Low Temperature High Efficiency Condensing Heat Engine for Propelling Road Vehicles

Abstract

A non-polluting, closed-cycle condensing heat engine and operating method is provided for propelling road vehicles at high efficiencies and high power densities by using a phase-changing working fluid having a critical temperature close to the natural ambient temperature of the surrounding atmosphere and shifting the high temperature heat reservoir downward by several hundred degrees by creating an artificial low temperature heat reservoir below ambient temperature by evaporating water. By isentropically compressing the liquefied working fluid at sub-ambient temperatures to very high pressure utilizing the fact that water has an unusually high latent heat of evaporation, and heating it to a compressed gas at a relatively low temperature in the high temperature heat reservoir by burning small amounts of fuel, it is possible for the engine to operate at high power densities by expanding the compressed gas back to the initial sub-ambient temperature where it is re-condensed to propel road vehicles several hundred miles on a tank of water holding only 40 gallons using a small fraction of the amount of fuel used by vehicles propelled by conventional internal combustion engines.

Description

BACKGROUND

It is a well known fact that the internal combustion engine used for propelling road vehicles has an energy efficiency of only about 12%. Unfortunately, this low efficiency is inherent in the basic design and operating principles of these engines. Most of the energy generated by the burning fuel (gasoline) is wasted to keep the engine from self-destructing. In order to understand this little known technical fact, it is important to understand the basic operating principles of these engines.

An internal combustion engine used for propelling conventional passenger-carrying automobiles consists primarily of a plurality of stationary cylinders containing a moving piston which form combustion chambers for generating compressed gas of variable volume. Both of these parts are constructed of metal. Since the gas temperatures generated inside the cylinders by the burning fuel are well above the ability of un-cooled metals to withstand, the cylinders of an internal combustion engine must be cooled by transferring a large amount of the heat of combustion through the cylinder walls. This is accomplished by surrounding the cylinders with a jacket of cooling water, and a system for cooling the water by circulating it through a radiator. The result is a huge waist of the heat energy generated by the burning gasoline while simultaneously generating enormous amounts of pollution that is toxic to life and harmful to the environment. Since the exhaust products have a temperature of about 400° F. (478° K), this represents another reason why internal combustion engines have inherently low efficiencies.

This inherently low operating efficiency of internal combustion engines can be theoretically described by Carnot's heat equation giving the maximum possible thermodynamic efficiency η of any heat engine operating between a high temperature heat reservoir at temperature T_H and a low temperature heat reservoir at temperature T_L. This equation is

$\eta =\frac{T_H-T_L}{T_H}$

Since both of these temperatures are fairly high in internal combustion engines, this equation shows why they are inherently inefficient.

The engine disclosed herein is basically a closed-cycle, condensing heat engine where the temperatures of the high and low temperature heat reservoirs are shifted downward a significant amount such that relatively small amounts of input heat energy is required to maintain the high temperature heat reservoir at temperature T_H. This downward shift in the high and low heat reservoirs is made possible by creating an artificial low temperature heat reservoir at temperature T_L that is maintained below the temperature of the natural environment by evaporating water utilizing the fact that water has an unusually high latent heat of evaporation. The engine is made possible by using a working fluid having a critical temperature close to natural ambient temperature that can be condensed at a sub-ambient temperature by a heat sink generated by continuously evaporating small amounts of water. This will result in an engine having a thermal efficiency more than twice that of internal combustion engines, generates no toxic pollution, is completely silent, and provides a miles per gallon fuel efficiency that is an order of magnitude greater than most internal combustion engines used for propelling conventional passenger carrying road vehicles. And, as in the design of cryogenic engines, the engine is also capable of generating bursts of accelerating power far exceeding that of conventional internal combustion engines. Thus, the invention disclosed herein, as in the case of my more advanced engine described in my U.S. Pat. No. 6,739,137 B2 entitled, “Magnetic Condensing System For Cryogenic Engines,” issued May 25, 2004, is presented as an ideal low-cost substitute for internal combustion engines used for propelling road vehicles.

BRIEF DESCRIPTION OF THE INVENTION

A closed-cycle, low-temperature, non-polluting, condensing heat engine and operating method is provided for propelling road vehicles at high efficiencies and high power densities. The working fluid comprises a phase changing gas having a critical temperature close to the temperature of the natural atmosphere at ambient temperature, but above the temperature of evaporating water. This enables the engine to operate cyclically as a closed-cycle condensing heat engine where the high temperature heat reservoir is maintained at relatively low temperatures not far above ambient temperature by burning small amounts of a clean burning liquified natural gas (LNG) such as Propane, Butane, or Pentane. The low temperature heat reservoir (heat sink) is created artificially at sub-ambient temperatures by evaporating small amounts of water.

