Patent Grant
10323603
(Not Applicable)
(Not Applicable)
(Not Applicable)
This invention relates to free-piston Stirling engines (FPSE) and more particularly relates to an improvement which causes the engine to be automatically depowered in the event that the engine load, as seen by the engine at its output, changes in a manner that the engine would become unstable, for example because of a failure of the engine's controller or wiring to the controller. This depowering prevents an increase of piston amplitude of reciprocation that would otherwise cause a runaway amplitude increase resulting in the piston having engine-damaging collisions with other internal engine components.
A problem with free-piston Stirling engines is that historically they have not been tolerant to loss of load. A kinematic Stirling machine that is adequately designed will, when its load is removed or reduced, often just run at a higher speed and the machine's internal heat exchanger pumping losses consume the power produced. However a FPSE is a resonant machine and so, if unloaded, the frequency will not change significantly. Instead, the piston and displacer will overstroke and collide with physical structures within the engine and with each other. The problem is made worse because the power increases not only with amplitude but also because of the resulting discontinuous motions resulting from collisions. The collisions often lead to failure of internal components and to the generation of debris which can lead to engine failure. The purpose of the invention is to provide a FPSE which is tolerant to loss of engine load because such collisions and damage are prevented by the invention if the engine's load is reduced or becomes zero.
Referring to
Reciprocating motion of the piston 28 and a displacer 30 cause the working gas to be alternately heated and cooled and alternately expanded and compressed in order to do work on the piston 28 that reciprocates in the cylinder 22. The piston 28 has a sidewall 32 that engages and slides along the cylinder 22 and the sidewall has an inward end 34. The terms “in”, “inward”, “out” and “outward” are used as a terminology convention to describe the opposite axial directions of motion of engine components including the piston 28 and the displacer 30. The terms “in” and “inward” indicate a direction or position toward or nearer the working space 8, which includes the compression space 12 part of the working space 8. The terms “out” and “outward” indicate a direction or position away from or farther from the working space 8. The piston 28 also has an annular cutout or relieved portion to form a central cap or boss 36 that is unrelated to the invention. Its purpose is to occupy a volume of the compression space 12 which would otherwise be an unswept volume.
The displacer 30 of a beta type Stirling engine typically reciprocates in the same cylinder 22. The displacer 30 is connected through a displacer connecting rod 38 to a planar spring 40 that is mounted to a casing 42. The casing 42 surrounds a relatively large volume back space 43 and also contains working gas. The reciprocating mass of the piston 28, the reciprocating mass of the displacer 30 and its connecting rod acting upon the planar spring 40 and the resiliently compressible and expansible working gas together form a resonant system which has been called a thermal oscillator.
The reciprocating displacer 30 cyclically shuttles the working gas between the compression space 12 and the expansion space 10 through the heat accepter 14, the regenerator 18 and the heat rejecter 16. This shuttling cyclically changes the relative proportion of working gas in each space. Gas that is in the expansion space 10, and gas that is flowing into or out of the expansion space 10 through the heat accepter 14 accepts heat from surrounding surfaces. Gas that is in the compression space 12 and gas that is flowing into or out of the compression space 12 through the heat rejecter 16 rejects heat to surrounding surfaces. The rejected heat is ordinarily transferred away by the cooling system. The gas pressure is essentially the same in both spaces 10 and 12 at any instant of time because the spaces 10 and 12 are interconnected through the working gas flow path between the expansion space 10 and the compression space 12 and that flow path has a relatively low flow resistance. However, the pressure of the working gas in the working space 8 as a whole varies cyclically and periodically. The periodic increase and decrease of the pressure of the working gas in the working space 8 drive both the piston 28 and the displacer 30 in reciprocation. The periodic pressure variations are caused by the resultant of two components that are out of phase with each other. The first component is the alternating net heating and cooling of the working gas in the workspace. When a majority of the working gas is in the compression space 12, there is a net heat rejection from the working gas and the first component of gas pressure variation decreases. When a majority of the working gas is in the expansion space 10, there is a net heat acceptance into the working gas and the first component of gas pressure variation increases. The second component of gas pressure variation is the result of piston motion which alternately compresses and expands working gas in the working space as a consequence of piston motion.
Gas Bearings.
