# Patent

## Title
Harmonic uniflow engine

## Abstract

A reciprocating-piston uniflow engine includes a harmonic oscillator inlet valve capable of oscillating at a resonant frequency for controlling the flow of working fluid into the engine. In particular, the inlet valve includes an inlet valve head and a spring arranged together as a harmonic oscillator so that the inlet valve head is moveable from an unbiased equilibrium position to a biased closed position occluding an inlet. When released, the inlet valve head undergoes a single oscillation past the equilibrium position to a maximum open position and returns to a biased return position close to the closed position to choke the flow and produce a pressure drop across the inlet valve causing the inlet valve to close. In other embodiments, the harmonic oscillator arrangement of the inlet valve enables the uniflow engine to be reversibly operated as a uniflow compressor.

## Background
CROSS-REFERENCE TO RELATED APPLICATIONS
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.


FEDERALLY SPONSORED RESEARCH OR DEVELOPMENT
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.


BACKGROUND OF THE INVENTION
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° F. 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.
SUMMARY OF THE INVENTION
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.



BRIEF DESCRIPTION OF THE DRAWINGS
The accompanying drawings, which are incorporated into and form a part of the disclosure, are as follows:
FIG. 1 is an overall cross-sectional view of the first exemplary embodiment of the present invention in its fully relaxed configuration.
FIG. 2 is a cross-sectional view of the top portion of the first embodiment showing both inlet and outlet valves fully closed.
FIG. 3 is a cross-sectional view of the top portion of the first embodiment showing the piston at TDC position.
FIG. 4 is a cross-sectional view of the top portion of the first embodiment showing the inlet valve at its maximally open position.
FIG. 5 is a cross-sectional view of the top portion of the first embodiment showing the inlet valve configuration just before it closes.
FIG. 6 is a timing diagram illustrating the relative positions of the inlet valve, outlet valve, the piston, and the pressure within the expansion chamber of the cylinder for nominal full power, full pressure operation.
FIG. 7 is a timing diagram for operation at reduced pressure and power.
FIG. 8 is a cross-sectional view of the upper section of the second embodiment showing both inlet and outlet valves in closed position.
FIG. 9 is a cross-sectional view of the upper section of the second embodiment showing the inlet and outlet valves in their relaxed, equilibrium positions.
FIG. 10 is a timing diagram for overdrive operation at high pressure.
FIG. 11 is a side cross-sectional view of the wobble-piston embodiment.
FIG. 12 is a top cross-sectional view showing the inlet and outlet valves with a portion of the cylinder and the outlet valve latch, taken along the line of sight 12-12 shown in FIG. 11.
FIG. 13 is an expanded view showing the details of the outlet valve latch mechanism, taken from the portion of FIG. 11 within circle 13.
FIG. 14 is a timing diagram for the wobble-piston embodiment.
FIG. 15 is a cross-sectional view of an aircraft embodiment of the present invention.
FIG. 16 is a cross-sectional view of a uniflow embodiment of the present invention corresponding to a crank angle 90° in advance of TDC.
FIG. 17 is a cross-sectional view of a uniflow embodiment of the present invention corresponding to a crank angle of 0° with the piston at TDC.
FIG. 18 is a cross-sectional view of a uniflow embodiment of the present invention corresponding to a crank angle of 900.
FIG. 19 is a cross-sectional view of a uniflow embodiment of the present invention corresponding to a crank angle of 180°.
FIG. 20 is a timing diagram illustrating the relative position of the inlet valve and piston and the pressure within the cylinder for normal operation.
FIG. 21 is a timing diagram illustrating the relative position of the inlet valve and the piston and the pressure within the cylinder for low speed operation.
FIG. 22 is a timing diagram illustrating the relative position of the inlet valve and the piston and the pressure within the cylinder for high-speed operation.
FIG. 23 is a plot showing the variation of engine torque (averaged over a cycle) as a function of operating speed.
FIG. 24 is a cross-sectional view of an atmospheric embodiment of the present invention corresponding to a crank angle 90° before TDC. This figure also illustrates the use of an electric car controller for operation of the harmonic engine.
FIG. 25 is a cross-sectional view of a double acting embodiment of the present invention corresponding to a crank angle 90° before TDC.
FIG. 26 is a timing diagram illustrating the relative position of the upper and lower inlet valves for the double acting embodiment of the present invention.
FIG. 27 is a cross-sectional view of a spherical wedge piston embodiment of the present invention corresponding to a crank angle 90° in advance of TDC.
FIG. 28 is a cross-sectional view of a spherical wedge piston embodiment of the present invention corresponding to a crank angle of 0° with the piston at TDC.
FIG. 29 is a cross-sectional view of a spherical wedge piston embodiment of the present invention corresponding to a crank angle of 90°.
FIG. 30 is a cross-sectional view of a spherical wedge piston embodiment of the present invention corresponding to a crank angle of 180°.
FIG. 31 is a cross-sectional view of a spherical wedge piston embodiment of the present invention operating as a compressor with a crank angle of 90° after TDC.
FIG. 32 is a cross-sectional view of a spherical wedge piston embodiment of the present invention operating as a compressor with a crank angle of 180°.
FIG. 33 is a cross-sectional view of a spherical wedge piston embodiment of the present invention operating as a compressor with a crank angle of 270°.
FIG. 34 is a cross-sectional view of a spherical wedge piston embodiment of the present invention operating as a compressor with a crank angle of 360°, i.e. TDC.



