at least one heater head defining a
working space
containing the working fluid, at least one piston cylinder containing a piston
for
compressing the working fluid, a cooler for cooling the working fluid, a
crankcase
comprising, a crankshaft for producing an engine output, a rocking beam
rotating about a
rocker pivot for driving the crankshaft, a piston rod connected to the piston,
a rocking beam
driven by the piston rod, a connecting rod connected at a first end to the
rocking beam and
at a second end to a crankshaft to convert rotary motion of the rocking beam
to rotary
motion of the crankshaft, and an airlock space separating the crankcase and
the working
space for maintaining a pressure differential between the crankcase housing
and the
working space housing.
A still further embodiment of the invention relate to one or more embodiments
of
a heating element for heating an external combustion engine or machine
comprising a
burner element for heating the working fluid of the engine, a blower providing
air or other
gas for facilitating ignition and combustion in the burner, a preheater
defining an incoming
air passage and an exhaust passage separated by an exhaust manifold wall for
heating
incoming air from the hot exhaust expelled from the heating element, a fuel
injector for
supplying fuel to mix with the incoming air, an igniter to ignite the fuel/air
mixture, a
prechamber defining an inlet for receiving the fuel/air mixture and promoting
ignition of the
mixture, a combustion chamber disposed linearly below the prechamber for
maintaining
supporting a flame developed and ignited in the prechamber, an electronic
control unit for
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controlling ignition and combustion operations of the burner, and wherein the
combustion
chamber is connected to the exhaust passage into which the exhausted
combustion gases are
pushed to heat the incoming air following combustion and heating of the engine
or machine.
These aspects of the invention are not meant to be exclusive and other
features, aspects, and advantages of the present invention will be readily
apparent to those
of ordinary skill in the art when read in conjunction with the appended claims
and
accompanying drawings.
BRIEF DESCRIPTION OF THE DRAWINGS
These and other features and advantages of the present invention will be
better
understood by reading the following detailed description, taken together with
the drawings
wherein:
FIGS. 1A-1E depict the principle of operation of a prior art Stirling cycle
machine;
FIG. 2 shows a view of a rocking beam drive in accordance with one embodiment;
FIG. 3 shows a view of a rocking beam drive in accordance with one embodiment;
FIG. 4 shows a view of an engine in accordance with one embodiment;
FIGS. 5A-5D depicts various views of a rocking beam drive in accordance with
one
embodiment;
FIG. 6 shows a bearing style rod connector in accordance with one embodiment;
FIGS. 7A-7B show a flexure in accordance with one embodiment;
FIG. 8 shows the operation of pistons of an engine in accordance with one
embodiment;
FIG. 9A shows an unwrapped schematic view of a working space and cylinders in
accordance with one embodiment;
FIG. 9B shows a schematic view of a cylinder, heater head, and regenerator in
accordance with one embodiment;
FIG. 9C shows a view of a cylinder head in accordance with one embodiment;
FIG. 10A shows a view of a rolling diaphragm, along with supporting top seal
piston
and bottom seal piston, in accordance with one embodiment;
FIG. 10B shows an exploded view of a rocking beam driven engine in accordance
with one embodiment;
FIG. 10C shows a view of a cylinder, heater head, regenerator, and rolling
diaphragm, in accordance with one embodiment;
FIG. 10D shows various views of a rolling diaphragm during operation, in
accordance with one embodiment;
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FIG. 11 shows a view of an external combustion engine in accordance with one;
FIGS. 12A-12E show views of various embodiments of a rolling diaphragm;
FIG. 13A shows a schematic of a rolling diaphragm identifying various load
regions;
FIG. 13B shows a schematic of the rolling diaphragm identifying the
convolution
region;
FIG. 14 shows a view of a piston and piston seal in accordance with one
embodiment;
FIG. 15 shows a view of a piston rod and piston rod seal in accordance with
one
embodiment;
FIGS. 16A-16B show views of a piston guide ring in accordance with one
embodiment;
FIG. 17 shows a view of a tube heat exchanger in accordance with one
embodiment;
FIG. 18 shows a portion of a cross section of a tube heat exchanger in
accordance
with one embodiment;
FIG. 19 shows a view of a heater head of an engine in accordance with one
embodiment;
FIG. 20A shows a view of a tube heat exchanger in accordance with one
embodiment;
FIG. 20B shows a view of a tube heat exchanger in accordance with one
embodiment;
FIG. 21A shows a view of a tube heat exchanger in accordance with one
embodiment;
FIG. 21B shows a view of a tube heat exchanger in accordance with one
embodiment;
FIG. 22A shows view of a tube heat exchanger in accordance with one
embodiment;
FIG. 22B shows a view of a tube heat exchanger in accordance with one
embodiment;
FIGS. 23A and 23B show a regenerator of a Stirling cycle engine in accordance
with
one embodiment;
FIGS. 24A-24E show various configurations of a regenerator of a Stirling cycle
engine in accordance with various embodiments;
FIGS. 25A-25C show various views of an engine in accordance with several
embodiments;
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FIGS. 26A and 26B show views of a cooler for an engine in accordance with some
embodiments;
FIGS. 27A and 27B show a view of a cooler for an engine in accordance with one
embodiment;
FIGS. 27C and 27D show a view of a cooler for an engine in accordance with one
embodiment;
FIGS. 28A-28C show views of an intake manifold for an engine in accordance
with
one embodiment;
FIGS. 29 is a gaseous fuel burner coupled to a Stirling cycle engine, where
the
ejector is a venturi, according to one embodiment;
FIGS. 30A is the burner of FIGS. 29 showing the air and fuel flow paths in
accordance with one embodiment;
FIGS. 30B is a graphical representation of the pressure across the burner in
accordance with one embodiment;
FIG. 31 shows a schematic of an embodiment of the burner with automated fuel
control for variable fuel properties;
FIG. 32 shows a schematic of another embodiment of the burner with temperature
sensor and engine speed control loop;
FIG. 33 shows a schematic of yet another embodiment of the burner with
temperature sensor and oxygen sensor control loop;
FIG. 34 shows an alternative embodiment of the ejector wherein the fuel is fed
directly into the ejector;
FIG. 35 shows a cross section of an engine in accordance with one embodiment;
FIGS. 36A and 36B show a cross-sectional view of a Stirling cycle machine
having
an inverted rocking beam design in accordance with one embodiment;
FIGS. 36C-36E show various views of a piston and piston rod assembly in
accordance with one embodiment;
FIG. 37A shows a view of an embodiment of the rocking beam with a conrod
bearing ratio of 1.6;
FIG. 37B shows a view of an embodiment of the rocking beam with a conrod
bearing ratio of 1.0;
FIG. 38A shows an oil pump according to one embodiment;
FIG. 38B shows a Gerotor displacement pumping unit according to one
embodiment;
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FIG. 39 shows an embodiment of a high pressure rod seal;
FIG. 40A shows another embodiment of a high pressure rod seal including a
spring
energized lip seal;
FIG. 40B is a hydraulic high pressure piston rod seal set inside the rod seal
cavity of
5 a test rig according to one embodiment;
FIGS. 41A and 41B show views of a rolling diaphragm in accordance with one
embodiment;
FIGS. 42A and 42B show views of a rolling diaphragm in accordance with another
embodiment;
10 FIG. 43 shows a view of a double bellows system in accordance with one
embodiment;
FIGS. 44A and 44B show views of an airlock pressure regulation system in
accordance with one embodiment;
FIG. 45 shows a bidirectional regulator according to one embodiment;
15 FIGS. 46A-46E show various positions of a spool valve in a bidirectional
regulator
in accordance with various embodiments;
FIG. 47 shows a view of an airlock pressure regulation system in accordance
with
one embodiment;
FIG. 48 shows a view of an airlock pressure regulation system in accordance
with
one embodiment;
FIG. 49 shows a view of a mechanical pump for regulating airlock pressure in
accordance with one embodiment;
FIGS. 50A and 50B show views of a heat exchanger in accordance with one
embodiment;
FIGS. 51A and 51B show views of a rocking beam mechanism in accordance with
one embodiment;
FIGS. 52A and 52B show views of a horizontally supported Stirling cycle engine
in
accordance with one embodiment;
FIGS. 53A and 53B show views of a tube-in-tube heat exchanger according to one
embodiment;
FIGS. 53C and 53D show views of a tube based heat exchanger according to one
embodiment;
FIGS. 54-59 show various views of a burner in accordance with one embodiment;
FIG. 60 show views of a burner in accordance with one embodiment;
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FIG. 61 is a diagram of a control burner scheme in accordance with one
embodiment;
FIGS. 62A ¨ 62D are a further embodiment of a venturi-type burner for use in
conjunction with the multiple heater head in accordance with one embodiment;
FIG. 63 shows a further embodiment of an airlock pressure regulation system;
FIGS. 64A ¨ 64F present an embodiment of an Airlock delta Pressure Regulation
(AdPR) block;
FIG. 65A ¨ 65D is an embodiment of an Airlock delta Pressure Regulation (AdPR)
block;
FIGS. 65E ¨ 651 is an embodiment of an Airlock delta Pressure Regulation
(AdPR)
block;
FIGS. 66A-66E show a further embodiment of various positions of a spool valve
in
a bidirectional regulator in accordance with various embodiments;
FIG. 67A shows a chart which shows the order of the embodiments as depicted in
FIGS. 67B-H;
FIGS. 67 B-H illustrate an embodiment of a Stirling Engine Controller;
FIG. 68A is a cross-section of one embodiment of a Stirling Engine with a
piston
rod seal unit;
FIG. 68B is a detailed view of the piston rod seal unit in FIG. 68A;
FIG. 69A is a cross-section of one embodiments of a piston rod seal unit;
FIG. 69B is a cross-section of one embodiment of a floating rod seal assembly;
FIG. 69C is an isometric view of one embodiment of a piston rod seal unit;
FIG. 69C is an isometric view of one embodiment of a piston rod seal unit;
FIG. 69D is a cross-section of an embodiment of a piston rod seal unit with
clearance seals;
FIG. 69E is a cross-section of an embodiment of a piston rod seal unit with
clearance seals;
FIG. 69F is a cross-section of an embodiment of a piston rod seal unit with a
hybrid
clearance and lip seal;
FIG. 70A is a cross-section of a further embodiment of a annular-venturi-type
burner
for use in conjunction with the multiple heater head in accordance with one
embodiment;
FIG. 70B is a cross-section of a the annular-venturi-type burner head;
FIG. 70C is a detailed view of the annular-venturi with fuel ports and an
ignitor;
17
FIG. 70D is a perspective view of a radial svvirler at the entrance to the
annular-
v enturi;
FIG. 71A is a cross-section of one embodiment of a Stirling Engine with an
linear-
cross-head bearing;
FIG, 71B is a cross-section of one embodiment of a piston, piston rod and
linear-
cross-head bearing;
FIG. 71C is a cross-section of one embodiment of a linear-cross-head bearing;
and
FIG. 72 is a cross-section of one embodiment of a piston with a clearance
seal.
DETAILED DESCRIPTION
Stirling cycle machines, including engines and refrigerators, have a long
technological heritage, described in detail in Walker, Stirling Engines,
Oxford University
Press (1980). The principle underlying the Stirling cycle engine is the
mechanical
realization of the Stirling thermodynamic cycle: isovolumetric heating of a
gas within a
cylinder, isothermal expansion of the gas (during which work is performed by
driving a
piston), isovolumetric cooling, and isothermal compression. Additional
background
regarding aspects of Stirling cycle machines and improvements thereto is
discussed in
Hargreaves, The Phillips Stirling Engine (Elsevier, Amsterdam, 1991).
The principle of operation of a Stirling cycle machine is readily described
with
reference to FIGS. 1A-1E, wherein identical numerals are used to identify the
same or
similar parts. Many mechanical layouts of Stirling cycle machines are known in
the art, and
the particular Stirling cycle machine designated generally by numeral 10 is
shown merely
for illustrative purposes. In FIGS. 1A to 1D, piston 12 and a displacer 14
move in phased
reciprocating motion within the cylinders 16 which, in some embodiments of the
Stirling
cycle machine, may be a single cylinder, but in other embodiments, may include
greater
than a single cylinder. A working fluid contained within cylinders 16 is
constrained by
seals from escaping around piston 12 and displacer 14. The working fluid is
chosen for its
thermodynamic properties, as discussed in the description below, and is
typically helium at
a pressure of several atmospheres, however, any gas, including any inert gas,
may be used,
including, but not limited to, hydrogen, argon, neon, nitrogen, air and any
mixtures thereof
The position of the displacer 14 governs whether the working fluid is in
contact with the hot
interface 18 or the cold interface 20, corresponding, respectively, to the
interfaces at which
heat is supplied to and extracted from the working fluid. The supply and
extraction of heat
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is discussed in further detail below. The volume of working fluid governed by
the position
of the piston 12 is referred to as the compression space 22.
During the first phase of the Stirling cycle, the starting condition of which
is
depicted in FIG. 1A, the piston 12 compresses the fluid in the compression
space 22. The
compression occurs at a substantially constant temperature because heat is
extracted from
the fluid to the ambient environment. The condition of the Stirling cycle
machine 10 after
compression is depicted in FIG. 1B. During the second phase of the cycle, the
displacer 14
moves in the direction of the cold interface 20, with the working fluid
displaced from the
region of the cold interface 20 to the region of the hot interface 18. This
phase may be
referred to as the transfer phase. At the end of the transfer phase, the fluid
is at a higher
pressure since the working fluid has been heated at constant volume. The
increased pressure
is depicted symbolically in FIG. 1C by the reading of the pressure gauge 24.
During the third phase (the expansion stroke) of the Stirling cycle machine,
the
volume of the compression space 22 increases as heat is drawn in from outside
the Stirling
cycle machine 10, thereby converting heat to work. In practice, heat is
provided to the fluid
by means of a heater head (not shown) which is discussed in greater detail in
the description
below. At the end of the expansion phase, the compression space 22 is full of
cold fluid, as
depicted in FIG. 1D. During the fourth phase of the Stirling cycle machine 10,
fluid is
transferred from the region of the hot interface 18 to the region of the cold
interface 20 by
motion of the displacer 14 in the opposing sense. At the end of this second
transfer phase,
the fluid fills the compression space 22 and cold interface 20, as depicted in
FIG. 1A, and is
ready for a repetition of the compression phase. The Stirling cycle is
depicted in a P-V
(pressure-volume) diagram as shown in FIG. 1E.
Additionally, on passing from the region of the hot interface 18 to the region
of the
cold interface 20. In some embodiments, the fluid may pass through a
regenerator (shown
as 408 in FIG. 4). A regenerator is a matrix of material having a large ratio
of surface area
to volume which serves to absorb heat from the fluid when it enters from the
region of the
hot interface 18 and to heat the fluid when it passes from the region of the
cold interface 20.
Stirling cycle machines have not generally been used in practical applications
due to
several daunting challenges to their development. These involve practical
considerations
such as efficiency and lifetime. Accordingly, there is a need for more
Stirling cycle
machines with minimal side loads on pistons, increased efficiency and
lifetime.
19
The principle of operation of a Stirling cycle machine or Stirling engine is
further
discussed in detail in U.S. Patent No. 6,381,958, issued May 7,2002, to Kamen
etal..
Rocking Beam Drive
Referring now to FIGS. 2-4, embodiments of a Stirling cycle machine, according
to
one embodiment, are shown in cross-section. The engine embodiment is
designated
generally by numeral 300. While the Stirling cycle machine will be described
generally with
reference to the Stirling engine 300 embodiments shown in FIGS. 2-4, it is to
be understood
that many types of machines and engines, including but not limited to
refrigerators and
compressors may similarly benefit from various embodiments and improvements
which are
described herein, including but not limited to, external combustion engines
and internal
combustion engines.
FIG. 2 depicts a cross-section of an embodiment of a rocking beam drive
mechanism
200 (the term -rocking beam drive" is used synonymously with the term "rocking
beam
drive mechanism") for an engine, such as a Stirling engine, having linearly
reciprocating
pistons 202 and 204 housed within cylinders 206 and 208, respectively. The
cylinders
include linear bearings 220. Rocking beam drive 200 converts linear motions of
pistons 202
and 204 into the rotary motion of a crankshaft 214. Rocking beam drive 200 has
a rocking
beam 216, rocker pivot 218, a first coupling assembly 210, and a second
coupling assembly
212. Pistons 202 and 204 are coupled to rocking beam drive 200, respectively,
via first
coupling assembly 210 and second coupling assembly 212. The rocking beam drive
is
coupled to crankshaft 214 via a connecting rod 222.
In some embodiments, the rocking beam and a first portion of the coupling
assembly
may be located in a crankcase, while the cylinders, pistons and a second
portion of the
coupling assembly is located in a workspace.
In FIG. 4 a crankcase 400 most of the rocking beam drive 200 is positioned
below
the cylinder housing 402. Crankcase 400 is a space to permit operation of
rocking beam
drive 200 having a crankshaft 214, rocking beam 216, linear bearings 220, a
connecting rod
222, and coupling assemblies 210 and 212. Crankcase 400 intersects cylinders
206 and 208
transverse to the plane of the axes of pistons 202 and 204. Pistons 202 and
204 reciprocate
in respective cylinders 206 and 208, as also shown in FIG 2. Cylinders 206 and
208 extend
above crankshaft housing 400. Crankshaft 214 is mounted in crankcase 400 below
cylinders 206 and 208.
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FIG. 2 shows one embodiment of rocking beam drive 200. Coupling assemblies 210
and 212 extend from pistons 202 and 204, respectively, to connect pistons 202
and 204 to
rocking beam 216. Coupling assembly 212 for piston 204, in some embodiments,
may
comprise a piston rod 224 and a link rod 226. Coupling assembly 210 for piston
202, in
5 some embodiments, may comprise a piston rod 228 and a link rod 230.
Piston 204 operates
in the cylinder 208 vertically and is connected by the coupling assembly 212
to the end
pivot 232 of the rocking beam 216. The cylinder 208 provides guidance for the
longitudinal
motion of piston 204. The piston rod 224 of the coupling assembly 212 attached
to the
lower portion of piston 204 is driven axially by its link rod 226 in a
substantially linear
10 reciprocating path along the axis of the cylinder 208. The distal end of
piston rod 224 and
the proximate end of link rod 226, in some embodiments, may be jointly hinged
via a
coupling means 234. The coupling means 234, may be any coupling means known in
the
art, including but not limited to, a flexible joint, roller bearing element,
hinge, journal
bearing joint (shown as 600 in FIG. 6), and flexure (shown as 700 in FIGS. 7A
and 7B).
15 The distal end of the link rod 226 may be coupled to one end pivot 232
of rocking beam
216, which is positioned vertically and perpendicularly under the proximate
end of the link
rod 226. A stationary linear bearing 220 may be positioned along coupling
assembly 212 to
further ensure substantially linear longitudinal motion of the piston rod 224
and thus
ensuring substantially linear longitudinal motion of the piston 204. In an
exemplary
20 embodiment, link rod 226 does not pass through linear bearing 220. This
ensures, among
other things, that piston rod 224 retains a substantially linear and
longitudinal motion.
In the exemplary embodiment, the link rods may be made from aluminum, and the
piston rods and connecting rod are made from D2 Tool Steel. Alternatively, the
link rods,
piston rods, connecting rods, and rocking beam may be made from 4340 steel.
Other
materials may be used for the components of the rocking beam drive, including,
but not
limited to, titanium, aluminum, steel or cast iron. In some embodiments, the
fatigue
strength of the material being used is above the actual load experienced by
the components
during operation.
Still referring to FIGS. 2-4, piston 202 operates vertically in the cylinder
206 and is
connected by the coupling assembly 210 to the end pivot 236 of the rocking
beam 216. The
cylinder 206 serves, amongst other functions, to provide guidance for
longitudinal motion
of piston 202. The piston rod 228 of the coupling assembly 210 is attached to
the lower
portion of piston 202 and is driven axially by its link rod 230 in a
substantially linear
reciprocating path along the axis of the cylinder 206. The distal end of the
piston rod 228
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and the proximate end of the link rod 230, in some embodiments, is jointly
hinged via a
coupling means 238. The coupling means 238, in various embodiments may
include, but
are not limited to, a flexure (shown as 700 in FIGS. 7A and 7B, roller bearing
element,
hinge, journal bearing (shown as 600 in FIG. 6), or coupling means as known in
the art.
The distal end of the link rod 230, in some embodiments, may be coupled to one
end pivot
236 of rocking beam 216, which is positioned vertically and perpendicularly
under the
proximate end of link rod 230. A stationary linear bearing 220 may be
positioned along
coupling assembly 210 to further ensure linear longitudinal motion of the
piston rod 228
and thus ensuring linear longitudinal motion of the piston 202. In an
exemplary
embodiment, link rod 230 does not pass through linear bearing 220 to ensure
that piston rod
228 retains a substantially linear and longitudinal motion.
The coupling assemblies 210 and 212 change the alternating longitudinal motion
of
respective pistons 202 and 204 to oscillatory motion of the rocking beam 216.
The
delivered oscillatory motion is changed to the rotational motion of the
crankshaft 214 by the
connecting rod 222, wherein one end of the connecting rod 222 is rotatably
coupled to a
connecting pivot 240 positioned between an end pivot 232 and a rocker pivot
218 in the
rocking beam 216, and another end of the connecting rod 222 is rotatably
coupled to
crankpin 246. The rocker pivot 218 may be positioned substantially at the
midpoint
between the end pivots 232 and 236 and oscillatorily support the rocking beam
216 as a
fulcrum, thus guiding the respective piston rods 224 and 228 to make
sufficient linear
motion. In the exemplary embodiment, the crankshaft 214 is located above the
rocking
beam 216, but in other embodiments, the crankshaft 214 may be positioned below
the
rocking beam 216 (as shown in FIGS. 5B and 5D) or in some embodiments, the
crankshaft
214 is positioned to the side of the rocking beam 216, such that it still has
a parallel axis to
the rocking beam 216.
Still referring to FIGS. 2-4, the rocking beam oscillates about the rocker
pivot 218,
the end pivots 232 and 236 follow an arc path. Since the distal ends of the
link rods 226 and
230 are connected to the rocking beam 216 at pivots 232 and 236, the distal
ends of the link
rods 226 and 230 also follow this arc path, resulting in an angular deviation
242 and 244
from the longitudinal axis of motion of their respective pistons 202 and 204.
The coupling
means 234 and 238 are configured such that any angular deviation 244 and 242
from the
link rods 226 and 230 experienced by the piston rods 224 and 228 is minimized.
Essentially, the angular deviation 244 and 242 is absorbed by the coupling
means 234 and
238 so that the piston rods 224 and 228 maintain substantially linear
longitudinal motion to
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reduce side loads on the pistons 204 and 202. A stationary linear bearing 220
may also be
placed inside the cylinder 208 or 206, or along coupling assemblies 212 or
210, to further
absorb any angular deviation 244 or 242 thus keeping the piston push rod 224
or 228 and
the piston 204 or 202 in linear motion along the longitudinal axis of the
piston 204 or 202.
Therefore, in view of reciprocating motion of pistons 202 and 204, it is
necessary to
keep the motion of pistons 202 and 204 as close to linear as possible because
the deviation
242 and 244 from longitudinal axis of reciprocating motion of pistons 202 and
204 causes
noise, reduction of efficiency, increase of friction to the wall of cylinder,
increase of side-
load, and low durability of the parts. The alignment of the cylinders 206 and
208 and the
arrangement of crankshaft 214, piston rods 224 and 228, link rods 226 and 230,
and
connecting rod 222, hence, may influence on, amongst other things, the
efficiency and/or
the volume of the device. For the purpose of increasing the linearity of the
piston motion as
mentioned, the pistons (shown as 202 and 204 in FIGS. 2-4) are preferably as
close to the
side of the respective cylinders 206 and 208 as possible.
In another embodiment reducing angular deviation of link rods, link rods 226
and
230 substantially linearly reciprocate along longitudinal axis of motion of
respective pistons
204 and 202 to decrease the angular deviation and thus to decrease the side
load applied to
each piston 204 and 202. The angular deviation defines the deviation of the
link rod 226 or
230 from the longitudinal axis of the piston 204 or 202. Numerals 244 and 242
designate
the angular deviation of the link rods 226 and 230, as shown in FIG. 2.
Therefore, the
position of coupling assembly 212 influences the angular displacement of the
link rod 226,
based on the length of the distance between the end pivot 232 and the rocker
pivot 218 of
the rocking beam 216. Thus, the position of the coupling assemblies may be
such that the
angular displacement of the link rod 226 is reduced. For the link rod 230, the
length of the
coupling assembly 210 also may be determined and placed to reduce the angular
displacement of the link rod 230, based on the length of the distance between
the end pivot
236 and the rocker pivot 218 of the rocking beam 216. Therefore, the length of
the link
rods 226 and 230, the length of coupling assemblies 212 and 210, and the
length of the
rocking beam 216 are significant parameters that greatly influence and/or
determine the
angular deviation of the link rods 226 and 230 as shown in FIG. 2.
The exemplary embodiment has a straight rocking beam 216 having the end points
232 and 236, the rocker pivot 218, and the connecting pivot 240 along the same
axis.
However, in other embodiments, the rocking beam 216 may be bent, such that
pistons may
be placed at angles to each other, as shown in FIGS. 5C and 5D.
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Referring now to FIGS. 2-4 and FIGS. 7A-7B, in some embodiments of the
coupling
assembly, the coupling assemblies 212 and 210, may include a flexible link rod
that is
axially stiff but flexible in the rocking beam 216 plane of motion between
link rods 226 and
230, and pistons 204 and 202, respectively. In this embodiment, at least one
portion, the
flexure (shown as 700 in FIGS. 7A and 7B), of link rods 226 and 230 is
elastic. The flexure
700 acts as a coupling means between the piston rod and the link rod. The
flexure 700 may
absorb the crank-induced side loads of the pistons more effectively, thus
allowing its
respective piston to maintain linear longitudinal movement inside the piston's
cylinder.
This flexure 700 allows small rotations in the plane of the rocking beam 216
between the
.. link rods 226 and 230 and pistons 204 or 202, respectively. Although
depicted in this
embodiment as flat, which increases the elasticity of the link rods 226 and
230, the flexure
700, in some embodiments, is not flat. The flexure 700 also may be constructed
near to the
lower portion of the pistons or near to the distal end of the link rods 226
and 230. The
flexure 700, in one embodiment, may be made of #D2 Tool Steel Hardened to 58-
62 RC. In
some embodiments, there may be more than one flexure (not shown) on the link
rod 226 or
230 to increase the elasticity of the link rods.
In alternate embodiment, the axes of the pistons in each cylinder housing may
extend in different directions, as depicted in FIGS. 5C and 5D. In the
exemplary
embodiment, the axes of the pistons in each cylinder housing are substantially
parallel and
preferably substantially vertical, as depicted in FIGS. 2 - 4, and FIGS. 5A
and 5B. FIGS.
5A-5D include various embodiments of the rocking beam drive mechanism
including like
numbers as those shown and described with respect to FIGS. 2-4. It will be
understood by
those skilled in that art that changing the relative position of the
connecting pivot 240 along
the rocking beam 216 will change the stroke of the pistons.
Accordingly, a change in the parameters of the relative position of the
connecting
pivot 240 in the rocking beam 216 and the length of the piston rods 224 and
228, link rods
230 and 226, rocking beam 216, and the position of rocker pivot 218 will
change the
angular deviation of the link rods 226 and 230, the phasing of the pistons 204
and 202, and
the size of the device 300 in a variety of manner. Therefore, in various
embodiments, a
wide range of piston phase angles and variable sizes of the engine may be
chosen based on
the modification of one or more of these parameters. In practice, the link
rods 224 and 228
of the exemplary embodiment have substantially lateral movement within from -
0.5 degree
to +0.5 degree from the longitudinal axis of the pistons 204 and 202. In
various other
embodiments, depending on the length of the link rod, the angle may vary
anywhere from
24
approaching 0 degrees to .75 degrees. However, in other embodiments, the angle
may be
higher including anywhere from approaching 0 to the approximately 20 degrees.
As the
link rod length increases, however, the crankcase/overall engine height
increases as well as
the weight of the engine.
One feature of the exemplary embodiment is that each piston has its link rod
extending substantially to the attached piston rod so that it is formed as a
coupling
assembly. In one embodiment, the coupling assembly 212 for the piston 204
includes a
piston rod 224, a link rod 226, and a coupling means 234 as shown in FIG. 2.
More
specifically, one proximal end of piston rod 224 is attached to the lower
portion of piston
204 and the distal end piston rod 224 is connected to the proximate end of the
link rod 226
by the coupling means 234. The distal end of the link rod 226 extends
vertically to the end
pivot 232 of the rocking beam 216. As described above, the coupling means 234
may be,
but is not limited to, a joint, hinge, coupling, or flexure or other means
known in the art. In
this embodiment, the ratio of the piston rod 224 and the link rod 226 may
determine the
angular deviation of the link rod 226 as mentioned above.
Referring now to FIG. 4, one embodiment of the engine is shown. Here the
pistons
202 and 204 of engine 300 operate between a hot chamber 404 and a cold chamber
406 of
cylinders 206 and 208 respectively. Between the two chambers there may be a
regenerator
408. The regenerator 408 may have variable density, variable area, and, in
some
embodiments, is made of wire. The varying density and area of the regenerator
may be
adjusted such that the working gas has substantially uniform flow across the
regenerator
408. Various embodiments of the regenerator 408 are discussed in detail below,
and in U.S.
Patents No. 6,591,609, issued July 17, 2003, to Kamen et al., and No.
6,862,883, issued
March 8, 2005, to Kamen et at.. When the working gas passes through the hot
chamber 404,
a heater head 410 may heat the gas causing the gas to expand and push pistons
202 and 204
towards the cold chamber 406, where the gas compresses. As the gas compresses
in the cold
chamber 406, pistons 202 and 204 may be guided back to the hot chamber to
undergo the
Stirling cycle again. The heater head 410 may have one of several forms
including a pin
head, a fin head, a folded fin head, or heater tubes as shown in FIG. 4 or any
other heater
head embodiment known, including, but not limited to, those described below.
Various
embodiments of heater head 410 are discussed in detail below, and in U.S.
Patents No.
6,381,958, issued May 7, 2002, to Kamen et al., No. 6,543,215, issued April 8,
2003, to
Langenfelcl et al.. No. 6,966,182, issued November 22, 2005, to Kamen et at,
and No.
7,308,787, issued December 18, 2007, to LaRocque etal., and in U.S. Patent
Application
CA 2942884 2017-12-19
25
Serial No. 13/447,990, filed April 16, 2012 and entitled Stirling Cycle
Machine (Attorney
Docket No. 184).
In some embodiments, a cooler 412 may be positioned alongside cylinders 206
and
208 to further cool the gas passing through to the cold chamber 406. Various
embodiments
of cooler 412 are discussed in detail in the proceeding sections, and in U.S.
Patent No.
7,325,399, issued Feb. 5, 2008, to Strimling et al.
In some embodiments, at least one piston seal 414 may be positioned on pistons
202
and 204 to seal the hot section 404 off from the cold section 406.
Additionally, at least one
piston guide ring 416 may be positioned on pistons 202 and 204 to help guide
the pistons'
motion in their respective cylinders. Various embodiments of piston seal 414
and guide ring
416 are described in detail below, and in U.S. Patent Application Ser. No.
10/175,502, filed
June 19, 2002, published February 6, 2003 (now abandoned).
In some embodiments, at least one piston rod seal 418 may be placed against
piston
rods 224 and 228 to prevent working gas from escaping into the crankcase 400,
or
alternatively into airlock space 420. The piston rod seal 418 may be an
elastomer seal, or a
spring-loaded seal. Various embodiments of the piston rod seal 418 are
discussed in detail
below.
In some embodiments, the airlock space may be eliminated, for example, in the
rolling diaphragm and/or bellows embodiments described in more detail below.
In those
cases, the piston rod seals 224 and 228 seal the working space from the
crankcase.
In some embodiments, at least one rolling diaphragm/bellows 422 may be located
along piston rods 224 and 228 to prevent airlock gas from escaping into the
crankcase 400.
Various embodiments of rolling diaphragm 422 are discussed in more detail
below.
Although FIG. 4 shows a cross section of engine 300 depicting only two pistons
and
one rocking beam drive, it is to be understood that the principles of
operation described
herein may apply to a four cylinder, double rocking beam drive engine, as
designated
generally by numeral 800 in FIG. 10B.
Piston Operation
Referring now to FIG. 8 that shows the operation of pistons 802, 804, 806, and
808
during one revolution of crankshaft 814. With a '/4 revolution of crankshaft
814, piston 802
is at the top of its cylinder, otherwise known as top dead center, piston 806
is in upward
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midstroke, piston 804 is at the bottom of its cylinder, otherwise known as
bottom dead
center, and piston 808 is in downward midstroke. With a 1/2 revolution of
crankshaft 814,
piston 802 is in downward midstroke, piston 806 is at top dead center, piston
804 is in
upward midstroke, and piston 808 is at bottom dead center. With 3/4 revolution
of crankshaft
814, piston 802 is at bottom dead center. piston 806 is in downward midstroke,
piston 804 is
at top dead center, and piston 808 is in upward midstroke. Finally, with a
full revolution of
crankshaft 814, piston 802 is in upward midstroke, piston 806 is at bottom
dead center,
piston 804 is in downward midstroke, and piston 808 is at top dead center.