Because liquid lubricants can foul the heat exchangers or vaporize in the hot regions, Stirling engines are provided with a gas bearing lubrication system. Working gas is cyclically pumped into a gas bearing cavity 44 through a gas bearing inlet passage 46. Although the bearing cavity 44 appears in the drawing as two separate cavities 44A and 44B, the gas bearing cavity 44 is a continuous annular space within the piston. A check valve 48 permits the working space 8 pressure variations in the compression space 12 to pump working gas into the bearing cavity 44 but prevents gas flow in the opposite direction. The working gas within the cavity 44 flows out of the cavity 44 through multiple gas bearing pads 50. The gas bearing pads 50 are chambers that are spaced at annular intervals around the piston with flow restrictive passages into the gas bearing cavity 44. Consequently, the interfacing surfaces of the piston 28 and the cylinder 22 are lubricated, and the piston is centered, by the flow of the pressurized working gas from the gas bearing pads 50 into the small clearance gap between those interfacing surfaces and then into the working space 8 and the back space 43.
Centering System.
FPSEs typically have a net flow of gas over the cycle from the working space to the back space. One cause is that gas passage through the piston/cylinder clearance gap has a net flow in the out direction. The reason is that, although the volume of gas flow is the same in both directions, the density of gas flowing out of the workspace is larger than the density of gas flowing into the workspace. The density is larger because the pressure of gas in the workspace, when gas flows out of the workspace, is greater than the pressure of gas in the back space when gas flows out of the back space. More importantly, for machines with gas bearings, the bearings tend to pump gas out of the working space to the back space such as by the flow through the gas bearing cavity 44 and out the gas bearing pads 50. The reason is that the entire input of gas into the gas bearing cavity 44 is from the workspace 8, but the gas passing out the gas bearing pads 50 is divided between returning to the workspace and flowing to the back space 43. The cumulative effect of this preferential blow-by over many cycles is that the mean position of the piston creeps in. The mean position of a piston is the center or mid-point between the farthest excursions of the piston in opposite directions. The distance between the farthest opposite excursions of a point on the piston is the piston stroke and one half of the stroke is the piston amplitude of reciprocation.
The engine is provided with a centering system that compensates for this preferential blow-by and prevents the inward creep by the piston 28. The centering system illustrated in
Inherent Instability of a FPSE
Most free-piston Stirling engines that are designed according to prior art principles have a typical engine power curve that relates engine power to piston amplitude.
In the absence of the invention and the absence of a controller, engine power is an increasing exponential function of piston amplitude over the engine's operating range. Typically engine power increases as the square of the engine amplitude. That makes the engine unstable with a linear load, such as a resistive electrical load which varies with voltage squared. Those skilled in the art of Stirling engines are familiar with the typical power curve of
Considering
The prior art uses an engine controller to overcome this instability and for additional reasons. The engine controller is commonly interposed between the output of the engine's alternator and input of the ultimate electrical load. Therefore, the controller's input terminals are seen by the output of the engine's alternator as the engine's load. In normal operation the controller prevents the instability and runaway increase in piston and displacer amplitude of reciprocation. Unfortunately, there are occasions when a malfunction of the controller or a disconnection or shorting of a connection between the controller and the FPSE or its alternator causes the load seen by the FPSE to appear as an open circuit or as a short circuit. In either instance there is no load to consume engine power and therefore the conditions for runaway piston amplitude exist. The purpose and object of the invention is to provide simple mechanical modifications of the free-piston Stirling engine that prevent the above-described runaway increase of piston amplitude and engine power despite the occurrence of a malfunction of the type described above.
The invention is a modification of prior art free-piston Stirling engines that causes piston amplitude to be limited and engine power to be reduced as the piston amplitude increases beyond the maximum power that the engine's designer selected when designing the engine. The power output is reduced by reducing the displacer phase with respect to the piston and is further reduced to essentially zero by increasing pumping losses through the engine's gas bearing system.
A first feature of the invention is that the inward edge of the heat rejecter cylinder port is located outward of the most inward excursion of the inward end of the piston sidewall during a part of the reciprocation cycle of the piston. Preferably, the inward edge of the heat rejecter cylinder port is located outward of the most inward excursion of the inward end of the piston sidewall when the engine is operating at a selected maximum engine power for which the engine was designed so that the heat rejecter cylinder port is entirely covered by the piston sidewall during an inward portion of the piston reciprocation when the engine is operating at the selected maximum engine power.
