Vibration absorber for a free piston Stirling engine

A free piston Stirling engine has a hermetic vessel containing a displacer piston and a power piston coupled to each other and to the vessel by gas springs. An additional mass is coupled to the vessel by a gas spring which is tuned to the operating frequency of the engine geometrically. The vibration absorber is effective at all engine frequencies because the absorber gas spring follows the engine working gas mean pressure and changes the absorber natural frequency as the engine frequency changes.

BACKGROUND OF THE INVENTION 
This invention relates to vibration cancellation in reciprocating 
machinery, and more particularly to a vibration absorber for a free piston 
Stirling engine. A free piston Stirling engine is a sealed power unit 
containing a piston and a displacer within a closed vessel. The piston and 
displacer reciprocate out-of-phase in the vessel to circulate a working 
fluid through a closed loop from a compression space, through a cooler, 
regenerator, and a heater to an expansion space and then back through the 
same loop cyclically to subject a working fluid to a thermodynamic cycle 
approximating the theoretical Stirling cycle. 
A simple and reliable form of free piston Stirling engine contains a single 
displacer and a single power piston. This form of engine presents the 
simplest control problems but is inherently unbalanced. That is, the 
reciprocating masses transmit an alternating force to the sealed vessel 
and this force must be absorbed by massive mounting structure or a 
sophisticated suspension arrangement. One such suspension arrangement is 
shown in U.S. application Ser. No. 153,839 entitled, "A Suspension and 
Vibration Oscillation System Incorporated Into the Mass Flow System for a 
Linear Reciprocating Machine" filed on May 27, 1980, by Peter Curwen, et. 
al now U.S. Pat. No. 4,360,087. This arrangement is extremely effective 
and produces a remarkable cancellation of force on the system so that the 
sinusoidal force exerted by the case is completely cancelled by the 
suspension system and the connection to the ground experiences an 
insignificant transmitting force, if any. 
Although this system works well, it would be desirable to provide a system 
more particularly designed for a free piston Stirling engine which 
automatically compensates for changing frequency and/or stroke of the 
reciprocating members in the engine. Moreover, certain installations of 
the free piston Stirling engine may necessitate the elimination of all 
substantial vibration by the vessel itself. 
SUMMARY OF THE INVENTION 
Accordingly, it is an object of this invention to provide an automatically 
compensating vibration absorption system for a free piston Stirling engine 
which automatically compensates for changing frequency and stroke of the 
reciprocating members of the engine. It is another object of this 
invention to provide a vibration absorption system that eliminates all 
substantial vibration of the engine vessel and whose parts are sealed 
within the vessel to protect from corrosion or mechanical injury. 
These objects are obtained in one embodiment of the invention wherein a 
fourth mass is sprung to the vessel by gas springs disposed in both 
directions of its travel and each having a stiffness that is tuned to 
provide, with the fourth mass, a spring mass system having a natural 
frequency equal to the operating frequency of the engine. A gas spring 
pressure balancing system is provided to maintain the pressure of the gas 
springs such that the tune of the absorber and the engine are maintained.

DESCRIPTION OF THE PREFERRED EMBODIMENT 
Referring now to the drawings wherein like reference characters designate 
identical or corresponding parts, and more particularly to FIG. 1 thereof, 
a free piston Stirling engine is shown having a sealed vessel 10 enclosing 
a working space including an expansion space 12 and a compression space 
14, respectively, above and below a displacer 16 mounted for oscillation 
in the working space. The oscillation of the displacer 16 circulates 
working fluid contained within the working space from the expansion space 
12 through a set of heater tubes 18, a regenerator 20, a cooler 22, and 
back into the compression space 14 below the displacer 16. The displacer 
16 causes the gas to circulate back and forth through the heat exchangers 
between the compression and the expansion spaces and subjects it to a 
thermodynamic cycle in which heat energy, entering the heater tubes 18 
from a combustor 24 burning fuel in a combustion space 26, is converted 
into mechanical energy of a power piston 28 which reciprocates under the 
influence of a pressure wave in the working space created by cyclically 
heating, expansion, cooling, and compression of the working gas. 
The power piston 28 includes an axially extending plunger 30 which 
reciprocates opposite a stationary stator 32 attached to the interior wall 
of the vessel 10. The plunger is made up of a pair of axially spaced 
annular pole faces 34 and the stator 32 includes a corresponding pair of 
pole faces 35. The alternator is in the form disclosed in U.S. Pat. No. 
3,891,874. 
The volume in the vessel below the power piston 28 is sealed and 
constitutes a bounce space 36 or gas spring which returns the power piston 
28 toward the top or engine end of the vessel after it has been displaced 
downwardly, as shown in FIG. 1, by the pressure wave in the working gas. 
The displacer 16 is suspended in the working space by a suspension system 
37. The suspension system includes a diaphram 38 connected at its outside 
peripheral edge 40 to the lower edge of the displacer shell 16 and 
connected at its center 42 to a partition 44 perforated by openings 46. 