During each 1/4
revolution, there is a 90 degree phase difference between pistons 802 and 806,
a 180 degree
phase difference between pistons 802 and 804, and a 270 degree phase
difference between
pistons 802 and 808. FIG. 9A illustrates the relationship of the pistons being
approximately
90 degrees out of phase with the preceding and succeeding piston.
Additionally, FIG. 8
shows the exemplary embodiment machine means of transferring work. Thus, work
is
transferred from piston 802 to piston 806 to piston 804 to piston 808 so that
with a full
.. revolution of crankshaft 814, all pistons have exerted work by moving from
the top to the
bottom of their respective cylinders.
Refen-ing now to FIG. 8, together with FIGS. 9A-9C, illustrate the 90 degree
phase
difference between the pistons in the exemplary embodiment. Referring now to
FIG. 9A,
although the cylinders are shown in a linear path, this is for illustration
purposes only. In
the exemplary embodiment of a four cylinder Stirling cycle machine, the flow
path of the
working gas contained within the cylinder working space follows a figure eight
pattern.
Thus, the working spaces of cylinders 1200, 1202, 1204, and 1206 are connected
in a figure
eight pattern, for example, from cylinder 1200 to cylinder 1202 to cylinder
1204 to cylinder
1208, the fluid flow pattern follows a figure eight. Still referring to FIG.
9A, an unwrapped
.. view of cylinders 1200. 1202, 1204, and 1206, taken along the line B-B
(shown in FIG. 9C)
is illustrated. The 90 degree phase difference between pistons as described
above allows for
the working gas in the warm section 1212 of cylinder 1204 to be delivered to
the cold
section 1222 of cylinder 1206. As piston 802 and 808 are 90 degrees out of
phase, the
working gas in the warm section 1214 of cylinder 1206 is delivered to the cold
section 1216
of cylinder 1200. As piston 802 and piston 806 are also 90 degrees out of
phase, the
working gas in the warm section 1208 of cylinder 1200 is delivered to the cold
section 1218
of cylinder 1202. And as piston 804 and piston 806 are also 90 degrees out of
phase, so the
working gas in the warm section 1210 of cylinder 1202 is delivered to the cold
section 1220
of cylinder 1204. Once the working gas of a warm section of a first cylinder
enters the cold
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section of a second cylinder, the working gas begins to compress, and the
piston within the
second cylinder, in its down stroke, thereafter forces the compressed working
gas back
through a regenerator 1224 and heater head 1226 (shown in FIG. 9B), and back
into the
warm section of the first cylinder. Once inside the warm section of the first
cylinder, the gas
.. expands and drives the piston within that cylinder downward, thus causing
the working gas
within the cold section of that first cylinder to be driven through the
preceding regenerator
and heater head, and into the cylinder. This cyclic transmigration
characteristic of working
gas between cylinders 1200, 1202, 1204, and 1206 is possible because pistons
802, 804,
806, and 808 are connected, via drives 810 and 812, to a common crankshaft 814
(shown in
FIG. 8), in such a way that the cyclical movement of each piston is
approximately 90
degrees in advance of the movement of the proceeding piston, as depicted in
FIG. 9A.
Rolling Diaphragm, Metal Bellows, Airlock, and Pressure Regulator
In some embodiments of the Stirling cycle machine, lubricating fluid is used.
To
prevent the lubricating fluid from escaping the crankcase, a seal is used.
Referring now to FIGS. 10A-13B, some embodiments of the Stirling cycle machine
include a fluid lubricated rocking beam drive that utilizes a rolling
diaphragm 1300
positioned along the piston rod 1302 to prevent lubricating fluid from
escaping the
crankcase, not shown, but the components that are housed in the crankcase are
represented
as 1304, and entering areas of the engine that may be damaged by the
lubricating fluid. It is
beneficial to contain the lubricating fluid for if lubricating fluid enters
the working space,
not shown, but the components that are housed in the working space are
represented as
1306, it would contaminate the working fluid, come into contact with the
regenerator 1308,
and may clog the regenerator 1308. The rolling diaphragm 1300 may be made of
an
elastomer material, such as rubber or rubber reinforced with woven fabric or
non-woven
fabric to provide rigidity. The rolling diaphragm 1300 may alternatively be
made of other
materials, such as fluorosilicone or nitrile with woven fabric or non-woven
fabric. The
rolling diaphragm 1300 may also be made of carbon nanotubes or chopped fabric,
which is
non-woven fabric with fibers of polyester or KEVLARO, for example, dispersed
in an
elastomer. In the some embodiments, the rolling diaphragm 1300 is supported by
the top
seal piston 1328 and the bottom seal piston 1310. In other embodiments, the
rolling
diaphragm 1300 as shown in FIG. 10A is supported via notches in the top seal
piston 1328.
In some embodiments, a pressure differential is placed across the rolling
diaphragm
1300 such that the pressure above the seal 1300 is different from the pressure
in the
crankcase 1304. This pressure differential inflates seal 1300 and allows seal
1300 to act as a
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dynamic seal as the pressure differential ensures that rolling diaphragm
maintains its form
throughout operation. FIGS. 10A, and FIGS. 10C-10D illustrate how the pressure
differential effects the rolling diaphragm. The pressure differential causes
the rolling
diaphragm 1300 to conform to the shape of the bottom seal piston 1310 as it
moves with the
piston rod 1302, and prevents separation of the seal 1300 from a surface of
the piston 1310
during operation. Such separation may cause seal failure. The pressure
differential causes
the rolling diaphragm 1300 to maintain constant contact with the bottom seal
piston 1310 as
it moves with the piston rod 1302. This occurs because one side of the seal
1300 will
always have pressure exerted on it thereby inflating the seal 1300 to conform
to the surface
of the bottom seal piston 1310. In some embodiments, the top seal piston 1328
'rolls over'
the corners of the rolling diaphragm 1300 that are in contact with the bottom
seal piston
1310, so as to further maintain the seal 1300 in contact with the bottom seal
piston 1310. In
the exemplary embodiment, the pressure differential is in the range of 10 to
15 PSI. The
smaller pressure in the pressure differential is preferably in crankcase 1304,
so that the
rolling diaphragm 1300 may be inflated into the crankcase 1304. However, in
other
embodiments, the pressure differential may have a greater or smaller range of
value.
The pressure differential may be created by various methods including, but not
limited to, the use of the following: a pressurized lubrication system, a
pneumatic pump,
sensors, an electric pump, by oscillating the rocking beam to create a
pressure rise in the
crankcase 1304, by creating an electrostatic charge on the rolling diaphragm
1300, or other
similar methods. In some embodiments, the pressure differential is created by
pressurizing
the crankcase 1304 to a pressure that is below the mean pressure of the
working space 1306.
In some embodiments the crankcase 1304 is pressurized to a pressure in the
range of 10 to
15 PSI below the mean pressure of the working space 1306, however, in various
other
embodiments, the pressure differential may be smaller or greater. Further
detail regarding
the rolling diaphragm is included below.
Referring now to FIGS. 10C, and 11, however, another embodiment of the
Stirling
machine is shown, wherein airlock space 1312 is located between working space
1306 and
crankcase 1304. Airlock space 1312 maintains a constant volume and pressure
necessary to
create the pressure differential necessary for the function of rolling
diaphragm 1300 as
described above. In one embodiment, airlock 1312 is not absolutely sealed off
from working
space 1306, so the pressure of airlock 1312 is equal to the mean pressure of
working space
1306. Thus, in some embodiments, the lack of an effective seal between the
working space
29
and the crankcase contributes to the need for an airlock space. Thus, the
airlock space, in
some embodiments, may be eliminated by a more efficient and effective seal.
During operation, the working space 1306 mean pressure may vary so as to cause
airlock 1312 mean pressure to vary as well. One reason the pressure may tend
to vary is that
during operation the working space may get hotter, which in turn may increase
the pressure
in the working space, and consequently in the airlock as well since the
airlock and working
space are in fluid communication. In such a case, the pressure differential
between airlock
1312 and crankcase 1304 will also vary, thereby causing unnecessary stresses
in rolling
diaphragms 1300 that may lead to seal failure. Therefore, some embodiments of
the
machine, the mean pressure within airlock 1312 is regulated so as to maintain
a constant
desired pressure differential between airlock 1312 and crankcase 1304, and
ensuring that
rolling diaphragms 1300 stay inflated and maintains their form. In some
embodiments, a
pressure transducer is used to monitor and manage the pressure differential
between the
airlock and the crankcase, and regulate the pressure accordingly so as to
maintain a constant
pressure differential between the airlock and the crankcase. Various
embodiments of the
pressure regulator that may be used are described in further detail below, and
in U.S. Patent
No. 7,310,945, issued Dec. 25, 2007, to Gursla et al..
A constant pressure differential between the airlock 1312 and crankcase 1304
may
be achieved by adding or removing working fluid from airlock 1312 via a pump
or a release
valve. Alternatively, a constant pressure differential between airlock 1312
and crankcase
1304 may be achieved by adding or removing working fluid from crankcase 1304
via a
pump or a release valve. The pump and release valve may be controlled by the
pressure
regulator. Working fluid may be added to airlock 1312 (or crankcase 1304) from
a separate
source, such as a working fluid container, or may be transferred over from
crankcase 1304.
Should working fluid be transferred from crankcase 1304 to airlock 1312, it
may be
desirable to filter the working fluid before passing it into airlock 1312 so
as to prevent any
lubricant from passing from crankcase 1304 into airlock 1312, and ultimately
into working
space 1306, as this may result in engine failure.
In some embodiments of the machine, crankcase 1304 may be charged with a fluid
having different thermal properties than the working fluid. For example, where
the working
gas is helium or hydrogen, the crankcase may be charged with argon. Thus, the
crankcase is
pressurized. In some embodiments, helium is used, but in other embodiments,
any inert
gas, as described herein, may be used. Thus, the crankcase is a wet
pressurized crankcase in
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the exemplary embodiment. In other embodiments where a lubricating fluid is
not used, the
crankcase is not wet.
In the exemplary embodiments, rolling diaphragms 1300 do not allow gas or
liquid
to pass through them, which allows working space 1306 to remain dry and
crankcase 1304
5 to be wet sumped with a lubricating fluid. Allowing a wet sump crankcase
1304 increases
the efficiency and life of the engine as there is less friction in rocking
beam drives 1316. In
some embodiments, the use of roller bearings or ball bearings in drives 1316
may also be
eliminated with the use of lubricating fluid and rolling diaphragms 1300. This
may further
reduce engine noise and increase engine life and efficiency.
10 FIGS. 12A-12E show cross sections of various embodiments of the rolling
diaphragm (shown as 1400, 1410, 1412, 1422 and 1424) configured to be mounted
between
top seal piston and bottom seal piston (shown as 1328 and 1310 in FIG. 10A),
and between
a top mounting surface and a bottom mounting surface (shown as 1320 and 1318
in FIG.
10A). In some embodiments, the top mounting surface may be the surface of an
airlock or
15 working space, and the bottom mounting surface may be the surface of a
crankcase.
FIG. 12A shows one embodiment of the rolling diaphragm 1400, where the rolling
diaphragm 1400 includes a flat inner end 1402 that may be positioned between a
top seal
piston and a bottom seal piston, so as to form a seal between the top seal
piston and the
bottom seal piston. The rolling diaphragm 1400 also includes a flat outer end
1404 that may
20 be positioned between a top mounting surface and a bottom mounting
surface, so as to form
a seal between the top mounting surface and the bottom mounting surface. FIG.
12B shows
another embodiment of the rolling diaphragm, wherein rolling diaphragm 1410
may include
a plurality of bends 1408 leading up to flat inner end 1406 to provide for
additional support
and sealing contact between the top seal piston and the bottom seal piston.
FIG. 12C shows
25 another embodiment of the rolling diaphragm, wherein rolling diaphragm
1412 includes a
plurality of bends 1416 leading up to flat outer end 1414 to provide for
additional support
and sealing contact between the top mounting surface and the bottom mounting
surface.
FIG. 12D shows another embodiment of the rolling diaphragm where rolling
diaphragm 1422 includes a bead along an inner end 1420 thereof, so as to form
an `o-ring'
30 type seal between a top seal piston and a bottom seal piston, and a bead
along an outer end
1418 thereof, so as to form an `o-ring' type seal between a bottom mounting
surface and a
top mounting surface. FIG. 12E shows another embodiment of the rolling
diaphragm,
wherein rolling diaphragm 1424 includes a plurality of bends 1428 leading up
to beaded
inner end 1426 to provide for additional support and sealing contact between
the top seal
31
piston and the bottom seal piston. Rolling diaphragm 1424 may also include a
plurality of
bends 1430 leading up to beaded outer end 1432 to provide for additional
support and
sealing contact between the top seal piston and the bottom seal piston.
Although FIGS. 12A through 12E depict various embodiments of the rolling
diaphragm, it is to be understood that rolling diaphragms may be held in place
by any other
mechanical means known in the art. Alternatively, the rolling diaphragm may be
replaced
by metal bellows as disclosed in U.S. Patent Application Serial No.
13/447,990, filed April
16, 2012 and entitled Stirling Cycle Machine (Attorney Docket No. 184).
Rolling Diaphragm and/or Bellows Embodiments
Various embodiments of the rolling diaphragm and/or bellows, which function to
seal, are described above. Further embodiments will be apparent to those of
skill in the art
based on the description above and the additional description below relating
to the
parameters of the rolling diaphragm and/or bellows.
In some embodiments, the pressure atop the rolling diaphragm or bellows, in
the
airlock space or airlock area (both terms are used interchangeably), is the
mean-working-gas
pressure for the machine, which, in some embodiments is an engine, while the
pressure
below the rolling diaphragm and/or bellows, in the crankcase area, is
ambient/atmospheric
pressure. In these embodiments, the rolling diaphragm and/or bellows is
required to operate
with as much as 3000psi across it (and in some embodiments, up to 1500psi or
higher). In
this case, the rolling diaphragm and/or bellows seal forms the working gas
(helium,
hydrogen, or otherwise) containment barrier for the machine (engine in the
exemplary
embodiment). Also, in these embodiments, the need for a heavy, pressure-rated,
structural
vessel to contain the bottom end of the engine is eliminated, since it is now
required to
simply contain lubricating fluid (oil is used as a lubricating fluid in the
exemplary
embodiment) and air at ambient pressure, like a conventional internal
combustion ("IC")
engine.
The capability to use a rolling diaphragm and/or bellows seal with such an
extreme
pressure across it depends on the interaction of several parameters. Referring
now to FIG.
13A, an illustration of the actual load on the rolling diaphragm or bellows
material is
shown. As shown, the load is a function of the pressure differential and the
annular gap
area for the installed rolling diaphragm or bellows seal.
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Region 1 represents the portions of the rolling diaphragm and/or bellows that
are in
contact with the walls formed by the piston and cylinder. The load is
essentially a tensile
load in the axial direction, due to the pressure differential across the
rolling diaphragm
and/or bellows. This tensile load due to the pressure across the rolling
diaphragm and/or
bellows can be expressed as:
= Pd * Aa
Where
Lt = Tensile Load and
Pa = Pressure Differential
A, = Annular Area
and
Aa = p / 4 * (D2- d2)
Where
D = Cylinder Bore and
d = Piston Diameter
The tensile component of stress in the bellows material can be approximated
as:
St = Lt / (p * (D+d) * tb)
Which reduces to:
St = Pd /4 * (D-d) / tb
Later, we will show the relationship of radius of convolution, Rc, to Cylinder
bore
(D) and Piston Diameter (d) to be defined as:
Rc = (D-d)/4
So, this formula for St reduces to its final form:
St = Pd * Rc / tb
Where
tb = thickness of bellows material
Still referring to FIG. 13A, Region 2 represents the convolution. As the
rolling
diaphragm and/or bellows material turns the corner, in the convolution, the
hoop stress
imposed on the rolling diaphragm and/or bellows material may be calculated.
For the
section of the bellows forming the convolution, the hoop component of stress
can be closely
approximated as:
Sh = Pd * Rc tb
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The annular gap that the rolling diaphragm and/or bellows rolls within is
generally
referred to as the convolution area. The rolling diaphragm and/or bellows
fatigue life is
generally limited by the combined stress from both the tensile (and hoop)
load, due to
pressure differential, as well as the fatigue due to the bending as the fabric
rolls through the
convolution. The radius that the fabric takes on during this 'rolling' is
defined here as the
radius of convolution, Rc.
(D-d)/4
The bending stress, Sb, in the rolling diaphragm and/or bellows material as it
rolls
through the radius of convolution, Rc, is a function of that radius, as well
as the thickness of
the materials in bending. For a fiber-reinforced material, the stress in the
fibers themselves
(during the prescribed deflection in the exemplary embodiments) is reduced as
the fiber
diameter decreases. The lower resultant stress for the same level of bending
allows for an
increased fatigue life limit. As the fiber diameter is further reduced,
flexibility to decrease
the radius of convolution Rc is achieved, while keeping the bending stress in
the fiber under
its endurance limit. At the same time, as Rc decreases, the tensile load on
the fabric is
reduced since there is less unsupported area in the annulus between the piston
and cylinder.
The smaller the fiber diameter, the smaller the minimum Rc , the smaller the
annular area,
which results in a higher allowable pressure differential.
For bending around a prescribed radius, the bending moment is approximated by:
M=E*I/R
Where:
M = Bending Moment
E = Elastic Modulus
I = Moment of Inertia
R = Radius of Bend
Classical bending stress, Sb , is calculated as:
Sb =Mflr/I
Where:
Y = Distance above neutral axis of bending
Substituting yields:
Sb = (E * I / R) * Y /
Sb =E*Y/R
Assuming bending is about a central neutral axis:
Ymax = tb / 2
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Sb = E * tb / (2 * R)
In some embodiments, rolling diaphragm and/or bellows designs for high cycle
life
are based on geometry where the bending stress imposed is kept about one order
of
magnitude less than the pressure-based loading (hoop and axial stresses).
Based on the
equation: Sb = E * tb / (2 * R), it is clear that minimizing tb in direct
proportion to Rc
should not increase the bending stress. The minimum thickness for the
exemplary
embodiments of the rolling diaphragm and/or bellows material or membrane is
directly
related to the minimum fiber diameter that is used in the reinforcement of the
elastomer.
The smaller the fibers used, the smaller resultant Rc for a given stress
level.
Another limiting component of load on the rolling diaphragm and/or bellows is
the
hoop stress in the convolution (which is theoretically the same in magnitude
as the axial
load while supported by the piston or cylinder). The governing equation for
that load is as
follows:
Sh = Pd * Rc/tb
Thus, if Rc is decreased in direct proportion to tb, then there is no increase
of stress
on the membrane in this region. However, if this ratio is reduced in a manner
that decreases
Rc to a greater ratio than tb then parameters must be balanced. Thus,
decreasing tb with
respect to Rc requires the rolling diaphragm and/or bellows to carry a heavier
stress due to
pressure, but makes for a reduced stress level due to bending. The pressure-
based load is
essentially constant, so this may be favorable---since the bending load is
cyclic, therefore it
is the bending load component that ultimately limits fatigue life.
For bending stress reduction, tb ideally should be at a minimum, and Rc
ideally
should be at a maximum. E ideally is also at a minimum. For hoop stress
reduction, Rc
ideally is small, and tb ideally is large.
Thus, the critical parameters for the rolling diaphragm and/or bellows
membrane
material are:
E, Elastic Modulus of the membrane material;
tb, membrane thickness (and/or fiber diameter);
Sut, Ultimate tensile strength of the rolling diaphragm and/or bellows; and
Slcf, The limiting fatigue strength of the rolling diaphragm and/or bellows.
Thus, from E, tb and Sut, the minimum acceptable Rc may be calculated. Next,
using Rc, Slcf, and tb, the maximum Pd may be calculates. Rc may be adjusted
to shift the
bias of load (stress) components between the steady state pressure stress and
the cyclic
35
bending stress. Thus, the ideal rolling diaphragm and/or bellows material is
extremely thin,
extremely strong in tension, and very limber in flexion.
Thus, in some embodiments, the rolling diaphragm and/or bellows material
(sometimes referred to as a "membrane"), is made from carbon fiber nanotubes.
However,
additional small fiber materials may also be used, including, but not limited
to nanotube
fibers that have been braided, nanotube untwisted yarn fibers, or any other
conventional
materials, including but not limited to KEVLAR, glass, polyester, synthetic
fibers and any
other material or fiber having a desirable diameter and/or other desired
parameters as
described in detail above.
Piston Seals and Piston Rod Seals
Referring now to FIG. 11, an embodiment of the machine is shown wherein an
engine 1326, such as a Stirling cycle engine, includes at least one piston rod
seal 1314, a
piston seal 1324, and a piston guide ring 1322, (shown as 1616 in FIG. 14).
Various
embodiments of the piston seal 1324 and the piston guide ring 1322 are further
discussed
below, and in U.S. Patent Application Ser. No. 10/175,502 (now abandoned).
FIG. 14 shows a partial cross section of the piston 1600, driven along the
central
axis 1602 of cylinder, or the cylinder 1604. The piston seal (shown as 1324 in
FIG. 11) may
include a seal ring 1606, which provides a seal against the contact surface
1608 of the
cylinder 1604. The contact surface 1608 is typically a hardened metal
(preferably 58-62
RC) with a surface finish of 12 RMS or smoother. The contact surface 1608 may
be metal
which has been case hardened, such as 8260 hardened steel, which may be easily
case
hardened and may be ground and/or honed to achieve a desired finish. The
piston seal may
also include a backing ring 1610, which is sprung to provide a thrust force
against the seal
ring 1606 thereby providing sufficient contact pressure to ensure sealing
around the entire
outward surface of the seal ring 1606. The seal ring 1606 and the backing ring
1610 may
together be referred to as a piston seal composite ring. In some embodiments,
the at least
one piston seal may seal off a warm portion of cylinder 1604 from a cold
portion of cylinder
1604.
Referring now to FIG. 15, some embodiments include a piston rod seal (shown as
1314 in FIG. 11) mounted in the piston rod cylinder wall 1700, which, in some
embodiments, may include a seal ring 1706, which provides a seal against the
contact
surface 1708 of the piston rod 1604 (shown as 1302 in FIG. 11). The contact
surface 1708
in some embodiments is a hardened metal (preferably 58-62 RC) with a surface
finish of 12
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RMS or smoother. The contact surface 1708 may be metal which has been case
hardened,
such as 8260 hardened steel, which may be easily case hardened and may be
ground and/or
honed to achieve a desired finish. The piston seal may also include a backing
ring 1710,
which is sprung to provide a radial or hoop force against the seal ring 1706
thereby
providing sufficient contact hoop stress to ensure sealing around the entire
inward surface of
seal ring 1706. The seal ring 1706 and the backing ring. 1710 may together be
referred to as
a piston rod seal composite ring.
In some embodiments, the seal ring and the backing ring may be positioned on a
piston rod, with the backing exerting an outward pressure on the seal ring,
and the seal ring
may come into contact with a piston rod cylinder wall 1702. These embodiments
require a
larger piston rod cylinder length than the previous embodiment. This is
because the contact
surface on the piston rod cylinder wall 1702 will be longer than in the
previous
embodiment, where the contact surface 1708 lies on the piston rod itself. In
yet another
embodiment, piston rod seals may be any functional seal known in the art
including, but not
limited to, an o-ring, a graphite clearance seal, graphite piston in a glass
cylinder, or any air
pot, or a spring energized lip seal. In some embodiments, anything having a
close clearance
may be used, in other embodiments, anything having interference, for example,
a seal, is
used. In the exemplary embodiment, a spring energized lip seal is used. Any
spring
energized lip seal may be used, including those made by BAL SEAL Engineering,
Inc..
Foothill Ranch, CA. In some embodiments, the seal used is a BAL SEAL Part
Number
X558604.
The material of the seal rings 1606 and 1706 is chosen by considering a
balance
between the coefficient of friction of the seal rings 1606 and 1706 against
the contact
surfaces 1608 and 1708, respectively, and the wear on the seal rings 1606 and
1706 it
engenders. In applications in which piston lubrication is not possible, such
as at the high
operating temperatures of a Stirling cycle engine, the use of engineering
plastic rings is
used. The embodiments of the composition include a nylon matrix loaded with a
lubricating
and wear-resistant material. Examples of such lubricating materials include
PTFE/silicone,
PTFE, graphite. etc. Examples of wear-resistant materials include glass fibers
and carbon
fibers. Examples of such engineering plastics are manufactured by LNP
Engineering
Plastics, Inc. of Exton, PA. Backing rings 1610 and 1710 is preferably metal.
The fit between the seal rings 1606 and 1706 and the seal ring grooves 1612
and
1712, respectively, is preferably a clearance fit (about 0.002"), while the
fit of the backing
rings 1610 and 1710 is preferably a looser fit, of the order of about 0.005"
in some
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embodiments. The seal rings 1606 and 1706 provide a pressure seal against the
contact
surfaces 1608 and 1708, respectively, and also one of the surfaces 1614 and
1714 of the seal
ring grooves 1612 and 1712, respectively, depending on the direction of the
pressure
difference across the rings 1606 and 1706 and the direction of the piston 1600
or the piston
rod 1704 travel.
Referring again to FIG. 14, at least one guide ring 1616 may also be provided,
in
accordance with some embodiments, for bearing any side load on piston 1600 as
it moves
up and down the cylinder 1604. Guide ring 1616 is also preferably fabricated
from an
engineering plastic material loaded with a lubricating material. A perspective
view of guide
ring 1616 is shown inFIGS. 16A, 16B. An overlapping joint 2100 is shown and
may be
diagonal to the central axis of guide ring 1616. The guide ring 1616 and other
guide
elements located on the moving piston or on matting stationary parts can bear
significant
side loads that are larger than the side loads supported by the pressure seal
sand the
associated backing ring. In a preferred embodiment, the backing ring 1610
exerts less than
15 psi, when fully compressed. The backing ring 1610 and pressure seal 1606
can bear a
significant side load and thus are not designed to guide the piston 1600.
Referring now to Fig. 68A, which shows a cross-section of an embodiment of a
Stirling
engine and burner, in some embodiments, the length of time of performance,
i.e., the life, of
the rod seals on the piston rod may be extended and/or maximized by reducing
and/or
minimizing the axial misalignment between the rod seals and the piston rod. In
some
embodiments, a rod seal may float radially when the pressure difference
between the
workspace and air lock is low, while forming a gas seal when the pressure
difference
between the workspace and air lock is high.
A rod seal assembly comprising a cylinder-gland housing between the working
space and
the air lock space is configured to receive the reciprocating piston rod that
is disposed
within the workspace and the airlock space. A floating bushing is configured
to move
axially and radially within the cylinder gland and is disposed coaxially
around the
reciprocating rod. A rod seal is configured to seal the outside diameter of
the reciprocating
rod relative to an inside surface of the floating bushing and at least one
stationary bushing
fixed within the housing that may form a seal with the floating bushing to
limit the flow of
gas between the workspace and the airlock. The rod seal may be a spring
energized seal.
The floating bushing may further comprises a circumferential flange on the
outside surface
that is configured to extend into the annular space formed by two stationary
bushings and
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form a seal with one of the stationary bushings in the presence of a pressure
difference
between the workspace and the airlock.
The rod seal assembly may also comprise a scraper ring located coaxially
around the
piston rod and located within the cylinder-gland housing between the floating
seal and the
workspace. The cylinder-gland housing may include a port connecting the
workspace to an
annular gap around the reciprocating rod between the scraper ring and the
floating seal, that
minimizes the pressure difference across the scraper ring. The rod seal
assembly may also
comprise a magnetic particle trap between the scraper ring and the floating
seal.
In another embodiment, the rod seal assembly comprising a cylinder-gland
housing between
the working space and the air lock space is configured to receive the
reciprocating piston
rod that is disposed within the workspace and the airlock space. A floating
clearance
bushing configured to move axially and radially within the housing and is
disposed
coaxially around the reciprocating piston rod and forms a clearance seal with
the piston rod.
The cylinder-gland contains at least one stationary annular element that is
fixed within the
housing and configured to form a face seal with the floating clearance
bushing.
Referring now to FIGS. 68A-69F, a piston rod seal unit 13750 is shown that
allows
the rod seal 13770 to move radially in order to minimize forces on and
maximize the life of
the rod seal 13770. In addition, the piston rod seal unit 13750 includes
components to
protect the rod seals 13770 from particles. In various embodiments, the piston
rod seal units
13750 mount in the duct plate 13715 of the Stirling engine drive and provides
gas seals
between the working spaces under the pistons 13738 from the airlock 13736 on
each piston
rod 13744. The piston rod seal unit 13750 is presented in FIGS. 69A-69C with
the piston
rod 13744 removed for clarity, wherein the rod seal 13770 is mounted in a
floating rod seal
assembly 13760 and one or more floating rod seal assemblies 13760 are mounted
in the
cylinder gland 13752. The pressure in the workspaces may vary + 300 psi about
the airlock
pressure which requires that the rod seals 13770 minimize gas leaks in both
directions. A
small amount of leakage between the workspace 13738 and the airlock 13736 is
tolerable
and even desirable to allow the average pressure of each of the multiple
workspaces 13738
to equalize with the pressure in the single airlock 13736 and thereby with
each other. A
large gas leak across any of the piston rod seal units 13750 will reduce the
pressure swing in
the workspace and thereby the power and efficiency of the Stirling engine. The
sliding
motion between the piston rod 13744 and the rod seal 13770 may result in wear
and
abrasion of the seal surface that leads to leaks, seal failure and reduced
engine power and
efficiency.
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Still referring to FIGS. 68A-69C, to minimize the wear, the rod seal 13770, in
some
embodiments, may be advantageously located concentric with the piston rod
13744 so that
the active sealing surfaces of rod seal 13770 are uniformly pressed against
piston rod 13744.
In some embodiments, and as shown in FIG. 68A, the location and motion path of
the
piston rod 13744 may be constrained by the crosshead bearing 13746 at one end
and the
piston guide ring 13742 in the cylinder at the other end. Misalignment or
radial movement
of either the crosshead bearing 13746 or the piston guide ring 13742 may
result in the piston
rod 13744 not being centered in mounting structure for the rod seal 13770
(e.g. the cylinder
gland 13752) or in slight movement of the piston rod 13744 during the stroke.
The floating
rod seal assembly 13760 (FIG. 69B) allows the rod seal 13770 to move radially
in order to
center itself on the piston rod 13744 or minimize the variation of radial
forces on the sealing
surfaces around the circumference of the rod seal 13770. The floating rod seal
assembly
13760 may also maintain a significant seal, but in some embodiments, not a
perfect seal, to
the flow of working gas between the workspace 13738 and airlock 13736, while
allowing
the radial movement of the piston rod seal 13370.
Still referring to FIGS. 68A-69C . the piston rod seal units 13750 including
one or
more floating rod seal assemblies 13760 mounted in the cylinder gland 13752,
may be
bolted to the duct plate 13715. In some embodiments, the piston rod seal units
13750 may
be located to align with the heater heads 13712 and or the engine block 13741
that guides
the crosshead bearings 13746 via structural features on the cylinder gland
13752 and
structural elements that assure proper alignment of duct plate 13715 with the
cooler plate
13717 and engine block 13741. In some embodiments, the diameter 13784 on the
cylinder
gland 13752 mates with a counter-bored diameter on the duct plate 13715.
Still referring to FIGS. 68A-69C, in some embodiments, the piston rod seal
unit
13750 includes: a housing 13754, into which one or more floating rod seal
assemblies
13760 are pressed and axially constrained therein, a scraper ring 13778, a
particle trap
13780 and a port 13782 to minimize the pressure difference across the scraper
ring 13778.
