Patent Application
20140318363
A. Technical Field
This invention generally relates to pressure activated engines and compressors. More particularly, this invention is a reciprocating-piston engine having a harmonic oscillator valve controlling the admission of a pressurized expansible fluid into an expansion chamber and an outlet controlled by the motion of the piston allowing the exhaust of working fluid from the expansion chamber. In some embodiments the present invention can also operate as a compressor.
B. Description of the Related Art
Engines that transform the internal energy within a high-pressure expansible fluid into useful mechanical energy, such as steam engines, are well known. Among reciprocating steam engines, the form having the greatest economy and greatest efficiency is the uniflow steam engine. In “Steam-Engine Principles and Practice”, published in 1922, a uniflow engine is disclosed operating with 461 psia superheated steam at a temperature of 1018 OF that produced an indicated efficiency of only 5.67 pounds of steam per indicated horsepower hour, or 37% thermal efficiency. In the uniflow steam engine, the valves controlling the admission of high-pressure supply steam to the cylinder are at an end of the cylinder, while all of the exhaust of expanded, low pressure steam, preferably at sub-atmospheric pressure for highest efficiency, is from vent ports placed around the circumference of the cylinder, and located nearly a full stroke distance away from the inlet valves. Double acting uniflow engines have inlet valves at both ends of the cylinder and vent ports at the middle of the cylinder, while single acting uniflow engines have inlet valves at one end of the cylinder, near the Top Dead Center position, TDC, of the piston with vent ports placed near the Bottom Dead Center position, BDC. The thermodynamic reason for the superior efficiency of the uniflow steam engine design, as opposed to the counter flow design, is that the detrimental phenomenon of hotter steam condensing on relatively colder cylinder walls, thus dropping the pressure on the power stroke, or colder steam vaporizing on relatively hotter cylinder walls, thus increasing the pressure on the recovery stroke, is greatly reduced.
There have been a number of technical challenges in the art of uniflow engines, such as the need to maintain a significant minimum cylinder clearance space, in order to avoid damage produced from overly recompressing steam as the piston approaches TDC. On the other hand, on the exhaust stroke, without recompressing steam all the way to the pressure of the incoming steam, and with a significant minimum cylinder clearance space, the resulting highly non isentropic, rapid inrush of high pressure steam at the time the inlet valve opens and the clearance space fills is detrimental to achieving high thermodynamic efficiency. This rapid inrush also leads to the problem with conventional steam engine inlet valves of “wire drawing”, which occurs when the high velocity flow of steam erodes or scores a pathway in the seating material that remains after the valve is closed, and can cause leakage. A similar “wire drawing” effect also happens as valves are closed, if they do not close quickly and a high velocity flow of steam is allowed to persist overly long. Inlet valves capable of rapid action, in order to enable high expansion ratios at high speed, and to avoid “wire drawing” problems, have been a challenge. Another historical challenge has been the choice of lubricant for the piston and valves, as common engine oils tend to degrade at high temperature.
One aspect of the present invention includes a harmonic uniflow engine comprising: a cylinder having a cylinder axis; a piston head reciprocable in the cylinder and together enclosing an expansion chamber, wherein the cylinder has an inlet at an inlet end fluidically connected to the expansion chamber and an outlet at a removed location from the inlet end; an intake header in fluidic communication with the inlet for channeling working fluid from a pressurized fluid source into the expansion chamber; an inlet valve for controlling the flow of working fluid from the intake header into the expansion chamber to effect a power stroke of the engine, said inlet valve comprising an inlet valve head and a resiliently biasing member arranged together as a harmonic oscillator so that the inlet valve head is moveable against an equilibrium restoring force of the resiliently biasing member from an unbiased equilibrium position located in the intake header to a biased closed position occluding the inlet, and so that upon releasing from the closed position the inlet valve head undergoes a single oscillation past the equilibrium position to an oppositely biased maximum open position and returns to a biased return position between the closed and equilibrium positions to choke the flow of working fluid and produce a pressure drop across the inlet valve causing the inlet valve to close, wherein the piston head is reciprocable to a venting position which fluidically connects the expansion chamber to the outlet for controlling the periodic venting of working fluid out from the expansion chamber, and periodic return means operably connected to the piston head to effect a return stroke of the engine after each power stroke.
Another aspect of the present invention includes a uniflow energy conversion system comprising: a cylinder having a cylinder axis; a piston head reciprocable in the cylinder and together enclosing a chamber, wherein the cylinder has a first port at a first end fluidically connected to the chamber and a second port at a second end opposite the first end; a valve for controlling the flow of working fluid between the first port and the chamber, said valve comprising a valve head and a resiliently biasing member arranged together so that the valve head is moveable against an equilibrium restoring force of the resiliently biasing member from an unbiased equilibrium position to a biased closed position occluding the first port; and periodic means operably connected to the piston head to effect at least one of two reciprocation strokes thereof, wherein the piston head is adapted to rotate about a pivot axis as it reciprocates in the cylinder so: that during one of the two reciprocation strokes the piston head maintains a seal with the cylinder to inhibit blow-by past the piston head; and during the other one of the two reciprocation strokes a pair of pivotable ends of the piston head on opposite sides of the pivot axis are radially displaced away from the cylinder so as to form blow-by channels between the pivotable ends and the cylinder which fluidically connect the chamber to the second port, for controlling the periodic flow of working fluid between the chamber and the second port.
Another aspect of the present invention includes a uniflow engine comprising: a cylinder having a cylinder axis; a piston head reciprocable in the cylinder and together enclosing an expansion chamber, wherein the cylinder has an inlet at an inlet end fluidically connected to the expansion chamber and an outlet at an outlet end opposite the inlet end; an intake header in fluidic communication with the inlet for channeling working fluid from a pressurized fluid source into the expansion chamber; an inlet valve for controlling the flow of working fluid from the intake header into the expansion chamber to effect a power stroke of the engine, said inlet valve comprising an inlet valve head and a resiliently biasing member arranged together as a harmonic oscillator so that the inlet valve head is moveable against an equilibrium restoring force of the resiliently biasing member from an unbiased equilibrium position located in the intake header to a biased closed position occluding the inlet, and so that upon releasing from the closed position the inlet valve head undergoes a single oscillation past the equilibrium position to an oppositely biased maximum open position and returns to a biased return position between the closed and equilibrium positions to choke the flow of working fluid and produce a pressure drop across the inlet valve causing the inlet valve to close; and periodic return means operably connected to the piston head to effect a return stroke of the engine after each power stroke, wherein the piston head is adapted to rotate about a pivot axis as it reciprocates in the cylinder so that: during the power stroke the piston head maintains a seal with the cylinder to inhibit blow-by past the piston head; and during the return stroke a pair of pivotable ends of the piston head on opposite sides of the pivot axis are radially displaced away from the cylinder so as to form venting channels between the pivotable ends and the cylinder which fluidically connect the expansion chamber to the outlet, for controlling the periodic venting of working fluid out from the expansion chamber.