DETAILED DESCRIPTION
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.
First Example Embodiment
Turning now to the drawings, FIGS. 1-5 show a first exemplary system of the engine of the present invention. In particular, FIG. 1 shows the harmonic engine in a static, non-operational state such as typically seen just prior to startup, and FIGS. 2-5 show the harmonic engine in various dynamic states of its two-stroke operation. The harmonic engine is shown comprising the following components and sub-assemblies. First a reciprocating-piston expander is shown comprising an expander cylinder 161 having an inlet and an outlet. The expander also includes a piston head 160 axially slidable in the expander cylinder and together enclosing an expansion chamber 162 accessible by the inlet and the outlet. Also the expander includes an intake header 125 in fluidic communication with the inlet for channeling working fluid from a pressurized fluid source into the expansion chamber, and an exhaust header 105 in fluidic communication with the outlet for channeling working fluid exhausted out from the expansion chamber. A harmonically oscillating inlet valve 101 is provided for controlling flow of high pressure working fluid into expansion chamber 162 from the intake header, and a resiliently biasing (e.g. spring-loaded) upper outlet valve 104 is also provided for controlling flow out of expansion chamber 162 to the exhaust header 105 capable of venting the expanded, low pressure working fluid. Finally, a crank assembly 155 as conventionally known in the art is shown operably connected to a piston 160 for converting reciprocating motion into rotary power output. In this embodiment the rotary power output is shown connected to a flywheel 185 and an induction motor/generator 188 further connected to an alternating current electric power grid. Each of these and other components is discussed in detail as follows.
Inlet Valve
The inlet valve 101 in the first exemplary configuration is shown in FIG. 1 and has an inlet valve head portion 103 (particularly shown as a convex frusto-spherical section) that is attached to a resiliently biasing member 107 such as a mono-leaf spring, or flexure bearing spring, that tends to hold the valve open by positioning the inlet valve head away from piston 160. In particular, the mono-leaf spring is shown cantilevered with the inlet valve head connected at one end. The end of the flexure spring 107 opposite the valve head 103 is attached to a wall of the inlet header duct 125. The inlet valve has a lower surface 109 of a lower portion that extends or protrudes into the interior of cylinder 161, i.e. expansion chamber, when valve 101 is in its fully closed position shown in FIG. 2 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 inlet valve preferably occludes when pushed toward piston 160, and opens when pulled away from the piston. The inlet valve seat 102 has a chamfered surface so that, when closed, the preferably convex spherical surface contour of the lower surface 109 of the inlet valve head 103 contacts the conical inlet valve seat 102 along a circle.
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 FIG. 1. This position is found in the de-energized, “cold-start” configuration of the engine, without pressurized working fluid supplied to the inlet manifold, and before the piston begins to reciprocate and produce power. The inlet valve is thus a normally open valve. When inlet valve 101 is closed, as shown in FIG. 2, the inlet valve is displaced from its equilibrium position, inlet valve spring 107 is flexed, and there is a restoring force produced by spring 107 that tends to open the inlet valve.
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.
Outlet Valve
The outlet valve 104 is also shown in FIG. 1 as a poppet valve connected to a resiliently biasing member, flexure or spring 106 that tends to hold the valve open and pushed toward piston 160. The outlet valve has a resiliently biasing member, such as outlet valve closing spring 108 attached to it that faces piston 160. The outlet valve also has a stopper 110 attached to it that prevents the outlet valve from moving past its designed fully open position in the direction towards the piston.
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 FIG. 2. By this construction, as the piston approaches the top of its travel, it is assured to encounter outlet valve closing spring 108 prior to contacting the bottom of inlet valve 109.
Outlet valve 104 is open (FIG. 1) when spring 106 is maximally extended towards its neutral position consistent with the constraint of stopper 110, and closed (FIG. 2) when the outlet valve is displaced maximally away from its neutral position. When outlet valve 104 is closed, spring 106 is compressed with respect to its neutral position and produces a restoring force tending to push outlet valve 104 into cylinder 161 towards piston 160. When outlet valve 104 is fully open, spring 106 is still compressed with respect to its neutral position, producing a force tending to keep outlet valve open as far as the stopper 110 allows.
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.
Crank Assembly for Power Conversion
FIG. 1 shows a crank assembly 155 connected to piston 160 by power rod 195 in a conventional manner known in the art for converting reciprocating piston motion to rotary power output. Piston rings 164 as known in the art are shown. As shown in FIG. 1, this crank assembly is connected to motor-generator 188 so as to further convert the rotary mechanical power into electrical power as is well known in the art. Flywheel 185 is shown connected to the crank assembly 155 as is known in the art. Thus it is appreciated that the flywheel is part of the crank assembly which is one type of periodic return means for effecting the return stroke of the expander after each power stroke. Because the flywheel is connected to the crankshaft that is operably connected to the piston head, rotational inertia of the flywheel may be transferred to the piston head via the crankshaft.
Starter Motor Generator
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 FIG. 1, motor-generator 188 tends to cause the flywheel to rotate at a frequency very nearly equal to an integer fraction (with the integer depending on the motor pole structure) of the power grid frequency. As is known in the art, the precise rotational speed depends on the load on the motor, and the slip rate. When under load, the rotational speed lags the grid frequency and the motor-generator acts as a motor, while when being driven, the rotational speed is greater than the grid frequency, and the motor-generator acts as a generator. Common and inexpensive single-phase and three-phase induction motors are readily available for operation at 60 Hz (near 3600 rpm), 30 Hz (near 1800 rpm) and 20 Hz (near 1200 rpm). Where three-phase power connections are available, it is preferable to exploit a three-phase induction motor as the motor-generator. Where only single-phase power is available, a capacitor-start capacitor-run single-phase induction motor is preferred. Induction motors connected to an alternating current electrical power grid, when overdriven, produce power that is automatically properly phased with the power grid. As such, the motor-generator is capable of drawing power from a power grid to initially drive the expander at startup, and supplying power back to the power grid once operational.
Normal Engine Operation
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 FIG. 6 as a function of the phase of the engine cycle. The pressure within the expansion chamber is indicated by dashed line 170 in FIG. 6. The normal engine cycle consists of a pair of strokes of the piston, starting at the TDC (Top Dead Center) position that corresponds to an angle of 0° for the crankshaft, with the piston at one extreme of its motion nearest the top of cylinder 161, followed by a downward power stroke to the opposite extreme piston position at BDC (Bottom Dead Center), indicated by dashed line 142 in FIG. 6, followed by an upward recovery stroke back to the beginning point of the cycle at TDC. Under normal operation, the motion of the piston and valves are strictly periodic, and every cycle is nominally identical. As the pressure in the upper expansion chamber is higher on the down stroke from 0° to 180° than on the up stroke from 180° to 360°, averaged over a full cycle, the upper expansion chamber of the engine delivers a net positive power to the crankshaft.
Starting the cycle arbitrarily at the TDC position, the configuration of the components and the state of their motion is shown in FIG. 3, and is described as follows. Expansion chamber 162 is at its minimum volume point, preferably as small as is feasible. Piston 160 is instantaneously in a state of zero velocity as it is in the process of turning around. Outlet valve closing spring 108 is compressed to its smallest position and is exerting its strongest closing force on the outlet valve. Outlet valve 104 is fully closed and stationary. Inlet valve 101 is partially open and is moving upwards, as indicated by the arrow. Working fluid within expansion chamber 162 is at a pressure that is approximately equal to that of the supply pressure in the inlet header duct 125. The pressure difference between the working fluid in expansion chamber 162 and outlet header duct 105 produces a force on the outlet valve that is much greater than the force of outlet valve opening spring 106, and thus the outlet valve is held shut even without the extra force of outlet valve closing spring 108.
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 FIG. 1, except that the outlet valve is closed and the inlet valve is moving upwards) to a maximally open position illustrated in FIG. 4. In the maximally open position, the inlet valve is instantaneously at rest before it returns towards its closed position. As the inlet valve completes the second half of its oscillating motion, it again passes through its neutral position (as seen in FIG. 1, except that the outlet valve is closed and the inlet valve is moving downwards), and finally, by virtue of the high Q of the inlet valve harmonic oscillation, just before the “closure point” indicated by arrow 140 in FIG. 6 arrives at the configuration illustrated in FIG. 5.