A second feature of the invention is the addition of a leaker port that extends from the gas bearing cavity and through the piston sidewall. The leaker port is positioned axially outward from the gas bearing pads of the engine's gas bearing system. The leaker port is covered by the cylinder when the amplitude of piston reciprocation is equal to or less than the piston's amplitude of reciprocation at maximum engine power and becomes uncovered and in fluid communication with the back space at a piston amplitude of reciprocation that exceeds the piston's amplitude of reciprocation at maximum engine power.
A third feature of the invention is a resilient bumper that is attached to the outward end of the piston or to the inward side of the displacer spring so it is located between the piston and the mechanical spring that is connected to the displacer connecting rod.
With the invention, if the engine load is reduced so that more power is produced by the engine than is consumed by the sum of the power delivered to the load plus the power consumed to drive the engine, then engine power is reduced and piston amplitude is limited as piston amplitude further increases beyond the piston amplitude at the designed maximum engine power.
In describing the preferred embodiment of the invention which is illustrated in the drawings, specific terminology will be resorted to for the sake of clarity. However, it is not intended that the invention be limited to the specific term so selected and it is to be understood that each specific term includes all technical equivalents which operate in a similar manner to accomplish a similar purpose.
Provisional patent application Ser. No. 62/410,987, filed Oct. 21, 2016 is incorporated in this application by reference.
Covering & Blocking the Heat Rejecter Cylinder Port
The first improvement of the invention is the positioning and location of the heat rejecter cylinder port 20. Unlike the prior art, the heat rejecter cylinder port 20 is positioned where it is covered and blocked by the piston sidewall 32 during a peak part of the piston's inward excursion when the engine power approaches near its maximum designed engine power. Stated another way, the heat rejecter cylinder port 20 is positioned so that, when the piston amplitude of reciprocation is near its amplitude at the engine's peak power, the heat rejecter cylinder port 20 becomes completely covered by the piston sidewall 32 and therefore the passage of gas through the heat rejecter cylinder port 20 becomes blocked. The result of this blockage is that the power curve (
Looking at this first improvement of the invention in more detail, the location of the heat rejecter cylinder port 20 is seen with reference to
Preferably, the inward edge 58 of the heat rejecter cylinder port 20 is located outward of the most inward excursion of the inward end 34 of the piston sidewall 32 when the engine is operating at a selected maximum engine power for which the engine was designed. That position assures that the heat rejecter cylinder port 20 is entirely covered by the piston sidewall 32 during an inward portion of the piston reciprocation when the engine is operating at its selected maximum engine power. I believe that, more preferably, the inward edge 58 of the heat rejecter cylinder port 20 should be located outward of the most inward excursion of the inward end 34 of the piston sidewall 32 by a distance that is within the range of 3% to 10% of the piston amplitude at maximum engine power. For example,
Covering the heat rejecter cylinder port 20 by the piston sidewall 32 during an inward excursion of the piston 28 traps working gas between the outward end of the displacer 30 and inward end of the piston 28. The trapped gas acts as a gas spring between the displacer 30 and the piston 28 because no significant quantity of gas can escape from the volume of space between the piston and displacer. The gas spring applies a relative force between the displacer and piston. When the piston just completes covering the port (i.e. still moving in but nearing the end of its inward excursion), the displacer is moving out. So the piston and displacer are moving closer together in opposite directions of motion. When the piston covers the port and makes the trapped working gas become an effective gas spring, that gas spring is pushing against the outward motion of the displacer which retards the displacer motion and therefore reduces the displacer phase lead ahead of the piston.
There is another effect from covering the heat rejecter cylinder port 20 by the piston sidewall 32 when the piston amplitude of reciprocation is sufficiently large. When the piston amplitude of reciprocation is less than an amplitude that is sufficient to cover the heat rejecter cylinder port 20, the mean position of the piston is maintained by the centering system described above. In that lower range of piston amplitude, the engine is running in the conventional prior art manner so that the mean piston position moves in slightly and increases in piston amplitude result in piston excursions that increase nearly equally in both the in direction and the out direction. However, when the rejecter cylinder port becomes covered and blocked during a part of each cycle, the above-described trapping of gas and the resulting creation of a gas spring applying opposite forces against the displacer and piston has an additional effect on the engine operation. The effect is that most of the further increase in piston amplitude occurs at the outward excursion of the piston and the mean piston position moves out.