And another aspect of the present invention includes a uniflow compressor comprising: a cylinder having a cylinder axis; a piston head reciprocable in the cylinder and together enclosing a compression chamber, wherein the cylinder has an outlet at an outlet end fluidically connected to the compression chamber and an inlet at an inlet end opposite the outlet end; an outlet header in fluidic communication with the outlet for channeling working fluid to a pressurized fluid reservoir from the compression chamber; an outlet valve for controlling the flow of working fluid from the compression chamber out through the outlet in a delivery stroke of the compressor, said outlet valve comprising a valve head and a resiliently biasing member arranged together so that the valve head is moveable against an equilibrium restoring force of the resiliently biasing member from an unbiased equilibrium position to a biased closed position occluding the outlet; and periodic means operably connected to the piston head to effect the delivery stroke and a reciprocal intake stroke after each delivery stroke, wherein the piston head is adapted to rotate about a pivot axis as it reciprocates in the cylinder so that: during the delivery stroke the piston head maintains a seal with the cylinder to inhibit blow-by past the piston head; and during the intake stroke a pair of pivotable ends of the piston head on opposite sides of the pivot axis are radially displaced away from the cylinder so as to form blow-by channels between the pivotable ends and the cylinder which fluidically connect the compression chamber to the inlet, for controlling the periodic replenishment of working fluid to the compression chamber.
Generally, the present invention is directed to a harmonic uniflow engine having a self-acting harmonic inlet valve capable of automatically relieving excess pressure as the piston approaches TDC, and capable of very rapid action without the need for oil lubrication. The use of the harmonic inlet valve in the uniflow engine enables inlet valve opening without the need for high-speed mechanical collision or contact by the piston. This feature, in the context of the uniflow engine, enables the minimum clearance space to be reduced to virtually nil, without the possibility of over pressure damage to the engine. Furthermore, there is a thermodynamic advantage as well, in that the recompression of working fluid is automatically limited to only the minimum pressure needed to open the inlet valve, and no more. As a result, the thermodynamic efficiency limit can be closely approached with the use of the harmonic inlet valve in a uniflow steam engine. Furthermore, the reciprocating-piston structure and resiliently biasing valve of certain embodiments of the harmonic uniflow engine may also be operated as a compressor, and therefore may be characterized generally as a uniflow energy conversion system.
The accompanying drawings, which are incorporated into and form a part of the disclosure, are as follows:
Generally, the present invention is an engine that converts the energy contained within a pressurized supply of a working fluid, such as steam or compressed air, into mechanical power, and is well suited for connection to an alternating current electrical generator. The engine generally comprises a reciprocating-piston expander assembly and a crank assembly or other periodic return mechanism or method operably connected to the piston for effecting the return stroke of the expander after each power stroke. The expander generally includes the following components and sub-assemblies: a harmonically oscillating inlet valve for controlling flow of high pressure working fluid into expansion chamber from an inlet header conduit, manifold or duct (hereinafter “intake header”) that is connectable to a source of pressurized working fluid; a resiliently biasing outlet valve for controlling flow out of expansion chamber to an exhaust header conduit, manifold or duct (hereinafter “exhaust header”) capable of venting the expanded, low pressure working fluid. In particular, the inlet valve includes an inlet valve head and a resiliently biasing member arranged together as a harmonic oscillator so that when the inlet valve head is displaced from a closed position (occluding the inlet to an expansion chamber) it undergoes a single oscillation to a maximum open position and returning to a return position where it chokes the flow of working fluid so as to close the inlet once again in a single two-stroke period of the engine. Because of this harmonic oscillation aspect of the inlet valve, the engine is characterized as a “harmonic engine.” And the crank assembly (for example of a type conventionally known in the art) is operably connected to the piston for converting reciprocating motion into rotary power output. For example the crank assembly may include a flywheel having rotational inertia that is transferred to the piston via the crankshaft.
Turning now to the drawings,
The inlet valve 101 in the first exemplary configuration is shown in
The inlet valve head 103 and its opening spring 107 form a spring-mass system of a harmonic oscillator which, when the inlet valve is displaced from its equilibrium position, experiences a restoring force proportional to the displacement according to Hooke's law, as known in the art. This oscillator preferably has a high quality factor Q value, so that, while freely oscillating, many cycles of oscillation occur before the amplitude of oscillation decays significantly. The significance of the high Q value in the context of this invention is that after a single oscillation, starting from a closed position, in the absence of other forces, the inlet valve returns almost all the way back to its closed position. In practice a Q value of at least 160 is preferred, as this returns the inlet valve to within 1% (relative to the full excursion of the valve) of its closed position after a single oscillation. With such close return to the closed position, the flow passageway from inlet duct 125 to expansion space 162 effectively forms a converging-diverging nozzle. With a sufficiently high Q, the narrowness of the throat of the converging-diverging nozzle section has the practical effect of choking the flow of working fluid between the inlet duct 125 and the expansion space 162. As is known in the art of converging-diverging nozzles, flow is choked by the limitation that the flow speed cannot exceed the speed of sound at the throat of the nozzle. With sufficiently high Q, and thus a sufficiently small throat area, even at the lowest practical engine operating speed, the flow at the throat reaches the speed of sound and is thus choked. The form of inlet valve shown is conducive to attaining very high Q values, as the frictional losses of flexure bearings, such as 107, constructed of high quality spring steel, are very low. Thus, in an example embodiment, the resiliently biasing member of the inlet valve has a high quality Q factor greater than about 160 so that the return position of the inlet valve head after undergoing the single oscillation is substantially near the closed position.
The completely relaxed neutral position of the inlet valve 101 is shown in
Thus the inlet valve is used for controlling the flow of working fluid from the intake header through the inlet to effect a power stroke of the expander. The inlet valve head and the resiliently biasing member of the inlet valve are arranged together as a harmonic oscillator so that the inlet valve head is moveable against an equilibrium restoring force of the resiliently biasing member from an unbiased equilibrium position located in the intake header to a biased closed position occluding the inlet. Furthermore, this arrangement enables the inlet valve head, upon being released from the closed position to undergo a single oscillation past the equilibrium position to an oppositely biased maximum open position and return to a biased return position between the closed and equilibrium positions. This chokes the flow of working fluid and produces a pressure drop across the inlet valve causing the inlet valve to close. Furthermore, the inlet valve head may be configured to protrude in part into the expansion chamber when in the closed position so as to enable the piston to bump open the inlet valve from the closed position and initiate the single oscillation of the inlet valve head.