The state of motion of the components in FIG. 5 is as follows. The narrowest passageway for the inflowing working fluid is located at the annular shaped throat defined by the smallest gap between the surface of the frusto-spherical inlet valve head at 100 and the nearest portion of the frusto-conical seat 102. At this time the piston is moving down the cylinder, and the flow of working fluid through this narrow passageway becomes choked, and the pressure within expansion chamber 162 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 pressure drops when the rate of increase of the volume within expansion chamber 162 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 140 in FIG. 6, the pressure in the expansion chamber drops approximately adiabatically as the volume of the working fluid within the chamber increases. Under nominal full power conditions, the pressure within the expansion chamber drops to very nearly equal the pressure of the working fluid in the exhaust header. At the moment that the force of the compressed outlet valve opening spring 106 exceeds the opposing force of the pressure drop across the outlet valve, the outlet valve begins to open. This point in the cycle is indicated by arrow 145 in FIG. 6. Once the outlet valve starts to open, the residual pressure drop between the expansion chamber and the exhaust header quickly decreases, and the outlet valve is forced by opening spring 106 to move rapidly toward its fully open position, indicated by dashed line 115 in FIG. 6.
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 FIG. 6. The impact of stopper 110 at the point that the outlet valve is fully closed is inelastic, and the outlet valve is brought to a sudden stop without bouncing. As an aid to prevent outlet valve bouncing, the outlet valve opening spring 106 is preferably still in compression at the fully open position of the outlet valve.
At the phase indicated in FIG. 6 by arrow 144, the piston makes initial contact with outlet valve closing spring 108. As the piston continues towards TDC, the outlet valve closing spring is compressed and the outlet valve accelerates towards its closed position. At the phase indicated by arrow 141, the outlet valve is closed. With proper choice of spring constants, the outlet valve closure occurs just before the piston makes contact with the bottom of the inlet valve 109 and forces open the inlet valve. This event is indicated by arrow 143 in FIG. 6.
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.
Startup of Engine
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 FIG. 6. With approximately equal area apertures for the inlet and outlet valves, and thus approximately equal flow resistance from the inlet duct 125 to the expansion chamber 162 as from the expansion chamber 162 to the exhaust duct 105, with both inlet and outlet valves initially open, the pressure drop across the inlet valve approximately matches the pressure drop across the outlet valve.
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.
Low Pressure Operation
The operation of the harmonic engine under conditions that the supply pressure is less than the nominal full power design pressure is shown in FIG. 7. Almost all of the events described in connection with full pressure operation in FIG. 6 are unchanged, except that, as the pressure of the working fluid admitted to the expansion chamber is less, following the closing of the inlet valve at the point indicated by arrow 140, the expansion of the working fluid causes its pressure 171 to drop to the level required to open the outlet valve at an earlier phase of the engine cycle. The earlier start of the opening of the outlet valve is indicated by arrow 147 in FIG. 7 and the earlier completion of the opening of the outlet valve is indicated by arrow 148. The start of the closing of the outlet valve, as it is driven by the position of the piston, still occurs at its normal phase point shown by arrow 144, and the complete closure, shown by arrow 141 occurs at its normal time. However, since the pressure in the inlet header duct 125 is less than the full power design pressure, the pressure within expansion chamber 162 tends to approach the full power design pressure, and may even “overshoot” the pressure in the inlet header duct as indicated by arrow 152 in FIG. 7, and thus tends to begin to force open inlet valve 101 very slightly earlier than normal, as indicated by arrow 151 in FIG. 7.
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.
Self-Governed Operation
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.
Excessive Pressure Operation
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.
Reversibility
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.
Second Example Embodiment
An embodiment that provides for greater accommodation to higher-pressure and higher speed operation is shown in FIGS. 8 and 9. FIG. 8 shows both inlet and outlet valves in their closed position. FIG. 9 shows both inlet and outlet valves in their quiescent neutral positions prior to startup of the engine. In this embodiment, there are two key modifications to the first embodiment described above. The first is that the function of the rigid inlet valve lower surface 109 of the first embodiment is replaced with a spring 409 or otherwise resilient, compliant or elastic member. The second is that the inlet valve is made sufficiently low in mass that the aerodynamic forces are no longer insignificant relative to the spring forces when the inlet valve is in its equilibrium position or higher, as is inherent in the discussion of the first embodiment above. Part of this mass reduction is affected by the removal of the lower frusto-spherical portion of the inlet valve head. In addition, the valve head may be made hollow to further lighten it.
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 FIGS. 8 and 9 show the use of a first protrusion carried by one of the inlet valve head and the piston head and positioned therebetween so that when the other non-carrying one of the inlet valve head and the piston head comes in contact with the protrusion the inlet valve is bumped open from the closed position to initiate the single oscillation of the inlet valve head. In particular the first protrusion in FIGS. 8 and 9 is shown as a spring carried by the piston head. For the outlet valve, an outlet valve closing spring is carried by one of the outlet valve head and the piston head and positioned therebetween 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. And while not shown in FIGS. 8 and 9, a periodic return means similar to FIGS. 1-5 is operably connected to the piston head for effecting the return stroke of the expander after each power stroke.
Normal Operation of Second Example Embodiment
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.
High Pressure Overdrive Operation
The operation of the second embodiment under high supply pressure conditions changes significantly, and the contrast with nominal pressure operation is shown in FIG. 10. In this figure, the motion of inlet valve 401, outlet valve 404, piston 460 and cylinder pressure 472 are shown as a function of crankshaft angle. The motion of the inlet valve for a single cycle of free oscillation (neglecting aerodynamic forces and assuming a sufficiently high Q that damping is negligible for a single cycle) starting from TDC is shown for comparison by dashed line 473, and the pressure in the expansion chamber for nominal conditions is shown for comparison by dashed line 470.
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 FIG. 10. With a higher-pressure supply of working fluid, the compression of inlet valve launching spring 409 is greater prior to the opening of inlet valve 401, and thus the stored potential energy is greater. Furthermore, as the density of higher-pressure supply of working fluid is greater, the aerodynamic force of in-rushing working fluid is greater than for normal operating conditions. Once inlet valve 401 opens, at the phase indicated by arrow 449 just before TDC, due to the combination of the upward force of compressed launching spring 409, flexed opening spring 407 and the compression of the working fluid in the expansion chamber produced after outlet valve 404 closes at the phase indicated by arrow 441, the pressure differential across inlet valve 401 rapidly disappears, and the potential energy stored in launching spring 409 is converted to kinetic energy of the upward velocity of valve 401. As a result, the motion of the inlet valve undergoes less than a full cycle of unrestrained, free oscillation, as can be seen by the comparison of the solid curve for valve 401 and the dashed curve 473 that represents a complete cycle of free oscillation, and the inlet valve moves further beyond neutral position 414 than in the nominal pressure case. As a result of the decreased time that inlet valve 401 is open, the initially higher-pressure working fluid experiences a greater degree of expansion. With careful design, it has been found that a factor of approximately two greater working fluid pressure beyond normal can be accommodated by this embodiment without significantly altering the phase that the outlet valve opens and thus without loss of nearly optimal expansion efficiency.
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.
Third Example Embodiment