An oil backing system is provided for the diaphram 38, including a rigid 
plate 48 attached at its outer peripheral edge to the shell of the 
displacer 16. An oil cavity 50 is defined between the diaphram 38 and the 
rigid backing plate 48. A cylinder 52 is connected to or integrally formed 
with the backing plate 48 and receives a piston 54 connected rigidly to 
the partition 44 by a piston rod 56 so that the piston 54 remains 
staionary relative to the vessel 10. As the displacer 16 oscillates, the 
cylinder 52 moves up and down relative to the piston 54 to compensate for 
volumetric changes in the oil cavity 50 so that the diaphram 38 is not 
subjected to pressure induced stresses but sees only displacement induced 
stresses. This structure is more particularly described and explained in 
my copending U.S. Pat. No. 4,372,115 for "Oil Backed Stirling Engine 
Displacer Diaphram," filed concurrently herewith. 
A separate housing 60 is attached to the vessel 10 by matching flanges 62 
and bolts 64, or it may be integrally formed with the vessel 10. A 
cylinder 66 is enclosed within the housing 60 and contains a piston 68 
which oscillates freely in the cylinder and is centered therein by weak 
centering springs 70. Top and bottom piston rings 72, or in their place 
labyrinth seals, prevent the leakage of gas passed the piston 68 from one 
side to the other. 
A gas line 76 extends from the compression space 14 of the engine to the 
bounce space 36 and to the top section 80A of an annular manifold 80 
connected to the axial midpoint of the cylinder 60. A set of midstroke 
ports 82 extend through the housing 60 and communicate between the 
cylinder 66 and the top section 80A of the manifold 80. A restriction 84 
is disposed in the portion of the gas line 76 extending between the 
compression space 14 and the bounce space 36 so that only the mean 
pressure in the working space is communicated through the gas line 76 to 
the bounce space 36 and the manifold 80. 
The annular manifold 80 includes a lower manifold section 80B which 
communicates with the cylinder 66 through a set of ports 83. A gas line 86 
connects the lower manifold section 80B to the cylinder volume 90 above 
the piston 68 and another gas line 92 connects the upper manifold section 
80A with the cylinder volume 94 below the piston 68. An annular groove 96 
formed in the axial midposition of the piston 68 provides a communication 
channel for gas flow between the ports 82 and 83 at the midstroke position 
of the piston 68 but is misaligned from the ports 82 and 83 at all 
positions other than the midstroke position and therefore closes off gas 
flow communication between the ports 82 and 83 at these other positions. 
In operation, the piston oscillates in phase position to the phasor 
addition of the piston and displacer inertia phasors so that the phasor 
addition of the inertia phasors of all the oscillating masses is virtually 
zero. The cancellation is not complete because there is some friction and 
hysteresis losses in the vibration absorber as the piston oscillates 
axially, and as the gas volumes 90 and 94 is alternately compressed and 
expands. This slight power loss by the absorber is made up by a slight 
phase lag from what the absorber mass motion would be without the absorber 
power loss. This causes a slight motion of the vessel 10 out of phase with 
the displacement of the piston 68 so that the gas spring volumes 90 and 94 
are compressed and expanded slightly more than that which would be 
achieved by the movement of the piston 68 alone. 
The midstroke porting arrangement provided by the manifold 80, the ports 82 
and 83, and the piston groove 96 enables the gas spring volumes 90 and 94 
to communicate through the gas lines 92 and 86 at the midstroke position 
of the piston 68. This ensures that the mean pressure in the gas spring 
volumes 90 and 94 will be equal so that the piston midstroke position will 
remain at the cylinder midstroke position. The gas line 76 communicating 
with the cylinder 60 through the restriction 84 ensures that the mean 
pressure in the gas spring volumes 90 and 94 will remain equal to the mean 
pressure in the engine working space. 
A schematic diagram of an engine and vibration absorber system is shown in 
FIG. 2 with the springs, masses, and dampers labeled, and a phasor diagram 
for the displacements and inertias of the elements in the system is shown 
in FIG. 3. The displacement phasor of the power piston 28 is indicated in 
FIG. 3 as phasor X.sub.P and its corresponding inertia phasor which is in 
phase with a displacement phasor is labeled I.sub.p where I.sub.p =M.sub.p 
.omega..sup.2 X.sub.p. The displacement phasor of the displacer 16 is 
labeled X.sub.d and leads the power piston phasor by some 
60.degree.-80.degree.. The displacer inertia phasor I.sub.d is in phase 
with the displacer displacement phasor X.sub.d where I.sub.d =M.sub.d 
.omega..sup.2 X.sub.d. The inertia displacer phasor is smaller than the 
power piston inertia phasor because the displacer mass is considerably 
less than the power piston mass. The phase addition of the power piston 
inertia phasor and the displacer inertia phasor is shown as I.sub.r ; it 
is the inertia phasor that is transmitted to the sealed vessel 10 in the 
absence of a cancellation or absorption system. 
The inertia phasor I.sub.r of the reciprocating engine components is 
substantially balanced by the inertia phasor I.sub.a of the reciprocating 
absorber piston 68. The inertia phasor I.sub.a lags the position of the 
inertia phasor I.sub.r by slightly more than 180.degree. depending on the 
friction and hysteresis losses in the absorber system, as mentioned 
previously. 