The basic thermodynamic operating steps consists of (1) compressing the liquefied working fluid discharged from the sub-ambient condenser isentropically to a high initial pressure; (2) feeding the compressed working fluid into a thermally insulated heating chamber where it is heated to relatively low temperatures by a clean burning fuel; (3) withdrawing the heated compressed working fluid from the heating chamber and feeding it into a large thermally insulated pressure vessel that serves as a load-leveling compressed gas energy storage system; (4) withdrawing the heated compressed working fluid from the load-leveling energy storage system at some desired mass flow rate {dot over (m)} depending upon the desired output power and feeding it into an isentropic multistage expansion system that generates the desired output power while simultaneously transforming the expanded gas into a saturated vapor at a temperature below the natural ambient temperature of the surrounding atmosphere; (5) feeding the expanded saturated vapor discharged from the expansion system into a thermally insulated condenser maintained at sub-ambient temperature by evaporating small amounts of water stored in a water tank where the expanded saturated vapor is re-condensed back into the liquid phase by transferring its latent heat of vaporization to the evaporating water; (6) withdrawing the condensed liquified working fluid from the condenser and feeding it into a thermally insulated liquified gas storage vessel at sub-ambient temperature; and (7) withdrawing the liquified gas from the storage vessel at some mass flow rate and feeding it back into the isentropic compressor system that re-compresses the condensed liquified working fluid back to the initial pressure and repeating the cycle.

The volume storage capacities of the compressed gaseous working fluid storage vessel, the liquified working fluid storage vessel, and the water storage vessel are sufficiently large to enable the engine to generate propulsive power to propel the vehicle for several hours without requiring any re-compression. The engine is designed such that the compression, heating, expanding, and condensing steps can operate independently of each other to maximize net power output while minimizing the rate water is consumed by taking advantage of the changing conditions of the surrounding atmosphere over a 24-hour period. The system is designed to operate automatically when the vehicle is parked to maintain the compressed gas storage system at maximum capacity.

DRAWINGS

These and other advantages and features of the invention will be apparent from the disclosure, which includes the specification with the foregoing and ongoing description, the claims and the accompanying drawings wherein:

FIG. 1 is a Psychrometric Chart for evaporating water in the open atmosphere giving the wet-bulb temperature for various dry-bulb temperatures corresponding to various values of the relative humidity (also known as percent saturation) illustrating how a sub-ambient low temperature heat reservoir can be created by evaporating water (Prior Art);

FIG. 2 is a schematic block diagram of the preferred embodiment of the invention illustrating its basic design and operating principles;

FIG. 3 is a Temperature—Entropy Diagram (TS Diagram) of a phase-changing working fluid used in the preferred embodiment of the invention describing its thermodynamic operating principles;

FIG. 4 is a schematic block diagram illustrating the design and operating principles of the engine's isentropic multistage expansion system;

FIG. 5 is a horizontal cross section through the thermally insulated compressed gas storage system that operates as a high-capacity, load-leveling, compressed gas energy storage system;

FIG. 6 is a schematic transverse cross section through the condenser that is maintained at sub-ambient temperatures by evaporating small amounts of water illustrating its basic design and construction;

FIG. 7 is a schematic transverse cross section through a vapor condensing tube mounted inside the condenser illustrating its internal rectangular geometry and external water cooling pads;

FIG. 8 is a longitudinal cross section through the condensing tube and cooling pads further illustrating the condenser's design and construction;

FIG. 9 is a schematic transverse cross section through the high-temperature heating chamber where the compressed gas is heated by burning small amounts of combustible fuel before it is fed into the thermally insulated compressed gas storage vessel illustrating its design and construction;

FIG. 10 is a schematic longitudinal cross section through the high-temperature heating chamber further illustrating its design and construction; and

FIG. 11 is a schematic longitudinal cross section of an automobile propelled by the invention illustrating the positions of the various engine components.

DESCRIPTION OF THE PREFERRED EMBODIMENT

The invention presented herein is a silent, non-polluting, closed-cycle condensing heat engine for propelling road vehicles operating at low temperatures designed to enable essentially all of the input heat energy to be absorbed by the working fluid to achieve a fuel efficiency (miles per gallon of fuel consumed) far exceeding that of conventional internal combustion engines that are currently used for propelling passenger carrying automobiles. This is achieved by shifting the operating temperatures significantly downward such that the low temperature heat reservoir is at a sub-ambient temperature that is maintained by evaporating small amounts of water utilizing the fact that water has an unusually high latent heat of vaporization. This enables the temperature of the high temperature heat reservoir to be shifted significantly downward that can be maintained by burning small amounts of a clean burning fuel where almost all of the heat of combustion is used for heating the working fluid. All of these operating features are made possible by the creation of the artificial low temperature heat reservoir achieved by evaporating small amounts of water. Thus, one of the most important components of the invention is the sub-ambient condensing system.

In order to understand how the artificial sub-ambient low temperature heat reservoir is created it is important to understand the underlying physics and operating features that distinguishes it from those used in conventional prior art condensing heat engines. Since the low temperature heat sink used by the present engine disclosed herein is created artificially by evaporating small amounts of water, it is important to understand the physical processes that take place during the evaporation of water that make it possible to create this artificial low temperature heat sink where a significant amount of latent heat can be extracted from a suitable working fluid so that it can be condensed into a liquid at sub-ambient temperatures without expending any mechanical work.