A third embodiment, shown in FIGS. 11 through 13, exploits the use of a wobble-piston 260, known in the art. The wobble-piston comprises a piston head 266 and a connecting rod 265 that are either rigidly attached or cast as a single unit. At one end of the connecting rod a crank pin 267, mounted off-center on eccentric drive 286 moves in a circle about the center of crankshaft 287. The piston head 266 is kept centered within cylinder 261 by a flexible ring, flange or cup 264 that both seals working fluid within expansion chamber 262 and provides a low-friction bearing between the piston and the cylinder. As the crankshaft turns, the wobble piston tilts back and forth as it moves up and down within the cylinder. The mechanism of the wobble piston described in this paragraph is known in the art.
In this embodiment of the harmonic engine, inlet valve 201 is in the form of a reed valve, shown from the side in FIG. 11, and from above in FIG. 12. Inlet valve 201 is open when in its fully relaxed position, as shown in FIG. 11, and closed when held down against inlet port 268 by the difference in pressure between inlet header duct 225 and the expansion space 262. Outlet valve 204 is also in the form of a reed, also normally open in its fully relaxed position. The outlet valve is closed when held up against outlet port 263 by either spring forces or the difference in pressure between outlet header duct 205 and the expansion chamber 262.
With the very low mass characteristic of reed valves, it is preferred to incorporate a latching mechanism, as shown in FIG. 13, that prevents the outlet reed valve 204 from closing prematurely in the face of the aerodynamic force of the outrushing working fluid from expansion space 262 past the outlet valve 204 and through the outlet port 263 to the exhaust header duct 205. The outlet valve latch 202 is configured to allow the outlet valve to easily pass the latch position while moving downwards, but requires a greater force to become released while moving upwards. This greater force is provided by outlet valve closing rigid protrusion 208 that presses against the outlet reed valve 204 as the wobble piston moves towards TDC. A rigid protrusion 209 on the piston presses against the inlet reed valve 201 to force it open just at TDC.
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.
Operation of Third Example Embodiment
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.
Timing of Third Example Embodiment
A timing diagram for the wobble-piston embodiment is displayed in FIG. 14. Curves in this figure show the relative positions of the inlet valve, the outlet valve and the center of the top of the wobble-piston. The pressure 270 within expansion chamber 262 is also displayed as a function of the crankshaft angle. Starting at 0° in this cycle, the wobble piston motion is such that the phasing of the opening of inlet valve 201, shown by arrow 243, may be designed to coincide precisely with the TDC position of the wobble piston. In contrast to a purely axially moving piston, in which the instantaneous velocity of the piston vanishes at TDC, for the wobble-piston, the instantaneous velocity of the right hand side, bearing the inlet valve opening protrusion, does not vanish at TDC. This important distinction allows precise and reliable timing of the opening of the inlet valve. Then, in the initial portion of the inlet valve opening cycle, just after TDC, the inlet valve is forced, by the continuing upward motion of inlet valve launching protrusion 209, to continue opening even against the strong aerodynamic force of the rapid inlet flow of high pressure working fluid. This turbulent aerodynamic flow could otherwise interact with the opening of a low mass inlet reed valve in a deleterious way. Once the expansion chamber is filled with high pressure working fluid, the force of the aerodynamic flow greatly lessens, and inlet reed valve 201 completes a cycle of oscillation without significant aerodynamic counter forces until just before closing. Immediately after inlet valve 201 opens, the pressure 270 rapidly increases to the level 213 of the pressurized supply of working fluid. While the inlet valve is far from its closed position, flow between inlet header duct 225 and expansion chamber 262 is relatively unrestricted and the expansion chamber pressure remains near the supply pressure. As the inlet valve returns to near its closed position, not only is the instantaneous velocity of piston 260 substantially negative, thus causing a rapid flow of working fluid past the inlet valve, but the area of the aperture, defined by the position of inlet valve 201, between inlet header duct 225 and expansion chamber 262 becomes small, leading to a pressure drop between the inlet header pressure and the expansion chamber pressure. By design, this pressure drop at the moment the inlet valve reaches its closed position, indicated by arrow 240 in FIG. 14, is sufficient to hold the inlet valve closed. After the inlet valve is closed, the pressure in the expansion chamber drops until the point indicated by arrow 245 where the pressure has reached the level indicated by arrow 217 for which the pressure differential across the outlet valve just matches the spring force tending to open the outlet valve. Once the outlet valve begins to open, this pressure differential rapidly decreases until it reaches the level indicated by arrow 216 corresponding to equalization with the pressure in the exit header duct 305, as shown in FIG. 14, and the outlet valve rapidly accelerates. Then, at the point indicated by arrow 247, the outlet reed passes latch 247, and may overshoot its designed open position 215, as shown in FIG. 14. However, by the design of latch 202, outlet reed 204 has insufficient momentum to be able to pass by the latch in the upward direction and is brought to a stop at the point indicated by arrow 248. The outlet valve remains open until the point indicated by arrow 246 that the outlet valve closing protrusion 208 mounted on the left hand side of the wobble piston encounters outlet reed valve 204, and the outlet reed begins to accelerate towards its closed position. Since the apogee of the left-hand side of the wobble-piston occurs before TDC, the outlet valve is readily forced closed by protrusion 208 at a phase prior to TDC, as shown by arrow 241. By virtue of the increase in the pressure shown by numeral 244 in FIG. 14, above the level 217, at the point indicated by double arrow 241, the outlet reed is held closed until TDC is reached, and the inlet reed valve is forced open and the cycle can begin again.
Irreversibility of Third Example Embodiment
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.
Aircraft Embodiment
An especially lightweight and efficient embodiment of the harmonic engine especially useful in the context of an aircraft engine, is shown in FIG. 15. Here a propeller 390 driven by eccentric drive 386 of a wobble-piston 360 not only serves in place of the flywheel for the harmonic engine by virtue of its natural large moment of inertia, but also is directly powered by the engine to provide aircraft propulsion. The use of a reed inlet valve 301, opened by a launching spring 309 and a reed outlet valve 304, closed by spring 308 provides such a desirably low-mass engine. Outlet valve stopper 302 implemented as a small protrusion on cylinder 361 is also very simple and lightweight. In the aircraft case, the exhaust duct 305 vents directly to ambient air, while the inlet header duct 325 is connected to a throttle valve 392 that provides a supply of gas from high-pressure gas cylinder 391. For small model aircraft, cartridges of CO2 provide a readily available and convenient source of high-pressure gas.
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.
Single Acting Harmonic Uniflow Engine
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. FIG. 16 shows the single acting harmonic uniflow engine in a stationary state with the inlet valve in its relaxed equilibrium position. The harmonic uniflow engine is shown comprising an expander cylinder 761 having an upper inlet at an inlet end and a lower outlet at a removed location from the inlet end. The engine includes a piston head 760, axially slidable and reciprocable in the expander cylinder in a coaxial direction of the cylinder axis, and together with the upper portion of the cylinder enclosing an expansion chamber 762U accessible by the inlet. The lower portion of the cylinder and the piston head together enclose a lower chamber 762L. If the exhaust is directly to the atmosphere, neither the exhaust manifold nor the lower chamber is required. Such an embodiment is shown in FIG. 24. The engine also includes an intake header 725 in fluidic communication with the inlet for channeling working fluid from a pressurized fluid source into the expansion chamber. A harmonic inlet valve 701 is provided for controlling flow of high-pressure working fluid into expansion chamber 762U from the intake header. A number of outlet ports 711 are provided for allowing flow out of expansion chamber 762U to exhaust manifold 705 when piston head 760 is reciprocated to a venting position. In particular, the outlet ports are positioned adjacent the venting position of the piston head, e.g. below the top of vent ports 711, as shown in FIG. 19, so that the outlet ports are uncovered and caused to be fluidically connected to the expansion chamber as the piston head passes the ports en route to the venting position. In this manner, the piston head operates to control the periodic venting of working fluid out from the expansion chamber. A crank pin 788 mounted on flywheel 789 is attached through a bearing at the big end of connecting rod 785 while cross head 754 is connected at the small end of the connecting rod. The cross head is connected to the piston head through piston rod 795 for converting reciprocating motion of piston head 760 into rotary motion of flywheel 789. Flywheel 789 is in turn operably connected to motor/generator 755. Piston rod 795 passes through the lower end of cylinder 761 through seal 796. In the single acting configuration, the maximum temperature that seal 796 experiences during the course of the engine cycle is much less than the inlet temperature of the working fluid, and thus a much wider variety of seal materials can be used in this application. Flexible rings, flanges or cups 764 attached to piston head 760 seal working fluid within the expansion chamber and provide a low-friction dry bearing between the piston head and the cylinder. Alternatively, conventional piston rings, as known in the art, can be used.