The floating rod seal assemblies 13760 include an outer ring 13762, at least
one bushing
13764, (wherein, in some embodiments, may include two bushings 13764), a
floating
bushing 13766 and the rod seal 13770. In some embodiments, the elements of the
floating
rod seal assembly 13760 may be assembled by pressing a first bushing 13764
into the outer
ring 13762 until the one end of the bushing 13764 is flush with one end of the
outer ring
13762; placing the floating bushing 13766 in the outer ring 13762 on the non-
flush end of
the first bushing 13764; pressing the second bushing 13764 into the outer ring
13762 unit
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one end of the second bushing 13764 is flush with one end of the outer ring
13762 and the
floating bushing 13766 is captured between the two bushings 13764; and
pressing the rod
seal 13770 into the floating bushing 13766
Still referring to FIGS. 68A-69C, in various embodiments, the elements of the
5 floating rod seal assembly 13760 may be sized to allow the floating
bushing 13766 to move
axially and radially, while rigidly holding the bushings 13764 relative to the
outer ring
13762. In various embodiments, the outside diameter of the bushings 13764 and
inner
diameter of the outer ring 13762 may be sized to provide a light interference
fit, so that the
bushings do not move relative outer ring 13762. In some embodiments, the
outside
10 diameter of the bushings 13764 and inner diameter of the outer ring
13762 may be sized to
provide a location fit. In various embodiments, the heights of the outer ring
13762 and the
two bushings 13764 and the axial thickness of the rib 13766A may be sized so
that the axial
gap between the assembled bushings 13764 is larger than the axial thickness of
the rib
13766A. The resulting axial gap 13768 may range from 0.001 to 0.02 inches. In
some
15 embodiments, the axial gap 13768 ranges from about 0.002 to about 0.004
inches. In
various embodiments, the outside diameter of the outer ring 13762 and the
inner diameter of
the housing 13754 are selected to provide leak tight seal. In various
embodiments, the
outside diameter of the outer ring 13762 and the inner diameter of housing
13754 may be
selected to have a light interference fit. In another embodiment, the outside
diameter of the
20 outer ring 13762 and the inner diameter of housing 13754 may be selected
to have a
location fit.
Still referring to FIGS. 68A-69C, in various embodiments, the outer ring 13762
and
housing 13754 may be made from metal. In various embodiments, the bushings
13764,
13766 may be made from metal with high wear resistance. In various
embodiments, the
25 cylinder gland 13752 may be made of metal or high strength plastics that
show resistance to
cracking and fatigue failure. In some embodiments, the cylinder gland 13752
may be 4140
steel. In various embodiments, the bushings 13764, 13766, may be made of a
wear resistant
steel such as P20 (i.e. vacuum treated 4140 steel) or a metal with similar
strength and wear
resistances. In various embodiments, the outer ring 13762 and housing may be
made from
30 4140 or 4340 steel that has been treated to achieve a Rockwell Hardness
of 28-32 or a metal
with similar strength and hardness. The surfaces on the floating bushing 13766
and the
bushings 13764 that contact each other (e.g. 13769 in FIG. 69B) provide a
metal to metal
seal. In various embodiments, the contacting surfaces of the floating bushing
13766 and
bushings 13764 may have a smooth surface finish. In some embodiments, the
surface finish
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may be less than 32 microinches RA. In another example the surface finish is
less than 16
microinches RA.
Still referring to FIGS. 68A-69C, in various embodiments, the rod seal 13770
may
be a spring energized seal. In some embodiments, coiled springs 13770A wrapped
around
the bearing surface may urge the lips of the seal 13770B toward the piston rod
13744. In
some embodiments, the bearing surface may be composed of a composite polymer.
In some
embodiments, the bearing material may be a PFTE composite. In some
embodiments, the
rod seal 13770 may include an o-ring 13770C around the outside diameter to
provide a seal
to axial flow of gas on the outside diameter of the rod seal 13370. In some
embodiments,
the seal may be graphite or a graphite impregnated with antimony and or may be
made of
any material and/or any combination and/or composite of material. In some
embodiments,
the seal may be made of antimony impregnated graphite which may be from the
SGL
Carbon Company of Germany. In some embodiments, the rod seal may be a part
supplied
by CoorsTek of Golden Colorado or Bal Seal of Foothill Ranch, California.
Still referring to FIGS. 68A-69C in some embodiments, the floating rod seal
assemblies may be assembled and axially constrained in the housing 13754
following a
method of: pressing the first floating rod seal assembly 13760 into the
housing 13754 from
the flanged end 13754A until it is flush with the flanged end 13754A; pressing
the washer
13758 into the housing 13754 from the flanged end 13754A until it is flush
with the flanged
end 13754A; placing a second washer 13758 on other end of the first floating
rod seal
assembly 13760; pressing the second floating rod seal assembly 13760 into the
housing
13754 from threaded end 13754B place; placing the third washer 13758 on the
second
floating rod seal assembly 13760; and threading the fitting 13776 into the
housing 13754 to
axially secure the assemblies and washer 13760, 13758. In some embodiments,
the housing
13754 may be secured to a rigid surface while the fitting 13776 is threaded
into the housing
13754 and tightened to a predetermined torque. In various embodiments, the
floating road
seal assemblies 13760, washers 13758 and fitting 13776 may be assembled in
using various
methods that would be evident to one skilled in the art.
Still referring to FIGS. 68A-69C in some embodiments, the piston rod seal unit
13750 may be assembled by pressing or inserting some or all of the following
into the
cylinder g1and13752: housing 13754 with the floating rod seal assemblies
13760; and/or the
particle trap 13780; and/or a scraper ring 13778. In various embodiments, an
end plate
13756 may be bolted to the bottom of the cylinder gland 13752 to capture the
floating rod
seal assembles 13760, particle trap 13780 and/or scraper ring 13778. In
various
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embodiments, the port 13782 may be threaded into the side of the cylinder
gland between
the scraper ring 13778 and the housing 13754 containing the floating seals, so
that both
sides of the scraper ring 13778 are exposed to the same work space pressure.
The port
13782 includes a port 13782A and a particle filter 13782B that, in some
embodiments, are
separate pieces. In various embodiments, one or more circumferential 0-rings
may be
located between the housing 13754 and the cylinder gland 13752, and between
the housing
13754 and the fitting 13776 to seal leak paths through piston rod seal unit
137750. In some
embodiments, there may be axial or face 0-rings between the end plate 13756
and the
housing 13754 and between the end plate 13756 and the first washer 13758 to
seal leak
paths through piston rod seal unit 137750.
In various embodiments, the floating rod seal assembly 13760 may be assembled
differently such as pressing the elements into the housing 13754. In another
embodiment
the outer ring 13762 may be incorporated into the housing 13754, so the
bushings 13764,
floating bushing 13766 and rod seal 13770 may be directly mounted into the
housing
13754. In another embodiment, the floating rod seal assemblies 13760 may be
pressed or
inserted directly into the cylinder gland 13752 or the duct plate 13715.
Still referring to FIGS. 68A-69C, the scraper ring 13778, particle trap 13780,
port
13782, and filter 13782B, may serve to protect the rod seal 13370 and floating
rod seal
assemblies from particles and/or may prevent particles from entering the
piston rod seal unit
.. 13750 with the movement of gas into the piston rod seal unit 13750.
Particles may be
produced in the workspace, for example, in some embodiments when pieces of the
fine wire
break off the regenerator. In some embodiments metal particles may be
generated when the
piston contacts the cylinder. In some embodiments, metal flakes may be
generated from
oxidation of metal in the heater head or other causes. Metal and/or oxidized
metal particles
that reach the rod seal 13370 may score the piston rod 13744 or damage the rod
seal 13770
leading to leakage and lower power or early seal failure. Particles may also
get trapped in
the axial gap 13768 between the bushing and the floating bushing 13766, which
may create
a leak path around the rod seal 13770 that may reduce engine power. Therefore,
in various
embodiments, it is desirable to minimize and/or prevent metal particles from
entering piston
rod seal space. In various embodiments, the scraper ring 13778 and/or the
filter l3782B
may minimize the number of particles that enter the piston rod seal unit
13750. The particle
trap 13780 may serve to attract and hold particles that pass by the scraper
ring 13778 or
filter 13782B. In some embodiments, the particle trap may include one or more
magnets
13780A that may attract steel particles and oxides of metal, which may include
iron.
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In various embodiments, the port 13782 may prevent and/or minimize and/or
reduce
a pressure difference across the scraper ring 13778 by fluidly connecting the
bottom side of
the scraper ring 13778 to the same workspace 13738 in which the top of the
scraper ring
13778 may be exposed. In various embodiments, the port may prevent and/or
minimize
.. and/or reduce a pressure difference across the scraper ring 13778 due to
changes in the
average workspace pressure which may occur in various stages, for example, but
not limited
to, during startup and/or due to the action of the scraper ring 13778 allowing
flow, for
example, in one direction along the piston rod 13744, but not allowing flow in
the other
direction. Still referring to FIGS. 68A-69C, in some embodiments, the scraper
ring 13778
may be a spring energized lip seal. In some embodiments, coiled springs
wrapped around
the bearing surface urge the ends of the bearing surface toward the piston rod
13744 and the
bearing surface. In various embodiments, the bearing surface may be made of a
composite
polymer. In some embodiments, the bearing surface may be made from a PFTE
composite.
In some embodiments, the rod seal 13370 may be one supplied by CoorsTek of
Golden
Colorado or Bal Seal Engineering Inc. of Foothill Ranch, California.
Still referring to FIGS. 68A-69C, in some embodiments, the floating rod seal
assembly 13760 allows the rod seals 13370 to self-center on the reciprocating
piston rod
13744. The floating bushing 13766, which is sealed to the outside diameter of
the rod seal
13770 may form a face seal against one of the stationary bearings 13764, 13765
to keep gas
from leaking around the rod seal. In some embodiments, the axial movement of
the
floating bushing 13766 may be constrained by the a small axial distance
between the two
stationary bushings 13764, 13765 that is 0.002 to 0.004 larger than the
thickness of the
circumferential rib 13766A of the floating bushing 13766.
Still referring to FIGS. 68A-69C, in some embodiments, the operation of the
floating bushing 13766 may be best understood by considering its motion as the
axial
pressure difference changes direction during every stroke. Herein, the axial
pressure
difference is the difference in pressure along the axis of the piston rod
13744 from one side
of the rod seal 13770 or floating bushing 13766 to the other. The sealing
action of the
floating rod seal assemblies 13760 may be understood by considering one piston
rod 13744
and the mated piston rod seal unit 13760, when the piston rod 13744 is at
either the top or
the bottom of the its stroke. At this point, there exists a large axial
pressure difference
across each of the floating rod seal assemblies 13760, which forces the
floating bushing
13766 against one of the stationary bushings 13764, 13765 and forming a metal
to metal
seal 13769. At some point during each stroke of the piston rod 13744, the
pressure
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difference between the workspace 13738 and the airlock 13736 will reverse. At
about the
same time, the pressure difference reverses, the axial pressure across each
floating rod seal
assembly 13760 is zero. The floating bushing 13766 may move radially within
floating rod
seal assembly 13760, when the pressure difference is near zero and self-
centered on the
piston rod 13744. A delay between a zero pressure difference across the piston
rod seal unit
13750 and zero axial pressure difference across the floating rod seal assembly
13760 may
be due to flow resistances and volumes within/across the piston rod seal unit
13750. As the
piston rod 13744 continues its stroke, the pressure difference across the
floating rod seal
assembly 13760 reforms in the opposite direction and forces the floating
bushing 13766
against the other bushing 13764 or 13765, reforming a metal seal. Thus, the
floating
bushing 13766 may move radially during part of each stroke to accommodate
changes in the
motion of the piston rod 13744 and form a seal against one of the bushings
13764, when
significant pressure difference occurs between the working space 13738 and the
airlock
13736.
Referring now to FIG. 69D, in some embodiments, the piston rod seal unit 13750
may include one or more floating clearance bushing 13870 that may provide a
clearance
seal with the piston rod 13744. The clearance seal is a long and narrow radial
gap between
the outside diameter of the piston rod 13744 and the inside diameter of the
floating
clearance bushing 13870 that creates enough flow resistance that an
insignificant amount of
working gas leaks past the clearance seal. The oscillating pressure of the
workspace about
the average pressure of the air lock assures that any gas that leaks from the
workspace when
the workspace pressure is relative to the air lock will leak back to the air
lock when the
workspace pressure is low.
In some embodiments, the inside diameter of the floating clearance bushing
13870 is
0.0005 to 0.001 inch larger than the outside diameter of the piston rod 13744
at room
temperature and the floating clearance bushing 13870 is 0.71 inches long. The
piston rod
outer diameter, in some embodiments, may increase, for example, by a few ten-
thousands of
an inch, due to thermal expansion during operation, when the rod may be
approximately 30
to 70 C above room temperature. In some embodiments, to seal the work space
13738
from the air lock 13736 (FIG. 68A), the floating clearance bushing 13870 (FIG.
69D) may
also form an axial seal to prevent gas from flowing around the outside
diameter of the
floating clearance bushing 13870. When the pressure is higher in the working
space than
the airlock, the floating clearance bushings 1370 may form axial face seals
with end plate
13756 and the washer 13858. When the working space pressure is lower than the
air lock
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pressure, the floating clearance bushings will form an axial face seals with
the threaded
fitting 13876 and the washer 13858.
In various embodiments, the outer rings 13862 may be sized to allow the
floating
clearance bushings 13870 to move radially and minimize axial movement. In some
5 embodiments, the outer ring 13862 may have an inner diameter that is,
e.g., 0.03 inch,
larger than the outside diameter of the floating clearance bushing 13870 and
the axial length
of the outer ring 13862 may be e.2. 0.0002 to 0.0005 inch, longer than the
floating clearance
bushing. In some embodiments the length of the outer rings 13862 may be
matched to the
length of the floating clearance bushing 13870. The floating clearance
bushings 13870,
10 washer 13858, outer rings 13862 may be assembled into the housing 13754
and this
assembly may be axially held in place between the end plate 13756 and the
threaded fitting
13876. The threaded fitting 13876 threads into the housing 13754 and contacts
the outer
rings 13862, which thereby set the axial spacing for the floating clearance
bushings 13870.
The embodiment in FIG. 69D may include the particle trap 13780, scrapper ring
13778, port
15 and filter described for FIG. 69A. In other embodiments, the two
floating clearance
bushings 13870 and washer 13858 in FIG. 69D may be replaced with a single
floating
clearance bushing to reduce part count and cost. The single floating clearance
bushing may
in some examples be as long as two of the floating clearance bushings 13870 in
FIG. 69D.
Referring now to FIG. 69E, in some embodiments, the piston rod seal unit 13750
20 .. includes one or more floating clearance bushings 13970 and a face seal
13972 on one end of
each floating clearance bushing 13970, where the floating clearance bushing
13970
provides a clearance seal with the piston rod 13744. The face seals 13972 may
take many
forms including but not limited to energized lip seals or o-rings. The face
seals 13972 may
provide improved axial sealing of the floating clearance bushing 13970 to the
threaded
25 fitting 13876 and end plate 13756. The use of the face seal may be
beneficial for many
reasons, including but not limited to, it may allow use of rougher surface
finishes on the
ends of the bushings and mating surfaces. The embodiment in FIG. 69E may
include the
particle trap 13780, scrapper ring 13778, port and filter described for Fig.
69A and the outer
rings 13862. washer 13858, and housing 13754 described in FIG. 69D.
30 Referring now to FIG. 69F, in some embodiments, the piston rod seal unit
13750
may include a hybrid seal that includes a floating clearance bushing 13971 and
an energized
lip seal 13752. In some embodiments, the hybrid seal may provide better
sealing and a
longer "life" or performance time than either a clearance seal or a lip seal
could by
themselves. The lip seal 13752 may provide a better seal than a clearance seal
as the lip
46
seal 13752 contacts the moving piston rod 13744. The clearance seal between
the floating
clearance bushing 13971 and the piston rod may reduce the pressure drop across
the lip seal
13752, which may extend the "life" or performance time of the lip seal. The
floating seals
13971 may include face seals 13972 in some embodiments. The embodiment in FIG.
69F
may include the particle trap 13780, scrapper ring 13778, port and filter
described for FIG.
69A and the outer rings 13862, washer 13858, and housing 13754 described in
FIG. 69D.
The floating clearance bushings in FIGS. 69D-69F in various embodiments may be
formed from a material with a low coefficient of friction, low wear and high
strength.
Materials for floating clearance bushings include, but are not limited to, one
or more of the
following: PTFE, Rulon, engineered plastics, graphite and graphite blends. In
some
embodiments, the material may be graphite impregnated with antimony, for
example, from
the SGL Carbon Company of Germany produces graphite impregnated with antimony
under
the grade EK3205.
Linear Cross-Head Bearing
Referring now to FIG. 71A, which shows a cross-section of an embodiment of a
Stirling engine with a drive 13920 to convert the linear motion of the piston
13930 into
rotary motion of the crankshaft and generator. Drive 13920 is one example of a
drive to
convert linear motion of the pistons to rotary motion of the generator.
Another example of a
drive to convert linear motion of the pistons to rotary motion of the
generator is in U.S.
Patent 6,253,550.
Still referring to FIG. 71A, the piston 13930 is fully guided by a cross-head
in the
form of linear bearing 13946. The linear cross-head bearing 13946 receives all
the side
loads in the lubricated drive of the engine and removes all side loads from
seals on the
piston and piston rod. The lubricated linear cross-head bearing may absorb the
side loads
with little friction and near zero wear. The linear cross-head bearing 13946
has a
sufficiently large length / diameter ratio, sufficient small radial clearance
and / or sufficient
accurate alignment with the heater head to assure that the piston 13930 will
not contact
cylinder walls of the heater head 13712 nor the cooler 13915 during operation
of the engine.
In addition, the linear cross-head bearing 13946 may receive side loads from
the link rod
13719. The linear cross-head bearing 13946 transfers these side loads to the
engine block
13949 without negligible wear or acceptable friction loads.
The linear cross-head bearing 13946 in FIGS. 71A ¨ 71C is configured to absorb
all
the side loads from the link rod 13719 and fully guide the piston 13930 and
piston rod
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13744 without creating side loads on the piston seals 13934 or piston rod seal
unit 13750.
The linear cross-head bearing 13946 is preferably a hydrodynamic bearing that
is fed
pressurized oil. In one example, pressurized oil is feed into the bearing
13946 at annulus
13948 from passages (not shown) in the engine block 13949. The ratio of length
over the
diameter of the linear bearing 13946 is at least 2.0 In a preferred
embodiment, length over
diameter ratio is 2.20. The diameter of the linear cross-head bearing 13946
has a diameter
of 1.45 inches or more than 63% of the heater head 13712 inner diameter. The
selection of
the linear bearing diameter balances axial stiffness which increases with
diameter and
frictional losses which also increase with diameter. Other embodiments of the
linear cross-
head bearing may have different diameters. The diametrical gap of the linear
cross-head
bearing 13946 is generally equal to the diameter of the bearing divided by
1000. In a
preferred embodiment the nominal diametrical gap for the linear cross-head
bearig is 0.0015
inches. The axial alignment of the linear cross-head bearings 13746 with the
heater head
13712, cooler 13915 and piston rod seal unit 13750 is improved by boring the 4
guide holes
for the 4 linear cross-head bearings 13946 in a single setup. The 4 guide
holes may be
linebored to provide holes that straight and perpendicular to the engine block
face 13949A.
Refen-ing now to Fig. 14, the linear cross-head bearing removes the need for
guide
ring 1616 on the piston 1600. The guide ring 1616 can be a wear component
that, in some
instances, has a shorter life than the pressure seals 1606, 1614. The guide
rings also occupy
axial space on the pistons that could be used for additional pressure seals
1606 1614. In
some instances, the life of the pressure seals 1606, 1614 increases with the
number of seals
used on a given piston.
Referring now to FIG. 71B which presents a cross-sectional view of a single
piston/cylinder in a multi-cylinder engine and a portion of the drive
assembly. The linear
cross-head bearing 13946 fully guides the motion of the piston rod 13744 and
the piston
assembly 13930. The linear cross-head bearing 13946 constrains the
piston/piston rod to
linear movement with minimal rotation so that the piston 13930 cannot contact
the cylinder
walls 13926. In one embodiment, the linear cross-section bearing 13946 has a
minimum
engaged length of 3.2 inches and a diametrical clearance of 0.0015 inches, so
that the
maximum radial movement of the piston base is less than 0.005 inches which is
half of the
radial gap between the piston base 13932 and the cylinder wall 13926.
In one embodiment, the linear cross-head bearing 13946 comprises a journal or
shaft
13947 that rides on a hydrodynamic layer of oil inside a hole or guide 13950
within the
engine block 13949. The journal 13947 may be connected at a first end via a
rotatable joint
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to the link rod 13719 which in turn is connected to the rocking beam 13916
(FIG. 71A).
The journal 13947 may be rigidly connected on the second end to the piston rod
13744. An
oil pump (not shown) supplies pressurized oil to the gap between the journal
13947 and the
guide 13950 via oil annulus 13948 in guide 13950. In a preferred embodiment
the oil
.. annulus is located approximately equal distant from both ends of the guide
13950. The
diametrical gap between the journal 13947 and the guild 13950 may be a value
between
0.001 inch and 0.002 inch. In a preferred embodiment, the diametrical gap is
0.0015 inch.
The guide 13950 is the section of the hole through the engine block 13949 that
has a nearly
constant diameter that is slightly larger than the journal diameter. In a
preferred
embodiment the guide diameter is 0.001 to 0.002 inch larger than the journal
diameter. The
linear cross-head bearing 13946 is located below the piston, the piston rod
seals and the
bellows that isolated the oil lubricated drive from the air lock and the
working space.
Referring now to FIG. 71C, in some embodiments the second end of the journal
13947 comprises an inner mount and inner support a rolling diaphragm or
bellows 13960
and a mounting point for the piston rod 13744. In one example, the inner bead
of the
bellows 13960 is compressed to create a gas tight seal by washer 13966 that is
held in place
by a hollow screw 13964 that threads into the top of the journal 13947 while
the piston rod
13744 pass through the hollow screw 13964. The outer bead of the bellows 13960
is
trapped between collar 13962 and bellows support 13968 that is mounted in the
engine
block 13949. In normal operation, the pressure in air lock 13766 is maintained
above the
crankcase pressure creating a positive pressure difference. The positive
pressure forms the
bellows as pictured in FIG. 71C. The positive pressure difference presses the
bellows
13960 against the outside diameter of the journal 13947 and the inside
diameter of the
support 13968, so that stress on the bellows is generally limited to a small
unsupported U
shape. In use. the pressure difference between the airlock 13736 and the
crankcase may
reverse so that the higher crank case will push the bellows 13960 up into the
airlock space.
The collar 13962 is sized and includes a smooth lower corner to support the
bellows 13960
out the outside diameter in this inverted position which will reduce the
stress on the bellows
and the likelihood of bellows failure. In a preferred embodiment, the collar
13962 has an
inner diameter that matches the inner diameter of the bellows support 13968
and a lower
corner radius of 0.10 inches.
Refen-ing now to FIG. 72, one embodiment of the piston 13930 includes a
clearance
seal 13938 that provides pressure isolation between the workspace above and
below the
piston without contacting the cylinder walls. A no-contact seal would not wear
and would
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have a much longer or even infinite life. As the piston seals can the first
element of a
Stirling engine to wear out or fail, a clearance seal could dramatically
increase the life and
reduce the time between failures or maintenance events of the whole Stirling
engine.
The outside diameter of the clearance seal 13938 nearly equal, but just
smaller than
the inside diameter of the mating cylinder 13926 (in FIG. 71B). In one
embodiment, the
clearance seal 13938 is fabricated from one or more non-metal materials
including but not
limited to graphite, PTFE, Ulta High Molecular Weight Polyethylene (UHMWPE),
ceramic,
composites containing graphite, composites containing antimony, composites
containing
PTFE, composites containing UHMWPE. The cylinder inner diameter may have a
hard
surface composed of material including but not limited to tool steel,
stainless steel, ceramic
coatings. In one embodiment, the outer diameter of the clearance seal 13938 is
0.0005 to
0.001 inch smaller than the matting cylinder inner diameter. In another
embodiment, the
outer diameter of the clearance seal 13938 is machined to be nominally equal
to the cylinder
inner diameter. The piston 13930 is assembled into the cylinder and operated
to wear down
interferences on the softer clearance seal 13938.
In some embodiments, the piston clearance seal 13938 may be graphite or a
graphite impregnated with antimony and or may be made of any material and/or
any
combination and/or composite of material. In some embodiments, the clearance
seal 13938
may be made of antimony impregnated graphite which may be from the SGL Carbon
Company of Germany.
In one embodiment, the clearance seal 13938 includes a plurality of radial
grooves
13938A in the outer diameter. The grooves increase axial flow resistance in
the gap
between the clearance seal 13938 and the cylinder wall. One theory, among
others,
suggests that the repeated grooves 13938A perpendicular to axial flow disrupt
that flow and
increase the flow resistance by creating repeated expansions and contractions.
In one
embodiment the clearance seal 13938 is mounted in the piston base 13932 and
held in place
by the piston dome 13934. Gas is prevented from leaking between the clearance
seal and
the piston base by one or more o-rings 13958.
The piston 13930 comprises a piston base 13942 that mounts to the piston rod
13744, the piston dome 13934 and one or more seals 13938. The piston may also
comprise
one or more guide rings 13940 that may be required in some embodiments of the
engine. In
other embodiments of the piston 13930, the guide rings 13940 are not included.
The piston
dome 13934 may further comprise a inner dome 13936 that reduces the transfer
of thermal
energy from the closed end of the piston dome 13934 to the piston base 13942.
The overall
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engine efficiency is improved by minimizing parasitic heat transfer from the
hot end to the
cold end of the engine. Parasitic heat transfer is the transfer of thermal
energy from the hot
end to the cold end that does not involve the working fluid. One example of
parasitic heat
transfer is axial heat transfer from the hot closed end of the piston dome
13934 that is near
5 the temperature of the heater to the cold piston base 13932 that is near
the temperature of
the cooler. Therefore, reducing axial heat transfer through the piston will
increase the
engine efficiency. In addition, the contact seals 13934 (FIG. 71B) last longer
if they are
kept cooler, so decreasing the axial heat transfer through the piston will
increase the life of
the contact seals. One theory, among others, states the inner dome 13936
reduces axial heat
10 transfer as a radiation shield between the closed end of the piston dome
13934 and the
piston base 13932. Another theory states inner dome reduces axial heat
transfer by reducing
convection in the dome created in part by the oscillating movement of the
piston. The mass
of gas in the piston dome 13934 and inner dome 13936 changes during engine
startup and
during some transient conditions. A port 13944 in the inner dome 13936 and a
small orifice
15 13942 in the piston base 13932 allow the gas to move into and out of the
piston 13930. The
orifice 13942 is small enough to act as a low pass filter on the pressure in
the piston. The
working pressure below the piston base 13932 may change by more than one
hundred psi
every engine rotation. The orifice 13942 is small enough to keep the pressure
in the piston
nearly constant over an engine rotation or cycle. In a preferred embodiment,
the orifice
20 13942 has an 0.004 inch diameter. The pressures in the inner dome and
the outer dome do
not oscilate, so the port 13944 between the two volumes can be ten times
larger.
In some Stirling engines, reciprocating seals on piston are wear items that
wear
away over time with normal use. The life of these piston seals determines the
time between
failures or the time between maintenance events of the entire engine. A long
time between
25 maintenance events and a long life are important to the economic success
of the Stirling.
The time between maintenance or failures could be dramatically increased if
the piston seals
and piston rod seals were not wear items such as clearance seals. The
decreased
maintenance would increase the value and utility of the Stirling engine. An
embodiment of
the Stirling engine without non-wearing seals comprises an oil lubricated
drive 13920 (FIG.
30 71A) and the linear cross-head bearing 13946 (FIG. 71A-C), a clearance
seal 13938 (FIG.
72) on the piston and a piston rod seal unit 13750 with clearance seals 13862
(FIG. 69D,
FIG. 69E). This embodiment with an oil lubricated drive and non-contact
recirprocating
seals may have a time between maintenance events that is many times longer
than a Stirling
engine with standard sliding seals.
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In an embodiment of the Stiring engine where the piston seal and or piston rod
seal
are contact seals, the linear cross-head bearing improves the life of the
seals and extends the
time between failures or the time between maintenance events for the engine.
The piston
and piston rod are fully guided or constrained by the linear cross-head
bearing so the
possible radial movement of the piston and or piston rod is limited to an
acceptably small
value as described above. In addition, because the linear cross-head bearing
does not wear,
the radial movement of the piston and piston rod does not change over time.
The reduced
radial movement limits the loads applied to the piston seals and the piston
rod seals and the
reduce loads increases the life of the these seals.
Lubricating Fluid Pump and Lubricating Fluid Passageways
In some embodiments, the lubricating fluid is oil. The lubricating fluid is
used to
lubricate engine parts in the crankcase 2206, such as hydrodynamic pressure
fed lubricated
bearings. Lubricating the moving parts of the engine 2200 serves to further
reduce friction
between engine parts and further increase engine efficiency and engine life.
In some
embodiments, lubricating fluid may be placed at the bottom of the engine, also
known as an
oil sump, and distributed throughout the crankcase. The lubricating fluid may
be distributed
to the different parts of the engine 2200 by way of a lubricating fluid pump,
wherein the
lubricating fluid pump may collect lubricating fluid from the sump via a
filtered inlet. In
the exemplary embodiment, the lubricating fluid is oil and thus, the
lubricating fluid pump
is herein referred to as an oil pump. However, the term "oil pump" is used
only to describe
the exemplary embodiment and other embodiments where oil is used as a
lubricating fluid,
and the term shall not be construed to limit the lubricating fluid or the
lubricating fluid
pump.
Tube Heat Exchanger
External combustion engines, such as, for example, Stirling cycle engines, may
use
tube heater heads to achieve high power. FIG. 19 is a cross-sectional view of
a cylinder and
tube heater head of an illustrative Stirling cycle engine. A typical
configuration of a tube
heater head 4200, as shown in FIG. 19, uses a cage of U-shaped heater tubes
4202
surrounding a combustion chamber 4204. A cylinder 4206 contains a working
fluid, such
as, for example, helium. The working fluid is displaced by the piston 4208 and
driven
through the heater tubes 4202. A burner 4210 combusts a combination of fuel
and air to
produce hot combustion gases that are used to heat the working fluid through
the heater
52
tubes 4202 by conduction. The heater tubes 4202 connect a regenerator 4212
with the
cylinder 4204. The regenerator 2812 may be a matrix of material having a large
ratio of
surface to area volume which serves to absorb heat from the working fluid or
to heat the
working fluid during the cycles of the engine. Heater tubes 2802 provide a
high surface area
and a high heat transfer coefficient for the flow of the combustion gases past
the heater
tubes 4202. Various embodiments of tube heater heads are discussed below, and
in U.S.
Patents No. 6,543,215 and No. 7,308,787.
FIG. 19 is a side view in cross section of a tube heater head and a cylinder.
The
heater head 4206 is substantially a cylinder having one closed end 4220
(otherwise referred
to as the cylinder head) and an open end 4222. Closed end 4220 includes a
plurality of U-
shaped heater tubes 4204 that are disposed in a burner 4200. Each U-shaped
tube 4204 has
an outer portion 4216 (otherwise referred to herein as an "outer heater tube")
and an inner
portion 4218 (otherwise referred to herein as an "inner heater tube"). The
heater tubes 4204
connect the cylinder 4202 to regenerator 4210. Cylinder 4202 is disposed
inside heater head
4206 and is also typically supported by the heater head 4206. A piston 4224
travels along
the interior of cylinder 4202. As the piston 4224 travels toward the closed
end 4220 of the
heater head 4206, working fluid within the cylinder 4202 is displaced and
caused to flow
through the heater tubes 4224 and regenerator 4210 as illustrated by arrows
4230 and 4232
in FIG. 19. Referring to FIG. 19, as mentioned above, the closed end of heater
head 4220,
including the heater tubes 4202, is disposed in a burner 4200 that includes a
combustion
chamber 4204. Hot combustion gases (otherwise referred to herein as "exhaust
gases") in
combustion chamber 4204 are in direct thermal contact with heater tubes 4202
of heater
head 4220. Thermal energy is transferred by conduction from the exhaust gases
to the heater
tubes 4202 and from the heater tubes 4202 to the working fluid of the engine,
typically
helium. Other gases, such as nitrogen, for example, or mixtures of gases, may
be used, with
a preferable working fluid having high thermal conductivity and low viscosity.
Non-
combustible gases are used in various embodiments. Heat is transferred from
the exhaust
gases to the heater tubes 4202 as the exhaust gases flow around the surfaces
of the heater
tubes 4202. Arrows 4230 show the general radial direction of flow of the
exhaust gases. The
exhaust gases exiting from the burner 4210 tend to overheat the upper part of
the heater
tubes 4202 (near the U-bend) because the flow of the exhaust gases is greater
near the upper
part of the heater tubes than at the bottom of the heater tubes (i.e., near
the bottom of the
burner 4200).
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The overall efficiency of an external combustion engine is dependent in part
on the
efficiency of heat transfer between the combustion gases and the working fluid
of the
engine.
Returning to FIG. 19, in general, the inner heater tubes 4218 are warmer than
the
outer heater tubes 4216 by several hundred degrees Celsius. The burner power
and thus the
amount of heating provided to the working fluid is therefore limited by the
inner heater tube
4218 temperatures. The maximum amount of heat will be transferred to the
working gas if
the inner and outer heater tubes are nearly the same temperature. Generally,
embodiments,
as described herein, either increase the heat transfer to the outer heater
tubes or decrease the
rate of heat transfer to the inner heater tubes.