The temperature of a given quantity of water is determined by the average kinetic energy of all the water molecules in the water. There will always be some molecules that move faster than others. If they are near the surface (i.e., the boundary separating the air and water) and have sufficient velocity to overcome the surface tension of the water, they will leave the water and enter the air as gaseous vapor. Thus, if the surface area of the air/water boundary is large relative to the amount of water, a large fraction of molecules moving with high kinetic energy will escape the water and enter the air as vapor. If there is no heat added to the water during this process by some external source in thermal contact with the water, the effect of these high-energy molecules leaving the water will lower the average kinetic energy of all the remaining water molecules which results in a lowering of the temperature of the water. This process is known as “evaporation” and the resulting lowering of the water temperature is known as “evaporative cooling.”

If the process is allowed to continue, the average kinetic energy of the remaining water molecules will become lower and lower. Thus, the water temperature will continue to fall. If the air containing the evaporated water vapor is continuously removed from the air/water boundary layer so that it never reaches the saturation point, the evaporation process will continue and the water temperature will continue to fall. Eventually, the average kinetic energy of the remaining water molecules become so low that all the water will eventually freeze at 32° F. But even in the solid state, there are some freely-moving molecules that still have sufficiently high velocity to escape from the frozen water. Thus, even the temperature of the frozen water will continue to decrease as long as there is an unlimited supply of unsaturated air blowing over the air/water boundary. Virtually all the kinetic energy of the high-energy water molecules that eventually escape from the water is obtained from intermolecular collisions with other water molecules. Since the density and thermal conductivity of the air blowing over the air/water boundary layer is so small compared to that of the body of water, very little of the high kinetic energy of the escaping water molecules will come from the kinetic energy of the air molecules. Additional details of the physics of evaporating water in air can be found in the book, Drying and Processing of Materials by means of Conditioned Air, Carrier Engineering Corporation, 1929, pages 26-42 by D. C. Lindsay.

By mounting a system of condensing tubes having very high thermal conductivity such as copper in thermal contact with evaporating water, and passing pressurized saturated vapor through the tubes at a slightly higher temperature than the evaporating water outside the tubes, the water temperature will stop falling and the heat energy that is given to the high energy water molecules that escape from the air/water boundary layer comes from the latent heat of condensation of the saturated vapor flowing through the condensing tubes. By adjusting the rate of mass flow through the condensing tubes and replenishing the evaporated water with a supply of new liquid water (pre-cooled to the water evaporation temperature), a desired vapor condensation temperature can be achieved that can be continued essentially indefinitely. And, by designing the system such that the area of the evaporating air/water boundary layer is relatively large, the mass flow rate of pressurized saturated vapor flowing through the tubes that is condensing at the desired condensing temperature can be increased to essentially any value desired. Pre-cooling the water before it is evaporated to produce the heat sink is not an essential operating feature of the engine because the specific heat of evaporating water (heat of vaporization) is many times greater than the specific heat of liquid water. Hence, the introduction of new water at a higher temperature to maintain the supply of liquid water being evaporated at constant temperature has almost no effect in raising the water temperature. This is how the artificial low temperature heat sink inside the condenser is created and sustained in the present invention. These are the basic operating principles of the sub-ambient condensing system that is artificially created without consuming any mechanical work. Since the latent heat of evaporating water is very high, this system will be capable of condensing large amounts of expanded saturated working fluid vapor by evaporating small amounts of water.

By compressing the liquified working fluid discharged from the sub-ambient condenser to very high pressures it is possible to heat the compressed working fluid to relatively low temperatures by burning small amounts of clean-burning combustible fuel such as Butane so that essentially 100% of the heat energy generated by the burning fuel is absorbed by the compressed working fluid. By expanding the heated compressed working fluid inside a thermally insulated isentropic expansion system, mechanical power can be generated at power densities sufficiently high to propel automobiles great distances by burning small amounts of fuel and evaporating small amounts of water that generates no pollution and no noise. The fuel efficiency is so great that a vehicle propelled by this engine can achieve a miles per gallon fuel efficiency an order of magnitude greater than vehicles propelled by conventional internal combustion engines.

It should be pointed out and emphasized that all of the basic thermodynamic operating principles of this closed-cycle condensing heat engine do not represent any violation or change in the basic theory and equations of classical thermodynamics. All the principles of thermodynamics still apply to the engine disclosed herein. What has changed is how the high and low temperature heat reservoirs are created, and what their operating temperatures are. In order to further describe the basic thermodynamic operating principles of the invention and its performance capabilities, a detailed quantitative thermodynamic analysis is presented which represents an important part of the disclosure.

Table 1 describes the critical temperatures T_C and the critical pressures P_C of several phase-changing working fluids that can be used as the phase-changing working fluids for the engine operating within the high and low temperature heat reservoirs maintained at low temperatures by burning small amounts of fuel and evaporating small amounts of water. Most are common refrigerants identified by the “R” designation. They all have critical temperatures close to the average temperature of the natural atmosphere which is taken to be 294° K (69.51° F.). It should be emphasized that since the engine is a “closed-cycle” condensing heat engine, no working fluid is discharged into the open atmosphere. The only substance that is discharged from the engine is water vapor that is generated by the evaporative cooling process used to generate the artificial low temperature heat sink inside the condenser, and relatively small amounts of exhaust products close to ambient temperature from a clean burning fuel. The fuel is burned quietly in a continuous process and not in a series of very loud and inefficient separate explosions as in internal combustion engines.