Inlet Valves
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.
Outlet Ports
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.
Normal Uniflow Engine Operation
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 FIG. 20 as a function of the crankshaft angle. The pressure within the expansion chamber is indicated by dashed line 770. The normal engine cycle consists of a pair of strokes of the piston head, starting at the TDC position that corresponds to an angle of 0° for the crankshaft, with the piston head at one extreme of its motion nearest the top of cylinder 761. There is first a downward power stroke to the opposite extreme piston head position at BDC, corresponding to 180° of crankshaft rotation. After passing BDC there follows an upward recovery stroke back to the beginning point of the cycle at TDC. Under steady operation, the motion of the piston head and valves are strictly periodic, and every cycle is nominally identical.
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 FIG. 17. Expansion chamber 762U is at its minimum volume point. This minimum volume is preferably made as small as practical, in consideration of dimensional tolerances and due accommodation for thermal expansion. Lower chamber 762L is at its maximum volume point, and in fluid communication with exhaust manifold 705 through vent ports 711. Piston head 760 is instantaneously in a state of zero velocity as it is in the process of turning around. Inlet valve 701 is partially open and valve head 703 is moving upwards, as indicated by the arrow in FIG. 17, away from its seat and piston head 760. Working fluid within expansion chamber 762U at the completion of the previous cycle is left approximately equal to that of the supply pressure in the inlet header duct 725.
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 FIG. 20 to a maximally open position. In the maximally open position, the inlet valve is instantaneously at rest before it returns towards its closed position. As the inlet valve completes the second half of its oscillating motion, it again passes through its neutral position, and finally, by virtue of the high Q of the inlet valve harmonic oscillation and the choked flow between inlet valve head 703 and seat 702, arrives and remains at the “closure point” indicated by arrow 740 in FIG. 20.
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 FIG. 20, the pressure in the expansion chamber drops approximately adiabatically as the volume of the working fluid within the chamber increases. This adiabatic expansion is characteristic of normal speed, high thermal efficiency operation of the engine. The configuration of the engine components at 90° of crankshaft angle is shown in FIG. 18, with inlet valve 701 shown in solid lines. (Another configuration of the inlet valve is shown in dashed lines, as is relevant for alternative modes of operation to be discussed later.)
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 FIG. 20. The total area of outlet ports is designed to be sufficient to allow the working fluid pressure to drop to substantial equality with the exhaust pressure before the piston head rises to the point that the exhaust ports are once again covered, at the phase indicated by arrow 746. The configuration of the components at BDC is shown in FIG. 19. As the piston head rises from BDC, and after vent ports 711 have again been covered, the expansion chamber pressure increases. By virtue of the nearly vanishing clearance space in the harmonic uniflow engine, there is a sharp spike in working fluid pressure, located near TDC, as indicated by arrow 747 that forces the inlet valve from its seat, as shown by arrow 743, to initiate a single cycle of oscillation. Because the pressure spike releases the inlet valve, it is automatically assured that the recompression of working fluid does not greatly exceed the supply pressure. After this, the engine cycle repeats from TDC again. With the uniflow embodiment, having a large ratio between the supply pressure and the exhaust pressure assures that the opening of the inlet valve occurs immediately before TDC, with only a slight dependence on the pressure ratio.
Low Speed Uniflow Engine Operation
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 FIG. 21 as a function of the angle of the crankshaft. The pressure within the upper expansion chamber is indicated by dashed line 780. For most of the power stroke, the inlet valve is harmonically oscillating about the equilibrium position indicated by dashed line 751, but the amplitude of oscillation is such that the inlet valve head does not reach the closed position indicated by dashed line 753, and thus the inlet valve is effectively in the OPEN state. The high quality Q factor of the oscillating inlet valve is such that each of the oscillations closely approaches the closed position. The quantitative boundary between “low speed” and normal speed operation is determined by the threshold speed at which in-flowing working fluid is unable to force the inlet valve to close after a single cycle of oscillation, and thus the inlet valve continues to oscillate about its equilibrium position, as shown in FIG. 21. As piston head 760 descends from its uppermost position, the pressure in the expansion chamber remains at the supply pressure level, while the inlet valve remains effectively open, until the point, indicated by arrow 758, at which the outlet ports are uncovered. Once the outlet ports are uncovered by the descent of the piston head, an extremely rapid flow of working fluid ensues, flowing from the supply source, through the open inlet valve, through the inlet, out through the exposed ports and finally out through the exhaust manifold. This rapid flushing flow of working fluid causes the inlet valve to close, as indicated by arrow 755, and once closed, the pressure within the expansion chamber rapidly drops even more rapidly to the level of the exhaust manifold pressure. On the upstroke, the pressure in the expansion chamber follows the same process as shown in FIG. 20 for normal speed operation. Because the average pressure in the expansion chamber on the power stroke is much greater than the average pressure on the recovery stroke, the torque produced at low speed by the harmonic uniflow engine is quite high. Indeed, the indicated MEP, or mean effective pressure, for low speed engine operation is nearly equal to the difference between the supply pressure and the exhaust pressure.
High Speed Uniflow Engine Operation
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 FIG. 22 as a function of the angle of the crankshaft. The pressure within the expansion chamber is indicated by dashed line 790. With the natural resonance period of the inlet valve being greater than the travel time of the piston head from TDC to BDC, at 900 of crankshaft phase, the position of inlet valve 701 may be as shown by the dashed representation of the inlet valve in FIG. 18. The function of range of motion limiter 726 is to prevent inlet valve 701 from being excessively stressed under high-speed operation. Without this range of motion limiter, under extreme high-speed operation, flexure 707 could potentially be strained beyond its breaking point. As piston head 760 descends from its uppermost position, the pressure in the expansion chamber remains near the supply pressure level, while the inlet valve remains open, until the point, indicated by arrow 758, at which the outlet ports are uncovered. Once the outlet ports are uncovered by the descent of the piston head, an extremely rapid flow of working fluid ensues, flowing from the supply source, through the open inlet valve, through the upper inlet, out through the exposed ports and finally out through the exhaust manifold. This rapid flow of working fluid causes the inlet valve to close, as indicated by arrow 755, and once closed, the pressure within the expansion chamber rapidly drops to the level of the exhaust manifold pressure. Because the average pressure in the expansion chamber on the power stroke is much greater than the average pressure on the recovery stroke, the torque produced at high speed by the harmonic uniflow engine is also quite high, and comparable to the torque produced at low speed. At either extreme of speed, the mean effective pressure for the harmonic uniflow engine is nearly as great as the working fluid supply pressure, while at normal speed, the thermal efficiency can be nearly equal to the ideal thermodynamic efficiency characteristic of the working fluid used in the engine.
Harmonic Engine Stability
By virtue of the changes in the operation of the harmonically acting inlet valves as a function of speed, as shown in FIGS. 20, 21 and 22, the harmonic engine mean torque per cycle varies as a function of operating speed as shown by solid line 710 in FIG. 23 for a given working fluid inlet supply pressure and exhaust back pressure. At both very low and very high speed, the engine torque is high, with an indicated MEP approaching the difference between the inlet supply pressure and the exhaust pressure, as described earlier. For normal operating conditions, between the high and low speed regimes, the harmonic engine torque increases approximately linearly with speed. The reason for this linear increase in torque with speed has to do with the fact that, with a fixed natural resonance period for the harmonic inlet valve, the time that the inlet valve is open remains approximately constant, while the phase interval that this time represents increases linearly with operating speed. As a result, the pressure in the cylinder remains high for longer phase duration, thus leading to more work done per cycle, and thus higher torque. The torque continues to increase until the speed at which the inlet valve is still open at the phase that the outlet ports are uncovered. For speeds even greater than this, eventually the higher working fluid flow speeds lead to pressure drops through the inlet and pressure increases through the outlet that decrease the work per cycle, and the torque decreases.