The outlet valve 104 is also shown in
Outlet valve closing spring 108 is constructed so that it extends farther into cylinder 161 than does the bottom of the inlet valve 109 when both the inlet and outlet valves are fully closed as shown in
Outlet valve 104 is open (
In this embodiment, outlet valve 104 penetrates outlet valve guide 112, and a support 113 for outlet opening spring 106 is positioned above the valve. The external location of support 113 permits modification of the strength of the restoring force produced by spring 104 in its fully compressed position even while the engine is in operation by adjustment of the position of support 113. The close fit of valve guide 112 suppresses leakage of working fluid to the outside of the engine. If needed, an optional valve stem seal (not shown) could be added.
Generally therefore the outlet valve operates to control the flow of working fluid exhausted out through the outlet to the exhaust header during a return stroke of the expander. To accomplish this, the outlet valve head, stopper, and the resiliently biasing member of the outlet valve are arranged so that the outlet valve head is moveable against an equilibrium restoring force of the resiliently biasing member from a maximum open position located in the expansion chamber and delimited by the stopper to a biased closed position occluding the outlet. And the outlet valve closing spring (which is carried by one of the outlet valve head and the piston head) is positioned between the outlet valve head and the piston head so that when the other non-carrying one of the outlet valve head and the piston head comes in contact with and resiliently biases the outlet valve closing spring, the outlet valve is moved by the outlet valve closing spring from the maximum open position to the closed position ahead of the bump opening of the inlet valve.
Motor-generator 188 is shown operably connected to the crankshaft and is preferably a squirrel cage induction motor compatible with the 60 Hz alternating current power in the United States. As is well known in the art, under low load conditions, such as when starting up, when connected directly to flywheel 185 as shown in
Operation of the preferred engine embodiment is now described for normal, steady running conditions. The variation in the positions of the inlet valve, the outlet valve and the piston are shown with solid lines in a timing diagram in
Starting the cycle arbitrarily at the TDC position, the configuration of the components and the state of their motion is shown in
As piston 160 initially descends from TDC and the outlet closing spring 108 extends to its fully relaxed position, the outlet valve is held closed by the pressure difference between the working fluid in expansion chamber 162 and outlet header duct 105. At the same time, the inlet valve undergoes a single oscillation, passing upwards through the neutral position of spring 107 (as seen in
The state of motion of the components in
After the inlet valve closes, at the phase indicated by arrow 140 in
As the piston reaches BDC, the outlet valve is in its initial stage of opening, and is moving downwards, towards the piston. Just after BDC, the piston is moving upwards, and the working fluid within the expansion chamber is forced out around the outlet valve. Near BDC, the piston speed is sufficiently small that the aerodynamic force of the outrushing working fluid produces only an insignificant fraction of the force produced by the outlet valve opening spring, and the outlet valve continues to open. The outlet valve is quickly brought to its fully open position 115, as determined by the location of stopper 110, at the point in the timing diagram indicated by arrow 146, and then remains there for most of the recovery stroke, as shown in
At the phase indicated in
During the portion of the cycle between the phases indicated by arrows 141 and 143, with the outlet valve closed and stationary and the inlet valve not yet open and also stationary, the working fluid is getting compressed, and its pressure increases due to the upward motion of the piston. With the proper choice of spring strengths, the working fluid pressure is preferably approximately equal to the full value of the pressure in the inlet header at the time that the bottom of the inlet valve 109 makes contact with piston 160. Once the inlet valve is forced open, however, as the remaining volume within the expansion chamber is minimal, whatever the pressure in the expansion chamber immediately prior to the opening of the inlet valve, the pressure in the expansion chamber very rapidly equalizes with the pressure of the supply.
If the working fluid pressure has not risen to match the supply pressure, the physical contact of piston 160 against the bottom of the inlet valve 109 provides sufficient impulse to force the opening of the inlet valve. On the other hand, if under off-nominal circumstances, the cylinder pressure has increased to well above the supply pressure, then the pressure force on the inlet valve, together with the inlet valve spring 107 act together to open the inlet valve and relieve the excess pressure. Because of this, the inlet valve acts as a safety valve, and this engine is quite tolerant of off-nominal conditions.
Under nominal, full power, steady operation, with the pressure in the expansion chamber nearly matching the supply pressure, the impact of piston 160 against the bottom of the inlet valve 109 is very mild or even non-existent in the case that the vanishing pressure drop allows inlet valve spring 107 to open the inlet valve prior to piston 160 making contact with the inlet valve. In any case, under steady running conditions, the state of all components at 360° of phase angle is identical to that described above for 0° of phase angle, and the engine cycle repeats.
With the application of high pressure working fluid to the inlet header manifold, and with a design choice that the outlet valve opening spring is stronger than the inlet valve opening spring, the aerodynamic force of working fluid flowing first past the inlet valve, then into the expansion chamber, and finally out past the outlet valve, the inlet valve is forced closed before the outlet valve has a chance to close. This aerodynamic force is much greater than the choked flow force that develops under normal running conditions just before the phase point indicated by arrow 140 in
As a result, the expansion chamber remains at the pressure of the exhaust header manifold, and there is no significant load on the piston. Because of this, with the induction motor/generator subsequently connected to a source of AC electrical power, it can rapidly come up to its unloaded rotational speed. As the piston encounters the open outlet valve on its first upstroke at less than full speed, outlet valve closing spring 108 assures that valve 104 will be closed prior to piston 160 making contact with inlet valve 101, and as a result the pressure within expansion chamber 162 will be brought to its nominal value under full speed conditions, and inlet valve 101 is forced to open. The high Q of the harmonic oscillator inlet valve assembly assures that inlet valve 101 returns very nearly to its fully closed position after the inlet valve undergoes a single cycle of oscillation. Because of the narrow opening after a single oscillation of the inlet valve, even the slower speed (at startup) descent of piston 160 suffices to produce a dynamic latching of inlet valve 101 in its closed position by virtue of the choked flow of the working fluid through the converging-diverging nozzle formed between the frusto-spherical surface of the inlet valve head 103 and the conical surface of the inlet valve seat 102. As a result, the pressure is assured to decrease to that of the outlet manifold, and outlet valve 104 is assured to open by the process described in the following paragraph. Thus after such a cycle, the rotational speed of flywheel 185 increases, until after one or more (depending on the moment of inertia of the flywheel) such startup cycles, the flywheel accelerates to its normal operating speed and the pressure and flow conditions are those of full running power conditions, and normal operational cycles begin. As the engine produces power, it overdrives motor 188, and instead generates electrical current that is forced to be in phase with the electric grid current by the nature of induction motors. With a sufficiently high moment of inertia flywheel, the angular velocity of the flywheel becomes almost constant, and the alternating current power generated is almost perfectly steady.