TABLE 1
Critical Thermodynamic Parameters of several Low-Temperature Phase-
Changing Working Fluids that can be used in the Invention
Working Fluid	Critical Temperature	Critical Pressure (Bar)
Ethane	305.32° K.	(89.89° F.)	48.72
R32	351.25° K.	(172.56° F.)	57.82
R152a	386.41° K.	(235.85° F.)	45.17
R41	317.28° K.	(111.42° F.)	58.97
R22	369.30° K.	(205.05° F.)	49.90

All of the detailed thermodynamic properties of these working fluids are very accurately given in a U.S. Department of Commerce, National Institute of Standards and Technology, (NIST) publication entitled, “NIST Reference Fluid Thermodynamic and Transport Properties—REFPROP Version 8.0,” April 2007. It is in the form of a computer program. By inputting the numerical values of any 2 of the four basic thermodynamic state parameters of a substance (Temperature T, Pressure P, Enthalpy E, Entropy S), this computer program determines the values of the two remaining parameters. By using this computer program it will be possible to accurately determine the performance of various embodiments of the engine.

In the field of “evaporative cooling” achieved by evaporating water, the temperature at which water evaporates in air is called the “wet bulb temperature.” It depends upon the local air temperature (local ambient atmospheric temperature) called the “dry-bulb temperature,” and the local relative humidity (also known as percent saturation). It can be determined from a chart called a “Psychrometric Chart.” FIG. 1 is a typical Psychrometric Chart covering a dry-bulb temperature range of 20° F. (266° K) to 105° F. (314° K). As an illustrative example of how to use the Psychrometric Chart to determine the temperature of evaporating water when the ambient air temperature (dry-bulb temperature) is 80° F. and the relative humidity is 10%, follow the 80° F. vertical line upward until it intersects the 10% relative humidity curve and follow the diagonal wet-bulb temperature curve until it intersects the wet-bulb saturation curve indicated on the chart. The value of the wet-bulb temperature of evaporating water for these conditions is 52.18° F. (284.37° K).

Tables 2-7 give the values of the dry-bulb and wet-bulb temperatures corresponding to dry-bulb temperatures (ambient air temperatures) ranging from 50° F. (283.16° K) to 110° F. (316.48° K) and relative humidity ranging from 0% to 50%. The tables also give the latent heat of evaporating water (sub-ambient heat sink) corresponding to the various wet-bulb temperatures.

TABLE 2
Measured Wet-Bulb Temperatures Corresponding to Various Dry-Bulb Temperatures of
Evaporating Water (Relative Humidity = 0%)
TH	TL	ΔT	QL (J/gm)
50° F.	(283.16° K.)	32.60° F.	(273.49° K.)	17.40° F.	(9.67° K.)	2500.11
55° F.	(285.94° K.)	35.50° F.	(275.10° K.)	19.50° F.	(10.84° K.)	2496.32
60° F.	(288.72° K.)	38.24° F.	(276.63° K.)	21.76° F.	(12.09° K.)	2492.68
65° F.	(291.49° K.)	40.95° F.	(278.13° K.)	24.05° F.	(13.36° K.)	2489.06
70° F.	(294.27° K.)	43.30° F.	(279.44° K.)	26.70° F.	(14.83° K.)	2485.96
75° F.	(297.05° K.)	45.92° F.	(280.89° K.)	29.08° F.	(16.16° K.)	2482.57
80° F.	(299.83° K.)	48.13° F.	(282.12° K.)	31.87° F.	(17.71° K.)	2479.60
85° F.	(302.60° K.)	50.57° F.	(283.48° K.)	34.43° F.	(19.13° K.)	2476.39
90° F.	(305.38° K.)	52.60° F.	(284.60° K.)	37.40° F.	(20.78° K.)	2473.80
95° F.	(308.16° K.)	54.73° F.	(285.79° K.)	40.27° F.	(22.37° K.)	2470.91
100° F.	(310.94° K.)	56.66° F.	(286.86° K.)	43.34° F.	(24.08° K.)	2468.42
105° F.	(313.72° K.)	58.62° F.	(287.95° K.)	46.38° F.	(25.77° K.)	2465.86
110° F.	(316.49° K.)	60.65° F.	(289.08° K.)	49.35° F.	(27.42° K.)	2463.12
TH = Dry Bulb Temperature,
TL = Wet Bulb Temperature,
ΔT = Temperature Drop,
QL = Heat of Evaporation