An alternative embodiment of flow diverter fins is shown in FIG. 17. FIG. 17
is a
top view of a section of a tube heater head including single flow diverter
fins in accordance
with an embodiment. In this embodiment, a single flow diverter fin 4002 is
connected to
each outer heater tube 4004. In some embodiments, the flow diverter fins 4002
are attached
to an outer heater tube 4004 using a nickel braze along the full length of the
heater tube.
Alternatively, the flow diverter fins may be brazed with other high
temperature materials,
welded or joined using other techniques known in the art that provide a
mechanical and
thermal bond between the flow diverter fin and the heater tube. Flow diverter
fins 4002 are
used to direct the exhaust gas flow around the heater tubes 4004, including
the downstream
side of the heater tubes 4004. In order to increase the heat transfer from the
exhaust gas to
the heater tubes 4004, flow diverter fins 4002 are thermally connected to the
heater tube
4004. Therefore, in addition to directing the flow of exhaust gas, flow
diverter fins 4002
increase the surface area for the transfer of heat by conduction to the heater
tubes 4004, and
thence to the working fluid.
FIG. 18 is a top view in cross-section of a section of a tube heater head
including the
single flow diverter fins as shown in FIG. 17 in accordance with an
embodiment. As shown
in FIG. 18, a flow diverter fin 4110 is placed on the upstream side of a
heater tube 4106.
The diverter fin 4110 is shaped so as to maintain a constant distance from the
downstream
side of the heater tube 4106 and therefore improve the transfer of heat to the
heater tube
4106. In an alternative embodiment, the flow diverter fins could be placed on
the inner
heater tubes 4108.
Engine performance, in terms of both power and efficiency, is highest at the
highest
possible temperature of the working gas in the expansion volume of the engine.
The
maximum working gas temperature, however, is typically limited by the
properties of the
54
heater head. For an external combustion engine with a tube heater head, the
maximum
temperature is limited by the metallurgical properties of the heater tubes. If
the heater tubes
become too hot, they may soften and fail resulting in engine shut down.
Alternatively, at too
high of a temperature the tubes will be severely oxidized and fail. It is,
therefore, important
to engine performance to control the temperature of the heater tubes. A
temperature sensing
device, such as a thermocouple, may be used to measure the temperature of the
heater tubes.
The temperature sensor mounting scheme may thermally bond the sensor to the
heater tube
and isolate the sensor from the much hotter combustion gases. The mounting
scheme should
be sufficiently robust to withstand the hot oxidizing environment of the
combustion-gas and
impinging flame that occur near the heater tubes for the life of the heater
head. One set of
mounting solutions include brazing or welding thermocouples directly to the
heater tubes.
The thermocouples would be mounted on the part of the heater tubes exposed to
the hottest
combustion gas. Other possible mounting schemes permit the replacement of the
temperature sensor. In one embodiment, the temperature sensor is in a
thermowell thermally
bonded to the heater tube. In another embodiment, the mounting scheme is a
mount, such as
a sleeve, that mechanically holds the temperature sensor against the heater
tube.
FIG. 19 is a side view in cross section of a cylinder 4204 and a burner 4210.
A
temperature sensor 4202 is used to monitor the temperature of the heater tubes
and provide
feedback to a fuel controller (not shown) of the engine in order to maintain
the heater tubes
at the desired temperature. In some embodiments, the heater tubes are
fabricated using
InconelTM 625 and the desired temperature is 930° C. The desired
temperature will
be different for other heater tube materials. The temperature sensor 4202
should be placed at
the hottest, and therefore the limiting, part of the heater tubes. Generally,
the hottest part of
the heater tubes will be the upstream side of an inner heater tube 4206 near
the top of the
heater tube. FIG. 19 shows the placement of the temperature sensor 4202 on the
upstream
side of an inner heater tube 4206. In some embodiments, as shown in FIG. 19,
the
temperature sensor 4202 is clamped to the heater tube with a strip of metal
4212 that is
welded to the heater tube in order to provide good thermal contact between the
temperature
sensor 4202 and the heater tube 4206. In one embodiment, both the heater tubes
4206 and
the metal strip 4212 may be InconelTM 625 or other heat resistant alloys such
as InconelTM
600, Stainless Steels 310 and 316 and HastellovTM X. The temperature sensor
4202 should
be in good thermal contact with the heater tube, otherwise it may read too
high a
temperature and the engine will not produce as much power as possible. In an
alternative
embodiment, the temperature sensor sheath may be welded directly to the heater
tube.
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In another embodiment, as shown in FIG. 20A-B, a temperature sensor mount 4320
is created with a formed strip or sheath of a refractory or high temperature
resistant metal
such as Inconel that is bonded to the exterior of the heater tube 4310. The
sensor mount
sheath 4320 is formed or shaped into a channel that when attached to the
heater tube creates
5 a void that accommodates a device. In a specific embodiment, the channel
is V-shaped to
accommodate the insertion of a thermal sensor such as a thermocouple device.
The shaped
channel is then bonded to the exterior of a heater tube 4310 as shown in FIG.
20A.
FIG. 20A shows a side view of the sensor mount sheath 4320 on the heater tube
4310, while FIG. 20B is a view along the axis of the sensor mount sheath 4320.
The metal
10 .. should be thin enough to form, yet thick enough to survive for the rated
life of the heater
head. In some embodiments, the metal is approximately between 0.005" and
0.020" thick.
The metal may be bent such that the bend is along the length of the strip.
This "V-channel"
sheath 4320 is then affixed to the exterior of the heater tube by high
temperature brazing.
Prior to brazing, the sheath may be tack welded in several places to insure
that the sheath
15 does not move during the brazing process, as shown in FIG. 20A.
Preferably, the braze
compound used during brazing is typically a high nickel alloy; however, any
compound
which will withstand the brazing temperature will work. Alternatively the
sheath may be
bonded to the heater tube by electron beam or laser welding.
Now referring to FIG. 20B, a cavity 4330 is formed by affixing the sheath to
the
20 heater tube. This cavity 4330 is formed such that it may accept a device
such as a
thermocouple. When formed and brazed, the cavity may advantageously be sized
to fit the
thermocouple. Preferably, the fit is such that the thermocouple is pressed
against the
exterior of the heater tube. Preferably, the sheath is thermally connected to
the heater tube.
If the sheath is not thermally connected to the heater tube, the sheath may
not be "cooled"
25 by the working gas. The lack of cooling may cause the sheath to operate
at or near the
combustion gas temperatures, which are typically high enough to eventually
burn through
any metal. Brazing the sensor mount to the heater tube leads to a good thermal
contact.
Alternatively, the sensor mount sheath 4320 could be continuously welded along
both sides
to provide sufficient thermal connection.
30 In another embodiment, as shown in FIGS. 21A-B, a second strip of metal
can be
formed to create a shield 4450 over the sensor mount 4420. The shield 4420 may
be used to
improve the thermal connection between the temperature sensor, in cavity 4430,
and the
heater tube 4410. The shield insulates the sensor mount sheath 4420 from the
convective
heating of the hot combustion gases and thus improves the thermal connection
to the heater
56
tube. Furthermore, there is preferably an insulating space 4440 to help
further insulate the
temperature sensor from the hot combustion gases as shown in FIG. 21B.
In another specific embodiment, as shown in FIGS. 22A and 22B, the temperature
sensor mount 4520 can be a small diameter tube or sleeve 4540 joined to the
leading edge of
the heater tube 4510. FIG. 22A shows a side view of the mount on the heater
tube 4510,
while FIG. 22B is a view along the axis of the tube 4540 or sleeve. The sensor
tube 4540 is
preferably brazed to the heater tube with a substantial braze fillet 4530. The
large braze
fillet 4530 will maximize the thermal bond between the heater tube and the
sensor mount. In
another embodiment, the tube or sleeve 4540 may have a shield. As described
supra, an
.. outer shield cover may help insulate the temperature sensor mount 4520 from
convective
heat transfer and improve the thermal connection to the heater tube.
In some embodiments of the heater head, inserts may be placed on the inside of
the
heater tubes to increase heat transfer between the working gas and the tube
walls as
disclosed in U.S. Patent Application Serial No. 13/447,990, filed April 16,
2012 and entitled
Stirling Cycle Machine (Attorney Docket No. 184).
Regenerator
A regenerator is used in a Stirling cycle machine, as discussed above and as
described in U.S. Patents No. 6,591,609, and No. 6,862,883, to add and remove
heat from
the working fluid during different phases of the Stirling cycle. The
regenerator used in a
Stirling cycle machine must be capable of high heat transfer rates which
typically suggests a
high heat transfer area and low flow resistance to the working fluid. Low flow
resistance
also contributes to the overall efficiency of the engine by reducing the
energy required to
pump the working fluid. Additionally, a regenerator must be fabricated in such
a manner as
to resist spalling or fragmentation because fragments may be entrained in the
working fluid
and transported to the compression or expansion cylinders and result in damage
to the
piston seals.
One regenerator design uses several hundred stacked metal screens. While
exhibiting a high heat transfer surface, low flow resistance and low spalling,
metal screens
may suffer the disadvantage that their cutting and handling may generate small
metal
fragments that must be removed before assembling the regenerator.
Additionally, stainless
steel woven wire mesh contributes appreciably to the cost of the Stirling
cycle engine.
A three dimensional random fiber network, such as stainless steel wool or
ceramic
fiber, for example, may be used as the regenerator, as now described with
reference to FIG.
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23A. Stainless steel wool regenerator 6300 advantageously provides a large
surface area to
volume ratio, thereby providing favorable heat transfer rates at low fluid
flow friction in a
compact form. Additionally, cumbersome manufacturing steps of cutting,
cleaning and
assembling large numbers of screens are advantageously eliminated. The low
mechanical
strength of steel wool and the tendency of steel wool to spall may both be
overcome as now
described. In some embodiments, the individual steel wires 6302 and 6304 are
"cross-
linked," into a unitary 3D wire matrix.
The starting material for the regenerator may be fibrilose and of random fiber
form
such as either steel or nickel wool. The composition of the fiber may be a
glass or a ceramic
or a metal such as steel, copper, or other high temperature materials. The
diameter of the
fiber is preferably in the range from 10 micrometers to 1 millimeter depending
on the size
of the regenerator and the properties of the metal. The starting material is
placed into a form
corresponding to the final shape of the regenerator which is depicted in cross-
section in
FIG. 23B. Inner canister cylindrical wall 6320, outer canister cylindrical
wall 6322, and
regenerator network 6300 are shown. The density of the regenerator is
controlled by the
amount of starting material placed in the form. The form may be porous to
allow fluids to
pass through the form.
In some embodiments, unsintered steel wool is employed as regenerator network
6300. Regenerator network 6300 is then retained within the regenerator
canister by
regenerator retaining screens 6324 or other filter, thereby comprising a
"basket" which may
advantageously capture steel wool fragments.
In yet other embodiments, knit or woven wire is employed in fabrication of a
regenerator as now described with reference to FIG. 24A. In accordance with
these
embodiments, knit or woven wire tube 6401 is flattened by rollers 6402 into
tape 6404, in
which form it is wound about mandrel 6406 into annular layers 6408. Stainless
steel is
advantageously used for knit wire tube 6401 because of its ability to
withstand elevated
temperature operation, and the diameter of the wire used is typically in the
range of 1-2
mils, however other materials and gauges may be used in various embodiments.
Alternatively, a plurality, typically 5-10, of the stainless steel wires may
be loosely wound
into a multi-filament thread prior to knitting into a wire tube. This process
advantageously
strengthens the resulting tube 6401. When mandrel 6406 is removed, annular
assembly
6410 may be used as a regenerator in a thermal cycle engine.
Still another embodiment is now described with reference to FIGS. 24B through
24E. Knit or woven wire tube 6401, shown in its right cylindrical form in FIG.
24B, is
58
shown scored and partially compressed in FIG. 24C. Alternatively, the scoring
may be at an
angle 6414 with respect to the central axis 6412 of the tube, as shown in FIG.
24D. Tube
6401 is then axially compressed along central axis 6412 to form the bellows
form 6416
shown in FIG. 24E that is then disposed as a regenerator within the
regenerator volume 408
(shown in FIG. 4) of a Stirling cycle engine.
It is to be understood that the various regenerator embodiments and methods
for
their manufacture described herein may be adapted to function in a multiple
cylinder
configuration.
Coolant Penetrating Cold-End Pressure Vessel
Referring now to FIGS. 25A-C, various cross-sections of an engine, such as a
Stirling cycle engine, are shown in accordance with some embodiments. Engine
6500 is
hermetically sealed. A crankcase 6502 serves as the cold-end pressure vessel
and contains a
charge gas in an interior volume 6504. Crankcase 6502 can be made arbitrarily
strong
without sacrificing thermal performance by using sufficiently thick steel or
other structural
material. A heater head 6506 serves as the hot-end pressure vessel and is
preferably
fabricated from a high temperature super-alloy such as InconelTM 625, GMR-235,
etc.
Heater head 6506 is used to transfer thermal energy by conduction from an
external thermal
source (not shown) to the working fluid. Thermal energy may be provided from
various heat
sources such as solar radiation or combustion gases. For example, a burner, as
previously
discussed, may be used to produce hot combustion gases (shown as 6507 in FIG.
25B) that
are used to heat the working fluid. An expansion area of cylinder (or warm
section) 6522 is
disposed inside the heater head 6506 and defines part of a working gas volume
as discussed
above with respect to FIG. 1. A piston 6528 is used to displace the working
fluid contained
in the expansion area of cylinder 6522.
In accordance with an embodiment, crankcase 6502 is welded directly to heater
head
6506 at joints 6508 to create a pressure vessel that can be designed to hold
any pressure
without being limited, as are other designs, by the requirements of heat
transfer in the
cooler. In an alternative embodiment, the crankcase 6502 and heater head 6506
are either
brazed or bolted together. The heater head 6506 has a flange or step 6510 that
axially
constrains the heater head and transfers the axial pressure force from the
heater head 6506
to the crankcase 6502, thereby relieving the pressure force from the welded or
brazed joints
6508. Joints 6508 serve to seal the crankcase 6502 (or cold-end pressure
vessel) and bear
the bending and planar stresses. In an alternative embodiment, the joints 6508
are
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59
mechanical joints with an elastomer seal. In yet another embodiment, step 6510
is replaced
with an internal weld in addition to the exterior weld at joints 6508.
Crankcase 6502 is assembled in two pieces, an upper crankcase 6512 and a lower
crankcase 6516. The heater head 6506 is first joined to the upper crankcase
6512. Second, a
cooler 6520 is installed with a coolant tubing (shown as 6514 in FIG. 25B)
passing through
holes in the upper crankcase 6512. Third, the double acting. pistons 6528 and
drive
components (designated generally as numeral 6540 in FIGS. 25A and 25C, not
shown in
FIG. 25B) are installed. In one embodiment, lower crankcase 6516 is assembled
in three
pieces, an upper section 6513, a middle section 6515, and a lower section
6517, as shown in
FIGS. 25A and 25C. Middle section 6515 is may be connected to upper and lower
sections
6513 and 6517 at joints 6519 and 6521, respectively, by any mechanical means
known in
the art, or by welding.
The lower crankcase 6516 is then joined to the upper crankcase 6512 at joints
6518.
Preferably, the upper crankcase 6512 and the lower crankcase 6516 are joined
by welding.
.. Alternatively, a bolted flange may be employed (as shown in FIGS. 25B and
25C).
In some embodiments a motor/generator (shown as 6501 in FIG. 25C), such as a
PM
generator, may be installed into motor/generator housing (shown as 6503 in
FIG. 25C),
which is attached to the lower crankcase 6516, as shown in FIG. 25C.
Motor/generator
housing 6503 may be attached to lower crankcase 6516 by any mechanical means
known in
the art, or may be welded to lower crankcase 6516. Motor/generator housing
6503 may
assembled in two pieces, a front section 6505, which is attached to lower
crankcase 6516,
and a rear section 6509, which may be welded or bolted to front section 6505.
In one
embodiment a seal 6511 may be positioned between the rear section 6509 and the
front
section 6505 of the motor/generator housing 6503. In some embodiments rear
section 6509
is removable attached to front section 6505, which serves, among other
functions, to allow
for easy removal and installation of motor/generator 6501 during engine 6500
assembly.
In order to allow direct coupling of the heater head 6506 to the upper
crankcase
6512, the cooling function of the thermal cycle is performed by a cooler 6520
that is
disposed within the crankcase 6502, thereby advantageously reducing the
pressure
containment requirements placed upon the cooler. By placing the cooler 6520
within
crankcase 6502, the pressure across the cooler is limited to the pressure
difference between
the working gas in the working gas volume, and the charge gas in the interior
volume 6504
of the crankcase. The difference in pressure is created by the compression and
expansion of
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the working gas, and is typically limited to a percentage of the operating
pressure. In one
embodiment, the pressure difference is limited to less than 30% of the
operating pressure.
Coolant tubing 6514 advantageously has a small diameter relative to the
diameter of
the cooler 6520. The small diameter of the coolant passages, such as provided
by coolant
5 tubing 6514, is key to achieving high heat transfer and supporting large
pressure
differences. The required wall thickness to withstand or support a given
pressure is
proportional to the tube or vessel diameter. The low stress on the tube walls
allows various
materials to be used for coolant tubing 6514 including, but not limited to,
thin-walled
stainless steel tubing or thicker-walled copper tubing.
10 An additional advantage of locating the cooler 6520 entirely within the
crankcase
6502 (or cold-end pressure vessel) volume is that any leaks of the working gas
through the
cooler 6520 will only result in a reduction of engine performance. In
contrast, if the cooler
were to interface with the external ambient environment, a leak of the working
gas through
the cooler would render the engine useless due to loss of the working gas
unless the mean
15 pressure of working gas is maintained by an external source. The reduced
requirement for a
leak-tight cooler allows for the use of less expensive fabrication techniques
including, but
not limited to, powder metal and die casting.
Cooler 6520 is used to transfer thermal energy by conduction from the working
gas
and thereby cool the working gas. A coolant, either water or another fluid, is
carried through
20 the crankcase 6502 and the cooler 6520 by coolant tubing 6514. The
feedthrough of the
coolant tubing 6514 through upper crankcase 6512 may be sealed by a soldered
or brazed
joint for copper tubes, welding, in the case of stainless steel and steel
tubing, or as otherwise
known in the art.
The charge gas in the interior volume 6504 may also require cooling due to
heating
25 resulting from heat dissipated in the motor/generator windings,
mechanical friction in the
drive, the non-reversible compression/expansion of the charge gas, and the
blow-by of hot
gases from the working gas volume. Cooling the charge gas in the crankcase
6502 increases
the power and efficiency of the engine as well as the longevity of bearings
used in the
engine.
30 In one embodiment, an additional length of coolant tubing (shown as 6530
in FIG.
25B) is disposed inside the crankcase 6502 to absorb heat from the charge gas
in the interior
volume 6504. The additional length of coolant tubing 6530 may include a set of
extended
heat transfer surfaces (shown as 6548 in FIG. 25B), such as fins, to provide
additional heat
transfer. As shown in FIG. 25B, the additional length of coolant tubing 6530
may be
61
attached to the coolant tubing 6514 between the crankcase 6502 and the cooler
6520. In an
alternative embodiment, the length of coolant tubing 6530 may be a separate
tube with its
own feedthrough of the crankcase 6502 that is connected to the cooling loop by
hoses
outside of the crankcase 6502.
In another embodiment the extended coolant tubing 6530 may be replaced with
extended surfaces on the exterior surface of the cooler 6520 or the drive
housing (shown as
6572 in FIGS. 25A and 25C). Alternatively, a fan (shown as 6534 in FIG. 25B)
may be
attached to the engine crankshaft (shown as 6542 in FIG. 25C) to circulate the
charge gas in
interior volume 6504. The fan 6534 may be used separately or in conjunction
with the
additional coolant tubing 6530 or the extended surfaces on the cooler 6520 or
drive housing
6572 to directly cool the charge gas in the interior volume 6504.
Preferably, coolant tubing 6514 is a continuous tube throughout the interior
volume
6504 of the crankcase and the cooler 6520. Alternatively, two pieces of tubing
could be
used between the crankcase and the feedthrough ports of the cooler. One tube
carries
coolant from outside the crankcase 6502 to the cooler 6520. A second tube
returns the
coolant from the cooler 6520 to the exterior of the crankcase 6502. In another
embodiment,
multiple pieces of tubing may be used between the crankcase 6502 and the
cooler in order
to add tubing with extended heat transfer surfaces inside the crankcase volume
6504 or to
facilitate fabrication. The tubing joints and joints between the tubing and
the cooler may be
brazed, soldered, welded or mechanical joints.
Various methods may be used to join coolant tubing 6514 to cooler 6520. Any
known method for joining the coolant tubing 6514 to the cooler 6520 may be
used in
various embodiments. In one embodiment, the coolant tubing 6514 may be
attached to the
wall of the cooler 6520 by brazing, soldering or gluing. Cooler 6520 is in the
form of a
cylinder placed around the cylinder 6522 and the annular flow path of the
working gas
outside of the cylinder 6522. Accordingly, the coolant tubing 6514 may be
wrapped around
the interior of the cooler cylinder wall and attached as mentioned above.
Various
embodiments of cooler configurations may be found in U.S. Patent Application
Serial No.
13/447,990, filed April 16,2012 and entitled Stirling Cycle Machine (Attorney
Docket No.
184).
Returning to FIG. 25B, one method for joining coolant tubing 6514 to cooler
6520 is
to overcast the cooler around the coolant tubing. This method is described,
with reference to
FIGS. 26A and 26B, and may be applied to a pressurized close-cycle machine as
well as in
other applications where it is advantageous to locate a cooler inside the
crankcase.
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Referring to FIG. 26A, a heat exchanger, for example, a cooler 6520 (shown in
FIGS. 25A and 25B) may be fabricated by forming a high-temperature metal
tubing 6602
into a desired shape. In one embodiment, the metal tubing 6602 is formed into
a coil using
copper. A lower temperature (relative to the melting temperature of the
tubing) casting
process is then used to overcast the tubing 6602 with a high thermal
conductivity material to
form a gas interface 6604 (and 6532 in FIG. 25B), seals 6606 (and 6524 in FIG.
25B) to the
rest of the engine and a structure to mechanically connect the drive housing
6572 (shown in
FIG. 25C) to the heater head 6506 (shown in FIG. 25B. In one embodiment, the
high
thermal conductivity material used to overcast the tubing is aluminum.
Overcasting the
tubing 6602 with a high thermal conductivity metal assures a good thermal
connection
between the tubing and the heat transfer surfaces in contact with the working
gas. A seal is
created around the tubing 6602 where the tubing exits the open mold at 6610.
This method
of fabricating a heat exchanger advantageously provides cooling passages in
cast metal
parts inexpensively.
FIG. 26B is a perspective view of a cooling assembly cast over the cooling
coil of
FIG. 26A. The casting process can include any of the following: die casting,
investment
casting, or sand casting. The tubing material is chosen from materials that
will not melt or
collapse during the casting process. Tubing materials include, but are not
limited to, copper,
stainless steel, nickel, and super-alloys such as InconelTM. The casting
material is chosen
among those that melt at a relatively low temperature compared to the tubing.
Typical
casting materials include aluminum and its various alloys, and zinc and its
various alloys.
The heat exchanger may also include extended heat transfer surfaces to
increase the
interfacial area 6604 (and 6532 shown in FIG. 25B) between the hot working gas
and the
heat exchanger so as to improve heat transfer between the working gas and the
coolant.
Extended heat transfer surfaces may be created on the working gas side of the
heat
exchanger 6520 by machining extended surfaces on the inside surface (or gas
interface)
6604. Referring to FIG. 25B, a cooler liner 6526 (shown in FIG. 25B) may be
pressed into
the heat exchanger to form a gas barrier on the inner diameter of the heat
exchanger. The
cooler liner 6526 directs the flow of the working gas past the inner surface
of the cooler.
The extended heat transfer surfaces can be created by any of the methods known
in
the art. In accordance some embodiments, longitudinal grooves 6704 are
broached into the
surface, as shown in detail in FIG. 27A. Alternatively, lateral grooves 6708
(also shown in
enlarged section view FIG. 27D) may be machined in addition to the
longitudinal grooves
6704 (also shown in enlarged section view FIG. 27B) thereby creating aligned
pins 6710 as
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shown in FIG. 27C. In some embodiments, grooves are cut at a helical angle to
increase the
heat exchange area.
In an alternative embodiment, the extended heat transfer surfaces on the gas
interface 6604 (as shown in 26B) of the cooler are formed from metal foam,
expanded metal
or other materials with high specific surface area. For example, a cylinder of
metal foam
may be soldered to the inside surface of the cooler 6604. As discussed above,
a cooler liner
6526 (shown in FIG. 25B) may be pressed in to form a gas barrier on the inner
diameter of
the metal foam. Other methods of forming and attaching heat transfer surfaces
to the body
of the cooler are described in U.S. patent No. 6,694,731, issued Feb. 24,
2004, entitled
Stirling Engine Thermal System Improvements.
Additional coolant penetrating cold-end pressure vessel embodiments are
described
in U.S. Patent No. 7,325,399. It is to be understood that the various coolant
penetrating
cold-end pressure vessel embodiments referred to herein may be adapted to
function in a
multiple cylinder engine configuration.
Intake Manifold
Referring now to FIGS. 28A-28C, an intake manifold 6899, is shown for
application
to a Stirling cycle engine or other combustion application in accordance with
some
embodiments. Various embodiments of intake manifold 6899 are further disclosed
in U.S.
Patent No. 6,381,958. In accordance with some embodiments, fuel is pre-mixed
with air that
may be heated above the fuel's auto-ignition temperature and a flame is
prevented from
forming until the fuel and air are well-mixed. FIG. 28A shows one embodiment
including
an intake manifold 6899 and a combustion chamber 6810. The intake manifold
6899 has an
axisymmetrical conduit 6801 with an inlet 6803 for receiving air 6800. Air
6800 is pre-
heated to a temperature, typically above 900 K, which may be above the auto-
ignition
temperature of the fuel. Conduit 6801 conveys air 6800 flowing inward radially
with respect
to combustion axis 6820 to a swirler 6802 disposed within the conduit 6801.
FIG. 28B shows a cross sectional view of the conduit 6801 including swirler
6802 in
accordance with some embodiments. In the embodiment of FIG. 28B, swirler 6802
has
several spiral-shaped vanes 6902 for directing the flow of air 6800 radially
inward and
imparting a rotational component on the air. The diameter of the swirler
section of the
conduit decreases from the inlet 6904 to the outlet 6906 of swirler 6802 as
defined by the
length of the swirler section conduit 6801. The decrease in diameter of
swirler vanes 6902
increases the flow rate of air 6800 in substantially inverse proportion to the
diameter. The
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flow rate is increased so that it is above the flame speed of the fuel. At
outlet 6906 of
swirler 6802, fuel 6806, which in a one embodiment is propane, is injected
into the inwardly
flowing air.
In some embodiments, fuel 6806 is injected by fuel injector 6804 through a
series of
nozzles 6900 as shown in FIG. 28C. More particularly. FIG. 28C shows a cross
sectional
view of conduit 6801 and includes the fuel jet nozzles 6900. Each of the
nozzles 6900 is
positioned at the exit of the swirler vanes 6902 and is centralized between
two adjacent
vanes. Nozzles 6900 are positioned in this way for increasing the efficiency
of mixing the
air and fuel. Nozzles 6900 simultaneously inject the fuel 6806 across the air
flow 6800.
Since the air flow is faster than the flame speed, a flame will not form at
that point even
though the temperature of the air and fuel mixture is above the fuel's auto-
ignition
temperature. In some embodiments, where propane is used, the preheat
temperature, as
governed by the temperature of the heater head, is approximately 900 K.
Referring again to FIG. 28A, the air and fuel, now mixed, referred to
hereafter as
"air-fuel mixture" 6809, is transitioned in direction through a throat 6808
which has a
contoured fairing 6822 and is attached to the outlet 6807 of the conduit 6801.
Fuel 6806 is
supplied via fuel regulator 6824.
Throat 6808 has an inner radius 6814 and an outer dimension 6816. The
transition of
the air-fuel mixture is from a direction which is substantially transverse and
radially inward
with respect to combustion axis 6820 to a direction which is substantially
parallel to the
combustion axis. The contour of the fairing 6822 of throat 6808 has the shape
of an inverted
bell such that the cross sectional area of throat 6808 with respect to the
combustion axis
remains constant from the inlet 6811 of the throat to outlet 6812 of the
throat. The contour
is smooth without steps and maintains the flow speed from the outlet of the
swirler to the
outlet of the throat 6808 to avoid separation and the resulting recirculation
along any of the
surfaces. The constant cross sectional area allows the air and fuel to
continue to mix without
decreasing the flow speed and causing a pressure drop. A smooth and constant
cross section
produces an efficient swirler, where swirler efficiency refers to the fraction
of static
pressure drop across the swirler that is converted to swirling flow dynamic
pressure. Swirl
efficiencies of better than 80% may typically be achieved in practice. Thus,
the parasitic
power drain of the combustion air fan may be minimized.
Outlet 6812 of the throat flares outward allowing the air-fuel mixture 6809 to
disperse into the chamber 6810 slowing the air-fuel mixture 6809 thereby
localizing and
containing the flame and causing a toroidal flame to form. The rotational
momentum
65
generated by the swirler 6802 produces a flame stabilizing ring vortex as well
known in the
art.
Gaseous Fuel Burner
Definitions: As used in this section of the detailed description, the
following terms
shall have the meanings indicated, unless the context otherwise requires: Fuel-
Air
Equivalence ratio (phi.)=-Actual Fuel-Air Mass Ratio/Stoichiometric Fuel-Air
Mass Ratio.
The stoichiometric fuel-air mass ratio is defined as the mass ratio needed to
balance the
fuel+air chemical equation. The stoichiometric fuel-air mass ratio is well
known for
common fuels such as propane (0.0638 g fuel/g air) and calculable for gases
such as biogas.
FIGS. 29 shows one embodiment of the engine 7212 embodiment having a gaseous
fuel burner 7201. Various embodiments of the gaseous fuel burner 7201 are also
disclosed
in U.S. Patent Application Ser. No. 11/122,447, filed May 5, 2005, published
Nov. 10, 2005. This embodiment may be used in the context of a Stirling cycle
engine,
however, other embodiments of the machine are not limited to such
applications. Those
skilled in the art will appreciate that the present machine may have
application in other
systems, such as, with other types of external combustion engines.
The use of an ejector in a gaseous fuel burner advantageously can solve some
of the
challenges faced by the traditional gaseous fuel burners. First, using an
ejector can eliminate
the need for additional equipment, controls, and space, such as, a gaseous
fuel pump, fuel
control circuitry, and the associated components. Furthermore, using an
ejector such as a
venturi simplifies the fuel control system by eliminating the need for a
separate fuel control
scheme. Based on the corresponding rise of the vacuum with the airflow, and
subsequently,
an increased fuel flow, the burner power can be regulated by regulating the
airflow.
Accordingly, removing separate fuel control simplifies the development and
implementation of automatic burner control in a gaseous fuel burner with an
ejector.
Secondly, the corresponding rise of the vacuum with airflow also results in an
approximately steady fuel-air ratio despite changes in temperature and airflow
rates. The
resulting steady fuel-air ratio simplifies the fuel control and operation of
the burner, by
eliminating the need for complex exhaust sensor/feedback fuel control
mechanisms.
Referring to FIGS. 29, a gaseous fuel burner 7201 comprises an ejector 7240, a
heat
exchanger 7220, a combustion chamber 7250, and a blower 7200 (shown as 7300 in
FIGS.
30A). The term ejector as used here includes eductors, siphons, or any device
that can use
the kinetic energy of one fluid to cause the flow of another fluid. Ejectors
are a reliable way
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of producing vacuum-based fuel flow systems with low initial cost, lack of
moving parts,
and simplicity of operation.
Referring again to FIGS. 29, in a some embodiments, the ejector 7240 is a
venturi.
The venturi 7240 is positioned downstream of the outlet of the air preheater
or heat
exchanger 7220, in a venturi plenum 7241 and proximal to the combustion
chamber 7250.
A blower 7200 forces air through the venturi 7240. The flow of air through the
venturi
draws in a proportional amount of fuel through the fuel inlet ports 7279. The
fuel inlet ports
7279 are placed at the venturi throat 7244 where the throat has the lowest
pressure. The
ports 7279 are sized to produce plumes of fuel across the airflow that promote
good mixing
within the venturi 7240. This fuel-air mixture exits the venturi 7240 and
forms a swirl-
stabilized flame in the combustion chamber 7250. The venturi 7240 draws in an
amount of
fuel that is substantially linearly proportional to the airflow regardless of
airflow rates and
temperature of the air entering the venturi 7240.