The reason is as follows. The force applied by the gas spring against the piston exists only when the heat rejecter cylinder port 20 is blocked. That force against the piston is in the outward direction because the displacer is moving out while the heat rejecter cylinder port 20 is blocked. This outward force on the piston causes the mean position of the piston to move outward. The mean piston position moves progressively further away from the working space as the piston amplitude increases. As a result of this outward creep of the mean piston position, as the piston amplitude increases, a greater proportion of the increased amplitude of reciprocation appears as increased excursions in the outward direction than appears as increased excursions in the inward direction. This effect is illustrated in
Losses Pumping Gas Through Gas Bearing Cavity
Although the above-described positioning of the heat rejecter cylinder port can be used alone to improve the stability of an FPSE for significantly reduced loads, it reduces the engine power only by at least one third and possibly as much as three fourths from the maximum power. For example, it reduces the engine power at least to approximately the point IV on the modified power curve of
However, this phase lead reduction does not reduce the engine power to zero as the piston amplitude increases still further and beyond (i.e. below) point IV in
In order to reduce the engine power to zero for further increases in piston amplitude beyond point IV, a port is provided in the piston that I have called a “leaker port” 64. As with the other ports, the leaker port 64 can be formed from multiple leaker ports spaced annularly around the piston. The leaker port 64 extends through the piston sidewall 32 and into the gas bearing cavity 44. Preferably the leaker port 64 includes diametrically opposite leaker ports 64 in order to balance side loads on the piston 28.
In summary, the leaker port 64 vents the gas bearing cavity 44 to the back space 43 during sufficiently distant outward piston excursions. This periodic venting causes the additional power, which results from further increases in piston amplitude, to be consumed by pumping losses from pumping working gas around a closed loop and also reduces the amount of power increase as a function of increased stroke by lowering the mean working space pressure as well as operating frequency.
Referring to
During each cycle of engine operation, the gas bearing cavity 44 is charged to peak workspace pressure through its gas bearing inlet passage 46. Consequently, there is a substantial pressure differential between the gas pressure in the gas bearing cavity 44 and the gas pressure in the back space 43. The gas bearing cavity 44 supplies gas out through the gas bearing pads 50 for lubrication purposes as described above. The leaker port 64 is located so that it is typically blocked by the cylinder in normal operation at and below maximum engine power. However, with a piston amplitude increase at least beyond the amplitude at maximum engine power, gas is leaked from the bearing cavity 44 through the leaker port 64 to the back space 43 during a part of each cycle when the piston 28 is at an outer part of its outward excursion. Whenever the leaker port 64 is uncovered, working gas flows directly out of the gas bearing cavity 44 into the back space 43. The substantial pressure differential results in a significant gas flow, during each cycle, from the gas bearing cavity 44 to the back space 43 when the leaker port 46 is not covered by the cylinder 22. As the piston amplitude progressively increases further, the leaker port 46 is uncovered for a longer time so more and more gas is leaked out of the gas bearing cavity 44. During a part of each inward motion of the piston 28 makeup gas is pumped into the gas bearing cavity 44 via the gas bearing inlet passage 46 to recharge the gas bearing cavity 44 to peak cycle workspace pressure.
Referring to
During each cycle of operation that the leaker port 64 becomes uncovered, the quantity of gas that recharges the gas bearing cavity 44 and flows to the back space 43 is considerably greater than the quantity of gas that recharges the gas bearing cavity 44 and flows to the back space 43 during cycles that the leaker port 64 does not become uncovered. Under the latter condition, the only gas flow out of the gas bearing cavity 44 is to supply the gas bearing pads 50. However, during each cycle of operation that the leaker port 64 becomes uncovered, the working gas flow from the working space 8 through the gas bearing cavity 44 and out the leaker port into the back space 43 is large enough that it substantially lowers the mean working space pressure and slightly increases the back space pressure. These pressure changes, which result from opening the leaker port to the back space, cause the averaged back space pressure to be greater than the averaged working space pressure at the times when the centering system passageways come into registration. Therefore, when the centering system passageways come into registration, gas flows out of the back space 43, through the centering system and is returned to the working space 8.
Consequently, when piston amplitude is large enough that the leaker port 64 is being uncovered during a part of each outward reciprocation of the piston, gas is being pumped around a loop. The loop consists of gas pumped by the engine from the working space 8 through the gas bearing cavity 44 and out the leaker port 64 into the back space 43 and gas pumped back in the opposite direction from the back space 43 through the centering system to the working space 8. Pumping the working gas around this loop causes pumping losses. The pumping losses consume energy (work is being done to transport the gas through the passages and their restrictions) thereby reducing engine power because some of the engine power is consumed by the pumping losses. As piston amplitude increases, the leaker port 64 is vented to the back space 43 for a greater angular interval of each cycle. That allows more gas venting which in turn causes more pumping loss until the engine power eventually goes to zero at point V on the modified power curve.