The operation of the harmonic engine under conditions that the supply pressure is less than the nominal full power design pressure is shown in
As a result, the outlet valve is open for a longer time during low-pressure operation and the inlet valve opens slightly earlier but stays open for approximately its normal duration. The ramification of operation at lower pressure is simply that the power output is less for a given speed of operation while the relative efficiency of operation is maintained. Lowering the supply pressure thus provides a convenient means to adjust to a lower power load requirement.
A very similar process is found during the startup of the harmonic engine described above, in that, at low rotational speed, the open period of the inlet valve, which is approximately a constant time interval, spans a shorter range of crankshaft phase angle, and thus the cylinder pressure drops to that of the outlet manifold earlier in phase, and the outlet valve is sprung open at an earlier phase angle as well. The outlet valve remains open for a longer span of crankshaft phase angle, but is closed by spring 108 at the normal phase by virtue of contact with piston 160 on its rise towards TDC.
It is by virtue of these processes that the valve timing is variable and self-adjusts to accommodate a wide range of supply pressure conditions in a nearly optimal way relative to what is thermodynamically possible.
In a variation of the first embodiment in which the crankshaft is not connected to a motor generator, but is instead used to supply rotational mechanical power to a load, the operational frequency is not held fixed by the induction motor/generator. In such applications, adjustment of the position of outlet valve opening spring support 113 allows adjustment of the crankshaft phase angle at which outlet valve 104 closes. Various devices or methods known in the art may be employed for adjusting the equilibrium restoring force exerted by the resiliently biasing member of the outlet valve so as to adjust a crankshaft phase angle at which the outlet valve closes. Specifically, with support 113 lowered, the compression of spring 106 is increased, and the closing phase for outlet valve 104 is delayed. Conversely, with support 113 raised, the closing phase for the outlet valve is advanced. Change in the phase of the outlet valve closure allows adjustment of the maximum pressurization within the expansion chamber as the piston approaches TDC. This adjustment enables adapting the engine for maximum efficiency operation even while running at a wide variety of speeds.
The operation of the harmonic engine of the first embodiment with excessive pressure leads to a tendency for a decrease in power output relative to normal operating conditions. If the cylinder pressure at BDC has not decreased sufficiently through the expansion process to allow the outlet valve to open, then the subsequent upstroke of the engine simply recompresses the working fluid in the cylinder, and positive work is not produced during such a cycle. However, during such a recompression stroke, the inlet valve then tends to open early, by virtue of the much greater than normal pressure in the expansion cylinder as the piston approaches TDC, and as a result the next cycle can produce some positive work. The work done under such a cycle is less than normal, as the pressure induced opening of the inlet valve tends to be early, which tends to lead to an early closure of the inlet valve. As a result, when driven by excessive pressure, the net power averaged over several cycles is decreased. This feature can be used to advantage under some circumstances, such as providing a self-governing operational mode.
The operation of the harmonic engine of the first embodiment is insensitive to the direction of rotation of the crankshaft, and thus it runs equally well with a clockwise or counter-clockwise rotation. Starting from rest, if the piston is just below TDC, with the crank at a positive angle of 10°, for example, and pressurized working fluid is supplied to the inlet port of the engine with a pressure sufficient to overcome static friction, the piston will begin to move downwards and the crankshaft will rotate in a positive direction and continue to run in a positive direction. On the other hand, if the crank starts at a negative angle of −10°, the piston will be the same distance from TDC, and will begin to move downwards, and the engine will run “backwards” with the crankshaft rotating in a negative direction.
An embodiment that provides for greater accommodation to higher-pressure and higher speed operation is shown in
Inlet valve launching spring 409 is shown mounted on piston 460 in order to help minimize the mass of the inlet valve assembly, although it could be mounted on valve 401 as well. Inlet valve 401 has a dished lower surface 400 that accommodates space for spring 409 to be compressed and allows piston 460 to rise to nearly contact the top surface 463 of cylinder 461 and thus minimize the minimum volume of expansion chamber 462. It is appreciated that a recess in the upper surface of the piston could serve this role as well. In contrast to the first embodiment, in which the lower surface of the inlet valve 109 forces the inlet valve to open immediately after surface 109 makes contact with piston 160, the compliance of spring 409 does not open inlet valve 401 immediately after contact.
In this embodiment, the inlet valve opening spring is implemented as a flexure spring 407 mounted to the internal wall of inlet duct 425. Outlet valve opening spring is also implemented as a flexure spring 406 mounted on the internal wall of outlet header duct 405, with the maximum opening position limited by a stopper 410. Outlet valve closing spring 408 is mounted to the upper surface of piston 460 and nestles within the dished surface 403 of outlet valve 404 when it is fully compressed. In this embodiment, the piston sealing element is preferably at least one unitary ring or flange 464 as known in the art, that not only provides for low friction bearing of the piston but also a hermetic seal against leakage of working fluid within expansion chamber 462 past piston 460. With all structures of the inlet valve located within the inlet header duct 425 and all structures of the outlet valve located within the outlet header duct 405, and with piston sealing ring 464 a unitary seal, the engine eliminates significant leakage of working fluid to the outside environment during its normal operation.
The relative heights of the relaxed outlet valve closing spring 408 and the relaxed inlet valve launching spring 409 together with the relative spring constants are chosen so that outlet valve 404 is closed by the compression of spring 408 prior to piston 460 reaching TDC and prior to the opening of inlet valve 401 by the compression of launching spring 409. The height of inlet valve opening spring 409 is chosen greater than the distance between the top of piston 460 and the top surface 463 of the cylinder at the time that outlet valve 404 just closes, so that launching spring 409 becomes compressed as piston 460 approaches TDC.
Generally therefore, this embodiment also includes the reciprocating-piston expander comprising: the expander cylinder having an inlet and an outlet; the piston head axially slidable in the expander cylinder and together enclosing an expansion chamber accessible by the inlet and the outlet, the intake header in fluidic communication with the inlet for channeling working fluid from a pressurized fluid source into the expansion chamber, and the exhaust header in fluidic communication with the outlet for channeling working fluid exhausted out from the expansion chamber. And also similar to the first embodiment, an inlet valve is also provided for controlling the flow of working fluid from the intake header through the inlet to effect a power stroke of the expander, with the inlet valve comprising an inlet valve head and a resiliently biasing member arranged together as a harmonic oscillator so that the inlet valve head is moveable against an equilibrium restoring force of the resiliently biasing member from an unbiased equilibrium position located in the intake header to a biased closed position occluding the inlet. Arranged in this manner, upon releasing the inlet valve head from the closed position, it undergoes a single oscillation past the equilibrium position to an oppositely biased maximum open position and returns to a biased return position between the closed and equilibrium positions to choke the flow of working fluid and produce a pressure drop across the inlet valve causing the inlet valve to close.