In a some embodiments as shown in FIGS. 30A and 30B, placing the venturi 7340
between the air preheater 7320 and the combustion chamber 7350 promotes a
substantially
steady air-fuel ratio over a wide range of airflows and venturi temperatures.
FIGS. 30A is a
schematic drawing of the burner including the components of the burner such as
a blower
7300, a preheater 7320, a venturi 7340, and fuel supply 7372. The drawing also
includes a
load heat exchanger or heater head 7390 (also shown in FIGS. 31-33 as 7290).
The load
heat exchanger 7390 is the heat exchanger of the engine or process that
absorbs the thermal
power of the hot gases leaving the combustion chamber 7350 in the burner at
some elevated
temperature. The partially cooled burned gases then enter the exhaust side of
the air
preheater, where they are further cooled by incoming combustion air. FIGS. 30B
shows the
pressure map of the same components arranged linearly. The air pressure
supplied by the
blower, the fuel supply pressure, and the ambient pressure are all indicated.
The mass flow
rate (m') of fuel into the burner is controlled by the difference between the
fuel supply
pressure at 7372 and the pressure in the venturi throat 7344 (shown in FIGS.
29 as 7244)
and the fuel temperature at the dominant restriction:
m'FUEL.varies.(PFUEL-
PTHROAT)0.5/TFUEL0.5
The pressure in the throat (PTHROAT) is set by the pressure drop through
the
exhaust side of the preheater 7320 plus the pressure drop through the heater
head tubes 7390
minus the suction generated by the venturi throat 7344. The pressure drops
7320, 7390 and
the throat suction pressure 7344 are all proportional to the airflow rate and
the venturi
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temperature.
PTHROAT.varies.m'AIR2*TVENTURI
Combining these equations shows that the fuel flow will vary approximately
linearly
with the airflow:
m'FUEL.varies.[PFUEL-
(m'AIR2*TVENTURI)]0.5/T- FUEL0.5
Regulating the fuel pressure to near ambient pressure, the fuel flow is
approximately
linear with airflow.
m'FUEL.varies.m.AIR*(TVENTURI/TFUEL)0.5
Thus, locating the dominant fuel restriction 7378 (shown as 7278 in FIGS. 29)
within the
venturi plenum (shown as 7241 in FIGS. 29) provides for an approximately
steady fuel-air
ratio over a wide range of airflow rates and venturi temperatures.
m'FUEL/m'AIR.varies.constant
FIG. 29 shows one embodiment of the ejector such as the venturi. In this
embodiment, the size of the opening of the venturi throat 7244 determines the
amount of
suction present at the throat 7244. In a specific embodiment, the venturi
throat is
approximately 0.24 inches in diameter. Referring back to FIG. 29, fuel
delivery means are
coupled to the venturi 7240. The fuel delivery means may be manifolds, fuel
lines or fuel
tubes. The fuel delivery means may include other components such as a fuel
restriction
7278, fuel inlet ports 7279 and fuel valves (not shown). Fuel supplied by a
pressure
regulator 7272 flows through a manifold 7273 and fuel inlet ports 7279 into
the relatively
lower pressure in the throat 7244. In one embodiment the fuel inlet ports 7279
provide the
largest portion of the pressure drop in the fuel delivery means. Preferably,
making the fuel
inlet ports the largest restriction in the fuel delivery means assures that
the restriction occurs
at the venturi temperature and maximizes fuel-air mixing by producing the
largest possible
fuel plumes. Referring back to FIGS. 29, the fuel and air flow into the
divergent cone or
diffuser 7248 of the venturi, where static pressure is recovered. In the
diffuser 7248, the
entrained fuel mixes with the air to form an ignitable fuel air mixture in the
combustion
chamber 7250. The ignitable fuel-air mixture then enters the combustion
chamber 7250,
where the igniter 7260 may ignite the mixture, and the tangential flow induced
by a swirler
7230 creates a swirl-stabilized flame. Using an ejector 7240 to draw the
gaseous fuel into
the combustion chamber eliminates the need for a high-pressure gaseous fuel
pump to
deliver the fuel.
68
In one embodiment, the venturi 7240 is constructed from high temperature
materials
to withstand high temperatures and maintain its structural integrity. For the
embodiment of
FIG. 29, the dimensions of the venturi can be approximately 0.9 inches
diameter inlet and
outlets with an approximately 0.24 inches diameter throat. The half angles of
the convergent
cone and divergent cones can be 21° and 7° respectively and the
throat can be
0.25 inches long. In this embodiment, the venturi can be constructed from
InconelTM 600.
Alternatively, other high temperature metals could be used including, but not
limited to
Stainless Steels 310, 316L, 409 and 439, HastelloyTM C76, Hastelloylm X,
InconelTm 625
and other super alloys.
In one embodiment, as shown in FIGS, 29, a swirler 7230 is located upstream of
the
venturi 7240 and advantageously creates a tangential flow of air through the
venturi. As is
well known in the art, the tangential flow from the swirler can create an
annular vortex in
the combustion chamber, which stabilizes the flame. Additionally, the swirler
7230
increases the suction pressure at the venturi throat 7244 by increasing the
local air velocity
over the fuel inlet ports 7279. Adding the swirler allows the venturi throat
7244 to be made
larger for a given suction pressure. Furthermore, the swirling action induced
by the swirler
7230 can suppress fluctuations in the combustion chamber pressure from
propagating
upstream to the venturi 7240. Such pressure fluctuations can temporarily slow
or stop the
flow of fuel gas into the venturi 7240. The swirler 7230 thereby facilitates a
steady fuel-air
ratio in the combustion chamber for steady airflows. The swirler 7230 may be a
radial
swirler.
FIG. 31 depicts an embodiment where an automated controller 7288 adjusts a
variable restriction 7292 such as a variable flow valve in the fuel delivery
means to hold the
exhaust oxygen constant as measured by a wide-range lambda sensor or UEGO
7286. In
this embodiment, the automated scheme allows any fuel from biogas to propane
to be
connected to the burner and the control system can compensate for the changing
fuel
density. In this embodiment, the automated controller can restrict the fuel
path for dense
fuels such as propane and open up the fuel path for low-density fuels such as
methane and
biogas. Ignition would be accomplished by starting the variable restrictor
7292 in the fully
open position, which will produce the richest mixture then closing it until
the fuel-air
mixture is ignited. After ignition, the controller can control the fuel flow
to achieve the
desired exhaust oxygen level. It is also envisioned that such an embodiment
would allow the
fuel air ratio to be adjusted during warm-up to optimize efficiency and burner
stability.
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Referring now also to FIGS. 33-34 the gaseous fuel burner 7201 may be a high
efficiency burner for an external combustion engine such as a Stirling cycle
engine. In this
embodiment, the burner includes an oxygen sensor 7286 located in the exhaust
stream 7284
and a microprocessor/controller 7288 to automatically restrict the fuel flow
with the
variable restrictor 7292. Additionally, the burner includes a blower
controller 7702. The
blower controller 7702 may be adjusted by the microprocessor/controller 7288
to match the
Stirling engine power output with the load. In this embodiment, the burner
temperature is
held constant by varying the engine speed and the engine power output is
automatically
adjusted by setting the blower speed. Accordingly, in this embodiment, the
burner may
burn most gaseous fuels, including fuels without constant properties such as
biogas.
Referring now also to FIG. 34, fuel may be delivered directly into the venturi
at a
point proximal to the venturi throat 7244. This embodiment may include a
swirler 7230 to
accommodate the fuel delivery means such as a fuel line or fuel tube. The
swirler 7230 may
be an axial swirler positioned in the venturi 7240 and upstream of the venturi
throat 7244. In
operation, the delivered fuel is entrained with the motive air to form the
fuel-air mixture. In
various embodiments, manual or automatic control mechanisms are adaptable to
this
alternate fuel delivery embodiment.
Referring back to FIG. 29, the gaseous fuel burner further comprises an
igniter 7260
and a flame-monitoring device 7210. Preferably, the igniter 7260 is an
excitable hot surface
igniter that may reach temperatures greater than 1150° C.
Alternatively, the igniter
7260 may be a ceramic hot surface igniter or an excitable glow pin.
With continuing reference to FIG. 29, other embodiments include a flame-
monitoring device 7210. The flame-monitoring device 7210 provides a signal in
the
presence of a flame. For the safe operation of the any burner, it is important
that the fuel be
shut-off in the event of a flameout. The monitoring device for flame sensing
is the flame
rectification method using a control circuit and a flame rod.
Flame rectification, well known in the art, is one flame sensing approach for
the
small, high efficiency gas burners. The device uses a single flame rod to
detect the flame.
The flame rod is relatively smaller than the grounded heater head and it is
positioned within
the combustion flame. In this flame rectification embodiment, the control unit
electronics
are manufactured by Kidde-Fenwal, Inc., and the flame rod is commercially
available from
International Ceramics and Heating Systems
70
Preferably, the flame-monitoring device uses the hot surface igniter as the
flame rod.
Alternatively, the flame-monitoring device may be either remote from the hot
surface
igniter, or packaged with the igniter as a single unit.
Alternatively, an optical sensor may be used to detect the presence of aflame.
A
preferred sensor is an ultraviolet sensor with a clear view of the flame brush
through an
ultraviolet transparent glass and a sight tube.
It is to be understood that the various fuel burner embodiments described
herein may be adapted to function in a multiple burner configuration as
disclosed in U.S.
Patent Application Ser. No. 13/447990, filed April 16, 2012.
Some embodiments my control or modulate the flow of gaseous fuel to the burner
with reciprocating pumps as disclosed in U.S. Patent Application Ser. No.
13/447990, filed
April 16, 2012.
Referring now to FIG. 35, a cross section of an engine 9200 is shown. The
engine
9200 is similar to the one described above with respect to FIG. 4, however,
includes another
embodiment of the rocking drive mechanism. The engine 9200 shown in FIG. 35
includes a
rocking drive mechanism including link rods 9210, 9210, a rocking beam 9214, a
rocking
pivot 9224, a connecting rod 9216, a connecting pivot 9218, end pivots 9220,
9222, and a
crankpin 9226. Although this engine 9200 is an example of another embodiment
of the
rocking drive mechanism as discussed above, the components function in a
similar fashion
.. however, this embodiment includes a number of additional benefits.
The configuration of the connecting rod 9216, rocking beam 9214, and
connecting
pivot 9218 limit the loads on the connecting rod. This configuration
additionally allows for
the use of larger bearings, including standard sized tri-metal bearings.
Additionally, the
increased distance between the rocking pivot 9224 and the connecting pivot
9218 increases
the mechanical advantage of the rocking pivot 9224, thus reducing the loads on
the
connecting rod bearings
In this embodiment of the engine 9200, the side loads on the link rods 9210,
9210
has been increased. However, as discussed above, the engine 9200 is an oil
lubricated
engine, thus, concern with limiting the side loads on the link rods 9210, 9210
has been
reduced. Thus, in the embodiment shown in, for example, FIG. 4, the link rods
are longer
and the loads on the connecting rod are higher. In the embodiment shown in
FIG. 35, the
link rods 9210, 9210 are shorter and the load on the connecting rod 921. 6 is
decreased.
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In some embodiments, the oil pump is a Gerotor pump driven by the crankshaft
through a spline connection. In some embodiments, the oil pump is driven by
the
crankshaft by a gear.
Referring now to FIG. 36A, an embodiment of a Stirling cycle machine is shown
in
cross-section and designated generally by numeral 9600. While the Stirling
cycle machine
9600 will be described generally with reference to the embodiment shown in
FIG. 36A and
36B, it is to be understood that many types of machines and engines, including
but not
limited to refrigerators and compressors may similarly benefit from various
embodiments
and improvements which are described herein, including but not limited to,
external
combustion engines and internal combustion engines. In particular, the present
embodiment
of the Stirling cycle machine is directed to improving the efficiency and
operation of a 10
Kilowatt (Kw) Stirling cycle machine, although any other power output level
are certainly
contemplated and encompassed within the following disclosure of a machine or
engine, that
achieves high efficiency, long durability and low cost targets based on
simultaneously
utilizing optimized mechanical and operational control systems with existing
Stirling cycle
machine platforms.
The engine 9600 shown in cross-section in FIG. 36A includes generally a
crankcase
9610 housing the drive components of the engine and a work space 9620
containing the
working gas and/or fluid and gas and/or fluid compression and expansion
related
components. Inside the crankcase 9610 is an embodiment of a rocking beam drive
mechanism 9601 (the term "rocking beam drive" is used synonymously with the
term
"rocking beam drive mechanism") for an engine, such as a Stirling engine,
having linearly
reciprocating pistons 9602 and 9604 housed within cylinders 9606 and 9608,
respectively.
As discussed previously, rocking beam drive 9601 converts linear motions of
pistons 9602
and 9604 into the rotary motion of a crankshaft 9614. Rocking beam drive 9601
has a
rocking beam 9616, rocker pivot 9618, a first coupling assembly 9619, and a
second
coupling assembly 9621. Pistons 9602 and 9604 are coupled to rocking beam
drive 9601,
respectively, via first coupling assembly 9619 and second coupling assembly
9621. The
rocking beam drive 9601 is coupled to and drives crankshaft 9614 via a
connecting rod
9622.
This embodiment shown in FIG. 36A and 36B is an inverted rocking beam design
similar to that disclosed in FIG. 35 and incorporates the same advantages and
benefits as
discussed therein. An important advantage of the inverted rocking beam
arrangement
having the crankshaft 9614 located relatively below the rocking beam mechanism
9601 of
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the machine ensures that the structural arrangement and alignment of the
piston rods 9624
and cross-head coupling means 9634 which connect the piston rods 9624 to the
rocking
beam drive 9601 do not have to account for the size of the crankshaft 9614 and
related
components. This arrangement facilitates a larger load carrying capacity
conrod bearing
9615 on the connecting rod 9622, better mechanical advantage developed by the
rocking
beam 9616 and space for such larger conrod bearings 9615. The arrangement also
relieves
space constraints of the pistons 9602 and 9604, piston rods 9624 and cylinders
9606 and
9608 which can occur with the crankshaft located above the rocking beam drive
and
between the piston shafts 9624. With the rocking beam 9616 now located above
the
.. crankshaft 9614, there are no longer space restrictions around the
crankshaft rocking beam
9616 and a larger wrist pin bearing 9628 can be provided to better support the
connecting
rod 9622 and rocking beam connection 9601.
Also, with the inverted design, the rocking beam 9616 can be designed to
reduce the
load on the connecting rod 9622 wrist pin bearing 9628 and conrod bearing 9615
by
adjusting the lever arm ratio A and B seen in FIGS. 37A and 37B between the
rocking shaft
pivot 9718, the connecting rod wrist pin bearing 9728 and between the rocker
shaft pivot
9718 and the lever arm of the piston acting at 9729. For example, as seen in
FIG. 37A, the
bearing load on the connection rod 9728 is greater where the conrod bearing
ratio A/B is 1.6
relative to the piston connection. In FIG. 37B, the rocking beam 9616 is shown
having a 1.0
ratio which essentially equates the distances of the two connection points
9728' and 9729
about the rocking shaft pivot 9718 and therefore correspondingly balances the
load on the
crankshaft 9614 to be the same or similar to that developed by the piston
shaft 9624 and
significantly lower than the load transmitted with the bearing ratio A/B of
1.6. It is to be
appreciated that other embodiments of a rocking beam drive besides the
inverted rocking
beam drive may also incorporate the benefits of the disclosed rocking beam
9616 as well.
Referring also to FIGS. 36A-36E, the alignment of the pistons 9602 and 9604,
piston rods 9624 and cylinders 9606 and 9608 in conjunction with the crankcase
9610 is of
critical importance for power transmission through the pistons 9602 and 9604
and piston
rods 9624, providing for reduced wear on the piston rings and to the dynamic
alignment and
reciprocating nature of the piston rods and high pressure piston rod bearings
9630. The
crankcase 9610 contains most of the rocking beam drive 9601 and is positioned
below the
cylinder housing 9631. The crankcase 9610 defines a space to permit operation
of rocking
beam drive 9601 having the crankshaft 9614 located below rocking beam 9616, a
connecting rod 9622, and furst and second coupling assemblies 9619 and 9621.
Pistons
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9602 and 9604 reciprocate in respective cylinders 9606 and 9608 as also shown
in FIG. 36
and cylinders 9606 and 9608 extend above crankcase 9610, through the cylinder
housing
9631, and into the heater heads.
The cross-heads 9634 and cross-head bores 9635 have a tolerance that is
difficult to
align with that of the mating cylinder liners in the cylinder housing 9631
during what is
referred to as "stack up" i.e. the joining of the separate parts of the
vessel, and therefore any
difference in concentricity between these two elements when they are assembled
together,
creates a potential for misalignment, where potentially, the piston rod 9624
could sit askew,
or at an angle and therefore the piston may reciprocate non-coaxially to the
cross-heads
9634. The cylinder liner bores 9606, 9608, the cylinder gland locating
diameter and the
cross-head locating diameter of the cross-head bore 9635 all must be in
alignment. To
alleviate this issue and potential for misalignment, all three of these
diameters are bored
together in the same set-up and essentially simultaneously in the same
operation resulting in
very close tolerances of the diameters and the concentricity of these elements
is maintained
as closely as possible based on machining tolerances. These elements may also
be
manufactured and bored in other ways as well including but not limited to with
alignment
jigs and separate boring process that can produce the requisite tolerances to
ensure that any
angular deviation of the piston is maintained within an acceptable range.
Also as shown in another embodiment in FIG. 36B, to improve the concentricity
of
the piston and piston rod 9624, each piston rod 9624 is provided with a
tapered end 9625 at
each end of the rod 9624 to wedge the first end of the piston rod into the
cross-head 9634.
The tapered end 9625 facilitates the location, resting and clamping (L,R,C)
between all
elements for a proper location of the piston rod 9624 with the diameter doing
the locating,
the taper 9625 doing the resting, and a nut 9633 of the end doing the
clamping. In the wedge
connection provided by the taper, the wedge can lock itself in place because
of the loads
developed by the piston. The wedge or tapers on the ends of the piston rods
are essentially
jammed more and more firmly into the crossheads 9634 at the lower end of the
piston rods
9624 and correspondingly into the piston at the upper end of the piston rod. A
nut 9633
may be used to facilitate the connection with the cross-head 9634 in case the
rod comes
loose, but in almost every case the wedge will maintain the appropriate
connection of the
piston rod 9624 to the piston above, and the cross-head 9634 below.
To facilitate the assembly of the tapered piston rods 9624 where the taper is
essentially a reduction in diameter of the ends of the piston rod 9624 along a
portion of the
piston rod, the piston 9602, 9604 is manufactured from two separate parts, a
piston base
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9643, and a piston shell 9645 better shown in FIGS. 36C, D and E. The piston
base and
shell can be matingly threaded where the piston base 9643 defines a threaded
inner diameter
surface wall 9647 corresponding to a threaded outer surface wall 9649 of the
piston shell
9645. Other connection arrangements between the base and shell are possible as
well to
facilitate the connection of the two piston elements. The piston base 9643 is
provided with
a receiving bore 9651 which may be a constant diameter bore, or a tapered bore
to receive
the tapered end of the piston rod 9624. To assemble these elements, the
tapered piston rod
9624 is inserted into the piston base 9643, clamped in place with a desired
pre-load, and
then the shell 9645 is threaded onto the base 9643 to complete the assembly.
The reason for
the two-part piston is that to appropriately clamp and pre-load the piston rod
9624 to the
base 9643, the assembly process necessitates access to the inside of the
piston, and hence
the two-part shell and based design facilitates the clamping process. Other
manufacturing
techniques may also be used to appropriately attach the tapered piston rod
9624 and piston
9602, 9604 without the necessity for a two-part piston as described above.
Another important aspect of the present embodiment is an increased volume of
the
combustion space in the heater head. To provide more volume for the combustion
of the
burner to take place and heat the tubes, an upper most portion 9655 of the
cylinders 9606,
9608 is provided with a larger diameter than the remainder lower portion of
the cylinders,
giving the cylinders 9606, 9608 to some extent a mushroom-shaped profile. The
benefit of
.. this includes but is not limited to the ability to move the heater tubes
9659 farther out from
an axial center of the cylinders 9606, 9608, thereby increasing the diameter
and combustion
volume above the cylinder inside the heater tubes 9659 and/or to accommodate a
larger
diameter tube to handle more working gas and fluid through the heater tubes
9659.
In another embodiment shown in FIG. 38B, a Gerotor displacement pumping unit
is
driven by the crankshaft 9814. The Gerotor pump uses an inner rotor 9844
having one less
gear tooth 9846 than the surrounding outer rotor 9848. During part of the
rotation cycle, the
area between the inner and outer rotor increases, creating a vacuum that draws
fluid through
an intake. The area between the rotors then decreases, causing compression
allowing oil to
be pumped out to the mechanical parts of the engine. The Gerotor pump is
driven coaxially
and directly from the crankshaft 9814 without the transmission losses of the
helical drive
gear, making the engine construction and assembly more efficient and less
expensive than
the construction and components of the helical drive gear to the gear pump.
The
construction and assembly is easier because the Gerotor pump is directly
driven by the
crankshaft 9814, whereas there are significant mechanical losses associated
with the
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previous gear pump. In other embodiments different pumps besides a Gerotor
pump may be
used, which include but are not limited to, gear pumps, piston pumps, rotary
gear pumps,
hydraulic pumps and diaphragm pumps for example and that other embodiments of
a
rocking beam drive besides the inverted rocking beam drive may incorporate the
benefits of
5 the Gerotor pump or similar direct drive pump.
High Pressure Rod Seals
The present embodiment of the Stirling cycle engine maintains the working
space
9620 and the working gas and/or fluid at a relatively high pressure, generally
in the range of
1200-1800 psi, and more preferably about 1500 psi. It is of course necessary
to ensure that
10 the working gas and/or fluid is essentially sealed in the working space
9620 so that it does
not escape into the crankcase 9610 and the environment. A critical place for
such leakage
of working fluid to occur is around the piston rods 9624, which extend and
reciprocate
between the working space 9620 and the crankcase 9610. To minimize such
leakage, a high
pressure piston rod seal 9630 is provided below the respective cylinders 9606
and 9608 and
15 between the working space 9620 and the crankcase 9610.
With a significantly higher pressure in the working space 9620 relative to the
crankcase 9610, a certain amount of working gas is anticipated to leak through
the high
pressure rod seals 9630. However, it is imperative to minimize the leakage
without
significantly affecting the reciprocating efficiency of the pistons and the
engine. Also, as
20 will be discussed in further detail below, an airlock and working fluid
recapture system may
be used in conjunction with the high pressure seals to capture certain amounts
of such
leaking working gas and/or fluid. Any working gas which leaks into the air
lock between
the working space 9620 and the crankcase 9610 can be drawn into an accumulator
and
supplied back into the workspace when necessary. Before more completely
discussing such
25 .. an airlock and recapture of working fluid, the present discussion is
focused on the use of the
high pressure rod seals 9630 between the working space 9620 and the crankcase
9610 to
ensure the most effective working fluid pressure and gas containment.
A mechanical embodiment of the high pressure rod seal 9930 is in FIG. 39. It
should be understood that such a rod seal is intended to be utilized not only
in the Stirling
30 engine embodiments described herein but also in other engines or
mechanisms with similar
reciprocating pistons.
In a mechanical embodiment of the high pressure piston rod seal 9930, shown in
better detail in FIG. 39, a substantially symmetrical hemispherical shaped
piston sleeve
9960 is supported by an upper seal support 9965 and a lower seal support 9966
inside a seal
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cavity defined inside a seal housing 9951. The symmetry of this hemispherical
shaped
piston sleeve 9960 provides more consistent wear across the length of the
sleeve 9960 as
compared to the wedge rod seal 9930 described above which focuses the radial
wear at one
end of the sleeve. The hemispherical surface 9963 of the piston sleeve 9960
bears on an
inner respective bearing surface of each of the upper and lower seal supports
9965, 9966. A
wear support clamp 9967 is provided axially disposed above the upper seal
support 9965
which forces the upper and lower seal supports 9965, 9966 into biased contact
with the
piston sleeve 9960. A gap G may be provided between the upper and lower seal
supports
9965, 9966 to accommodate any wear that may occur on abutting surfaces in the
seal. As
wear occurs, the abutting surfaces in the seal may be reduced so that as the
sleeve bearing
wears, the upper and lower seal supports 9965, 9966 are biased towards one
another by the
support clamp 9967. The gap G permits the upper and lower seal supports 9965,
9966 to
move closer to one another as the seal wears without interfering with one
another and so
maintaining contact with the hemispherical shaped outer surface 9963 of the
piston sleeve
9960.
A still further embodiment of a high pressure rod seal shown in FIG. 40A
includes a
spring energized lip seal 10003 generally comprising a seal jacket, made from
PTFE or
graphite for example, and a spring (not shown) circumferentially secured
within a groove or
between lips 10007 of the seal 10003. When the spring energized lip seal
100003 is seated
.. in the housing, the spring lip seal 10003 is under compression, forcing the
jacket lips 10007
against the respective adjacent walls of the seal block 10011 and the surface
of the
reciprocating piston 10024, thereby creating a leak free seal. The lip seal
10003 provides
permanent resilience to the seal jacket 10005 and compensates for jacket wear
and hardware
misalignment or eccentricity. System pressure also assists in energizing the
seal jacket
10005. Spring loading assisted by system pressure provides effective sealing
at both high
and low pressures. Spring energized lip seals are highly durable and designed
for static,
rotary and reciprocating applications in temperatures from cryogenic to +600F
as well as
pressures from vacuum to 25,000 psi, and to survive most corrosive
environments.
A spring cup retaining cylinder 10008 is set around the piston rod 10024 and
supported on a lower collar 10006. The retaining cylinder 10008 maintains a
circumferential space about the piston rod 10024 in which the lip seal 10003
is maintained.
The lip seal 10003 can be a PTFE and graphite ring supported around an outer
circumference by the retaining cylinder 10008 and frictionally slidably
engages the piston
rod 10024 to create the high pressure spring energized lip seal. The spring
(not shown)
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inside the lip seal 10003, along with the higher pressure of the working
space, forces the lip
seal 10003 against the respective piston rod 10024 and retaining cylinder
wall, and also
maintains the lip seal 10003 set down in the retaining cylinder 10008
generally against the
lower collar 10006.
A hydraulic embodiment of a high pressure piston rod seal can facilitate an
efficient
and long term seal between the working space and the airlock. FIG. 40B
discloses a
hydraulic high pressure piston rod seal 10021 set inside the rod seal cavity
of a test rig. A
rod seal sleeve 10023 circumferentially encompasses the piston rod 10024 and
defines a
pressure space 10025 between a wall of the test rig and an outer surface of
the rod seal
sleeve 10023. A hydraulic fluid pressure line 10027 communicates with pressure
space
10025 to provide the appropriate fluid pressure to maintain the rod seal
sleeve 10023 in
sealing engagement with the piston rod 10024. A sensor (not shown), such as a
piezo-
electric pressure sensor, can be provided in the pressure space 10025 and on
the rod seal
sleeve 10023 to ensure that the appropriate pressure and flexure is actuating
the rod seal
sleeve 10023 and providing the appropriate sealing pressure against the piston
rod 10024.
The inner surface of the rod seal sleeve 10023 slidably engages along the
piston rod 10024
as the rod reciprocates and the rod seal sleeve 10023 is influenced radially
inwards by the
hydraulic pressure fluid in the pressure space 10025. As the rod seal sleeve
10023 wears,
the hydraulic fluid pressure in the pressure space 10025 can be increased to
ensure that the
rod seal sleeve 10023 is motivated radially towards the piston rod 10024 to
maintain
slidable engagement with the piston rod.
It is to be appreciated that the above disclosed embodiment of high pressure
rod
seals are intended only as examples and that the machines described herein are
not limited
to these examples, and that other embodiments of high pressure rod seals may
also be used
to ensure that the high pressures used in Stirling engines, or any other
engine for that matter,
are maintained in the appropriate working space, crankcase and other engine
compartments
as necessary.
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Rolling Diaphragm Seal
Turning to FIGS. 41A and 41B, and referring back also to FIGS. 10A-D, in
certain
embodiments a rolling diaphragm 10190 is used in conjunction with the piston
rods 10124
to prevent the escape of lubricating fluid from the crankcase 10110 up past
the rods 10124
and into the working space 10120 and regenerator. If the lubricating fluid
necessary for the
rocking drive can bypass the piston rod seals, it can potentially damage the
working space,
clog the regenerator and contaminate the working fluid or gas of the engine in
the cylinders.
To facilitate the appropriate rolling and flexing of the diaphragm 10190, a
pressure
differential is maintained across the rolling diaphragm 10190 so that
preferably the pressure
above the diaphragm 10190 is slightly greater than the pressure in the
crankcase. The seal
is thus essentially inflated into the crankcase, which facilitates the
diaphragm 10190
maintaining its desired form as it rolls and flexes with the reciprocating
piston rod 10124.
This alleviates stresses on the circumferential sealing points so the seal is
not compromised.
It is generally necessary to place a differential of approximately 15 PSI
across the
diaphragm 10124 to properly inflate the seal so that it conforms to the shape
of the bottom
seal piston 10195 as it moves with the piston rod 10124. It is to be
appreciated that the
pressure differential maintained across the rolling diaphragm 10190 is not
limited to 15 PSI.
Rolling diaphragms made of stronger materials or having a particular shape may
be able to
sustain a higher differential or operate at a lower differential as the case
may be. In
embodiments of the stirling cycle engine where the working space 10120 is at a
relatively
high pressure 1500 PSI ¨ 1800 PSI, the crankcase 10110 must be charged with a
pressure
for instance of 1485 PSI, which is approximately 10-15 PSI less than that of
the working
space at 1500 PSI. Although it is possible to regulate these larger pressures
to maintain the
10-15 PSI difference across the diaphragm, it is difficult and adds to the
complexity of the
machine.
The rolling diaphragm 10190 may be manufactured by injection molding or hot
compression molding. In hot compression molding of the rolling diaphragm
10190, it can
be more difficult to control material properties but injection molded
diaphragms have
shown in testing a better transition of dynamic stresses across the profile of
the rolling
diaphragm 10190 as it transitions and rolls with the reciprocation of the
piston rod 10124.
Testing on the materials used to fabricate the rolling diaphragm 10190
indicate chopped
fiber is most successful for example but not limited to, nitrile with Kevlar
fiber or Fab-
Air .
79
FIGS. 41A and 41B disclose an embodiment of the rolling seal or diaphragm
10190
having a profile which facilitates the dynamic rolling translation of the
diaphragm. As
discussed in previously herein, the pressure differential that is placed
across the seal allows
the seal to act dynamically to ensure that the rolling diaphragm 10190
maintains its form
throughout its dynamic range of motion. As previously discussed, the pressure
differential
causes the rolling diaphragm 10190 to conform to the shape of the bottom seal
piston 1310
with reference to FIG. 10A as it moves with the piston rod10124 , and prevents
separation
of the diaphragm 10190 from the surface of the piston rod 10324 during
operationit is
desirable to lower the amount of inflation of the rolling diaphragm 10190
without the
diaphragm buckling or separating, i.e., deviating from a consistent dynamic
axial and radial
rolling of the diaphragm 10190 along the diaphragm profile with the axial
reciprocation of
the piston rod 10124. As discussed above, the inflation of the diaphragm is
provided by the
pressure differential across the rolling diaphragm 10190. To accomplish this,
it has been
found that particular structural profiles facilitate the conservation of
material and consistent
rolling of the diaphragm.
The cross-section in FIG. 41A-B shows a profile view of the molded form of the
diaphragm of the present embodiment about a diaphragm axis L. For purposes of
describing the diaphragm structure the inner edge 10192 as being the top 10194
of the
diaphragm and the outer edge 10193 is the bottom of the diaphragm as shown in
the figures.
The diaphragm has a lateral wall 10190 extending axially and radially relative
to axis L
from the inner edge 10192 to the outer edge 10193; the lateral wall is
composed of several
sections. A top fillet section 10198 turns the material approximately 90
degrees from the
top of the diaphragm 10190 as shown, to a sidewall section 10196 substantially
parallel to
the piston rod 10124 and axis L. The sidewall section 10196 in turn then turns
towards the
outer edge10193. Before reaching the outer edge 10193, the sidewall section
merges
contiguously into a chamfer section 10199, which while still depending axially
from the
sidewall section 10196, extends from the sidewall 10190 in a greater radial
degree relative
to axis L to connect with the outer edge of the diaphragm 10193. The sidewall
section
10196 may be parallel to the axis L or may also have a radial component which
slightly
angles the sidewall section 10196 radially away from the axis L. In either
case the chamfer
section 10199 extends to a greater radial degree from axis L than the sidewall
section
10196. A bottom fillet 10197 connects to the outer edge 10193 defining the
bottom of the
diaphragm as drawn. The outer edge 10193 like the inner edge 10192 is provided
with a
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thickened circumferential lip, which can be secured inside a matching groove
formed in the
vessel joint.