TABLE 3
Measured Wet-Bulb Temperatures Corresponding to Various Dry-Bulb Temperatures of
Evaporating Water (Relative Humidity = 10%)
TH	TL	ΔT	QL (J/gm)
50° F.	(283.16° K.)	34.40° F.	(274.48° K.)	15.60° F.	(8.67° K.)	2497.73
55° F.	(285.94° K.)	37.60° F.	(276.26° K.)	17.40° F.	(9.67° K.)	2493.53
60° F.	(288.72° K.)	39.40° F.	(277.26° K.)	20.60° F.	(11.44° K.)	2491.12
65° F.	(291.49° K.)	43.38° F.	(279.47° K.)	21.62° F.	(12.01° K.)	2485.93
70° F.	(294.27° K.)	46.35° F.	(281.12° K.)	23.65° F.	(13.14° K.)	2482.00
75° F.	(297.05° K.)	49.51° F.	(282.88° K.)	25.49° F.	(14.16° K.)	2477.81
80° F.	(299.83° K.)	52.18° F.	(284.36° K.)	27.82° F.	(15.46° K.)	2474.30
85° F.	(302.60° K.)	55.02° F.	(285.94° K.)	29.98° F.	(16.66° K.)	2470.58
90° F.	(305.38° K.)	57.96° F.	(287.57° K.)	32.04° F.	(17.80° K.)	2466.75
95° F.	(308.16° K.)	60.73° F.	(289.11° K.)	34.27° F.	(19.04° K.)	2463.10
100° F.	(310.94° K.)	63.45° F.	(290.62° K.)	36.55° F.	(20.31° K.)	2459.48
105° F.	(313.72° K.)	66.00° F.	(292.04° K.)	39.00° F.	(21.67° K.)	2456.13
110° F.	(316.49° K.)	68.50° F.	(293.43° K.)	41.50° F.	(23.06° K.)	2452.81
TH = Dry Bulb Temperature,
TL = Wet Bulb Temperature,
ΔT = Temperature Drop,
QL = Heat of Evaporation

TABLE 4
Measured Wet-Bulb Temperatures Corresponding to Various Dry-Bulb Temperatures of
Evaporating Water (Relative Humidity = 20%)
TH	TL	ΔT	QL(J/gm)
50° F.	(283.16° K.)	36.45° F.	(275.62° K.)	13.55° F.	(7.53° K.)	2495.72
55° F.	(285.94° K.)	40.00° F.	(277.59° K.)	15.00° F.	(8.33° K.)	2491.04
60° F.	(288.72° K.)	43.30° F.	(279.43° K.)	16.70° F.	(9.28° K.)	2486.73
65° F.	(291.49° K.)	46.70° F.	(281.33° K.)	18.30° F.	(10.17° K.)	2482.21
70° F.	(294.27° K.)	50.00° F.	(283.15° K.)	20.00° F.	(11.11° K.)	2477.98
75° F.	(297.05° K.)	53.00° F.	(284.82° K.)	22.00° F.	(12.22° K.)	2473.27
80° F.	(299.83° K.)	56.36° F.	(286.68° K.)	23.64° F.	(13.13° K.)	2468.88
85° F.	(302.60° K.)	59.80° F.	(288.59° K.)	25.20° F.	(14.00° K.)	2464.28
90° F.	(305.38° K.)	62.90° F.	(290.32° K.)	27.10° F.	(15.06° K.)	2460.23
95° F.	(308.16° K.)	66.06° F.	(292.07° K.)	28.94° F.	(16.08° K.)	2456.11
100° F.	(310.94° K.)	69.20° F.	(293.82° K.)	30.80° F.	(17.11° K.)	2451.98
105° F.	(313.72° K.)	72.35° F.	(295.57° K.)	32.65° F.	(18.14° K.)	2447.76
110° F.	(316.49° K.)	75.45° F.	(297.29° K.)	34.55° F.	(19.19° K.)	2443.67
TH = Dry Bulb Temperature,
TL = Wet Bulb Temperature,
ΔT = Temperature Drop,
QL = Heat of Evaporation

TABLE 5
Measured Wet-Bulb Temperatures Corresponding to Various Dry-Bulb Temperatures of
Evaporating Water (Relative Humidity = 30%)
TH	TL	ΔT	QL (J/gm)
50° F.	(283.16° K.)	38.30° F.	(276.65° K.)	11.70° F.	(6.50° K.)	2492.59
55° F.	(285.94° K.)	42.00° F.	(278.71° K.)	13.00° F.	(7.22° K.)	2487.73
60° F.	(288.72° K.)	45.30° F.	(280.54° K.)	14.70° F.	(8.17° K.)	2483.34
65° F.	(291.49° K.)	49.55° F.	(282.90° K.)	15.45° F.	(8.58° K.)	2477.83
70° F.	(294.27° K.)	52.95° F.	(284.79° K.)	17.05° F.	(9.47° K.)	2473.30
75° F.	(297.05° K.)	56.61° F.	(286.82° K.)	18.39° F.	(10.22° K.)	2468.49
80° F.	(299.83° K.)	60.62° F.	(288.72° K.)	19.98° F.	(11.10° K.)	2464.03
85° F.	(302.60° K.)	63.70° F.	(290.76° K.)	21.30° F.	(11.88° K.)	2459.19
90° F.	(305.38° K.)	67.20° F.	(292.71° K.)	22.80° F.	(12.67° K.)	2454.53
95° F.	(308.16° K.)	70.75° F.	(294.68° K.)	24.25° F.	(13.47° K.)	2449.88
100° F.	(310.94° K.)	74.40° F.	(296.71° K.)	25.60° F.	(14.22° K.)	2445.09
105° F.	(313.72° K.)	77.95° F.	(298.68° K.)	27.05° F.	(15.03° K.)	2440.45
110° F.	(316.49° K.)	80.42° F.	(300.05° K.)	29.58° F.	(16.43° K.)	2437.13
TH = Dry Bulb Temperature,
TL = Wet Bulb Temperature,
ΔT = Temperature Drop,
QL = Heat of Evaporation