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 FIG. 23. It is generally the case, as is known in the art, that an electric induction motor, acting as a generator, has a load torque that increases quite rapidly with increasing generator speed beyond the generator's synchronous operating speed. It is for this reason that an electric induction motor used as motor/generator 755 is so well suited for use with the harmonic engine. A particularly advantageous configuration is the use of an induction motor/generator controlled by a variable frequency drive controller 756, as shown in FIG. 24, in which the synchronous frequency of the induction machine is adjusted by the frequency output from the controller. Just such variable frequency drive controllers are used in electric cars that feature regenerative braking. In the known art of the electric car, electrical energy is generally stored in a rechargeable battery, and when the accelerator pedal is pushed, DC electric power is converted to a variable frequency three-phase alternating current waveform that flows to a three-phase induction motor to propel the car. When the brake pedal is pushed, kinetic energy of motion instead drives the three-phase induction motor above the synchronous speed of the controller to generate electric power. The electric car controller then rectifies this alternating current to direct current that recharges the battery. In the harmonic engine case, the dynamic braking feature of the electric car controller is used to generate electric power, which may be three-phase alternating current, or single-phase alternating current, or direct current depending on the nature of the variable frequency controller. The harmonic engine operating speed may thus be adjusted electronically by changing the controller's frequency, so that any point along the torque curve 710 can be used as a stable operating point. The controller drive startup of the harmonic engine is exactly analogous to the controller driven acceleration of an electric car. Furthermore, by appropriately choosing the direction of rotation of the magnetic fields in the three-phase induction motor/generator 755, startup may be in either forward or reverse directions, and by virtue of the symmetry of the harmonic engine operation in this embodiment, its performance in either forward or reverse is equally good. In addition, by electronically changing the direction of rotation of the magnetic fields in motor/generator 755, this embodiment of the harmonic uniflow engine while operating at full speed in one direction may be slowed down, reversed in direction and sped up to full speed in the opposite direction, entirely by command of controller 756.
Double Acting Harmonic Uniflow Engine
FIG. 25 shows the double acting harmonic uniflow engine in a stationary state. Here the numbering of the similar elements corresponding to the single acting embodiment is generally shown with a number 100 less for each element. This embodiment is shown comprising an expander cylinder 661 having an upper inlet and a lower inlet. The engine also includes a piston head 660, having an upper face 660U and a lower face 660L, axially slidable in the expander cylinder, with upper face 660U and cylinder 661 together enclosing an upper expansion chamber 662U accessible by the upper inlet and with lower face 660L and cylinder 661 together enclosing a lower expansion chamber 662L accessible by the lower inlet. Also the engine includes an intake header 625 in fluidic communication with both the upper inlet and the lower inlet for channeling working fluid from a pressurized fluid source into the upper and lower expansion chambers. An upper harmonic inlet valve 601U is provided for controlling flow of high-pressure working fluid into upper expansion chamber 662U from the intake header, and lower harmonic inlet valve 601L is provided for controlling flow of high-pressure working fluid into lower expansion chamber 662L from the intake header. A number of outlet ports 611 are provided for allowing flow out of upper expansion chamber 662U to exhaust manifold 605 when upper piston face 660U is below the top of vent ports 611, and also for allowing flow out of lower expansion chamber 662L to exhaust manifold 605 when lower piston face 660L is above the bottom of vent ports 611. A crank pin 688 is connected at the big end of connecting rod 685 while cross head 654 is connected at the small end of the connecting rod. The cross head is connected to the piston head through piston rod 695 for converting reciprocating motion of piston head 660 into rotary motion of flywheel 689. Flywheel 689 may in turn be operably connected to motor/generator 655. Piston rod 695 passes through the lower end of cylinder 661 through seal 696. Flexible piston rings, flanges or cups 664 attached near the top and bottom ends of piston head 660 both seal working fluid within the upper and lower expansion chambers and provide a low-friction bearing between the piston head and the cylinder.
Resonant Ratio of Upper and Lower Inlet Valves
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.
Outlet Ports
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.
Normal Double Acting Uniflow Engine Operation
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 FIG. 26 as a function of the crankshaft angle. The pressure within the upper expansion chamber is indicated by dashed line 670. The pressure within the lower expansion chamber is indicated by dotted line 672. The normal engine cycle consists of a pair of strokes of the piston head, starting at the TDC position that corresponds to an angle of 0° for the crankshaft, with the piston head at one extreme of its motion nearest the top of cylinder 661, there is first a downward power stroke to the opposite extreme piston head position at BDC. After passing BDC there follows an upward recovery stroke back to the beginning point of the cycle at TDC. The power stroke of the upper section corresponds to the recovery stroke of the lower section, and vice versa. Under normal operation, the motion of the piston head and valves are strictly periodic, and every cycle is nominally identical.
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 FIG. 26 to a maximally open position. In the maximally open position, the inlet valve is instantaneously at rest before it returns towards its closed position. As the inlet valve completes the second half of its oscillating motion, it again passes through its neutral position, and finally, by virtue of the high Q of the inlet valve harmonic oscillation and the choked flow between inlet valve head 603U and seat 602, arrives and remains at the “closure point” indicated by arrow 640 in FIG. 26. Note that the period of time that the upper inlet valve is open is less than the period of time that the lower inlet valve is open, as is shown by the greater width of the “OPEN” bump for 601L compared to 601U. This difference between the upper and lower systems enables having equal pressure vs. volume expansion curves for the upper and lower chambers, and thus smoother operation and higher efficiency of the engine.
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 FIG. 25 that the upper and lower expansion volumes are quite different, even though the crank angle is halfway between TDC and BDC. At the moment that the upper surface 660U of the piston head first clears outlet ports 611, with the pressure in expansion chamber 662U being greater than the pressure in exhaust manifold 605, flow of working fluid from the upper expansion chamber to the exhaust manifold causes a rapid drop of the working fluid pressure in the upper expansion chamber towards the exhaust manifold pressure. This point in the cycle is indicated by arrow 645 in FIG. 26. The total area of outlet ports is designed to be sufficient to allow the working fluid pressure to drop to substantial equality with the exhaust pressure before the piston head rises to the point that the exhaust ports are once again covered, as indicated by arrow 646. At a time that may be either earlier or later than the uncovering of the outlet ports, as the compression of working fluid in lower expansion chamber 662L causes the pressure to exceed the supply pressure in the inlet header 625, there is a sharp pressure spike indicated by arrow 644 in FIG. 26, that causes lower valve 601L to open. Immediately after starting to open, the pressure spike dissipates, and this removes the pressure force driven acceleration of the lower inlet valve head 603L. As a result, the lower inlet valve ends up being “gently” released from its seat, and can undergo a single harmonic oscillation before returning to its seat, and being held closed, just as described above for the upper inlet valve. As the piston head rises from BDC, the upper expansion chamber pressure increases, with a sharp spike indicated by arrow 647 that releases the upper inlet valve to initiate a single cycle of oscillation. After this, the engine cycle repeats from TDC again. With the uniflow embodiment, having a large ratio between the supply pressure and the exhaust pressure assures that the opening of the upper inlet valve occurs immediately before TDC, and the opening of the lower inlet valve occurs immediately before BDC.
Atmospheric Harmonic Uniflow Engine
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 FIG. 24.