The cross-section shown in FIGS. 42A and 42B is a profile view of the molded
form
of another embodiment of the rolling diaphragm 10290 of the present embodiment
about a
5 diaphragm axis L. Like reference numbers for this embodiment correspond
to the same or
similar elements in the previous rolling diaphragm embodiment. For purposes of
describing
the diaphragm structure, the inner edge 10292 is the top of the diaphragm and
the outer edge
10293 is the bottom of the diaphragm 10290 as shown in the figures. The
diaphragm 10290
has a lateral wall 10296 extending axially and radially relative to axis L
from the inner edge
10 10292 to the outer edge10293; the lateral wall here is again composed of
several sections. A
top fillet 10294 section turns the material approximately 90 degrees from the
top of the
diaphragm 10290 as shown, to the sidewall section 10296, which depends both
axially and
radially outwards towards the bottom of the diaphragm along the axis L. Before
reaching
the outer edge 10293, the sidewall section 10296 merges contiguously into a
bottom fillet
15 10299 to extend towards the outer edge 10293 of the bottom of the
diaphragm as drawn.
An outer lip 10197 similar to the thickened circumferential lip 10295 of the
inner edge
10192 is provided, which are secured inside a matching groove formed in the
vessel or
crankcase joint which secures and seals the outer edge 10293 of the diaphragm.
The injection molding of the diaphragm is important because the gating methods
and
20 other molding techniques, characteristics, methods and specifications
can affect the fiber
alignment and molecular alignment of the diaphragm material during the molding
process.
These material characteristics are important because this can affect the hoop
stress of the
diaphragm. For example, if the material is gated at one end and overruns an
opposing end
of the mold, the fibers can be aligned in a particular direction to optimize
the hoop strength
25 of the diaphragm while potentially enhancing the flexible and rolling
characteristics of the
final diaphragm element.
It is very important in the dynamic rolling actuation of the diaphragms 10190,
10290
that no imperfections or particles including fluid particles such as oil
droplets are disposed
on the surfaces of the bottom seal piston or on the adjacent vessel wall
surrounding the
30 bottom seal piston. Such fluid particles, most likely oil, are
detrimental to the rolling
actuation of the diaphragm against the respective crankcase surfaces, because
they cause
stress points on the diaphragm.
Turning to FIG. 43 another embodiment of the rolling diaphragm includes a
first and
second rolling diaphragm 10391, 10393 to make what is essentially a double
bellows
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system 10392. A double bellows system 10392 can facilitate the elimination of
the 10-15
PSI pressure differential between working space and airlock and/or crankcase
by providing
the appropriate expansion pressure between the double bellows themselves. The
double
bellows include first and second rolling diaphragms 10391. 10393 which are
oppositely and
axially aligned along the piston rod, and define a space therebetween with a
light oil
contained between the diaphragms and pressure charged between the double
bellows. The
incompressible oil prestresses the diaphragms and facilitates the consistent
rolling of the
diaphragm as the piston rod 10324 reciprocates along its axis.
Airlock and Working Fluid Recapture System
The power, life and value of a Stirling engine can be maximized by building an
oil
lubricated drive and sealing the work-space from the oil. Oil lubricated
drives allow high
powers and are inexpensive compared to drives based on rolling elements. It is
essential to
isolate the oil in the drive from the workspace. Even oil mist will migrate to
the hot end of
the working space, where the oil will breakdown and the resulting carbon will
clog the heat
exchanger. Flexible membranes or bellows such as the rolling diaphragms
discussed above
that attach to the moving piston rod and the structure provide an oil and gas
tight seal
between the oil filled crankcase and the workspace, ensuring that the
lubricant is maintained
in the crankcase. In order to function for thousands and millions of cycles, a
small pressure
difference must be maintained across the bellows.
An important aspect of the rolling diaphragm and oil lubricated crankcase
relates to
the use of an airlock 10401 and an airlock pressure regulation system 10411 as
shown in
FIGS. 44A and 44B. The airlock pressure regulation system 10411 provides the
benefit of
ensuring working gas escaping from the working space 10403 is returned to the
working
space, provided that the working gas does not leak into the environment or
atmosphere.
which would require replenishment of the working gas, and that an appropriate
pressure
differential is maintained across the rolling diaphragms as described above.
The airlock
pressure regulation system 10411 permits an easily serviceable bottom end i.e.
crankcase
10410 if, as in the embodiment disclosed in FIG. 48, the crankcase is intended
to be
maintained essentially at atmospheric pressure.
As shown in FIG. 44A relating to a pressurized crankcase 10410 at
approximately
1485 PSI, in order to maintain an appropriate working space pressure and
airlock pressure
regulation, an airlock space 10401 is provided between the working space 10403
and the
crankcase 10410 at a pressure of, for example 1500 PSI, so that the
substantially greater
pressures in the working space 10403 should not significantly influence the
air lock space
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10401 and any pressure and working gas leaking from the working space 10403
into the air
lock can be captured and accumulated as described below with respect to the
airlock
pressure regulator and returned to the airlock and working space and not
merely escape into
the crankcase and environment.
It is to be understood that airlock space 10401 is intended to maintain a
constant
volume and pressure necessary to create the pressure differential necessary
for the function
of rolling diaphragm 10490 as previously described. In the present embodiment
the airlock
10401 may or may not be sealed off from working space 10403 with high pressure
rod seals
10430. In any case, the pressure of airlock space is desired to be maintained
at essentially
1500 PSI and equal to the mean pressure of working space 10403. The pressure
in the
working space 10403 may vary at least +/- 300 PSI so the intention of the
airlock space
10401 is to insulate the diaphragms from such fluctuations and maintain itself
at around the
necessary pressure, by way of example here 1500 PSI, relative to the 1485 PSI
charged in
the crankcase 10410. To facilitate the equalization of pressures between the
working space
10403 and the airlock space 10401, a small opening or pressure equalization
orifice 10404
communicates between the working space 10403 and the airlock space 10401. The
crankcase 10410 must be charged to 1485 PSI, and be maintained at
approximately 15 PSI
less than the airlock space 10401 so that the appropriate pressure is applied
to the rolling
diaphragm 10490 to ensure the proper dynamic movement of the diaphragm.
In this pressurized crankcase 10410 embodiment an airlock pressure regulator
10411, a pump 10412 and relief valve system is provided between the crankcase
10410 and
the air lock space 10401 to maintain the exemplary 15 PSI pressure
differential
therebetween. Other predetermined pressure differentials may also be
maintained depending
on the diaphragm material and the design of the entire airlock pressure
regulation system.
In its most general form, an uptake line 10416 communicates from the
pressurized
crankcase 10410 to a a filter 10418, a pump 10412 (having a check valve on its
outlet), and
a pressure regulator 10413 in parallel with the pump 10412 and filter 10418
for returning
pressurized working gas back to the air lock 10401 and so maintains the
pressure
differential between the airlock space 10401 and the crankcase 10410 and
consequently
across the rolling diaphragm 10490. This airlock pressure regulator system
10411 is
described more completely with respect to FIG. 44B.
The airlock pressure regulator 10411 regulates the pressure difference between
the
airlock 10401 and the crankcase 10410. When the engine is turning, the airlock
pressure
regulator 10411 keeps the airlock pressure preferably 10 to 14 PSI above the
crankcase
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pressure although a range of 5 to 20 PSI is possible and other pressure
differentials can be
accomplished by the regulator as well. When the engine is off, the airlock
pressure
regulator 10411 keeps the airlock pressure preferably less than 15 PSI above
the crankcase
pressure and not more than 5 PSI below crankcase pressure. It is permissible
to have a
greater fluctuation of pressure differential when the engine is off since
there is little or no
dynamic forces being applied to the rolling, diaphragms 10490 via moving
pistons.
The airlock pressure regulator 10411 performs several important functions. The
airlock pressure regulator 10411 uses a pump 10412 to move pressurized gas
from the lower
pressure crankcase 10410 into the airlock 10401, thereby maintaining the
airlock 10401 at a
higher pressure. The airlock pressure regulator 10411 relieves excess pressure
between the
airlock 10401 and crankcase 10410 volumes. A bidirectional regulator 10413
vents some of
the airlock gas into the crankcase 10410, when the airlock pressure is
preferably 15 PSI
above the crankcase and vents in the opposite direction, venting gas from the
crankcase
10410 to airlock 10401, when the airlock pressure is more than 5 PSI below the
crankcase
pressure. Also, a filter 10418 in the airlock pressure regulator 10411 filters
out the oil from
the crankcase gas before it enters the airlock volume.
The components of the preferred embodiment are the mechanical pump 10412, the
bidirectional pressure regulator 10413, an oil filter 10418, a pump pressure
switch 10417 to
control the pump 10412 and a controller pressure switch 10419 to signal the
engine
controller C. An example of the mechanical pump is the Linear AC 0410A pump by
Medo.
Other pumps could certainly be used as well. The important qualities of the
pump are the
ability to operate in a high pressure inert environment, long life, no
maintenance and quiet.
Solberg Mfg. produces a line of oil-mist eliminators, i.e. filters, that are
compact, effective
and can hold enough oil for several thousand hours of operation. In a
preferred
embodiment, the bidirectional regulator 10413 allows pressure flow when the
design
pressure difference has been exceeded in either direction. Pump pressure
switch 10417
operates the pump when the pressure difference between the airlock 10401 and
the
crankcase 10410 is preferably less than 10 PSI for example. Pump pressure
switch 10417
includes a predetermined dead band, or range, that keeps the pump 10412 on
until the
airlock pressure is for example 14 PSI above the crankcase pressure.
Controller pressure
switch 10419 signals to the controller C that the airlock pressure is at least
5 PSI, for
example, above the crankcase pressure. This insures that the engine will not
turn until the
airlock pressure is sufficiently greater than the crankcase pressure. The
rolling diaphragms
10490 could tear if moved without such pressure difference across them. A fill
source
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10414 may be connected with the airlock to replenish the pressurized vessel
charging and
working gas/fluid if necessary.
FIG. 45 is a specific embodiment of the bidirectional regulator 10413 showing
the
pump 10412, oil filter 10418 and a spool valve 10441 which operates between an
airlock
port 10449, a crankcase port 10451 and a pump port 10453 according to the
pressure
differentials between the crankcase pressure and airlock pressure. In this
case, alternative to
the pressure switches 10417, 10419 described above a proximity sensor 10425
for
determining location of the spool 10441 via a target magnet 10426 is used to
control the
pump 10412 and if necessary to signal the engine controller C. The spool valve
10441 is
biased by a primary spring 10443 against the airlock over-pressure and an
underpressure
relief valve 10445 is biased by an inner spool spring 10447. Observing FIGS.
46A-A6E the
spool is shown in certain positions: in (1) is shown the spool influenced open
by the spring
where the airlock pressure is low so that the airlock port 10449 now
communicates to pump
port 10453 to receive pressurized gas from the pump 10412, in (2) the spool
10441 is shown
where the airlock pressure is within normal limits so the airlock port 10449
is closed by
spool 10441 and the spool is still displaced enough according to the proximity
sensor 10425
to cause operation of the pump 10412, even without flow from the pump to the
airlock. In
(3) the spool 10441 is shown where the airlock pressure is again within normal
limits so the
airlock port 10449 is closed by spool 10441 and the spool is now displaced so
that the
proximity sensor 10425 does not turn on the pump. Either one or two proximity
sensors
10425 are shown in FIGS. 45-D, however any desired number and type of
proximity
sensors may be used in normal operation in other embodiments. In (4) the spool
10441 is
shown with the airlock pressure is high so that the airlock port 10449 is
connected to
crankcase port 1045 land pump is disabled while airlock pressure is reduced.
(5) is a case
where the engine is shut down so there is no power to the pump and the airlock
pressure is
extremely low and to keep the diaphragms from being damaged, the airlock port
10449 is
connected directly to the crankcase through the underpres sure relief valve
10445 which
opens to provide direct pressure relief through the spool 10441 so that the
crankcase
pressure and airlock pressure are at least equalized.
In another embodiment of the pressure regulator 10401 shown in FIG. 47, the
bidirectional regulator 10413 is replaced with a back pressure regulator 10431
which
provides for one way pressure flow from the airlock 10401 into the crankcase
10410 should
the pressure differential exceed for instance 15 PSI. To accommodate flow in
the other
direction from the crankcase to the airlock, a check valve, or pair of check
valves 10433,
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10435 can be provided in a separate path. This ensures that the crankcase will
not be
pressurized higher than the airlock. In FIG. 48, the crankcase 10510 is
intended to be
maintained at atmospheric pressure. This is a critical improvement as it
provides for a more
easily serviceable lower unit on the vessel without the need to recharge the
crankcase 10510
5 should work need to be done inside the crankcase 10510 and also provides
that a
significantly lighter crankcase housing is necessary to contain the drive
components. In this
embodiment of the airlock pressure regulator system 10511 the airlock space
10501 is
maintained essentially at atmospheric plus 15 PSI and therefore any
pressurized working
gas which escapes from the working space 10503 into the airlock 10501 needs to
be
10 removed from the airlock 10501 and returned to the working space 10503.
To accomplish
this, in its simplest form a first relief valve 10520 means is provided in an
uptake line 10522
communicating with the airlock space 10501 so that any pressure greater than
15 PSI is
relieved from the airlock 10501 and passed via a pump 10512 into an
accumulator 10523
outside the working space 10503, airlock space 10501 and crankcase 10510. From
the
15 .. accumulator 10523 a return line 10525 includes a second relief valve
10521 means which
opens to permit recharging of the working space 10503 with pressurized gas
from the
accumulator 10523 should the pressurized working gas in the working space
10503 fall
below 1500 PSI. It is to be appreciated that the balancing of this pressurized
system may
include other pressure considerations across the first and second relief
valves 10520, 10521,
20 particularly with respect to the variation which can occur in the
working space 10503 where
the pressure can swing plus or minus 300 PSI during the Stirling cycle itself.
When the engine is running, a mechanical pump 10612 defined by a cavity 10608
in the piston rod may be utilized to reduce the load and work done by the
above described
airlock pressure regulator system. As seen in FIG. 49, the mechanical pump
10612 of the
25 piston rod 10624 is added to the airlock pressure regulation system to
reduce the load on the
electrical system during operation. A check valve 10605 receives crankcase
pressure
through an intermediate passage 10607 from the crankcase. The check valve
10605 opens
when the airlock pressure has dropped too low relative to the crankcase
pressure and
pressurized gas from the crank case is drawn into the piston cavity 10608 as
the piston rod
30 10624 reciprocates. The piston rod cavity 10608 is defined by a reduced
diameter portion
of the piston rod which essentially defines the mechanical pump 10612 itself.
As the piston
rod 10624 reciprocates the piston rod cavity 10608 is reduced in size as shown
by the right-
hand piston, pumping the pressurized gas into the airlock space 10609. In this
way during
engine operation the airlock can be efficiently replenished with sufficient
pressurized gas
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should its pressure drop too low. An outlet check valve 10611 is provided
between the
airlock and the crankcase so that pressure in the airlock which exceeds the
desired
differential can be reduced from the airlock space 10609 into the crankcase.
The mechanical
pump 10612 defined by movement of the piston rod 10624 does not operate at
engine
startup because there is no mechanical operation of the engine, however the
airlock pressure
regulator system must be operational during startup operations.
Cooler Liner Diameter Reduction
As explained previously with respect to FIGS. 1, 4-9C and 19, the heater tubes
communicate with a heat exchanger which circumferentially surrounds each
cylinder. The
heat exchanger of the present embodiment described in FIG. 50A provides
cooling for the
working gas/fluid during the appropriate portion of the stifling cycle. The
heat exchanger
10705 is supplied with cooling water through coolant tubing which communicates
with a
heat sink such as the environment via a radiator (not shown). Generally, the
coolant water
picks up heat through the heat exchanger in the vessel from the hot working
gas, and the
coolant water then is pumped to the radiator where the heat is transferred to
the
environment.
The heat exchanger 10705 shown in FIG. 50A surrounding each respective
cylinder
is provided with a water jacket sleeve 10704 having an inner surface defining
a channel to
allow passage of the cooling water through an interfacial area 10706 between
the inner
surface of the water jacket sleeve 10704 and an outer surface of a cooler
liner 10702. The
cooler liner 10702 also has an inner surface 10708 which directs the flow of
hot working
gas along the inner surface to facilitate the transfer of the heat through the
cooler liner
10702 to the coolant water in the interfacial area. A goal of the described
structure is to
increase the heat transfer surfaces within the interfacial area 10706 for
absorbing heat from
the hot working gas and the heat exchanger so as to improve heat transfer
between the
working gas and the coolant water.
The water jacket sleeve 10704 surrounds the cooler liner 10702 and forms the
heat
exchanger 10705 which cools the working fluid during the appropriate portion
of the
Stirling cycle. The cooler liner 10702 directs the flow of the working gas
along the inner
surface of the cooler liner 10702. An improvement of the presently described
engine is an
increased heat transfer surface area in the heat exchanger where the outer
diameter of the
cooler liner 10702 is reduced to increase the interfacial area 10706 with a
plurality of
extended surfaces, for instance, longitudinal arranged fins 10707, or pins,
provided around
the outside diameter of the cooler liner 10702 and extending into the
interfacial area 10706
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between the inner surface of the water jacket sleeve 10704 to increase the
surface area of
the heat exchanging surfaces and provide more efficient cooling of the working
gas/fluid.
The outer diameter of the cooler liner wall 10708 can be reduced to an extent
so that the
cooler liner wall 10708 is relatively thin, as compared to the radial length
of the longitudinal
.. fins 10707, or pins in the interfacial region 10706 between the cooler
liner 10702 and the
inner surface of the heat exchanger 10705. The inner wall 10708 of the cooler
liner is
generally maintained an appropriate diameter to accommodate the working gas
flow from
the heater head and cylinder. The inner diameter of the liner is provided with
axially
arranged fins 10707 to direct the flow of gas along the inner wall of the
liner and facilitate
.. the transfer of heat out of the working gas, through the cooler liner and
into the coolant
water.
It is also important to ensure that the stationary seals utilized in the heat
exchanger
are to the extent possible redundant and not compromised, particularly where
the water
jacket sleeve 10704 and cooler liner 10702 should sufficiently maintain the
coolant water in
.. the interfacial region 10706 between these elements and the working fluid
inside the coolant
line 10702r. As shown in FIG. 50B, the heat exchanger 10705 in the present
embodiment
has an outer surface which abuts the inner surface of the vessel and is sealed
with respect to
the vessel by an upper stationary seal 10710 and a lower stationary seal
10711. Similarly
the cooler liner 10702 inside the heat exchanger 10705 is sealed with respect
to the inner
surface of the heat exchanger by an upper seal 10713 and a lower seal 10715. A
top surface
of each of the cooler liner 10702 and the heat exchanger 10705 are formed and
both support
the base of the heater heads 10703 and provide the communicating interface for
the working
gas between the heater tubes 10709 and the heat exchanger 10705. An additional
or
redundant seal can be added at the intersection between the cooler liner 10702
and the heat
exchanger 10705 adjacent the top surface of each element which supports the
heater heads
10703. This redundant seal 10712, for example a 45 degree o-ring, is located
axially spaced
above the upper stationary cooler liner sealer 10710 and extends
circumferentially around
the entire joint between the heat exchanger 10705 and the cooler liner 10702.
The addition
of the heater head base as it is supported on the top surfaces of the liner
10702 and heat
exchanger 10705 compresses the redundant seal into the joint and adds
redundancy to the
system to prevent the escape of cooling water and/or working gas/fluid from
the working
space.
In a further embodiment, as seen in FIGS. 53C and 53D, the heat exchanger
10720
has a circumferentially disposed interfacial area 10722 comprising a plurality
of
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longitudinally extending tubes 10724 provided around the outside diameter of
the heat
exchanger body 10726. The tubes 10724 extend through the interfacial area
10722 to
increase the sutface area of the heat exchanging surfaces and provide more
efficient cooling
of the working gas/fluid. This is accomplished by routing the working fluid
through the
tubes 10724 that are surrounded by the cooling fluid in the interfacial area
10722 instead of
on the opposite side of the cooler liner as in the above described
embodiments. The tubes
10724 may be assembled onto the heat exchanger 10720 through a brazing process
wherein
the completed assembly is run through a brazing oven to solidify the
connection between
the tubes 10724 and the heat exchanger body 10726.
In the current embodiment, the tubes 10724 and the heat exchanger body 10726
are
constructed of the same material to simplify the assembly process. In one
example, the
tubes 10724 and heat exchanger body 10726 are fabricated from 300 series
stainless steel.
In another example, the tubes 10724 and exchanger body 10726 are fabricated
from an
aluminum alloy such as one of the following but not limited to AL7075-T6,
In an alternate embodiment, the tubes 10724 are constructed of a 300 series
stainless
steel while the body 10726 is constructed of a 400 series stainless steel,
such that the tubes
10724 have a lower coefficient of thermal expansion. The tubes' 10724 lower
coefficient of
thermal expansion may cause the tubes to not expand as much during the heating
period as
the body 10726. The braze will join the tubes 10724 to the body 10726 at the
highest cycle
temperature, when the tubes have not lengthened as much as the body.
Subsequently during
cooling the heater tubes will be compressed as the length of the body is
reduced by more
than the tubes. However, as the tubes are soft due to the high temperature,
some or all of
the tubes may buckle slightly due to the compression force. The buckling is
not enough to
weaken the tube or restrict flow through the tube. One benefit is that during
operation,
when the body 10726 and tubes 10724 are heated to a lower temperature, the
greater
thermal expansion of the body will not break tubing/body braze joint because
the buckled
tubes provide structural elasticity. In other words, during use, unbuckled
tubes 10724 are
colder, less elastic and thus can apply a greater and repeated load on the
braze joint which
can lead to failure. The buckled tube is less stiff, due to the new shape of
the tube and
applies a lower load on the braze joint during use, which may lead to less
failures and
longer periods between failures. and during use at lower temperatures, the now
stronger and
less elastic tubes .. In an embodiment, the heat exchanger assembly may go
through a
second heating phase and possibly a second brazing after the initial brazing
process. This
second heating phase or brazing will again join the tubes 10724 to the
body10726. when the
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tubes 10724 have not extended as much as the body 10726. The cooling process
from a
brazing temperature which is above the annealing temperature may allow
longitudinally
compressed tubes 10724 to slightly deform away from a completely vertical
structure, as in
a slight bend, such that the deformation eliminates any pre-compression in
each tube 10724.
In a further improvement to the drive system a more easily constructed and
easy to
maintain connection between the link rod 10826 and the rocking beam 10816 is
described.
Referring now to FIG. 51A, a rocking beam drive mechanism 10801 is shown. In
this
embodiment, the rocking beam drive mechanism has pistons 10802 and 10804
coupled to
two rocking beam drives 10801. In the exemplary embodiment shown more clearly
in FIG,
51B, the link rod 10826 is coupled at a first end to the piston rod via a link
rod upper pin
10832, and a second end of the link rod 10826 may be coupled to one end of a
link rod
lower pin10832 attached to the yolk of the rocking beam 10816. The link rod
lower
pin10832 had been previously accomplished by press fitting a pin 10823 into a
passage of
the link rod 10826, and with bearings provided on either side of the link rod
10826 and
around the pin 10823, the second end of the link rod is secured to the rocking
beam drive
10801 in a yolk 10825. The pin 10823 extends into respective pin passages in
the yolk
10825 of the rocking beam 10816 in order to complete the link rod lower
pin10832
structure. A bearing is also provided between the pin passages in the rocking
beam 10816
and the pin 10823 to facilitate the pivoting of the link rod and pin relative
to the rocking
beam 10816.
The present embodiment eliminates the need for a press fit of the pin into the
passage in the second end of the link rod 10825. The press fit made it
difficult to maintain,
fix assemble and disassemble in any manner this end pivot structure during
maintenance of
the engine. As seen in FIG. 51B the link rod lower pin10823 is provided to be
inserted
with a loose fit into and through the passage in the second end of the link
rod 10826. A
bearing 10822 may be provided around the link rod lower pin10823 on either
side of the
link rod 10826, and the width of the bearing 10822 is reduced in order to fit
a retaining ring
10828 onto the pin 10823 adjacent each bearing to retain the pin 10823 and
bearings axially
aligned in the passage of the link rod 10826 and in the yolk 10826. With the
pin 10823 and
bearings essentially axially fixed by the retaining rings 10828 to the link
rod 10826, the pin
10823 and link rod passage can have a loose fit so that the pin 10823 can be
easily removed
from the link rod 10826, when disassembly is necessary, merely by removing the
retaining
rings 10828 and sliding the pin 10823 out of the link rod passage. A
lubricating oil passage
10829 may be provided in the link rod lower pin 10823to communicate with a oil
passage
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10838 in the link rod 10826 and provide oil to the bearings 10822 and the
respective surface
of these pivoting components.
The link rod upper pin 10832 is similarly arranged with a loose fit with the
first or
upper end of the link rod 10826. A bearing 10834 in this case is provided
directly between
5 the bearing surface of the upper pin 10832 and an inner surface of the
upper link rod
passage. A pair of retaining rings 10836 are applied to grooves in the ends of
the upper pin
10832 to maintain the pin in its axial placement in the cross head 10840. The
bearing
10834 and respective bearing surfaces can be supplied with lubricating oil via
the oil
passage 10838 in the link rod 10826
10 The arrangement of the Stirling machine discussed above is generally
referred to and
shown as having a vertical orientation, i.e. with the pistons reciprocating
generally
perpendicularly aligned relative to a horizontal support surface or ground
surface. In another
embodiment of the present Stirling cycle engine 10903 shown in FIGS. 52A and
52B the
engine may be horizontally arranged, i.e. with the pistons 10905, piston rods
10907, heater
15 heads 10911, cross heads 10913 etc., being arranged and reciprocating in
a horizontal
orientation relative to a ground support surface as opposed to the vertical
orientation
discussed above. One of the significant challenges in such a design is the
arrangement and
structure of the oil cooling system in the crankcase 10915 where it imperative
to ensure that
the mechanical elements of such a horizontal crankcase such as the cross heads
10913,
20 rocking beam 10919 and other crankcase components and drive elements are
sufficiently
supplied with a free flow of oil through the crankcase and back to the oil
sump and pump.
As seen in FIG. 52B and by way of general example, the oil cooling system
comprises a central oil supply line 10921 disseminating a flow of oil directly
to each of the
cross head bores 10923 through radial oil passages 10925. Oil drains down by
gravity in
25 the crankcase 10915 into oil sump 10931 which can then be re-circulated
back to the
central oil supply line 10921 via a pump 10935 through main line 10937 which
communicates eventually with central oil supply line 10921. It is to be
appreciated that
other oil supply arrangements and orientations can also be accomplished, and
that the
embodiment described with respect to FIGS. 52A-B and the horizontal
arrangement of the
30 engine and crankcase components is merely exemplary with respect to
these figures.
In another embodiment of the engine it is also beneficial to cool the
crankcase by
cooling the oil in the crankcase. An oil cooler 10941 shown diagrammatically
in FIG. 52B
is designed to pick up a substantial amount of the heat generated in the
crankcase, and with
a co-axial (or a tube-in-tube) heat exchanger 11043 shown specifically in
FIGS. 53A-B, oil
91
from the crankcase passes through an outer oil channel 11045 over a series of
fins 10947
positioned along the outer surface of a cooling tube 11049 containing flowing
cooling water
from a cool water source 11046. The fins 11047 can be radial fins or axially
aligned fins
relative to the cooling tube 11049 depending upon the necessity for a desired
oil flow along
the outer surface of the coolant tube. After taking up heat from the oil, the
cooled oil returns
to the main line 11037 and the heated water can be dumped to a heat sink
11051. Methods
or apparatus are disclosed for external starters, both manual and powered, in
U.S. Patent
Application Serial No. 13/447,990, filed April 16, 2012 and entitled Stirling
Cycle Machine
(Attorney Docket No. 184).
B-Burner
FIGS. 54 ¨ 60 disclose a further embodiment of a burner 11201 for use in
conjunction with a multiple heater head and piston engine described previously
in FIGS. 4,
35, 36A. The present burner 11201 is specifically directed to the independent
heating of
multiple heater heads, in this case four (4) heater heads, each heated by an
individual burner
and flame and having a single air inlet 11223, single outer wall 11212, and
two exhaust
openings 11225.
Turning to FIG. 55 the four burner design 11301 of the present embodiment
includes
a single blower B providing air for the fuel/air mixture in the ignition
process of all the
burner head assemblies 11305 as shown in FIG. 55. The heater heads 11303 as
also
discussed above, may be any of the various embodiments described in the
preceding
sections, including, but not limited to, tube heater heads, or pin or fin
heater heads as
disclosed in U.S. Patent Application Serial No. 13/447,990, filed April 16,
2012 and entitled
Stirling Cycle Machine (Attorney Docket No. 184). By way of example, the
present
embodiment is contemplated utilizing heater tubes 11309 through which flow the
working
gas, for example helium, which must be heated by the burner head assemblies
11305 during
the appropriate portion of the Stirling cycle.
By way of more detailed description and referring back to FIG. 54 as well, the
burner 11301 includes multiple burner head assemblies 11305, one for each of
the heater
heads 11303, in the case of the present embodiment there are four (4) heater
heads and
hence four (4) burner head assemblies 11305. The cross-section of FIG. 55
shows three (3)
of the burner head assemblies 11305. Generally, the burner 11301 is defined by
a burner
housing as shown in FIG. 54 having a substantially cylindrical outer wall
11312, although
other geometrical configurations could be imagined. The blower B pumps air
into the
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burner 11301 through air intake 11223 for purposes of ignition and combustion,
and exhaust
gases are ejected from the burner via the two exhaust outlets 11225 adjacent
the base of the
burner.
Turning to FIGS. 56 and 57 a top surface 11413 of the burner housing includes
a
number of ports 11415 for receiving fuel inputs 11416. igniters 11427, flame
and possibly
temperature sensors or flame viewing elements. The ports 11415 also facilitate
access to a
particular burner head 11405, as discussed in detail below, without having to
remove the
entire burner 11401 from the vessel stack-up for maintenance. As seen in FIG.
57,
associated with each burner port 11515 on the top surface 11513 of the burner
is a
secondary port 11517 which can serve a number of purposes for instance a flame
viewing
element such as a viewing window for viewing the flame of the burner head, or
alternatively
a spark plug 11520 for igniting the fuel/air mixture and/or a sensor for
sensing UV light
used in flame detection.
The base of the burner 11601 best seen in FIGS. 58 and 59, is provided with
heater
head openings 11619 to accommodate the entrance of the heater heads and
respective heater
tubes since the burner as a whole is set over and stacked up on a cooling
plate 11604 of the
vessel so that the heater heads 11603 are within and substantially sealed
inside or
encompassed by a lower region of the burner 11601. The base of the burner
11701 is
secured to the mounting plate 11704 in the vessel stack-up by a
circumferential band clamp
11710, such as a Marmon clamp, which is provided for securing and critically
circumferentially centering the burner relative to the cooling plate and lower
stack-up of the
pressure vessel. The centering of each burner head assemblies 11705 relative
to each of the
associated heater heads 11703 is critical because if the flame from the burner
head assembly
is nearer one side of the heater heads 11703 and heater tubes 11709 than
another, there will
be not only inefficient heating of the working gas/fluid in the heater tubes
11709, where one
set of heater tubes is heated to a higher temperature than other tubes.
The clamp 11710 extends circumferentially and radially around the entire base
of
the burner 11701 and provides both radial and axial compressive forces between
the burner
base plate and the mounting plate to ensure that there is both a critical
axial sealing pressure
to contain the hot exhaust gases in the burner and a radial circumferential
alignment of the
burner heads with the heater heads. The base of the burner housing 11711 may
be provided
in this regard with a circular sealing edge 11721 as shown in FIG. 59 which is
angled
relative to the vertical axial arrangement of the vessel stack-up to create
the axial
compressive force and for mateably engaging with an oppositely angled circular
sealing
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edge 11722 of the mounting plate. The circumferential clamp 11710 and the
mating angled
circular sealing edges 11721,11722 of both the burner and the cooling plate
ensure the
critical circumferential, i.e. radial alignment of the burner housing 11711
and burner head
assemblies 11705 with the mounting plate and heater heads 11703 in the vessel
stack-up so
that the burner head assemblies 11705 are appropriately aligned with the
heater heads 11703
and there is sufficient axial force between the burner 11701 and cooling plate
11704 to
contain the hot exhaust gas generated in the burner 11701. The circular
sealing edge 11721
may include a graphite seal (not shown) between the burner and cooling plate
to ensure that
the hot gases which are at around 1000 C, where the flame temperature is
around 1200 C,
do not leak out between the burner 11701 and the mounting plate 11704.
The single blower B as shown in FIG. 59 provides air into the burner housing
11711
adjacent the top surface 11713 for the fuel/air mixture to the burner head
assemblies. The
air intake 11723 is provided at essentially a normal angle to the circular
burner housing
11711 and provides air inside the burner 11701 for combustion as described in
detail below.