TABLE 6
Measured Wet-Bulb Temperatures Corresponding to Various Dry-Bulb Temperatures of
Evaporating Water (Relative Humidity = 40%)
TH	TL	ΔT	QL (J/gm)
50° F.	(283.16° K.)	40.20° F.	(277.71° K.)	9.80° F.	(5.44° K.)	2490.13
55° F.	(285.94° K.)	44.00° F.	(279.82° K.)	11.00° F.	(6.11° K.)	2485.06
60° F.	(288.72° K.)	47.90° F.	(281.98° K.)	12.10° F.	(6.72° K.)	2480.69
65° F.	(291.49° K.)	51.85° F.	(284.18° K.)	13.15° F.	(7.31° K.)	2475.43
70° F.	(294.27° K.)	55.80° F.	(286.37° K.)	14.20° F.	(7.89° K.)	2469.58
75° F.	(297.05° K.)	59.60° F.	(288.48° K.)	15.40° F.	(8.56° K.)	2464.56
80° F.	(299.83° K.)	63.50° F.	(290.65° K.)	16.50° F.	(9.17° K.)	2459.45
85° F.	(302.60° K.)	67.30° F.	(292.71° K.)	17.70° F.	(9.83° K.)	2454.53
90° F.	(305.38° K.)	71.15° F.	(294.90° K.)	18.85° F.	(10.47° K.)	2449.36
95° F.	(308.16° K.)	75.20° F.	(297.15° K.)	19.80° F.	(11.00° K.)	2444.05
100° F.	(310.94° K.)	79.30° F.	(299.43° K.)	20.70° F.	(11.50° K.)	2438.62
105° F.	(313.72° K.)	82.94° F.	(301.45° K.)	22.06° F.	(12.26° K.)	2433.87
110° F.	(316.49° K.)	86.60° F.	(303.48° K.)	23.40° F.	(13.00° K.)	2428.99
TH = Dry Bulb Temperature,
TL = Wet Bulb Temperature,
ΔT = Temperature Drop,
QL = Heat of Evaporation

TABLE 7
Measured Wet-Bulb Temperatures Corresponding to Various Dry-Bulb Temperatures of
Evaporating Water (Relative Humidity = 50%)
TH	TL	ΔT	QL (J/gm)
50° F.	(283.16° K.)	41.96° F.	(278.68° K.)	8.04° F.	(4.47° K.)	2487.75
55° F.	(285.94° K.)	45.98° F.	(280.93° K.)	9.02° F.	(5.01° K.)	2482.40
60° F.	(288.72° K.)	50.20° F.	(283.26° K.)	9.80° F.	(5.44° K.)	2476.92
65° F.	(291.49° K.)	54.30° F.	(285.54° K.)	11.00° F.	(6.11° K.)	2471.56
70° F.	(294.27° K.)	58.40° F.	(287.82° K.)	11.60° F.	(6.44° K.)	2466.10
75° F.	(297.05° K.)	62.70° F.	(290.21° K.)	12.30° F.	(6.83° K.)	2460.49
80° F.	(299.83° K.)	66.70° F.	(292.43° K.)	13.30° F.	(7.39° K.)	2455.20
85° F.	(302.60° K.)	70.80° F.	(294.71° K.)	14.20° F.	(7.89° K.)	2449.86
90° F.	(305.38° K.)	74.83° F.	(296.94° K.)	15.17° F.	(8.43° K.)	2444.53
95° F.	(308.16° K.)	78.90° F.	(299.21° K.)	16.10° F.	(8.94° K.)	2439.14
100° F.	(310.94° K.)	83.00° F.	(301.48° K.)	17.00° F.	(9.44° K.)	2433.75
105° F.	(313.72° K.)	87.20° F.	(303.82° K.)	17.80° F.	(9.89° K.)	2428.27
110° F.	(316.49° K.)	91.40° F.	(306.15° K.)	18.60° F.	(10.33° K.)	2422.73
TH = Dry Bulb Temperature,
TL = Wet Bulb Temperature,
ΔT = Temperature Drop,
QL = Heat of Evaporation

The above tables of measured wet-bulb temperatures corresponding to various dry-bulb temperatures (ambient air temperatures) and various humidity demonstrate the relatively large temperature drop that can be achieved by evaporating water. The large latent heats of evaporation for water {circumflex over (Q)}_L show that the creation of a significant low temperature artificial heat sink capable of absorbing large amounts of latent heat of condensation (heat of vaporization) from a working fluid at sub-ambient temperatures will be possible by evaporating water in a flowing stream of unsaturated air. This fact establishes the basic theoretical operability of the invention and hence represents an important part of the theoretical basis for the invention.