It is appreciated that the single acting Harmonic Uniflow Engine may operate with a wobble piston mechanism of the form shown in FIG. 11. It is appreciated that there may be a protrusion on the inlet valve or on the piston head, as described for earlier embodiments. It is appreciated that inlet valve head 703 could be a poppet valve, with resiliently biasing member 707 being a coiled spring. It is appreciated that any outlet port control or outlet valve mechanism that closes, or restricts the flow of working fluid out from the expansion chamber on the recovery stroke as the piston head approaches TDC, and thus causes a pressure rise sufficient to initiate the opening of the normally open, resiliently biased inlet valve, can suffice to produce timing phase diagrams for the inlet valve similar to those shown by the lines labeled 601U or 601L in FIG. 26 or the line labeled 701 in FIG. 20, and thus enable a pressure driven engine according to the present invention.
Spherical Wedge Piston Harmonic Uniflow Engine
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 FIGS. 27 through 30. The numbering of the corresponding components in these figures is generally 200 less than the numbering of the components in FIG. 16. This embodiment is shown comprising a circular cylinder 561 with central axis 522 having a port opening 502 (i.e. a first port) at its upper end and is open (i.e. has a second port) at its lower end. The harmonic valve is comprised of resilient flexure 507 and head 503. The engine also includes a piston head 560, whose outer sealing surface has the shape of a pair of opposing spherical lunes, with the left spherical lune surface 523 and the right spherical lune surface 524 joined along their common diameter 520, i.e. a pivot axis, and that has its center located on the cylinder axis 522, and is slidable in the cylinder and rigidly attached to the small end of connecting rod 595, with the piston head upper surface and cylinder 561 inner surface together enclosing a chamber 562 accessible by port opening 502. Also the engine includes a manifold 525 in fluidic communication with the port opening for channeling working fluid from a pressurized fluid source into the chamber while operating as an expander. A crank pin 588 is connected to a bearing at the big end of connecting rod 595 for converting the reciprocating and tilting motion of piston head 560 into rotary motion of the crankshaft. As crank pin 588 follows a circular orbit, the rigid connection of rod 595 to piston head 560 causes the piston head to tilt back and forth through the course of the engine cycle with the pivot axis coinciding with and identical to diameter 520. Pivot axis 520 is constrained to be perpendicular to and to pass through cylinder axis 522, so that the intersection point of pivot axis 520 and cylinder axis 522 follows a purely axial motion along the centerline of cylinder 561 over the course of an engine cycle. This intersection point has a range of reciprocating motion between its uppermost position illustrated in FIG. 28 and its lowermost position illustrated in FIG. 30. Pivot axis 520 remains parallel to the axis of the crankshaft throughout the engine cycle. An induction motor/generator 555 is preferably directly connected to the crankshaft. Although an asynchronous induction motor/generator 555 is preferred here, by virtue of the generally very small variation in operating speed of induction machines whether they are motoring or generating, and the ease of controlling them, it is also feasible to use synchronous motor/generators as well. It is appreciated that while FIGS. 27-30 particularly show the piston head embodied as a pair of opposing spherical wedges, the piston head may be hollow, or of more complex shape, as long as the outer load bearing and sealing surface of the piston head contains left spherical lune surface 523 and right spherical lune surface 524.
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 FIG. 29. Here, the circular seal lies within plane 527 perpendicular to cylinder axis 522 and containing pivot axis 520. This circular seal coincides with the lower edge of the right spherical lune surface 524 of piston head 560, and the upper edge of the left spherical lune surface 523. At both 0° (in FIG. 28) and 180° (in FIG. 30) of crankshaft rotation, the circular line of contact coincides with the upper edge of the right spherical lune surface 524 and the lower edge of the left spherical lune surface 523. At 0° (in FIG. 28) the circular seal lies within plane 521 perpendicular to cylinder axis 522 and containing pivot axis 520. At 180° (in FIG. 30) the circular seal lies within plane 526 perpendicular to cylinder axis 522 and containing pivot axis 520. For crankshaft angles between 0° and 180°, the circular seal remains in a plane perpendicular to cylinder axis 522, and lies at intermediate positions between planes 521 and 526. This continuous contact between the spherical lunes and the cylinder bore maintains a seal of the chamber for the full duration of the power stroke.
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 FIG. 27 at 270° of crankshaft rotation, as two gaps develop between the surface of the piston head and the inner wall of the cylinder. These crescent shaped gaps allow expanded working fluid to exhaust from chamber 562, as shown by the arrows in FIG. 27. As the piston head pivot axis approaches its uppermost position at 360° of crankshaft rotation, the vanishing crescent area tends to cause the pressure of working fluid within chamber 562 to increase, and as described for earlier embodiments, with sufficient engine speed, this pressure rise forces the valve to open as 360° of crankshaft rotation is reached, as indicated by the position of valve head 503 and reed 507 in FIG. 28. Alternatively, at low speed having insufficient pressure rise before the uppermost position of the pivot axis is reached, the valve may instead remain closed at 0° of crankshaft angle. However, in the cases that the inlet valve is closed at 0°, very shortly afterwards, the continued rise of the right hand spherical piston head segment optionally may or may not (depending on whether or not mechanical contact between the piston head and the inlet valve head is designed to occur at some point in the crankshaft rotation cycle) assure that piston head 503 is nudged away from its closed position, just as was discussed earlier in connection with the wobble piston case shown in FIG. 11. With the choice of geometry shown in FIG. 27 through 30, there is sufficient protrusion of the surface of valve head 503 through port 502 that it will assure that it is opened for the power stroke. Also as described for earlier embodiments, although the time that the harmonic valve remains open is governed by the natural resonance period of the harmonically oscillating valve, the range of crankshaft rotation angle corresponding to this fixed period varies linearly with the speed of operation of the engine. This leads to a mean torque per cycle that varies as shown in FIG. 23 as described earlier.
The shape of the ceiling of cylinder 561, i.e. the cylinder inner wall surface above plane 521 in FIG. 28, is contoured so that at no time during the reciprocation of piston head 560 is there any contact with the cylinder ceiling. It is, however, preferable that the ceiling contour be chosen to be no higher than necessary to avoid contact with the piston, in order to minimize the clearance volume of chamber 562 when the piston is at its TDC position, as shown in FIG. 28.
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 FIG. 21.
Spherical Wedge Piston Operation as Compressor
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 FIG. 24. Without change in the structure of the Spherical Wedge Piston Harmonic Engine, but merely by rotating the crankshaft in the reverse direction, this embodiment functions as a compressor rather than an expander. This is illustrated by the sequence of FIGS. 31, 32, 33, and 34, corresponding to successive 90° of rotation of the crankshaft while being driven counterclockwise by motor/generator 555. On the entire down stroke from TDC to BDC, the crescent shaped gaps between the walls of cylinder 561 and the surface of piston head 560 allow working fluid to be drawn in to chamber 562. The widest gaps are seen in FIG. 31 at 90° after TDC, and the upward arrows illustrate the flow of inrushing working fluid through the gaps. On the entire up stroke from BDC to TDC, the continuous contact between the spherical lune surface of the spherical wedge piston head and the cylinder walls maintains a seal for the compression stroke. As the volume within chamber 562 decreases in this mode of operation, the pressure of expansible working fluid within the chamber increases, until it closely approaches the pressure in the manifold on the opposite side of valve head 503. Once the pressure in chamber 562 is sufficiently high, flexure 507 and valve head 503 move to the open position illustrated in FIG. 34, and compressed working fluid leaves manifold 525 as indicated by the leftward directed arrow. In contrast to the engine case, the time that the harmonic valve remains open is not determined by the natural resonance period of the valve, but instead the closing of the valve happens in response to the reversal of the flow direction of working fluid. In short, the harmonic valve in this case functions as a check valve. Just after passing TDC, as the volume of chamber 562 begins to increase, the reversal of the pressure differential across the valve forces it to close, even against the biasing force of flexure 507, and for the remainder of the down stroke, working fluid flows through the crescent shaped gaps between piston head 560 and the inner walls of cylinder 561, into chamber 562.
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 FIG. 27 through 30, the harmonic valve meters the inflow of working fluid for a time that is determined by the natural resonance period of the valve capped by the engine half-cycle period, but that when operating as a compressor, corresponding to the counter-clockwise rotation of the crankshaft shown in FIG. 31 through 34, the harmonic valve instead functions as a check valve to readily allow the outward flow of compressed working fluid but prevents the inward return flow of high pressure working fluid to the expansible chamber.
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.