This arrangement of the air intake 11723 at a normal angle to the cylindrical
burner housing
11711, better seen in FIG. 58, facilitates air entering the burner 11701 with
a designed
pressure drop which is important for incoming air to the burner to maintain a
desired air
velocity for maximizing heat transfer efficiencies as the air passes through
the air intake
manifoldand into a preheater 11726 where the incoming (cold) air is warmed by
the exiting
(hot) exhaust gasses. A single blower B is placed in communication with the
single air
intake 11723 to provide air to all the burner head assemblies 11705 in the
burner 11701. A
pair of exhaust outlets 11725 are also connected normal to the substantially
cylindrical
burner 11701 and spaced approximately 180 degrees apart around the base of the
burner
11701. Prior to exiting through the exhaust outlets 11725 the exiting exhaust
from the
burner 11701 preheats the incoming air in the preheater 11726 described in
detail below,
then exits the burner 11701 from one of the two exhaust outlets 11725.
Observing FIG. 59, each burner head assembly 11705 has a fuel injector 11724,
an
igniter 11727 of one kind or another, for instance a sparkplug or glow-plug, a
flame
detection device 11729 which may also be provided in the secondary port 11717
as shown.
Fuel, either liquid fuel or gaseous fuel is fed to the fuel injector 11724 via
a fuel line 11731
from a fuel source F and is dispersed as a fine mist or vapor by the nozzle
11734 of the fuel
injector 11724 into a prechamber 11728 of the burner head assembly 11705. In
the
prechamber 11728 the dispersed fuel is combined with a desired volumetric flow
of air from
the preheater 11726, preferably preheated to a desired ignition temperature by
the exhaust
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as discussed in detail below, to form a desirable fuel/air mixture for
ignition. The fuel/air
mixture then is ignited by the igniter 11727 and combusts at least partly
inside the
prechamber 11728, but more complete combustion may occur after the fuel/air
mixture
exits, or is pushed, from the prechamber 11728 through the prechamber nozzle
11730 of the
.. prechamber 11728 to form a flame which extends from the prechamber 11728
and is
directed into a center combustion chamber inside the heater tube arrangement
of each
respective heater head 11703. Exhaust from the combustion in the burner 11701
exits the
burner via the preheater 11726 and exhaust manifold 11714 described in detail
below.
In the present embodiment of the burner, the single blower B, shown here
diagrammatically, may be incorporated to maintain a consistent average air
ratio supplied to
the burner 11701 and hence to each of the individual burner head assemblies
11705. The
blower B pumps air at a desired velocity depending on instructions from a
controller C for
purposes of ignition, then once ignition has occurred, the desired air flow
rate may be
regulated by the controller C dependent on data received from sensors
including but not
limited to an oxygen sensor. A more complete description of the burner control
algorithm is
provided below. The single blower B is also controlled dependent on the data
from
individual burner head assemblies, for example in the case of at least one
burner head
assembly being extinguished or not igniting the controller may decrease the
blower rate to
facilitate ignition in the extinguished burner head assembly. The fuel input
may be
correspondingly controlled in the remaining burner head assemblys 11705 to
accommodate
such an air velocity decrease. In any event, the blower B is intended to
provide a consistent
flow rate to each of the multiple burner head assemblys 11705 in the burner
after passing
through the preheater 11726. An important aspect of the present embodiment is
the
consistent flow and velocity of cold air developed by the blower B and the
efficient heating
.. of the incoming air through the extraction of waste heat from in the
preheater 11726, to
raise the cold air temperature thereby improving the efficiency of combustion
processes and
the burner unit.
The blower B connects through the air intake 11723 in the outer wall 11712 of
the
burner housing into a cold air channel 11735 of the preheater 11726. The cold
air channel
.. 11735 extends circumferentially around the burner inside the outer wall
11712 of the burner
11701 and directs the cold air developed by the blower B down around an
insulated
intermediate baffle 11739 and up into the preheater. The intermediate baffle
11739 is
insulated to protect the outer wall 11712 of the burner 11701, and anyone or
thing that
comes in contact with the outer wall, from the intense high temperatures
inside the burner
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11701. Also, the insulated baffle 11739 ensures that heat captured by the
incoming air in the
preheater 11726 is not lost directly to the outer wall 11712 of the housing
11711.
The preheater 11726 essentially begins where the cold air from the blower B
drops
down through the cold air channel 11735 and enters into a preheater channel
11741 in
5 which the cold air is preheated in order to raise its mean temperature
which increases the
efficiency of the burner combustion. The preheater channel 11741 is defined by
the
intermediate baffle 11739 on one side, and on the inner side, an exhaust
manifold wall
11743. The exhaust manifold wall 11743 directly separates the incoming cold
air from the
exhaust air exiting the burner and provides for the heat transfer from the
exiting exhaust to
10 the incoming cold air in the preheater channel 11741. The heat transfer
efficiency through
the manifold wall 11743 in the preheater 11726 is critical because the hotter
the incoming
cold air can be raised by the preheater 11726, the less fuel is necessary to
get the gas up to
desired ignition and combustion temperatures. The preheater channel 11741 also
extends
circumferentially around the entire burner 11701 which provides for a maximum
surface are
15 which in some embodiments may produce better heat exchange with the
exhaust flowing
out of the burner through an exhaust channel 11744. Inside the preheater
channel 11741 are
a series of radially extending fins 11745 which are directly connected to the
exhaust
manifold wall 11743 and assist in efficient heat transfer from the exiting
exhaust air through
the manifold wall 11743 into the air in the preheater channel 11741. Exhaust
side fins
20 11746 may also be connected to the exhaust manifold wall 11743 extending
into the exhaust
channel 11744.
The cold incoming air is preheated to a desired temperature, for example, but
not
limited to 600- 750 C, in the preheater 11726 which facilitates ignition and
combustion as
the air is directed to the burner head assemblys. The amount of preheating
which may be
25 accomplished is primarily based on the efficiency of heat transfer from
the exiting exhaust
so that as the exhaust temperature is raised during operation of the engine,
the incoming
cold air can be accordingly preheated to a higher temperature. The preheated
air exits the
preheater channel 11741 and is directed radially into a hot air chamber 11747
which
communicates with each of the multiple burner head assemblies 11705. It is to
be
30 appreciated that the preheated air enters the hot air chamber 11747
through a substantially
360 degree circumferential opening around the exit of the preheater channel
11741 so that a
consistent flow rate of preheated air is delivered to each of the burner head
assemblies
11705. While additional channels or passageways (not shown) may be provided in
the hot
air chamber 11747 to direct the preheated air in the hot air chamber to a
specific burner
96
head, the 360 degree output from the preheater channel of the present
embodiment is
important since there is only one blower B developing the air flow into the
engine. In
previous engines a multitude of blowers delivered a desired air flow to each
of the burner
head assemblies, for instance where there were four (4) burner head
assemblies, there were
four (4) blowers, one directed to each burner head. However, having a blower
associated
with each burner head 11705 on a multiple burner head engine is expensive and
adds a
significant amount of weight to the engine. In any event, a single preheater
is much less
expensive and less complicated from a control standpoint than separate
preheaters fo each
heater head.
The preheated air is directed in the hot air chamber 11747 to the individual
burner
head assemblies 11705 and specifically to an intersection with a nozzle 11734
of each fuel
injector 11724 in each burner head 11705 and an igniter 11727. The fuel
injectors 11724
may use either liquid fuel or gaseous fuel but in either case the fuel is
ejected from the
injector into the prechamber 11728 where the fuel mixes with the preheated air
to attain a
desired fuel/air ratio or mixture for either ignition of the burner head
11705, or, combustion
where the burner head 11705 is currently supporting a flame. The fuel
injectors 00024 inject
the fuel into the prechamber 11728 directly below the fuel injector 11724 and
the preheated
air is combined in the prechamber 11728 with the liquid or gaseous fuel. The
fuel may be
delivered as a mist or vapor, combined with the preheated air and ignited in
the prechamber
11728 by the igniter 11727. While ignition of the fuel/air mixture may occur
to some extent
in the prechamber 11728, the flame derived from the ignition and combustion of
fuel/air in
the prechamber 11728 needs to be pushed out of the prechamber 11728 to be more
efficient
and provide the requisite thermal output. It is preferable that the constant
combustion flame
which heats the heater heads 11703 and heater tubes 11709 be pushed out of the
prechamber
.. 11728 and actually extend beyond the end cone 11730 of the prechamber 11728
and into the
combustion chamber 11750. This is accomplished by providing appropriate
adjustment to
the fuel/air mixture by the controller and by the prechamber and nozzle
geometry to
properly control the shape of the flame. Structural elements may be added to
the
prechamber to improve the shape of the flame in the combustion chamber as
disclosed in
U.S. Patent Application Serial No. 13/447,990, filed April 16, 2012 and
entitled Stirling
Cycle Machine (Attorney Docket No. 184).
Another aspect of the present embodiment shown in FIG. 60 is the prechamber
support which extends around the outermost wall of the prechamber 11928
adjacent the exit
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cone 11930and includes a heater head restrictor 11937 restricting the flame
exhaust gases
from passing over the heater head 11903. The restriction increases the
pressure drop at the
top of the heater head 11903 and improves heat transfer from the hot gas to
the helium
inside the heater tubes 11909 by encouraging the hot gas to flow through the
heater tubes.
A substantial amount of the heat not used to heat the working fluid remains in
the
exhaust gases and thus the efficiency of the entire engine can be increased by
using the
excess exhaust gas heat to preheat the incoming air. After heating the heater
head 11903
and heater tubes 11909 the hot combustion gases are forced out an exhaust
inlet 11953 by
newer combusted gases into the exhaust channel 11944 defined partly by an
inner wall
11942 of the exhaust manifold. The exhaust passes down along the exhaust
channel 11944
exchanging a substantial amount of heat through the preheater wall 11943to the
incoming
cold air entering the preheater 11926 via the preheater channel 11941. The
exhaust gases
should flow as quickly as possible through the preheater 11926 as the heat
transfer from the
exhaust gases is dependent upon the velocity of the exhaust. Another aspect of
the present
embodiment limits the pressure drop of the exhaust gases by allowing the
exhaust gases
flow out from two exhaust outlets 11925, as opposed to one exhaust outlet,
arranged around
the bottom of the exhaust manifold 11914. The shorter flow path provided by
the two
exhaust outlets 11925 for the exhaust leaving the exhaust manifold 11914
lowers the
pressure drop for the exhaust, the blower does not have to work as hard, and
thus the blower
load is reduced on the engine.
FIG. 60 also discloses the use of spark plugs 11960 in the secondary port
11917
(rather than a UV viewing window) through the burner housing. In certain cases
a flame
sensor may also be inserted through the secondary port 11917 which extends
into the burner
adjacent each of the prechamber nozzle so that flame detection can occur. In
any event, the
secondary port 11917 provides for access to the burner head 11905 so that a
sensor or
window, or ignition components for instance glo-plugs or spark plugs as shown
here can be
inserted down into the burner head 11905 to ignite different gases and fuels.
Gaseous fuel
use may necessitate the spark plug 11960 to ignite the fuel air mixture
adjacent the nozzle
of the prechamber whereas liquid fuel uses glow plugs and are generally
located closer to
the fuel injector itself. In the embodiment shown here a high voltage
conductive element
11961 is encased within and insulative layer 11963 and a ground layer 11965
and inserted
through the secondary port so that the exposed conductive element 11961 is
exposed in the
combustion chamber to ignite the fuel/air mixture exiting the prechamber
11928.
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The ability to see and/or detect each flame is important so that each of the
four
individual burner head assemblies 11905 and respective flame can be
appropriately adjusted
by the controller. It is to be appreciated that such flame detection and
viewing may be
accomplished by many embodiments, including but not limited to an actual
viewing
window for example having appropriate lenses in the tube which allow a human
operator to
look through the tube and visually identify a flame within the range of
visible wavelengths
in the combustion chamber. Alternatively, the viewing window may include a
camera or
other image data receiving and recording device such as a UV light sensor and
display for
visually displaying a received representation of the flame in the combustion
chamber.
Other types of heat sensors including but not limited to thermocouples,
infrared
thermometers, and thermisters, may be used to identify and quantify the flame
and flame
characteristics in the combustion chamber.
With only a single blower providing air to four burner head assemblies 11905,
generally a variable in addition to air, such as fuel, must be altered to
obtain a desired flame
quality. Keeping one blower providing air to all four burner heads is
especially helpful for
cost and for blower power consumption.
With liquid, diesel or other gaseous fuel, the UV viewing window will be
compromised because the fuel vapor tends to absorb the UV radiation from the
flame.
Without the UV window as in the previous embodiment it may still be important
to detect
the flame and the temperatures in the combustion chamber. The electrode of the
sparkplug
may be utilized as a sensor in some cases to detect the flame. Such data can
be forwarded
to the controller to determine the flame and combustion disposition in the
combustion
chamber 11950. Another method of flame detection obtains temperatures with
temperature
sensors inside the heater head, for example a thermocouple attached on the
walls of the
heater tubes can provide data to the controller to alter the operational
conditions of the
engine. This temperature data is used to judge the temperature and/or flame
quality based on
temperature/flame data and helps the controller decide what operational mode,
as discussed
in further detail below, to set for each burner head 11905 and for the engine
as a whole.
Burner Control
The burner may be operated in several modes as shown in FIG. 61, and referring
in
part to FIG. 60, according to predetermined electronics and software programs
embodied in
an electronic controller. The operation modes evaluated by the controller
include at least a
start-up mode 12002, normal operation mode 12004, shut-down mode 12006 and a
stop
mode 12008. The start-up mode includes the initial ignition of a richer fuel
mixture in the
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prechamber 11928 to ease ignition as colder mixtures have a narrower ignition
range of
fuel/air ratio that are ignitable compared to the range of fuel/air ratios
that maintain
combustion. With a desired fuel/air ignition mixture present in the prechamber
11928, the
igniter 11927 is actuated and the ignition mixture is ignited. A thermocouple
(not shown) in
the prechamber 11928 detects what is referred to as a diffusion flame in the
prechamber
11928 and once the incoming air is hot enough from the preheater 11926, the
flame is
pushed out of the prechamber 11928 by either increasing the air flow from the
blower B, or
increasing fuel so the flame travels out of the prechamber 11928 and forms in
the
combustion area adjacent the heater head 11903.
Generally in the start-up mode 12002 as shown in FIG.61 a user sets a desired
blower speed 12003 and fuel/air ratio 12005 for a certain time period 12007,
for example 30
seconds. After the predetermined time period the blower shuts off and resets
12011 the
start-up phase which may include blowing out 12013 any remaining fuel in the
engine and
exhaust system so that there are no backfires or other damaging events from
residual fuel.
The start-up phase may also include for instance a number of ignition attempts
12009 before
resetting and providing the user wih an error A sensor (not shown) within the
prechamber
11928 or a visual sensor using the secondary port 11917 detects if a flame
12010 is present
within the prechamber 11928 or the combustion chamber 11950. If a flame is not
detected
the system is reset 12011 or if a flame is detected the temperature readings
are taken 12015
from the heater head and oxygen levels are measured 12017 from the exhaust
gases. The
fuel/air ratio is then adjusted 12019 based on these readings.
Once the flame is supportable out of the prechamber 11928 and is heating the
heater
head 11903, the control system and operation mode 12004 include a number of
failsafe
triggers 12023 based on sensor data and controller evaluation algorithms which
evaluate the
system and determine if the system should be turned to the shut-down or stop
mode. The
operation mode 12004 monitors levels of heat, power and oxygen for example and
perform
shut-down or stopping of the engine, or other modifications to the system and
engine if a
temperature reading is too high, or exhaust oxygen level is too high or if
engine speed
exceeds a desired value, or the differential pressure within the air lock is
too low. These are
just exemplary triggers for starting shut-down or stop procedures, other
triggers could be
used as well or in combination with these examples.
During normal engine operation, the blower is operated at least partially by a
control
loop which measures the excess oxygen 12017 in the exhaust to determine blower
speed.
The failsafe triggers 12023 shown in the flowchart and operation analysis
table 12021 in
100
FIG. 61 include: Engine speed exceeds predetermined range; Oxygen levels in
exhaust
exceed a predetermined range; Generator temperature exceeds a predetermined
range;
Burner temperature exceeds a predetermined range; Cooler temperature exceeds a
predetermined range; Flame/Ignition failure; repeatable Failure of flame
ignition. It is to be
appreciated that the described control method is not limited to the disclosed
triggers 12023
and that other triggers, factors and variables may also be analyzed by the
controller under
the start-up and operation modes 12002 and 12004.
A failure of the engine in one of these failsafe triggers 12023 directs the
controller C
to adjust the fuel/air ratio 12019 and continue acquisition of sensor
readings. A preset
number of a repeated failure 12025 of the engine to run within a predetermined
range for
any of these triggers leads to a shutdown sequence with an immediate fuel turn
off 12029.
The engine however can continue to run in the shut-down mode 12006 in many
cases. On
the other hand, certain events may cause complete engine stoppage (i.e. shut-
off as opposed
to shut down) so that damage to the engine is minimized. A status check 12037
on system
components is repeatedly run. These shut-off triggers 12034 are for example,
low oil
pressure, low airlock pressure differential, and low engine power levels will
ensure
complete engine stoppage to prevent damage.
During a shut-down mode 12006, the fuel and burner is turned off but the
engine keeps
running until the heater head 11903 is cooled to a desired temperature. A
system shut-down
may also be caused by excessive heat measurements in a number of components
such as the
Generator, the burner, or a cooler, or a system shut down may occur if there
is a failure to
ignite. A shut down due to system failure may trigger a safe mode where fuel
is pumped out
of the system. Any fault or system failure or trigger, will kill the fuel
delivery immediately
00036, but the engine will continue to run to cool down the system. The engine
runs until it
reaches a predetermined power level 12035 in the shut down mode 12006, or in
the event of
the more dangerous fail safe triggers the engine is stopped 12008, i.e. the
RPMs are set to 0.
The shut-down mode helps engine efficiency since the engine, burner and heater
heads
remain hot for a while, even while there is no fuel supplied, the engine will
continue to run
producing power until the predetermined low power level is reached. This
recovers some of
the energy put in at start-up mode which improves efficiency. Methods to
concentrate
exhaust gases near heater tubes is disclosed in U.S. Patent Application Serial
No.
13/447,990, filed April 16, 2012 and entitled Stirling Cycle Machine (Attorney
Docket No.
184).
Venturi Burner
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FIGS. 62A - 62D disclose a further embodiment of a burner 12901 for use in
conjunction
with a multiple heater head and piston engine as described previously in FIGS.
54 ¨ 60.
This embodiment of the burner 12901 is also specifically directed to the
independent
heating of multiple heater heads 12903. In this embodiment there are four (4)
heater heads
12903 and respective burner assemblies 12907, although only two (2) are
visible in the
cross-section of FIG. 62A, and more or less heater heads are of course also
possible for the
engine. As in the previously discussed embodiments the heater heads 12903 and
burner
assemblies 12907 are encompassed by a burner housing 12911 and each heater
head 12903
is heated by an individual burner assembly 12907 and flame supplied with a
fuel/air mixture
for combustion via a blower B and air inlet 12923 and a fuel injector 12927.
As described in
further detail below the exhausting combustion gases are used to pre-heat the
incoming air
and, following combustion, one or more exhaust outlets are provided through
exhaust
opening(s) 12925 (not shown) to finally exhaust the combustion gases from the
burner
housing 12911.
More specifically observing FIG. 62A, the four burner design 12901 of the
present
embodiment includes the single blower B connected to the housing 12911
providing air for
the fuel/air mixture in the ignition process of each the burner assemblies
12907 as shown.
The heater heads 12903 themselves may be any of the various embodiments of
tube heater
heads described in the preceding sections, including, but not limited to,
straight tube heater
heads or helical tube heater heads as disclosed in U.S. Patent Application
Ser. No.
13/447990, filed April 16, 2012. By way of example, the present embodiment is
contemplated utilizing healer tubes 12909 through which flow the working gas,
for example
helium, which is heated by the respective burner assemblies 12907.
By way of more detailed description, the burner 12901 includes multiple burner
assemblies 12907, one for each of the heater heads 12903. In the case of the
present
embodiment there are four (4) heater heads 12903 and hence four (4) burner
assemblies
12907. FIG. 62C discloses the burner housing 12911 having a substantially
cylindrical
outer wall 12912, although other geometrical configurations could be
accomplished. The
blower B pumps air into the burner 12901 through air intake 12923 for purposes
of ignition
and combustion, and exhaust gases are ejected from the burner via the exhaust
outlet 12925
adjacent the base of the burner.
As shown in FIG. 62C, the top of the burner housing provides one or more ports
12915 in the top of the burner 1212A for receiving fuel injector assemblies
12916. The
injector assembly 12916 may provide a port for fuel 12916A, a port to mount an
ignitor
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12916B and a sensor port 12916C to monitor the hot combustion air. The ignitor
12918
may be a spark plug and may function as the high voltage electrode of a flame
ionization
flame detector. In other embodiments, the ignitor 12918 may be a hot surface
ignitor. In
other embodiments the port 12916B may be used for a vision based flame
detection circuit
including but not limited to one or more of the following: IR flame detector,
visible light
flame detector and/ or UV flame detector. Injector assembly 12916 may be
removably
connected to burner 12901. This arrangement may be beneficial for many
reasons,
including, but not limited to, to allow the injector assembly 12916 to be
changed, cleaned
and / or modified. Fuels with different energy densities may require a
different sized port
12945. Alternatively, the fuel ports 12945 may be made very small or fitted
with nozzles
for liquid fuels. In some embodiments, liquid fuels may be atomized by forcing
the liquid
through very small orifices. In some embodiments, the liquid fuel may be
atomized by a
number of nozzle geometries including, but not limited to, one or more of
pressure nozzles,
and/or air blast nozzles. In some embodiments, the injector assembly may be
sealed to the
burner housing 12911 with exhaust gaskets 12922. In some embodiments, the fuel
injector
assembly 12916 may be brazed or welded to the burner top 12912A.
In some embodiments, the burner as a whole is set over and stacked up on a
cooling
plate 12904 of the vessel so that the heater heads 12903 are within and/ or
are substantially
sealed inside and/or are encompassed by a lower region of the burner 12901. In
some
embodiments, the burner base 12902 may be secured to the cooling plate mount
12904A in
the vessel stack-up by a circumferential band clamp 12910, such as, in some
embodiments,
a Marmon clamp, which is provided for securing and critically,
circumferentially centering
the burner relative to the cooling plate 12904. However, in various
embodiments, the
burner base 12902 may be secured using another securing apparatus/device. In
some
embodiments, the burner base 12902 may include one or more pins (not shown)
that mate to
holes in the cooling plate mount 12904A to orient the burner such that each
burner assembly
12907 may be substantially centered over each heater head 12903. As previously
discussed,
the centering of each burner assembly 12907 relative to each of the associated
heater heads
12903 is critical for many reasons, including, but not limited to, if the
flame from the burner
assembly is nearer one side of the heater heads 12903 and heater tubes 12909
than another,
there may be inefficient heating of the working gas/fluid in the heater tubes
12909, and a
potential for certain heater tubes to be heated to a higher temperature than
other tubes. As
discussed in further detail below, the efficiency of the engine may be
improved when even
103
and consistent heating of the heater tubes and the working gas/fluid is
accomplished in the
engine.
Referring again to FIG. 62A, in various embodiments, each burner assembly
12907
has a fuel injector assembly 12916, an igniter, which, in some embodiments,
may be a
sparkplug 12918 or glow-plug or another igniter, and a flame detection device
which may
also be provided in a secondary port. Fuel, which may be, in various
embodiments, either
liquid fuel or gaseous fuel, is fed to the fuel port 1216A via a fuel line
from a fuel source F
and is dispersed as a fine mist / mist or vapor through the multiple fuel
ports into an ejector
12941 of the burner assembly 12907. The ejector 12941 in this embodiment is a
venturi
type ejector as disclosed for example in U.S. Patent Application Serial No.
12/829,320 filed
July 1,2010, now U.S. Publication No. US-2011-0011078-A1 published January 20,
2011
and entitled Stirling Cycle Machine (Attorney Docket No. 178). In some
embodiments the
venturi 12941 may be beneficial for many reasons, including, but not limited
to, providing
the benefit of reducing or eliminating the need for a completely separate fuel
control
.. scheme as regulation of the airflow changes the vacuum which, in turn,
correspondingly
affects the fuel flow and regulates the burner power. In some embodiments, the
venturi
allows use of typical gas pressures in buildings, e.g., 7 inches of water
column, without a
compressor. The blower B forces air into an initial swirler portion 12943 of
the venturi
12941. The flow of swirled air through the venturi draws in a proportional
amount of fuel
through the fuel inlet ports 12945 on the fuel injector assembly 12916 which,
in various
embodiments, may be positioned at, or in, the venturi throat 12947. This
swirling fuel/air
mixture exits the venturi and forms a swirl stabilized flame in the combustion
chamber
12931.
The fuel/air mixture is ignited by the igniter or spark plug 12918 and may
combust
.. partly inside the diverging section of the venturi 12949. More complete
combustion may
occur in the combustion chamber 12931 extending down inside the heater tube
12909
arrangement of each respective heater head 12903. The hot combustion gas then
passes
between the heater tubes 12909 and collects on the outside of the heater heads
12903. In
various embodiments, exhaust shields 12908 direct the hot combustion gases to
flow
through the outer row of heater tubes 12909. The now cooled exhaust exits the
burner via
the preheater and exhaust manifold described in greater detail below.
In various embodiments of the burner 12901, the single blower B, shown here
diagrammatically, may be incorporated to maintain a consistent average air
flow supplied to
the burner 12901 and hence to each of the individual burner assemblies 12907.
In some
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embodiments, the blower B may also produce a variable air flow when necessary
to control
the fuel/air mixture in the venturi 12941. The blower B may pump air at a
desired velocity
depending on instructions from a controller for purposes of ignition, then
once ignition has
occurred, the desired air flow rate may be regulated by the controller
dependent on data
received from sensors, which may include, but not limited to, an oxygen sensor
sampling
the cooled exhaust gas. In some embodiments, a fuel system with variable flow
valves for
each heater head, which, in some embodiments, may be a MaxitrolTM EXA-40
valves for
example, as manufactured by the Maxitrol Company, Southfield, Michigan, USA,
may
control the fuel flow to achieve the commanded temperature on each heater
head. The
blower may be adjusted to achieve the desired fuel/ air ratio.
An important aspect of the present embodiment is the efficient heating of the
incoming air through the extraction of waste heat in the exhaust to raise the
incoming air
temperature thereby improving the efficiency of combustion processes and the
burner
12901. The blower B connects through the air inlet 12923 in the outer wall
12912 of the
burner housing into an air channel 12951. The air channel 12951 extends
circumferentially
around the burner inside the outer wall 12912 of the burner 12901 and directs
the air
developed by the blower B across an intermediate baffle 12953 and up into the
hot manifold
12957 before entering the swirlers 12943 of the venttiri ejectors 12941. The
intermediate
baffle 12953 directly separates the incoming air from the exhaust gases
exiting the burner
through an exhaust channel 12955 and provides for the heat transfer from the
exiting
exhaust to the incoming air in the air channel 12951. The heat transfer
efficiency across the
intermediate baffle 12953 is critical because the hotter the incoming air can
be heated, the
less fuel is necessary to reach the desired combustion temperatures.
The incoming air is preheated to a desired temperature, for example, but not
limited
to 600-750 C. These are many reasons preheating the incoming air may be
beneficial and
these include, but are not limited to, facilitating ignition and combustion as
the air is
directed to the burner assemblies 12907 and/or increasing the thermal
efficiency of the
burner by capturing some of the thermal energy in the combustion gases exiting
the heater
heads. In some embodiments, preheating of the air may reduce the hot exhaust
temperature
from 900 C to 300 C. In some embodiments, the amount of preheating which may
be
accomplished may be related to the efficiency of heat transfer from the
exiting exhaust to
the incoming air. The heat transfer across the intermediate baffle 12953 may
be improved
by adding rows of folded fins 12952 on the air side and folded fins 12952a on
the exhaust
side of the intermediate baffle 12953. The folded fins may be brazed to the
intermediate
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baffle 12953 to assure good thermal attachment. In various embodiments, the
material
properties of the folded fins may be optimized for the operating temperature.
For example,
in some embodiments, the rows of folded fins near the top may be heat
resistant metals,
which may include, but is not limited to, INCONELThi 625, while lower and
cooler folded
fins may have higher thermal conductivity but lower operating temperature. In
various
embodiments, the materials for these folded fins may include, but is not
limited to, stainless
steel 409 or Ni 201 for example. The preheated air exits the air channel 12951
and is
directed radially into a hot air manifold 12957 which communicates with each
of the
multiple burner assemblies 12907 specifically directing the preheated air to
the swirler
.. portion 12943 of the venturi 12941. In various embodiments, the preheated
air enters the
hot air chamber 12957 through a substantially 360 degree circumferential
opening around
the exit of the air channel 12951. In some embodiments, this may result in a
consistent flow
rate of preheated air delivered to each of the burner assemblies 12907. In
various
embodiments, additional channels or passageway's (not shown) may be provided
in the hot
air chamber 12957 to direct the preheated air in the hot air chamber to a
specific burner
head. In various embodiments, the 360 degree output from the air channel 12951
is used
when there is only one blower B developing the air flow into the engine.
Airlock and Working Fluid Repressurization System
As described previously in this application, the power, life and value of a
Stirling
engine may be maximized, in some embodiments, by building an oil lubricated
drive
contained in a pressure vessel, generally referred to herein as the crankcase,
and sealing the
working space of the Stirling engine which contains the working fluid, for
example helium,
from the crankcase oil with flexible membranes or bellows such as the rolling
diaphragms
also discussed above. The rolling diaphragms attach to the moving piston rod
and the
engine casing structure enable the piston rod to move relative to the casing
and to provide
an oil tight seal between the oil filled crankcase and the workspace, ensuring
that the
lubricant is maintained in the crankcase and does not disperse into the
working fluid of the
Stirling. Dispersion of the oil from the crankcase into the working fluid
would lead to
engine failure. In order for the bellows to function for thousands and
millions of cycles, as
necessary, a small pressure difference must be maintained across the bellows.
An airlock is
provided between the constant pressure crankcase and ossilating pressure
workspace to
create a volume at the mean pressure of the workspace. The pressure of this
airlock may be
controlled to provide a constant pressure difference across the bellows. This
is described in
more detail above.
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An important aspect of the rolling diaphragm and oil lubricated crankcase
relates to
the use of an airlock 10401 and an airlock pressure regulation system 10411 as
shown
previously in FIGS. 44A and 44B. By way of review, the airlock pressure
regulation
system 10411 provides the benefit of ensuring that an appropriate / desired
pressure
differential is maintained across the rolling diaphragms 10490 and that
working gas
escaping into the crankcase is cleaned of lubricating oil returned to the
working space.
Referring now to FIG. 63, as previously discussed, the pressure of airlock
space
13101 is desired to be maintained at essentially 1500 PSI and equal to the
mean pressure of
working space 13103. Other pressures are of course possible with 1500 psi
being an
example of one embodiment of the airlock. In various embodiments, the pressure
in the
working space 13103 may vary approximately +/- 300 psi so the function of the
airlock
space 13101 is to insulate the diaphragms 13190 from such fluctuations and
maintain itself
at around the necessary pressure, by way of example here 1500 psi, relative to
the 1485 PSI
charged in the crankcase 13110 so that there is approximately about a 15 psi
difference
.. between airlock space 13101 and crankcase space 13110. In various
embodiments, the
crankcase, workspace and air lock are initially charged, at room temperature,
to a pressure
well below the desired operating pressures because the pressure rises as the
gas in the three
volumes, i.e., workspace, crankcase and air lock, heat up during engine
startup. In addition,
during different operating conditions, the temperature of one or more of the
three volumes
.. may change causing changes in pressure across the bellows. In some
embodiments, an
Airlock delta Pressure Regulation (AdPR) block 13111 is provided between the
crankcase
13110 and the airlock space 13101 to create and maintain the exemplary 15 psi
pressure
differential (and/or the desired pressure differential) therebetween in all
operating
conditions including but not limited to engine starup, engine shutdown,
changes in
temperature or speed, leaks from one volume to another or leaks from one
volume to
ambient.
The embodiment as shown in FIGS. 63 and 132 A-C is referred to hereinafter in
general as the "AdPR block" or "AdPR system". In these embodiments the AdPR
block
13111 is connected between the crankcase 13110 and the airlock 13101. The AdPR
block
regulates the pressure difference between the airlock 13101 and the crankcase
13110.
When the reciprocating pistons 13124 of the Stirling cycle machine are movingõ
the AdPR
block 13111 keeps the airlock pressure preferably 10 to 15 PSI above the
crankcase
pressure so that the rolling diaphragms are maintained in a desired
anangement, essentially
bellowing into the crankcase. It is to be appreciated that a range of 5 to 20
PSI is possible
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and other pressure differentials can be accomplished by the regulator as well.