Also the outlet valve of the second embodiment is provided to control the flow of working fluid exhausted out through the outlet to the exhaust header during a return stroke of the expander. As discussed the outlet valve includes an outlet valve head, stopper, and a resiliently biasing member arranged so that the outlet valve head is moveable against an equilibrium restoring force of the resiliently biasing member from a maximum open position located in the expansion chamber and delimited by the stopper to a biased closed position occluding the outlet.
And generally, the second embodiment of
The operation of the second embodiment under nominal or lower pressure conditions is very much as described above for the first embodiment. The minimal volume of the expansion chamber at TDC, dictates that the amount of working fluid that must be admitted through the inlet valve to raise the pressure within the cylinder to that of the supply is minimal, and the pressure jump as the piston approaches TDC can be achieved in minimal time. This is advantageous for achieving higher efficiency and power.
The operation of the second embodiment under high supply pressure conditions changes significantly, and the contrast with nominal pressure operation is shown in
The highest efficiency in the extraction of the energy of the supplied pressurized working fluid is obtained with the complete expansion down to the pressure of the working fluid in the exhaust manifold 405 occurring just as the piston reaches BDC. Under these conditions, the outlet valve opens at a phase point shown by arrow 445 just before BDC shown by dashed line 442, and is completely open at phase point indicated by arrow 446. The pressure in the expansion chamber after the closure of the inlet valve and before the re-opening of the inlet valve is virtually the same for both high pressure and normal pressure conditions, as shown in
Finally, with even greater than a factor of two overdrive pressure, if the mass of inlet valve 401 is made sufficiently light, the aerodynamic force of the inrushing working fluid can drive inlet valve closed in even less time than half the natural resonance period of the freely oscillating inlet valve. In practice a prototype engine has achieved as much as a factor of four decrease in the open period of a harmonic engine inlet valve with respect to its natural resonance period. This prototype engine had a cylinder bore of 7 cm, a stroke of 4.4 cm, both inlet and outlet valve port diameters of 1.5 cm, a mass of 11 g for the outlet valve, a mass of 7 g for the inlet valve, a spring constant of 590 N/m for the outlet valve opening spring, and a spring constant of 170 N/m for the outlet valve closing spring. The natural resonance period of the inlet valve was approximately 0.02 s, with the engine operating at low pressure and a cycle time of 0.05 s. As the pressure increased, the open period of the inlet valve decreased to as little as 0.005 s. This prototype engine was able to run satisfactorily over the range of supply pressures from 3 psig to 43 psig.
The advantage of such overdriven inlet valve operation is that much higher efficiency of use of the pressurized working fluid over a wider range of supply pressures is made possible relative to the case without overdrive.
A third embodiment, shown in
In this embodiment of the harmonic engine, inlet valve 201 is in the form of a reed valve, shown from the side in
With the very low mass characteristic of reed valves, it is preferred to incorporate a latching mechanism, as shown in
Reed valves, firmly supported, have low friction, and thus readily provide the high Q resonant behavior desirable in the present engine. Reed valves are also naturally low in mass, which is conducive to high-speed operation as well. The springiness of the reeds provides the resilient action described for the prior embodiments without the need for a separate resilient member.
The feature of the wobble-piston that is exploited here is that the left hand side of the piston (as shown in the drawings herein) reaches the apogee of its motion towards the top of the cylinder before the right hand side of the piston reaches its apogee, and before the middle of the piston reaches its apogee. Furthermore, the right hand side of the piston reaches its apogee after the middle of the piston. Note that the height of apogee of the left hand side of the wobble piston is above the height of apogee of the center of the wobble piston.
In normal operation, the protrusion 209 on the wobble-piston serves to force inlet valve 201 to open at a phase angle just at or slightly after TDC. Although it is appreciated that this protrusion could be compliant or elastic, as described for the second embodiment, with a rigid protrusion, the phase of opening of the inlet valve is well defined, and independent of the magnitude of the supply pressure. Once forced open, and with the pressure in expansion chamber 262 equalized with the supply pressure, inlet valve 201 undergoes a single oscillation, and is then held closed by the pressure differential that develops across it, just as described above for the first embodiment. Having a rigid protrusion 209 helps keep the low mass inlet valve 201 from being unduly influenced by the rapidly inrushing working fluid just after it opens.
In normal operation, the outlet valve remains closed from TDC to just before BDC, until the pressure within the expansion chamber decreases to nearly that of the outlet manifold, at which point the outlet reed rapidly snaps opens and is stopped at its fully relaxed, neutral position by latch 202.
Both the inlet and outlet valves remain in these positions, the inlet closed and the outlet opened, for most of the up-stroke of the wobble-piston. As the wobble-piston approaches TDC, it is tilted, and its left hand side is closer to the top of the expansion chamber than its right hand side. Thus, the preferred time ordering of the closing of the outlet valve before the inlet valve is opened is easily achieved by positioning the outlet valve over the portion of the wobble-piston that arrives at the upper extreme of its travel earlier. A protrusion 208, that may be rigid, elastic, compliant or springy, is located on the left hand side of the wobble-piston. As the outlet valve closing protrusion 208 makes contact with the outlet valve and begins to close it, the piston has not yet reached TDC, and thus the volume of the expansion space is decreasing. As the outlet valve is forced closed by protrusion 208, the increasing pressure (by virtue of the decreasing volume) within the expansion chamber in combination with the compression (if compliant) of protrusion 208 serve to hold the outlet valve closed. With the outlet valve closed, the piston continues to TDC and the cycle repeats. A particular virtue of the wobble-piston embodiment is the natural enforcement of the closure of the outlet valve prior to TDC, and the opening of the inlet valve after TDC by the natural wobbling nature of the motion of the piston.
A timing diagram for the wobble-piston embodiment is displayed in
In contrast to the first two embodiments, the wobble-piston embodiment is not symmetrical in its operation with respect to the direction of rotation of the crankshaft. Since the inlet valve is forced open after the inlet valve is forced closed for one direction of rotation but not the other, the wobble-piston engine operates best in that direction, and may not work at all in the opposite direction. Also, as the inlet valve is forced open by protrusion 209 over a wider range of crankshaft angles, and with proper design these may all be positive angles, the startup conditions for the wobble-piston embodiment are more tolerant of variations in the engine speed and operating pressure. For example, with a wobble-piston connecting rod length of 12.7 cm, an eccentric radius of 1.8 cm, and a piston width of 7.2 cm, then the height of the right hand side of the piston increases from its position at 0° where it is at the same height as the center of the piston at TDC, reaches a maximum position that is higher by 0.64 mm at a crankshaft angle of 14° and then returns to the height of the piston at TDC when the crankshaft is at 28°. Thus if the protrusion 209 makes initial contact to open the inlet valve at 0°, then it will force the inlet valve to remain open over the range of angles from 0° to 28°, regardless of the engine speed or supply pressure.