## Claims

1. 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.

2. The harmonic uniflow engine of claim 1,
wherein the outlet is positioned adjacent the venting position of the piston head so that the piston head passes and uncovers the outlet to the expansion chamber en route to the venting position.

3. The harmonic uniflow engine of claim 2,
wherein the outlet comprises a plurality of ports.

4. The harmonic uniflow engine of claim 1,
wherein the resiliently biasing member of the inlet valve is a mono-leaf spring cantilevered with the inlet valve head connected at one end.

5. The harmonic uniflow engine of claim 1,
wherein the periodic return means for effecting the return stroke of the engine after each power stroke is a crank assembly having a crankshaft operably connected to the piston head and a flywheel connected to the crankshaft to transfer rotational inertia to the piston head via the crankshaft.

6. The harmonic uniflow engine of claim 5,
further comprising an induction motor operably connected to the crankshaft and capable of drawing power from an electric energy supply to drive the engine, or supplying power back to the electric energy supply when driven by the engine.

7. The harmonic uniflow engine of claim 1,
wherein the outlet is at an outlet end of the cylinder opposite the inlet end, and 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 at the venting position 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.

8. The harmonic uniflow engine of claim 7,
wherein the pair of pivotable ends comprise a pair of spherical-surface sidewall sections each having a radius of curvature substantially equal to a radius of the cylinder, with a maximum distance between the sections substantially equal to a diameter of the cylinder.

9. The harmonic uniflow engine of claim 8,
wherein the piston head is adapted to rotate about the pivot axis as it reciprocates in the cylinder so that at upper and lower limits of a reciprocation range of the pivot axis the piston head is tilted about the pivot axis from a plane orthogonal to the cylinder axis.

10. 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.

11. The uniflow energy conversion system of claim 10,
wherein the pair of pivotable ends are a pair of spherical-surface sidewall sections each having a radius of curvature substantially equal to a radius of the cylinder, with a maximum distance between the sections substantially equal to a diameter of the cylinder.

12. The uniflow energy conversion system of claim 11,
wherein the piston head is adapted to rotate about the pivot axis as it reciprocates in the cylinder so that at upper and lower limits of a reciprocation range of the pivot axis the piston head is tilted about the pivot axis from a plane orthogonal to the cylinder axis.

13. 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.

14. The uniflow engine of claim 13,
wherein the pair of pivotable ends are a pair of spherical-surface sidewall sections each having a radius of curvature substantially equal to a radius of the cylinder, with a maximum distance between the sections substantially equal to a diameter of the cylinder.

15. The uniflow engine of claim 14,
wherein the piston head is adapted to rotate about the pivot axis as it reciprocates in the cylinder so that at upper and lower limits of a reciprocation range of the pivot axis the piston head is tilted about the pivot axis from a plane orthogonal to a cylinder axis.

16. 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.

17. The uniflow compressor of claim 16,
wherein the pair of pivotable ends are a pair of spherical-surface sidewall sections each having a radius of curvature substantially equal to a radius of the cylinder, with a maximum distance between the sections substantially equal to a diameter of the cylinder.

18. The uniflow compressor of claim 17,
wherein the piston head is adapted to rotate about the pivot axis as it reciprocates in the cylinder so that at upper and lower limits of a reciprocation range of the pivot axis the piston head is tilted about the pivot axis from a plane orthogonal to the cylinder axis.