In order to determine the actual operating performance of the preferred embodiment of the invention, a detailed quantitative thermodynamic analysis will be conducted based on an assumed ambient air temperature, humidity, and corresponding temperature of evaporating water (which represents the low temperature heat reservoir), for the working fluid which will be assumed to be R32. (The analysis is identical for all the other possible working fluids given in Table 1.) The corresponding sub-ambient temperature (i.e., the artificial sub-ambient heat sink of the low-temperature condensing system) resulting from the evaporation of water at various humidities are determined from Tables 2-7. The detailed thermodynamic analysis using the R32 working fluid will be based on the assumption that the relative humidity is 10% and the natural ambient air temperature (dry-bulb temperature) is 60° F. (288.72° K) which is close to the average temperature of the natural atmosphere (294° K). Therefore, assuming an ambient air temperature of 60° F. (288.72° K) and a relative humidity of 10%, it follows from Table 3 that the temperature of the evaporating water (i.e., the artificially created low temperature heat sink) inside the sub-ambient condensing system will be 39.40° F. (277.26° K).

FIG. 2 is a schematic block diagram of the preferred embodiment of the invention illustrating its design and operating principles. The Temperature-Entropy Diagram (TS Diagram) of the R32 working fluid circulating around the closed cycle of the engine corresponding to the thermodynamic analysis of the numerical example is given in FIG. 3. This TS Diagram describes the basic thermodynamic operating principles and features of the invention. The points A, B, C, and D on the TS Diagram with arrows showing the direction of flow correspond to the values of the thermodynamic state parameters of the working fluid as it flows through the engine passing the various flow points around the closed cycle. Referring to FIG. 2, the water storage vessel 10 is designed with a double wall construction. The inner wall, in thermal contact with the water 12 inside the storage vessel 10, is made of material having high thermal conductivity such as aluminum or copper. The outer wall is made of material having low thermal conductivity such as fiberglass. Water absorbing padding material is mounted on the external surfaces of the inner wall and thermal insulation is mounted on the external surfaces of the outer wall. The water 12 inside the water vessel 10 is cooled by evaporating small amounts of water on the evaporation pads mounted on the inner walls. Air is drawn into the space between the walls over the water pads through an inlet duct via a small electric fan which is discharged through a small outlet duct. When the water inside the storage vessel 10 reaches a temperature close to the temperature of the evaporating water, no further evaporation from the pads is required. Small thermally insulated shutters are closed over the inlet and outlet air ducts thereby thermally insulating the water 12 inside the vessel 10 at the low evaporation temperature of the cooling water. In the preferred embodiment, the water vessel 10 is designed to have a capacity of 150 liters (39.6 gal) which is regarded as one of the “fuel” components that is consumed by the engine to sustain the low temperature heat sink in the condenser achieved by the evaporation process. Since the engine stops running when all the water is consumed in the water storage vessel 10, water can be viewed as a “fuel” that powers the engine. It is fuel because it is required to generate the engine's artificial low temperature heat reservoir analogous to the fuel burned to generate the high temperature heat reservoir. In the preferred embodiment, the fuel that is used to generate the engine's high temperature heat reservoir is a clean-burning liquified natural gas (LNG) such as Propane or Butane.

Referring to FIG. 2, a small electric pump 14 pumps the pre-cooled water from the water tank 10 into a thermally insulated low temperature sub-ambient condenser 16 via a conduit 18 where it is evaporated in a flowing air stream to create the engine's artificial low temperature heat reservoir (heat sink) via the process of “evaporative cooling.” After the working fluid is condensed (at the condensing temperature and pressure) by passing through the condenser 16, it is fed into a large thermally insulated pressurized holding tank 20 having a capacity of 100 kg of liquified working fluid where it is accumulated at a certain mass flow rate and held as a large reservoir of liquified working fluid 22 at the sub-ambient condensing temperature. (It is regarded as another energy storage system.) At a time determined by the engine's control computer 24, liquified working fluid 22 is withdrawn from the thermally insulated holding vessel 20 at a certain mass flow rate and fed into a multistage isentropic compression system 26 where it is compressed to 300 Bar (4,351 lbs/in²).