An especially lightweight and efficient embodiment of the harmonic engine especially useful in the context of an aircraft engine, is shown in
Thus this example embodiment is also an engine having a reciprocating-piston expander operably connected to a crank assembly. In particular, and similar to the other embodiments discussed herein, the expander includes an expander cylinder having an inlet and an outlet, an intake header in fluidic communication with the inlet for channeling working fluid from a pressurized fluid source into the expansion chamber, and an exhaust header in fluidic communication with the outlet for channeling working fluid exhausted out from the expansion chamber. In this embodiment, however, a wobble piston is used having a piston head with a flexible flange positioned between the piston head and the expander cylinder so as to seal an expansion chamber enclosed by the piston head and the expander cylinder and which is accessible by the inlet and the outlet. The piston head is connected to the crank assembly via a fixed connected piston rod. Also, an inlet reed valve is used for controlling the flow of working fluid from the intake header through the inlet to effect a power stroke of the expander.
Here too, the inlet reed valve is a harmonic oscillator with a first end connected to a wall of the intake header and a second end moveable to a closed position by resiliently biasing the inlet reed valve against an equilibrium restoring force thereof from an unbiased equilibrium position located in the intake header to a biased closed position occluding the inlet. In this manner, and upon releasing from the closed position, the second end of the inlet reed valve undergoes a single oscillation past the equilibrium position to an oppositely biased maximum open position and returns to a biased return position between the closed and equilibrium positions to choke the flow and produce a pressure drop across the inlet valve causing the inlet valve to close.
And an outlet reed valve is used for controlling the flow of working fluid exhausted out through the outlet to the exhaust header during a return stroke of the expander. Like the inlet reed valve, the outlet reed valve has a first end connected to a wall of the expansion cylinder and a second end moveable to a biased closed position occluding the outlet by resiliently biasing the outlet reed valve against an equilibrium restoring force thereof from an open position located in the expansion chamber. As discussed above, the outlet valve latch operates to latch the second end of the outlet reed valve in the open position. And two protrusions are carried by the piston head, which are positioned to bump open the inlet valve from the closed position to initiate the single oscillation of the second end of the inlet reed valve, and to release the second end of the outlet reed valve from the outlet valve latch and move the second end of the outlet reed valve from the open position to the closed position ahead of the bump opening of the inlet valve
The crank assembly of the third example embodiment has a crankshaft operably connected to the piston rod for effecting the return stroke of the expander after each power stroke, and inducing wobble motion of the piston head as it reciprocates in the expansion cylinder. When a propeller is connected to the crankshaft, it can provide the rotational inertia to transfer to the piston head via the crankshaft to effect the return stroke.
The embodiments described above are illustrative of the present invention, but it is appreciated that many other variations have utility in a variety of applications. It is appreciated that any of the variations discussed in each of the embodiments could be used in the other embodiments.
It is appreciated that a hinged member and spring could be used for either the inlet or outlet valves. It is appreciated that a variety of working fluids may be used to provide the pressure that drives this engine, including compressed air, steam, or other expansible fluids or the pressurized exhaust from an internal combustion engine. It is appreciated that combinations of reed valves and poppet valves, such as a reed valve for the inlet and a poppet valve for the outlet, are advantageous in some applications. It is appreciated that a double acting configuration with a substantially identical duplicate set of inlet and outlet valves placed in a complementary expansion chamber below the piston could be used to effectively double the power for a given engine bore, stroke and speed. It is appreciated that this engine may be used as a key component in a heat powered engine, either open cycle or closed cycle. It is appreciated that a linear induction motor, driven by a magnetic or magnetized piston, could be used to advantage, and especially in the context of a completely hermetically sealed double acting embodiment. It is appreciated that multiple cylinders may be employed together to provide dynamic balancing and smoother operation. It is appreciated that with proper phasing of multiple cylinders, the engine may be started with the provision of pressurized working fluid regardless of the initial angle of the crankshaft. It is appreciated that the addition of an overpressure relief port that is exposed as the piston approaches BDC may be useful for some applications.
In a uniflow embodiment of the harmonic engine, rather than outlet valves, a number of outlet ports are placed around the circumference of the cylinder, so that in the course of the reciprocation of the piston head, pressurized working fluid is able to vent to the exhaust manifold as the piston head uncovers the ports. As in the known art of condensing uniflow steam engines, it is preferable for the exhaust manifold pressure to be sub-atmospheric, but not necessary.
Inlet valve 701 has a head portion 703 that is attached to a resiliently biasing member 707 such as a mono-leaf spring, or flexure, or reed, that tends to hold the valve open by positioning the inlet valve head away from seat 702 and away from piston head 760. In a simple reed configuration, valve head 703 is merely the end of flexure element 707. The end of flexure 707 opposite valve head 703 is attached to a wall of inlet header duct 725. The inlet valve preferably occludes when pushed toward piston head 760, and opens when pulled away from the piston head. In particular, the inlet valve head is moveable against an equilibrium restoring force of the resiliently biasing member from an unbiased equilibrium position located in the intake header to a biased closed position occluding the inlet. The inlet valve seat 702 has a flat surface so that, when the inlet valve is nearly closed, the flat lower surface of inlet valve head 703 becomes parallel to inlet valve seat 702. Alternatively, both the lower surface of the inlet valve head and the inlet valve seat may be curved, with equal and opposite curvature, in order to tolerate a higher pressure without excessive strain in the inlet valve. The relative dimensions of the valve head and flexure length are chosen so that a pressure force acting on the valve head predominantly excites the lowest vibrational mode of oscillation of resiliently biasing member 707. For example, in the simple reed configuration the vibrational modes of the inlet valve will be those of a cantilevered beam, and since the second lowest vibrational mode of a cantilevered beam has a node at 78% of its length from the fixed end, with the center of head 703 at 78% of the free length of the flexure from the attachment point, the lowest vibrational mode will be most strongly excited, with no excitation of the second vibrational mode. Furthermore, since the third lowest vibrational mode of a cantilevered beam has a node at 87% of its length from the fixed end, this mode is also only weakly excited by an impulsive pressure force impact on head 703. Since the resonant frequency of the third mode for a cantilevered beam is nearly 18 times greater than the resonant frequency of the fundamental mode, at modest speed impact, not only is the third mode barely excited, but it will also decay away extremely quickly compared to the fundamental mode of vibration.