In addition to the pumping losses, the reduction of work space mean pressure (because working gas is flowing from the working space 8 through the gas bearing cavity 44 and out the leaker port into the back space 43) also reduces engine power. The reduced mass of gas in the working space means that the amplitude of gas pressure variations in the working space is reduced so the power to drive the piston and displacer is reduced.
In the description of the first feature of the invention, which allows the heat rejecter cylinder port to be covered, it was explained how the mean piston position moves outward and piston excursions in the outward direction increase more than piston excursions in the inward direction. It can now be seen that moving the piston's mean position in the out direction and increasing the piston's excursions in the out direction also increases the angular interval during each cycle that the leaker port 64 is uncovered. The increased angular interval means that more gas is leaked from the gas bearing cavity 44 to the back space 43 which means that more power is consumed by pumping losses. Increasing the angular interval that the leaker port 64 is uncovered also causes additional lowering of workspace mean pressure and therefore further lowers the power produced by the engine. A still further power reduction also occurs because, with the lowered mean workspace pressure, the frequency decreases in the loss of load case and this also reduces the power produced.
Referring to
As stated above, when piston amplitude increases beyond the piston amplitude at maximum power, about one third to one half of the power reduction from the present invention is the result of covering and blocking the rejecter cylinder port with the piston sidewall. Therefore, the leaker port 64 should start to be uncovered at a piston amplitude that is about one third of the way down (point IV) on the modified power curve. As piston amplitude increases further, power reduction, from pumping losses and from lowering the workspace pressure, increases until its maximum reduction when the engine power goes to zero at piston amplitude DD at point V.
It is not necessary that the leaker port 64 be vented to the back space 43 by moving below the end of the cylinder 22. Alternatively, the cylinder can extend further to cylinder extension 22A, as illustrated in dashed lines in
Also it is not necessary that the leaker port have a particular configuration. It is, of course, desirable to maintain lubrication of the piston sidewall 32. So a designer would want to maintain a number and placement of the gas bearing pads that provide appropriate lubrication according to prior art engine design principles. If a bearing pad, which is constructed to provide adequate lubrication according to those principles, moves immediately beyond the end of the cylinder 22 or otherwise opens a gas passage between the gas bearing cavity 44 and the back space 43, its lubrication function is lost. In fact, as a bearing pad approaches close to the end of the cylinder, its lubrication function is somewhat degraded. For that reason it is undesirable to have a gas bearing pad, which is used to provide lubrication, move beyond the end of the cylinder 22. However, it is not necessary that a leaker port be a simple cylindrical hole. A leaker port can have other shapes that provide a gas passage extending through the piston sidewall 32 and into the gas bearing cavity 44. Among the other possible configurations of the leaker port is the configuration of a gas bearing pad. In other words, a leaker port can be provided that is made to look like a gas bearing pad but is included to function as a leaker port.
Bumper.
A third feature of the invention that improves loss of load operation is to include a bumper 70 that limits the relative inward motion of the displacer with respect to the piston. The bumper limits displacer relative motion (motion relative to the piston) by striking the planar spring 40 and thereby pushing the displacer connecting rod, and therefore the displacer, in the out direction. The bumper 70 is a soft or resilient material that is attached to the outward end of the piston or the inward side (70B) of the planar spring 40 and cushions and dampens any collision between the piston and the planar spring 40 that is fixed to the end of the displacer connecting rod 38. Any contact of the bumper 70 with the planar spring 40 would be relatively soft or glancing in nature because the displacer phase angle has been reduced greatly by power limiting effects of one or both of the first two above-described features of the invention. The bumper is intended to contact the displacer spring and thereby limit displacer motion relative to the piston. This limit of relative motion of the displacer away from the piston also helps to reduce power growth during overstroke. Typically this bumper is not needed but in certain arrangements, such as operation with a tuned vibration absorber, it provides added protection. As described above, the closure of the heat rejecter cylinder port by the piston sidewall effectively limits the relative motion of displacer toward the piston. The bumper has the same effect in the opposite direction making it desirable for use with the tuned vibration absorber. With the absorber, casing motion will increase when the engine frequency changes. This in turn inputs energy to drive the displacer to a larger amplitude.