The desired
pressure difference across the bellow seal may depend on the material and
physical
dimensions of the bellows, so that other bellows may require other pressure
differences.
While the Stirling cycle machine engine is off, in various embodiments, the
AdPR block
13111 keeps the airlock pressure preferably less than 15 PSI above the
crankcase pressure
and not more than 5 PSI below crankcase pressure. It is permissible to have a
greater
fluctuation of pressure differential when the engine is off since there are
little or no dynamic
forces being applied to the rolling diaphragms 13190 via moving piston rods.
In various
embodiments, the desired pressure difference may vary.
In some embodiments of the AdPR block 13111, for example, as shown in FIGS.
64A-C, the AdPR block 13111 has an outer housing 13207 having mounting
brackets 13229
or other attachment fixtures to secure the AdPR block 13211 on or near to the
Stirling cycle
machine. Along the center of the housing 13207 one or more ports may be
positioned with
at least one port connected to the crankcase designated as crankcase port
13251 and at least
one port connected to the airlock designated as airlock port 13249. The
crankcase port
13251 is connected to the oil filter volume 13219. The airlock port 13249 is
connected to
the AdPR airlock space 13202 surrounding the pump 13221 (FIGS. 64D, 64E, 64F).
A port 13457 in FIGS. 65A, 65B connects the AdPR airlock space with the
airlock
side of the spool valve 13342. A working gas fill port 13214 and drain port
13215 may be
positioned as shown at either end of the block 13211 with a first end of the
housing having
an electrical conduit and wire feed-thru 13317 necessary to drive a pump motor
13313
shown in FIGS. 65C, 65D.
Turning to the cross-section seen in FIG. 65B, in various embodiments, the
AdPR
block 13311 may include a spool valve regulator 13341 with appropriate
passages 13353,
13344 to the pump (not shown) and components of a linear position sensor
13352. The
passage labeled 13344 operates at approximately the crankcase pressure and
connects the
spool valve to the pump inlet port 13439 as shown in FIGS. 65G, 65H. The
passage labeled
13353 operates at approximately the airlock pressure and connects the spool
valve to the
pump outlet port 13438 as shown in FIGS. 65G, 65J. An AdPR pump controller
13350 is
provided in some embodiments and in some embodiments, the AdPR pump controller
13350 may prevent premature wear of the pump components and reduce airlock
pressure
variations. The AdPR pump controller 13350 communicates with at least the
spool position
sensor elements 13352 and may operate the pump only when needed to increase
the
pressure difference between the crankcase volume 13110 and the airlock volume
13101 as
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sensed by the position of the spool valve. In some embodiments, the AdPR
controller
13350 may stop the pump 13312 when the spool position indicates that the air
lock pressure
is sufficiently above the crankcase pressure. In one example, the AdPR
controller 13350
may change the speed of the pump 13321 proportionally to the spool valve
position in order
to achieve a more constant pressure difference across the bellows 13190. In an
example, the
AdPR controller 13350 may run the pump at maximum speed for spool valve
positions
beyond a given value. In an example the AdPR may command the engine to zero
speed for
spool positions beyond a second given value where that level indicates the
airlock pressure
is not sufficiently greater than the crankcase pressure. In various
embodiments, the position
.. sensor 13352 may be a proximity sensor that senses an absolute position or,
may be a
relative, i.e. differential, position sensor. In various embodiments, the
sensor may be any
other type of sensors. Depending on the location of the spool 13342, the
position sensor
13352 may sense the position of the spool 13342 and transmits that data to the
controller
13350.
In some embodiments, the position sensor 13352, shown herein, may be an LVDT
(Linear Variable Differential Transducer) linear position sensor which is a
type of electrical
transformer used for measuring linear displacement. Various embodiments of an
LVDT
generally have three solenoidal coils (not shown) placed end-to-end around a
tube. The
center coil is the primary, and the two outer coils are the secondaries. A
cylindrical
ferromagnetic core 13354, attached to the object whose position is to be
measured, slides
along the axis of the tube 13356. An alternating current is driven through the
primary,
causing a voltage to be induced in each secondary proportional to its mutual
inductance
with the primary. The frequency is usually in the range of about 1 to 10 kHz.
As the core
13354 moves, these mutual inductances change, causing the voltages induced in
the
secondaries to change. The coils are connected in reverse series, so that the
output voltage is
the difference (hence "differential") between the two secondary voltages.
By way of further explanation, in some embodiments, when the core 13354 is in
its
central position, e.g., equidistant between the two secondaries, equal but
opposite voltages
are induced in these two coils, so the output voltage is zero. If the core
13354 is displaced
.. in one direction, the voltage in one coil increases as the other decreases,
causing the output
voltage to increase from zero to a maximum. This voltage is in phase with the
primary
voltage. When the core 13354 moves in the other direction, the output voltage
also
increases from zero to a maximum, but its phase is opposite to that of the
primary. The
magnitude of the output voltage is proportional to the distance moved by the
core (up to its
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limit of travel), which is why the device may be described as "linear". The
phase of the
voltage indicates the direction of the displacement. Because the sliding core
13354 does not
touch the inside of the tube 13356, it may move without friction, making the
LVDT 13352 a
highly reliable device. The absence of any sliding or rotating contacts allows
the LVDT
13352 to be completely sealed against the environment.
Turning to cross-section shown in FIG. 65C, in various embodiments, an
electric
motor 13313 may be used to drive the diaphragm pump 13321 that, depending on
the
position of the spool 13342, may either port to the airlock 13101 or be
deadheading. In
some embodiments, the seals on the spool are removed so that the pump always
pumps to
the air lock 13101. In these alternative embodiments, the AdPR controller
varies the speed
of the pump to produce the same effect on the airlock pressure as deadheading
the pump.
The spool valve 13341 comprises again similar components to those described
previously in
relation to FIGS. 45 and 46A-E. Here the spool 13342 is balanced against the
pressure of
the airlock by a spring 13343. A porting manifold 13302 is shown here to port
the pump
13321 from the crankcase side of the spool valve 13342, and into the airlock
side of the
spool valve 13342. Additionally, the porting manifold 13302 mounts the pump
head 13321
and LVDT position sensor 13352. The crankshaft 13322 and connecting rod
assembly
13323 connects the electric motor 13313 to drive the pump 13321 and an oil
scrubbing filter
13318 is provided adjacent the crankcase port 13351 to ensure that any oil
which makes its
way into the AdPR block 13311 via the crankcase port 13351 is filtered from
the working
fluid being pumped through the regulator 13341 to the airlock 13101.
FIGS. 65E-F disclose a still further cross-section of an embodiment detailing
the
drain port 13415 which connects the volumes of the AdPR block 13411 and also
connects
the engine crankcase 13110 through the oil scrubbing filter 13418 to release
pressure in the
crankcase13110, workspace 13103, airlock 13101 and AdPR block 13411. Gas from
the
crankcase, which may carry oil and other particles, enters the AdPR via port
13451 and
flows through the oil filter 13418 before entering any other volumes of the
AdPR block
13411. The crankcase gas is filtered to prevent oil and contaminants from
entering the
pump 1342 and airlock 13101 and workspace 13103. The filtered gas at the
crankcase
pressure flows into the crankcase side of the spool valve 13442 via line
13446. Filtering the
gas from the crankcase allows the use of gas from the crankcase 13110 to
maintain the
airlock pressure above the crankcase pressure. In some embodiments, an
internal valve may
be positioned within the drain line 13446 to directly connect the airlock
13401 and the
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crankcase 13110 through the oil filter 13418 to reduce differential pressure
fluctuations
during fill and drain operations and when the piston rods 13124 are not
moving.
Although in some embodiments the AdPR may regulate differential pressure, the
rolling diaphragms 13190 may, in some embodiments, still experience
fluctuations, which
may be especially large during normal fill or drain operation. In some
embodiments,
alternatively to an internal valve, a crossover valve (not shown) may be
located externally to
the AdPR block. The valve may connect the two sides of the AdPR between the
airlock
13401 and the crankcase 13110 to avoid a large pressure differential during a
fill or drain
cycle of the Stirling engine. The valve may be opened during a fill or drain
cycle to greatly
reduce the magnitude of these fluctuations and be closed during normal
operation of the
Stirling engine.
In the cross-section in FIG. 66A, the pump port 13555, connecting the
diaphragm
pump directly to the airlock 13101 via port 13449, may be controlled by the
spool valve
regulator 13541. A port 13553 connects to the pump outlet port 13438 via a
check valve
13445 that redundantly prevents a leak from the airlock 13101 to the crankcase
pressure
side of the AdPR 13411. A port 13544 connecting the crankcase side of the
spool valve to
the diaphragm pump inlet 13448 is also shown.
The spool valve regulator 13541 operation is now described with reference to
FIGS.
66A-E. In various embodiments, the spool valve 13542 is biased against the
airlock
pressure so that where the pressure difference between the crankcase and the
airlock is
within normal limits, e.g. 15 PSI, as shown in FIG. 63; the airlock port 13553
is closed by
the spool 13542. Additionally, the position sensor 13552 tells the pump 13421
that no
pumping operation is necessary and the pump does not operate. If the airlock
pressure
drops too low as shown by the valve spool position in FIG. 66A, the spool
13542 is biased
by the spring 13543 to the left, which permits communication between the
airlock port
13553 and the pump port 13553. The position sensor 13552 tells the pump 13421
to turn on
and pump working fluid from the crankcase 13110 to the airlock 13101. The
fluid travels
first through port 13451, then through passages 13446 and 13444, through the
pump 13421,
past the check valve 13445, into the AdPR airlock 13402 via the passage 13453,
and finally
into the engine airlock 13101 through port 13449
In some embodiments, in order to get back into the Stirling cycle engine
helium
flows from the crankcase 13110 through the filter 13518 and the small amounts
of oil in the
helium are scrubbed out to keep any oil from getting into the AdPR block
13511, the pump
13421 and the Stirling engine where oil can damage the Stirling engine. In
some
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embodiments, the Stirling engine itself may be disabled to ensure that until
the pressure
between the crankcase and the airlock is better equalized so that the rolling
diaphragms will
not be stressed. In some embodiments, the engine may be stopped if the
pressure difference
between the airlock and the crankcase is too low as measured by the position
of the spool
position sensor 13552.
In some embodiments, where the airlock pressure is too high as in FIG. 66D,
the
airlock port 13553 may be connected to the crankcase port 13544 and the pump
13421 may
be disabled while the airlock pressure is reduced. FIGS. 66B and 66C show the
airlock
pressure within normal limits. In FIG. 66B, the airlock port 13553 is closed
by the spool
13541 and the spool is still displaced enough according to the LVDT sensor
13552 to cause
operation of the pump 13421 even without flow from the pump to the airlock. In
some
embodiments, within normal limits, FIG. 66C shows the airlock port 13553
closed by the
spool 13541 and the spool displaced so that the LVDT sensor 13552 does not
turn on the
pump 13421.
In some embodiments, as shown in FIG. 66E, where the Stirling cycle engine is
shut
down or in the case of an airlock leak, the crankcase 13110 may be pressurized
higher than
airlock 13101 and workspace 13103, forcing the rolling diaphragms in a way
opposite from
their intended use. In some embodiments, in the event of such a leak, where an
internal
pressure measurement within the spool is 0-5 PSI higher than the airlock
pressure, an
internal spool valve seated within the main spool valve 13541 may open to
equalize the
pressure. As with the outer spool valve, in some embodiments, a spring
balances against
the pressure differential (against the crankcase pressure in this case) and
would only open
during a time when the engine is pressurized and off in order to reduce damage
to the
rolling diaphragms.
In some embodiments, the control of small pressure changes within the airlock
and
maintaining pressure differential of 5 to 20 PSI above the crankcase pressure
may be
achieved using a pump controller that may accurately vary the speed of the
pump to run
faster-slower when necessary. In some embodiments, using the LVDT sensor and a
suitable
pump controller, a desired range and/or threshold range may be determined. In
some
embodiments, where the pressure differential is outside the desired /
threshold range, the
pump may cycle at a higher rate. In some embodiments, where the differential
is closer to
or within the desired / threshold range, the pump may cycle at a slower rate
thereby
accurately controlling pressure within the airlock.
112
In some embodiments, one or more of the functional components of the ADPR
module may be located within the Stirling Engine pressure vessel. In one
example, the oil
filter may be located between the four piston rods and either in the airlock
or just below the
airlock in the crankcase. The inlet line to the pump would run from the clean
side of the oil
filter to the pump location either in an external ADPR or to an electric pump
located within
the pressure vessel.
Stirling Engine Controller
Another important aspect of the present embodiment from the standpoint of
controlling the actuation of the above described airlock and AdPR block, as
well as the rest
of the Stirling engine, is the Stirling engine controller 13660 shown
diagrammatically in
FIGS. 674B-67H. The engine controller 13660 itself, in some embodiments, may
be
separate from but connected to and in communication with a power electronics
software and
hardware scheme which facilitates conversion of mechanical to electrical
energy essentially
downstream from the Stirling. Some embodiments of the power electronics may be
those
described in U.S. Patent Application No. 13/447897 filed on April 16, 2012 and
entitled
"MODULAR POWER CONVERSION SYSTEM. While the two systems may share data
and communications of numerous system variables in some embodiments, the
engine
controller 13660 is generally understood as a separate system from the power
electronics.
The engine controller 13660 retains responsibility for all aspects of the
Stirling engine
operational control including, but not limited to, regulation of the airlock
as described above
as well as for example the four burners 13674 in the current embodiment of the
Stirling
engine. Clearly a controller such as described here can accommodate control of
other
Stirling designs as well.
In some embodiments, and as shown in FIGS. 67B-67H, a fuel source 13662
provides fuel to each of four separate gas trains including valves 13666 and
variable flow
element 13668 through a main regulator 13664. In some embodiments, each valve
13666 is
a dual gas valve, each shown with a combination regulator. In some embodiments
the
valves 13666 may be MAXITROLTN4 CV-300 valves for example, as manufactured by
the
Maxitrol Company, Southfield, Michigan, USA or White-Rodgers 36H32-423 valves,
for
example, as manufactured by White-Rodgers, Saint Louis, Missouri, USA. The
variable
flow element 13668 provides a variable flow resistance to vary the flow of
fuel to each
burner assembly 12907 independently of the air flow through the burner
assembly 12907.
In some embodiments, the modulating valve may be a MAXITROLTm EXA-40 as
manufactured by the Maxitrol Company, Southfield, Michigan, USA. Another
embodiment
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may forgo the main regulator 13664 and regulate the gas pressure with a
combination
valves 13666 such the MAXITROL' CV-300 as manufactured by the Maxitrol
Company,
Southfield, Michigan, USA. In various embodiments, the fuel delivery may be
controlled in
part by a variable flow element 13668 which in turn may be controlled and may
be
monitored by the engine controller 13660 over four respective control signal
lines 13670. It
is to be appreciated that in some embodiments, the variable flow element may
be a rotary
actuator and a throttle plate. In some embodiments, the actuator 13668 may be
a
modulating valve such as a MAX1TROLTm EXA-40 as manufactured by the Maxitrol
Company, Southfield, Michigan, USA. A blower 13672 provides the air flow for
combustion in the burner(s) 13674, as well as cooling of a burner enclosure.
The engine
controller 13660 controls the air flow via a speed command 13678 passed to the
variable
frequency drive 13676. A blower speed signal 13677 may provide a feedback
signal to the
controller 13660 which permits, amongst other, evaluation and control of the
blower drive
by the engine controller 13660. The blower 13672 is shown here as a single
blower but
may also be a plurality of blowers. A main valve enable relay 13680 and
pressure switch
13682 to enable/disable the fuel valves 13666 is provided in conjunction with
the blower
13672 and with a series of heater head temperature sensors 13685 and a
Programmable
Logic Controller (PLC) 13684. An oxygen sensor 13690 as discussed previously
in this
application may also be provided in the burner as well to communicate oxygen
data to the
controller which may also facilitate the fuel and ignition control.
In some embodiments, for each of the four burners 13674 a flame detection
sensor
13692 is provided which, as described previously in this application, is
critical for safety,
temperature control and the ignition process among other things. In some
embodiments,
the ignitor circuit 13694 for each of the burners are directly influenced by
the flame sensors
.. 13692 through the controller 13660 and the ignitor circuit 13694 are
controlled via the
igniter signal lines 13696 based on the flame sensor data and other data from
the engine
such as oxygen sensors for example. In various embodiments, ignition may be
accomplished by either spark or hot surface ignitors and flame detection
sensors may
include a flame rod and flame rectification, as well as optical sensing of the
flame or
alternatively other methods of known flame detection.
In some embodiments, the ignitor circuits 13694 are commercial combustion
control
circuits that open the fuel valves and attempt to ignite a flame, then monitor
the flame,
attempt to relight the flame if it fails and closes if the fuel valves if
unable to establish a
flame within a given amount of amount of time. In some embodiment, the
combustion
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control circuits are customized variants of Series 35-53 made by Fenwal
Controls of
Ashland Massachusetts. In these embodiments, the engine controller 13660
enables the
ignitor circuit 13694, which each close a relay to power the associated fuel
valves 13666,
and initiate an ignition sequence. In some embodiments, if at any time the
ignitor circuit
13694 is unable to ignite and detect a flame, it will open the fuel valve
relay, thereby
closing the fuel valve 13666 for that burner. In some embodiments, if the
pressure switch
13682 does not detect air flow, the system may interrupt power to all the fuel
valves 13664
thereby ending combustion and preventing a safety hazard. In some embodiments,
if the
overtemp circuit 13684 detects excessive temperatures in a given heater head,
then the fuel
valve 13666 associated with that heater head may be closed to prevent damage
to that heater
head and allow it to cool.
In various embodiments, coolant flow and temperature are also inputs to the
controller 13560 to control the coolant flow pump 13698 and ensure that
appropriate
coolant temperature is maintained in the Stirling cycle. In some embodiments,
the Airlock
delta Pressure Regulator (AdPR) 13611 is also directly connected and
controlled via the
engine controller 13660. In some embodiments, the engine controller 13660
receives the
airlock pressure data from the AdPR 13611 as described above and activates the
pump in
the AdPR to maintain the appropriate pressure differential between the
crankcase and the
airlock.
In some embodiments, the engine controller 13660 may also communicate with the
power electronics (not shown) over CAN bus but could also, in some
embodiments, rely on
wireless communications or other communications protocols such as USB. The
engine
controller 13660 may, in some embodiments, command the speed of a permanent
magnet
synchronous motor ("PMSM") motor. Embodiments of power electronics as they
relate to
control and monitoring of the PMSM motor may be those described in U.S. Patent
Application No. 13/447,897 filed on April 16, 2012 and entitled "MODULAR POWER
CONVERSION SYSTEM. For purposes of this discussion with regards to one
embodiment
discussed in the present application, the engine controller 13660 and power
electronics may
exchange data and commands including, but not limited to, motor drive velocity
command,
generator velocity, Bus voltage, Bus current, motor drive IGBT bridge temp.,
shunt control,
shunt active, battery voltage, battery temperature, inverter power, inverter
enable, inverter
PWM, inverter voltage inverter current, inverter temperature, converter power,
converter
enable, converter PWM, converter voltage, converter current and converter
temperature.
Here converter
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refers to one or more DC/DC converter circuits. Certain direct inputs into the
engine
controller 13660 may also be necessary and can include but are not limited to
oil
temperature from the crankcase, battery temperature, motor temperature and
shunt
temperature.
Annular Venturi Burner
FIGS. 70A - 70D disclose a further embodiment of an annular-venturi burner
13801 for use
in conjunction with a multiple heater head and piston engine as described
previously in
FIGS. 54¨ 62. This embodiment of the burner 13801 is also specifically
directed to the
independent heating of multiple heater heads 13803. In this embodiment there
are four (4)
heater heads 13803 and respective burner assemblies 13807, although only two
(2) are
visible in the cross-section of FIG. 70A, and more or less heater heads are of
course also
possible for the engine. As in the previously discussed embodiments the heater
heads
13803 and burner assemblies 13807 are encompassed by a burner housing 13811
and each
heater head 13803 is heated by an individual burner assembly 13807. The burner
assembly
13807 is supplied a fuel/air mixture for combustion via a blower (not shown)
that supplies
air through the air inlet 13823 and a fuel system (not shown) that provides
fuel through the
fuel inlets 13816. The flame may form within the ventuii body 13847 and / or
in the
combustion chamber 13831 formed by an annular arrangement of the heater tubes
13809.
The hot combustion gases then flow past the heater tubes 13809 before entering
a
recuperative preheater 13851-13855. Burner fairings 13808 in the form of metal
rings
mounted on a mid-burner plate 13805 direct the hot combustion gases across the
heater
tubes 13809 and diverts the hot combustion gases from exiting axially from the
heater tubes
13809. As described previously in relation to FIGS 59 and 62 the exhausting
combustion
gases pre-heat the incoming air in the recuperative heat exchanger,13851-1385
and then exit
the burner housing 13811 through exhaust outlets (not shown).
More specifically, observing FIG. 70A, The burner housing 13811 is comprised
of
an outer wall 13811A and a manifold 13811B that connects the inlet 13823 port
to the air
side 13852 of the recuperative heat exchanger and the exhaust port to the
exhaust side
13852A of the recuperative heat exchanger. In some embodiments, the manifold
13811B
may include an interface 13804A to the stirling engine. The combustion gases
exiting from
the heater heads 13803 may be sealed at the interface with o-rings or other
airtight seals and
the manifold 13811B may be mechanically attached to the cooler plate 13810 of
the Stirling
engine with a marmin clamp. In other embodiments, the burner 13801 may be
mechanically attached to the Stirling engine with a bolted flange or other
mechanical
116
means. The heater heads 13803 themselves may be any of the various embodiments
of tube
heater heads described in the preceding sections, including, but not limited
to, straight tube
heater heads or helical tube heater heads as disclosed in U.S. Patent
Application Ser. No.
13/447990, filed April 16, 2012. By way of example, the present embodiment is
contemplated utilizing heater tubes 13809 through which flow the working gas,
for example
helium, which is heated by the respective burner assemblies 13807.
In various embodiments of the burner 13801, the single blower (not shown) may
be
incorporated to maintain a consistent average air flow supplied to the burner
13801 and
hence to each of the individual burner assemblies 13807. In some embodiments,
the blower
may also produce a variable air flow when necessary to control the fuel/air
mixture in the
venturi 13841. The blower may supply combustion air at a desired velocity
depending on
instructions from a controller for purposes of ignition, then once ignition
has occurred, the
desired air flow rate may be regulated by the controller dependent on data
received from
sensors, which may include, but not limited to, an oxygen sensor sampling the
cooled
exhaust gas. In some embodiments, a fuel system with variable flow valves for
each heater
head, which, in some embodiments, may be a MaxitrolTM EXA-40 valves for
example, as
manufactured by the Maxitrol Company, Southfield, Michigan, USA, may control
the fuel
flow to achieve the commanded temperature on each heater head. The blower may
be
adjusted to achieve the desired fuel/ air ratio.
An important aspect of the present embodiment is the efficient heating of the
incoming air through the extraction of waste heat in the exhaust to raise the
incoming air
temperature thereby improving the efficiency of combustion processes and the
burner
13801. The blower connects through the air inlet 13823 in the outer wall
13811A of the
burner housing into an air channel 13851. The air channel 13851 extends
circumferentially
around the burner inside the outer wall 13812 of the burner 13801 and directs
the air
developed by the blower B across an intermediate baffle 13853 and up into the
hot manifold
13857 before entering the swirlers 13882 of the venturi ejectors 13841. The
intermediate
baffle 13853 directly separates the incoming air from the exhaust gases
exiting the burner
through an exhaust channel 13855 and provides for the heat transfer from the
exiting
exhaust to the incoming air in the air channel 13851. The heat transfer
efficiency across the
intermediate baffle 13853 is critical because the hotter the incoming air can
be heated, the
less fuel is necessary to reach the desired combustion temperatures.
The incoming air is preheated to a desired temperature, for example, but not
limited
to 600-750 C. These are many reasons preheating the incoming air may be
beneficial and
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these include, but are not limited to, facilitating ignition and combustion as
the air is
directed to the burner assemblies 13807 and/or increasing the thermal
efficiency of the
burner by capturing some of the thermal energy in the combustion gases exiting
the heater
heads. In some embodiments, preheating of the air may reduce the hot exhaust
temperature
from 900 C to 300 C. In some embodiments, the amount of preheating which may
be
accomplished may be related to the efficiency of heat transfer from the
exiting exhaust to
the incoming air. The heat transfer across the intermediate baffle 13853 may
be improved
by adding rows of folded fins 13852 on the air side and folded fins 13852a on
the exhaust
side of the intermediate baffle 13853. The folded fins may be brazed to the
intermediate
baffle 13853 to assure good thermal attachment. In various embodiments, the
material
properties of the folded fins may be optimized for the operating temperature.
For example,
in some embodiments, the rows of folded fins near the top may be heat
resistant metals,
which may include, but is not limited to, INCONELTm 625, while lower and
cooler folded
fins may have higher thermal conductivity but lower operating temperature. In
various
embodiments, the materials for these folded fins may include, but is not
limited to, stainless
steel 409 or Ni 201 for example. The preheated air exits the air channel 13851
and is
directed radially into a hot air manifold 13857 which communicates with each
of the
multiple burner assemblies 13807 specifically directing the preheated air to
the swirler
portion 13882 of the venturi 13841. In various embodiments, the preheated air
enters the
hot air chamber 13857 through a substantially 360 degree circumferential
opening around
the exit of the air channel 13851. In some embodiments, this may result in a
consistent flow
rate of preheated air delivered to each of the burner assemblies 13807. In
various
embodiments, additional channels or passageways (not shown) may be provided in
the hot
air chamber 13857 to direct the preheated air in the hot air chamber to a
specific burner
head. In various embodiments, the 360 degree output from the air channel 13851
is used
when there is only one blower B developing the air flow into the engine.
The burner assembly 13807 may be further understood by referring to FIGS. 70B-
70D. Several elements of the burner assembly 13807 are shown in FIG. 70B
include a
venturi body 13841, a fuel inlet 13816, an air swirler 13882, an ignitor
13818, and a flame
detector 13860. The venturi 13841 in this embodiment is a venturi type ejector
as disclosed
for example in U.S. Patent Application Serial No. 12/829,320 filed July 1,
2010, now U.S.
Publication No. US-2011-0011078-Al published January 20, 2011 and entitled
Stirling
Cycle Machine (Attorney Docket No. 178). In some embodiments the venturi 13841
may
be beneficial for many reasons, including, but not limited to, providing the
benefit of
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reducing or eliminating the need for a completely separate fuel control scheme
as regulation
of the airflow changes the vacuum which, in turn, correspondingly affects the
fuel flow and
regulates the burner power. In some embodiments, the venturi allows use of
typical gas
pressures in buildings, e.g., 7 inches of water column, without a compressor.
The blower
forces air through radial swirler vanes 13882. The flow of swirling air mixes
with fuel in
the venturi throat 13847 and froms a swirl stailized flame in the expanding
section of the
venturi 13849 and/or in the combustion chamber 13831 of FIG. 70C.
Referring again to Fig. 70B, the ignitor 13818 ignites the fuel-air mixture in
the
venturi throat 13847. The ignitor 13818 may be a spark plug and may function
as the high
voltage electrode of a flame ionization flame detector. In other embodiments,
the ignitor
12918 may be a hot surface ignitor. In one embodiment the ignitor are Silicon
Nitride Hot
Surface ignitors produced by Crystal-Technica of South Grafton, MA. The
ignitor 13818
may located advantageously near the center of the venturi 13841 to promote a
uniform
composition and flow from the venturi 13841. The ignitor 13818 may be mounted
in the
ignitor port 13817B.
The flame detector 13860 mounts in the 13817C port and provides a signal to
the
controller indicative of the presence or absence of a flame from the burner
assembly 13807.
In one embodiement, the flame detector 13860 comprises a temperature sensor
13862 inside
a heat resistant tube 13861. The temperature sensor 13862 may be a type K
thermocouple
in an inconel sheath, a type B or R thermocouple or other high temperature
sensor. The heat
resistant tube 13861 may be a heat resistant metal such as InconelTm 625, Mar-
M or it
may be a ceramic tube formed from zirconia or other high temperature ceramics.
In other
embodiments the port 13817B may be used for a vision based flame detection
circuit
including but not limited to one or more of the following: IR flame detector,
visible light
flame detector and/ or UV flame detector. In another embodiment, the flame
sensor may be
a flame rod connected to a Series 35-52 Ingition controller as manufactured by
by Kidde-
Fenwal Inc. of Ashland MA.
Still referring to FIG. 70B, the fuel enters the burner assembly 13816 through
the
fuel inlet 13816, that is mounted in port 13817A. The fuel flows into a plenum
or manifold
13871 before flowing through the fuel ports into the venturi throat 13847. The
mixing of
the fuel and air in the venturi throat is best shown in FIG 70C, which is a
detailed view of
the venturi throat and fuel injectors. The air enters the venturi throat 13847
with an
induced swirl created by radial vanes 13882 that are described in detail
below. The air
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initial flows in radially toward the ignitor 13818 and then is directed into
an axial flow by
an axisymetric protrusion 13880 around the ignitor and the inlet to the
venturi 13877A. In
one embodiment pictured in FIG 70C, the inlet to the venturi 13877A is part of
the venturi
bushing 13877. In other embodiements the inlet 13877 may be an integral part
of the
venturi body 13841 or the a conical or a separate piece. In some embodiments,
the inlet
13877A and the protrusion 13880 may be formed to provide an approximately
constant
cross-section flow area as the flow changes from radial to axial. One possible
benefit of an
approximately constant flow area is a minimization of pressure drop through
the burner
assembly 13807. The protrusion 13880 may be conical or may have a surface with
an
increasing slope from near horizontal to approaching vertical.
Still referring to FIG. 70C, the fuel enters the fuel plenum 13873 from the
fuel inlet
tube 13816 that is connected to the fuel system that was described above in
reference to FIG
62. The fuel plenum forms an annular space around the outside diameter of the
venturi
body 13841 and supplies fuel to a plurality of fuel ports 13875. The fuel
ports 13875
provide fuel to an annular space between the venturi bushing 13877 and the
venturi body
13841. This annular space is mostly filled with fuel and is herein referred to
as the fuel
annulus. The fuel exits the fuel annulus flowing axially along the walls of
the venturi
throat 13847. The air flowing axially through center of the bushing 13877 and
into the
venturi throat 13847 creates a region of low pressure downstream of the fuel
annulus that
draws the fuel into venturi throat 13847 where it mixes with air. The exit of
the fuel
annulus is peferrably upstream of the ignitor 13818 by a sufficient distance
to create an
ignitable mixture of fuel-air next to the hot surface or spark of the ignitor
13818. The fuel
annulus may have a very thin annular opening to maximize fuel flow velocities.
In other
embodiments, the annulus is larger. In general, the annular exit of the fuel
annulus has a
constant radial gap to maximize fuel flow uniformily around the venturi throat
13847. In a
preferred design, the gap is 1/20 of the venturi throat diameter or has a
radial gap of 0.035".
The fuel ports 13875 may be radial or may enter the fuel annulus at an angle
to induce a
swirl in the fuel.
FIG. 70D presents an isometric view of the swirl vane plate 13882 mounted on
the
.. venturi body 13841, where the venturi entrance 13877A is visible. The
radial vanes
13882A impart a tangential velocity or swirl motion to the radially flow air.
In one
embodiment, the radial vanes 13882A are straight and not curved. In another
embodiment
the vanes are curved sothat the vanes are radial at the outside diameter of
the radial vane
plate, the vanes curve until the flow leaves the inward edge of the vane with
the desired
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swirl. In FIG. 70D, the vanes are curved and aerodynamic to minimize pressure
drop
through the burner. The vanes are essentially airfoils that are initially
thick near the ouside
diameter of the radial venturi plate and increasingly thin toward the venturi
inlet.
One theory on the advantage of injecting the the fuel through and the fuel
annulus is
that the annulus provides a more uniform provision of fuel around the venturi
throat by
providing a plenum for the fuel from the plurality of fuel jets to mix and
flow uniformly into
the venturi throat 13847. Another theory on the advantage of the fuel annulus
is that it
advantageously places the fuel next to the wall, where the local air flow may
be more
uniform than in the center. Still another theory is that the lack of fuel jets
across the airflow
avoid disturbing the airflow and result in a more uniform air flow exiting
from the venturi
throat 13847.
While the principles of the invention have been described herein, it is to be
understood by those skilled in the art that this description is made only by
way of example
and not as a limitation as to the scope of the invention. Other embodiments
are
contemplated within the scope of the present invention in addition to the
exemplary
embodiments shown and described herein. Modifications and substitutions by one
of
ordinary skill in the art are considered to be within the scope of the present
invention.