The inlet valve head 703 and its opening flexure 707 effectively form a spring-mass system of a simple harmonic oscillator which, when the inlet valve is displaced from its unbiased equilibrium position, experiences a restoring force proportional to the displacement according to Hooke's law, as known in the art. This oscillator preferably has a high quality factor Q value, so that, while freely oscillating, many cycles of oscillation occur before the amplitude of oscillation decays significantly. The significance of the high Q value in the context of this invention is that after a single oscillation, when released at rest from a closed position, in the absence of other forces, the inlet valve returns almost all the way back to its closed position. The higher the Q factor, the more closely the inlet valve returns, after a single oscillation, to its closed position. For example, with a Q value of 160 the inlet valve returns to within 1% (relative to the full excursion of the valve) of its closed position after a single oscillation, and with a Q value of 14, the inlet valve returns to within 10% of its closed position after a single oscillation. With close return to the closed position, the narrowest effective waist of the flow passageway from inlet duct 725 to expansion space 762U effectively forms a converging-diverging nozzle. With a sufficiently high Q, the narrowness of the throat of the converging-diverging nozzle section has the practical effect of choking the flow of working fluid between the inlet duct 725 and the expansion space 762U for only a modest pressure differential across the throat. As is known in the art of converging-diverging nozzles, flow is choked by the limitation that the flow speed cannot exceed the speed of sound at the throat of the nozzle. With sufficiently high Q, and thus a sufficiently small throat area, even at the lowest practical engine operating speed, the flow at the throat reaches the speed of sound and is thus choked, while at higher engine operating speeds, the rapidly narrowing throat facilitates extremely rapid inlet valve closure. The form of inlet valve shown is conducive to attaining very high Q values, as the frictional losses of flexure bearings, such as 707, constructed of high quality spring steel, are very low.
The outlet from expansion chamber 762U is comprised of a number of outlet ports 711 placed around the circumference of cylinder 761 at a distance approximately 85% to 95% of the full stroke of piston head 760 below its uppermost position. With this placement, fluidic communication between expansion chamber 762U and outlet manifold 705 is only available at times that the piston head is within approximately 5% to 15% of its lowermost position. Also, lower chamber 762L is in fluidic communication through outlet ports 711 with outlet manifold 705 for all but 5% to 15% of the piston stroke, so that the pressure within lower chamber 762L varies only slightly throughout the stroke of the piston head.
Operation of the harmonic uniflow engine embodiment is now described for normal, steady running conditions. The variation in the positions of the inlet valve and the piston head are shown with solid lines in a timing diagram in
Starting the cycle arbitrarily at the TDC position, the configuration of the components and the state of their motion is described as follows, and best shown in
As piston head 760 descends from TDC the inlet valve undergoes a single oscillation, passing upwards through the neutral position of flexure 707 indicated by dashed line 741 in
The state of motion of the components near the closure point is as follows.
The narrowest passageway for the inflowing working fluid is defined by the smallest area gap between the lower surface of the inlet valve head and the upper surface of seat 702. At this time the piston head is moving down the cylinder, and the flow of working fluid through this narrow passageway becomes choked, and the pressure within expansion chamber 762U begins to drop significantly below the supply pressure in the inlet manifold. The pressure drop produces a force that urges the inlet valve to close. As the inlet valve gets very close to closing, under normal operating conditions, this pressure drop ensures that the inlet valve closes without bouncing, and remains closed for the remainder of the downward power stroke. This phenomenon is referred to as “dynamic latching” in this specification. The cylinder pressure drops when the rate of increase of the volume within expansion chamber 762U overwhelms the choked mass flow rate of working fluid through the narrow annular throat.
After the inlet valve closes, at the phase indicated by arrow 740 in
At the moment that seal 764 first drops below the top of outlet ports 711, with the pressure in expansion chamber 762U being greater than the pressure in exhaust manifold 705, flow of working fluid from the expansion chamber to the exhaust manifold causes a rapid drop of the working fluid pressure in the expansion chamber towards the exhaust manifold pressure. This point in the cycle is indicated by arrow 745 in
Operation of the uniflow engine embodiment is now described for low speed conditions, as at startup, or under heavy load. The variation in the positions of the inlet valve and the piston head are shown with solid lines in a timing diagram in
Operation of the uniflow engine embodiment is now described for high-speed conditions. The onset of this high-speed mode occurs when the natural period of the inlet valve exceeds the time from the inlet valve being forced open by the pressure spike prior to TDC to the time the piston head uncovers the outlet ports. The variation in the positions of the inlet valve and the piston head are shown with solid lines in a timing diagram in
By virtue of the changes in the operation of the harmonically acting inlet valves as a function of speed, as shown in
The speed of stable engine operation depends on the nature of the load. For loads that have a torque that is constant, as a function of engine speed, such as shown by dashed line 716 for example, there is a stable operating point indicated by arrow 712 located at the intersection of the engine torque curve 710 and the load curve 716. Alternatively, for a higher load and higher speed, such as shown by dashed line 718 for example, there is a stable operating point indicated by arrow 715 located at the intersection of the engine torque and load torque curves. In contrast, although there is an intersection of the constant load line 718 with the harmonic engine torque 710 at the point indicated by arrow 714, this is an unstable operating point, because a small increase in engine speed, for example, leads to a further increase in engine speed, while a small decrease in engine speed would lead to a further decrease in engine speed.
The pull up torque for the harmonic engine occurs at the bottom of the torque valley, indicated by arrow 719, located at relatively low operating speed. The significance of the pull up torque, in analogy with the pull up torque characteristic of electric motors, is that in order to get beyond the torque valley without stalling, the pull up torque must exceed the load torque at that point.
Since the high efficiency expansive mode of operation of the harmonic engine lies within the region of increasing torque with speed, it is preferred to have a load that increases with speed faster than the rate that the harmonic engine torque increases with speed. The portion of a load curve indicated by dashed line 717 has this character, and the stable operating point is indicated by arrow 713 in
The construction and functioning of both the upper and lower inlet valves in the double acting engine is precisely as described for the single acting engine, except that they need not be identical to each other. For a given crank angular velocity, for example corresponding to the generation of electric power at a specific alternating current frequency, and for a given inlet valve cutoff fraction as a function of the full stroke, the dwell time from upper piston head 660U being at its uppermost position to the closing of upper inlet valve 601U is less than the dwell time from lower piston head 660L being at its lowermost position to the closing of lower inlet valve 601L. As a specific example, with the length of connecting rod 685 being four times as long as the radius from the center of the flywheel to crank pin 688, then, for a cutoff of 10% of stroke, the upper piston head dwell time would be 9.2% of the engine period, while the lower piston head dwell time would be 11.6% of the engine period. For 10% cutoff, it is thus preferred to have the resonant period of the lower inlet valve 601L be 26% greater than the resonant period of the upper inlet valve 601. As another example, for a cutoff of 20%, the upper dwell is 13.3%, the lower dwell is 16.6%, and the lower period should be 24% greater than the lower resonant period. Thus a close matching of the upper and lower inlet valve cutoffs is readily achieved by providing that upper flexure 607U is stiffer than lower flexure 607L. The desired relative stiffness is easily achieved by a choice of the relative width and/or thickness of the upper and lower flexures. Although it is also possible to adjust the resonant frequency of the upper and lower valves by having a different mass for the upper and lower inlet valve heads, it is preferred that the upper and lower valve heads 603 be identical.