Those skilled in the art are capable of designing a free-piston Stirling engine to have a selected amplitude under the operating conditions of their choice. Of course engineering design is not perfected to the extent that a prototype always operates exactly according to its design parameters. So persons skilled in the art can build a prototype engine, test it and then modify its design to obtain the design parameters they want. Repetition of the design, build, test and modify procedure is a common iterative process that eventually leads to a desired operation.
One way to design an engine using one or more of the features of the invention is to begin with a graph of an engine's typical power curve known in the prior art. The designer would then estimate, on the same graph, what the modified power curve created by the invention would be for a particular engine design and its chosen parameters. Engine amplitude at zero power on the modified power curve is the allowed amount of piston amplitude. The designer can estimate or choose the piston amplitude CC at the peak of the modified power curve (
The leaker port location is then chosen so that the leaker port opens to the back space at a piston amplitude on the down side of the estimated power curve, preferably at least one third of the way down. A prototype is then constructed and tested and the power curve for the prototype can be generated. From that design modifications are made, such as relocation of the centering system, the rejecter port and/or the leaker port.
Displacer Gas Cushion.
The cushion cylinder 74 has a larger diameter than the end portion of the displacer 30 in order to provide a clearance gap between the displacer 30 and the cushion cylinder 74 that is sufficient to prevent the displacer 30 from striking or rubbing the cushion cylinder 74 and for permitting gas flow blow-by to provide pumping losses for damping displacer motion. The displacer 30 typically has a Heylandt crown (hot cap) which is smaller than the engine cylinder 22 to provide the clearance gap. However, the cushion cylinder 74 diameter should be larger than the displacer diameter if the displacer diameter is equal to the diameter of the engine cylinder 22. Preferably, the axial length 78 of the cushion cylinder 74 from the head end 26 of the expansion space 10 to the heat accepter cylinder port 24 is in the range of 5% to 10% of the displacer stroke.
The cushion cylinder 82 also has a larger diameter than the end portion of the displacer 30 in order to provide a clearance gap between the displacer 30 and the cushion cylinder 82 that is sufficient to prevent the displacer 30 from striking or rubbing the cushion cylinder 82 and for permitting gas flow blow-by to provide pumping losses for damping displacer motion. As with the embodiment of
The displacer gas cushion operates to close off a space, which is a portion of the expansion space 10, by blocking the heat acceptor cylinder ports 24. The space within the cushion cylinder 76 or 82 is sufficiently sealed so that, when the heat acceptor cylinder ports 24 are blocked by the displacer, the space within the cushion cylinder 76 or 82 functions as a gas spring. For example, the cushion cylinder 82 does not have to be perfectly or completely sealed against the head end 26 and a small amount of leakage could be desirable to provide additional pumping losses. The cushion cylinder 76 or 82 needs only to be sufficiently sealed so that, when the end 88 of the displacer 30 covers the heat acceptor cylinder ports 24 and reciprocates into the cushion cylinder 74 or 82, as shown in phantom as displacer end 88A, the working gas within the cushion cylinder 76 or 82 is compressed and applies a retarding force against the displacer end 88. The retarding force is a combination of a damping component and a spring component, although primarily spring component. The retarding force prevents the displacer end 88 from colliding with the head end 26. This also has a limiting effect on the displacer stroke and therefore has a limiting effect on the mass of working gas that is periodically shuttled back and forth through the regenerator. The consequent result is that the displacer cushion also has some limiting effect on the piston stroke.
Previously explained is the manner in which covering and blocking the heat rejecter cylinder port and causing losses pumping gas through gas bearing cavity are used to limit engine power and prevent the instability of and the runaway increase in piston and displacer amplitude of reciprocation. The displacer gas cushion can further assist in that purpose. Therefore, the displacer gas cushion is desirably used with either or both embodiments of those previously explained concepts. The displacer gas cushion can also be used alone, especially where it is desired to prevent the displacer from colliding with the engine head or dome.
Although a displacer can have a uniform diameter along its entire axial length, as seen in
This detailed description in connection with the drawings is intended principally as a description of the presently preferred embodiments of the invention, and is not intended to represent the only form in which the present invention may be constructed or utilized. The description sets forth the designs, functions, means, and methods of implementing the invention in connection with the illustrated embodiments. It is to be understood, however, that the same or equivalent functions and features may be accomplished by different embodiments that are also intended to be encompassed within the spirit and scope of the invention and that various modifications may be adopted without departing from the invention or scope of the following claims.
This application claims the benefit of U.S. Provisional Application No. 62/410,987 filed Oct. 21, 2016.
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