The cancellation or virtual cancellation of inertia by the absorber piston 
68 under all operating conditions is achieved by self adjusting of the 
frequency and stroke of the absorber piston 68 as the conditions change. 
The frequency of the piston 68 is maintained by adjustment of the pressure 
in the gas spring volumes 90 and 94 when the mean pressure of the engine 
working space changes. The engine frequency is determined by the mean 
pressure of the working space. Thus, as the engine mean pressure 
increases, the frequency of the engine increases correspondly. 
Simultaneously, the mean pressure in the gas spring volumes 90 and 94 
increase correspondly and the resonant frequency of the spring-mass system 
represented by the mass of the absorber piston 68 and the gas springs 90 
and 94 increases equally. This frequency equality can be demonstrated by a 
derivation based on the inertia equality equation. This is represented by 
the equation M.sub.a .omega..sup.2 X.sub.a =M.sub.p .omega..sup.2 X.sub.p 
+M.sub.d .omega..sup.2 X.sub.d (this equation neglects the small movement 
of the vessel 10 which is necessary to impart sufficient work into the 
absorber spring mass system to overcome the damping effects of friction 
and hysteresis losses). The frequency of the absorber system is determined 
by the spring constant and the mass of the reciprocating member so that 
.omega..sup.2 =K.sub.a /M.sub.a where K.sub.a and M.sub.a are the spring 
constant and mass of the absorber, respectively, and where: 
##EQU1## 
in which expression: .gamma.=the ratio of constant volume to constant 
pressure specific heats of the working gas; 
P.sub.m =mean pressure; 
A.sub.a =absorber piston area; 
V.sub.a =absorber gas spring mean volume; 
K.sub.a =spring constant (of each gas spring). 
Since the only thing that changes in the frequency expression is the value 
of P.sub.m and since the value of P.sub.m remains equal for the engine and 
for the absorber system under all conditions because of the mean pressure 
equalization system, therefore the resonant frequency of the two systems 
remain equal to each other, and the phase position of the absorber mass 
remains close to the 180.degree. position relative to the inertia phasor 
I.sub.r of the reciprocating engine members. 
The stroke of the absorber mass is also self adjusting to enable the 
magnitude of the absorber mass inertia phasor to match the magnitude of 
the resultant inertia phasor I.sub.r of the reciprocating parts of the 
engine. When some change occurs in the magnitude of the resultant inertia 
phasor of the moving parts of the engine, the difference will be 
transmitted to the vessel 10 for which purpose spring feet 99 are 
provided, and will be quickly eliminated by either increasing or 
decreasing the stored energy (i.e. stroke) of the absorber mass to provide 
an inertia phasor substantially equal and opposite to the resultant 
inertia phasor I.sub.p of the engine moving masses. 
In operation assume the engine is running in a steady state condition under 
a steady electrical load on the alternator, and then the load on the 
alternator increases. The damping effect of the alternator load on the 
power piston tends to decrease the piston stroke and decrease the stroke 
and phase of the displacer, but this tendency is offset by the engine 
power control (not shown) which senses the incipient of initial power 
piston stroke change by direct stroke measurement, pressure change 
measurement, or sensing of the electrical load. The engine power control 
(one example of which is disclosed in U.S. Pat. No. 4,345,437 for 
"Stirling Engine Control System" filed by John J. Dineen on July 14, 1980) 
causes the engine working gas mean pressure to increase, which increases 
the capacity of the engine to convert input heat energy to mechanical work 
of the power piston. 
The increase in the engine working gas mean pressure causes a nearly 
instantaneous corresponding increase in the mean pressure of the absorber 
gas springs. This increases the gas spring stiffness and increases the 
natural frequency of the absorber spring-mass system to match the engine 
frequency change, so that the absorber remains tuned to the engine under 
the operating conditions. 
The vibration absorber of this invention thus provides the free piston 
Stirling engine with a self adjusting or self regulating mechanism for 
substantially cancelling or absorbing the inertia forces normally 
transmitted to the engine vessel. It eliminates the need for complicated, 
sophisticated, expensive, delicate, and injury prone mechanisms for 
suspending the engine vessel to prevent transmission of the vessel 
vibration to ground because the vessel itself is virtually stationary with 
the use of this invention. The only support mechanisms that the vessel 
requires are the usual simple and inexpensive spring feet on which 
compressors are normally mounted. The invention requires virtually no 
maintenance, has only one moving part and its life should exceed the life 
of the engine. It requires only simple machining techniques and 
inexpensive materials and therefore the cost of the system is low. It is 
itself quite in operation and, because it eliminates the vibration of the 
engine housing, the engine itself runs quieter with the use of this 
invention. It uses only miniscule amounts of energy and, in the unlikely 
event that service is ever required, can be serviced easily. 
Obviously, numerous modifications and variations of the disclosed 
embodiment are possible will occur to those skilled in the art in light of 
this disclosure.