The outlet from upper expansion chamber 662U is comprised of a number of outlet ports 611 placed around the circumference of cylinder 661 at a distance approximately 85% to 95% of the full stroke of piston face 660U below its uppermost position. With this placement, fluidic communication between upper expansion chamber 662U and outlet manifold 605 is only available at times that piston head 660U is within approximately 5% to 15% of its lowermost position. Also, the separation between upper piston head 660U and lower piston head 660L is chosen so that when the upper piston head is within approximately 5% to 15% of its uppermost position, fluidic communication between lower expansion space 662L and outlet manifold 605 is enabled.
Operation of the double acting harmonic uniflow engine embodiment is now described for normal, steady running conditions. The variation in the positions of the upper inlet valve, the lower inlet valve and the piston head are shown with solid lines in a timing diagram in
As piston head 660 descends from TDC the inlet valve undergoes a single oscillation, passing upwards through the neutral position of flexure 607U indicated by dashed line 641 in
The configuration of the engine components at 90° of crankshaft angle is that both upper and lower inlet valves are closed (for normal speed operation) and the working fluid in lower expansion chamber 662L is being adiabatically compressed, while the working fluid in upper expansion chamber 662U is being adiabatically expanded. It is apparent from
With the single acting uniflow engine embodiment, it is not necessary to have a lower expansion chamber 762L or seal 796, if the crankcase itself is connected to the low-pressure working fluid exhaust manifold. Alternatively, if the exhaust is directly to atmospheric pressure, a configuration without lower expansion chamber or seal is feasible. The structure of either of these cases is as shown in
It is appreciated that the single acting Harmonic Uniflow Engine may operate with a wobble piston mechanism of the form shown in
An embodiment of the Harmonic Uniflow Engine in which the shape of the piston head itself provides the mechanism for the release of expanded working fluid from the expansion chamber is shown in
The radii of curvature of each of the two spherical lune surface segments 523 and 524 of piston head 560 are nominally equal to the bore radius of the cylinder, and the angular extent of the spherical segments chosen so that as the piston head is tilted by connecting rod 595 as crank pin 588 traverses a circular orbit about the crankshaft, a horizontal circular seal between the piston head surface and the cylinder wall is maintained from 0° to 180° of crankshaft rotation. The greatest counter-clockwise piston head tilt angle is attained at 90° of crankshaft rotation, as shown in
In contrast, for the entire recovery stroke between 180° and 360° of crankshaft rotation, a pair of crescent shaped passageways develop, breaking the seal between the surface of piston head 560 and cylinder 561, allowing working fluid to flow past the head. Only two small areas of piston head 560 at either end of pivot axis 520 remain in contact with the walls of the cylinder on the recovery stroke, but this is sufficient to keep the center of piston head 560 positioned along axis 522. In effect, the piston head acts as a partially open butterfly valve for the upward recovery stroke. This is readily seen in
The shape of the ceiling of cylinder 561, i.e. the cylinder inner wall surface above plane 521 in
With a liquid lubricating fluid, the radius of curvature of the spherical segments in piston head 560 may be smaller than the bore radius of cylinder 561 by a value equal to the desired minimum gap corresponding to the thickness of a lubricating film layer between the outer surface of the piston head and the inner surface of the cylinder wall, so that a low friction seal may be formed between the surface of the piston head and the inner wall of the cylinder. Since piston head 560 is tilting as it is reciprocating, lubricating fluid is drawn into the region of contact between the right and left spherical surface lunes and the cylinder wall by the relative motion between them, and such hydrodynamic lubrication promotes low frictional power loss and low wear.
With a pressurized phase changing working fluid such as saturated steam, a thin liquid film naturally develops on the inner walls of the cylinder, and this liquid film itself can provide aquaplaning lubrication without the need for supplemental lubricants, such as oil. On the other hand, with conventional oil lubrication, a somewhat thicker oil film may be expected, and a somewhat larger minimum gap exploited. In the case that the spherical lune surfaces of the piston head are made of, or coated with, a compliant, low friction material, such as pure PTFE, brass filled PTFE to enhance heat transfer, or MoS2 and glass filled PTFE, it is preferable to have no gap between the piston head surface and the cylinder walls and to exploit the dry lubricating properties of PTFE itself for both the piston head bearing and the piston head seal mechanism.
It is appreciated that the angular width of the spherical wedge segments can be adjusted to provide venting for more or less than 180° of crankshaft rotation. It is appreciated that placing the crankshaft axis slightly off center with respect to the centerline of the cylinder allows venting to begin slightly prior to 180° of crankshaft rotation, in order to increase engine power while allowing cessation of venting slightly prior to 360° of crankshaft rotation, in order to promote earlier recompression. With such an off axis crankshaft, the pivot axis 520 still remains on the centerline of the cylinder and parallel to the crankshaft axis. It is appreciated that the sharp edges of piston 560 may be slightly rounded or chamfered, in order to prevent digging or gouging at the point that piston 560 first reengages or contacts the walls of cylinder 561. It is appreciated that the engine described here may be operated as a hydraulic motor by taking advantage of the non-expansive mode of operation described in connection with
A distinctive feature of the spherical wedge piston embodiment of the harmonic uniflow engine is that it has the ability to function as a compressor rather than as an engine or expander. This is in contrast to the fully reversible embodiment discussed earlier in connection with
In summary, the advantage of the harmonic valve in combination with the spherical wedge piston, is that when operating as an engine, corresponding to the clockwise rotation of the crankshaft shown in
Further modifications and changes may become apparent to those skilled in the art, and it is intended that the invention be limited only by the scope of the claims.
This application is a continuation-in-part of U.S. patent application Ser. No. 13/221,783 filed Aug. 30, 2011, which claims the benefit of U.S. provisional application No. 61/378,327 filed Aug. 30, 2010, both of which are incorporated by reference herein.
The United States Government has rights in this invention pursuant to Contract No. DE-AC52-07NA27344 between the United States Department of Energy and Lawrence Livermore National Security, LLC for the operation of Lawrence Livermore National Laboratory.
| Number | Date | Country | |
|---|---|---|---|
| Parent | 13221783 | Aug 2011 | US |
| Child | 14243729 | US |