After the working fluid is compressed, it is fed into a thermally insulated heating system 28 where it is heated by burning small amounts of the clean-burning fuel. The heating system 28 is designed to be nearly 100% efficient in that essentially all of the heat generated by the burning fuel is absorbed by the compressed working fluid. This is achieved by designing the heating system using two serially connected thermally insulated heating chambers 30, 32. The first chamber 30 is a pre-heating chamber 30 that initially heats the compressed working fluid to some intermediate temperature T₁ not far above the ambient temperature of the natural atmosphere by absorbing the heat in the combustion gases discharged from the main heating chamber 32 at the temperature T_C indicated by point C on the TS Diagram (FIG. 3) thereby initially heating the working fluid in the pre-heating chamber 30 by utilizing the high-temperature combustion gases discharged from the main heating chamber 32 as input heat energy for the pre-heating chamber 30. After the combustion gases are circulated through the main heating chamber 32 thereby heating the working fluid to a temperature T_C at point C on the TS Diagram, the gases are discharged from this chamber 32 at this temperature T_C and fed into the pre-heating chamber 30 where it is utilized to heat the working fluid discharged from the compressor 26 to the intermediate temperature T₁. Thus, after circulating through the pre-heating chamber 30, the combustion gases are discharged into the open atmosphere just a few degrees above ambient temperature. Consequently, essentially all of the heat of combustion of the burning fuel is used to heat the compressed working fluid. (In internal combustion engines, the combustion gases are discharged into the open atmosphere at several hundred degrees above ambient atmospheric temperature thereby contributing to the very low efficiency of these engines.)

The heated compressed working fluid is withdrawn from the heating system 28 and fed into a large, thermally insulated high pressure compressed gas storage system 34 (comprising several serially-connected thermally insulated high-pressure compressed-gas storage cylinders) where it is accumulated at a certain mass flow rate determined by the engine's control computer 24.

The engine's control computer 24 is designed to keep the pressurized gas storage system 34 full to nearly maximum capacity even when the vehicle is parked and not being used by automatically withdrawing liquified working fluid 22 from the holding vessel 20, feeding it into the isentropic compression system 26, withdrawing the compressed working fluid from the compressor system 26, feeding it into the heating system 28 where it is heated, and feeding the heated compressed gas 36 into the energy storage system 34 via electrically operated servo control valves. An automatic environmental sensing system 38 that measures atmospheric temperature, humidity, barometric pressure, etc., feeds this information continuously into the engine's control computer 24 that determines the best time to operate all the automatic subsystems when the vehicle is parked and not being driven by the driver. When the vehicle is moving under a certain desired power setting, compressed heated working fluid 36 stored in the storage system 34 is withdrawn at a certain mass flow rate {dot over (m)} and fed into a multistage isentropic expansion system 40 thereby generating the motive power to propel the vehicle. The expansion system 40 is designed to transform the expanded gaseous working fluid into a saturated vapor at sub-ambient temperature which is fed back into the condenser 16 and re-condensed back into a liquid to repeat the cycle.

Referring to FIGS. 2 and 3, a detailed quantitative thermodynamic analysis of the above cycle will now be presented to further describe the basic operating principles and features of the invention, and to determine its performance corresponding to the atmospheric conditions where ambient air temperature=60° F., relative humidity=10%, using the R32 working fluid described in Table 1.

In this thermodynamic analysis using the R32 working fluid, the operating cycle of the engine, represented by the working fluid circulating around the block diagram of FIG. 2, and circulating around the corresponding Temperature-Entropy (TS)Diagram of the working fluid shown in FIG. 3, begins by withdrawing liquified working fluid 22 from the holding tank 20 at point A at the sub-ambient condensing temperature T having some initial values of pressure P, specific enthalpy H, and specific entropy S. These four thermodynamic state parameters T, P, H, S, completely determine the thermodynamic state of the working fluid in the temperature range considered herein. The detailed thermodynamic analysis of the engine will proceed with the usual assumptions of isobaric heating, isentropic compression, and isentropic expansion steps by determining the numerical values of the four thermodynamic state parameters of the working fluid at all the flow points indicated on the Temperature—Entropy Diagram of FIG. 3. For simplicity, it will be assumed that the mass flow rate {dot over (m)} of the working fluid at all the flow points (A, B, C, D) indicated in FIG. 3 will be equal. With this information, it will be possible to accurately determine the engine's continuous, steady-state, net specific output work W_Net corresponding to any given values of dry and wet bulb temperatures, relative humidity, and the effective heating temperature generated by the heating system 28 which will be assumed, for definiteness, to be the heat generated by the fuel burning in the heating system 28 according to the preferred embodiment.

The thermodynamic analysis begins by assuming that the local atmospheric temperature is 60° F. (288.72° K) and the relative humidity is 10%. Consequently, in view of Table 3, the corresponding dry bulb and wet bulb temperatures are 288.72° K (60° F.) and 277.26° K (39.40° F.), respectively. Thus, the sub-ambient condensing temperature inside the condenser 16 corresponding to these assumed atmospheric conditions will be 277.26° K (39.40° F.). The latent heat of vaporization of water {circumflex over (Q)}_L at this condensing temperature is equal to 2,491.12 J/gm.

Referring to FIGS. 2 and 3, the values of the four thermodynamic state parameters of the expanded working fluid entering and leaving the condenser 16 at points D and A, respectively, can be calculated by designing the isentropic expansion step from point C to point D, represented by the vertical line segment CD in the TS Diagram (FIG. 3), such that the temperature of the expanded working fluid is reduced from some as yet undetermined high temperature T_C (at a pressure of 300 Bar) at point C. on the TS Diagram down to the condensing temperature of 277.26° K (39.40° F.) such that when the expanded gaseous working fluid reaches the low temperature at the end of the expansion (at point D on the TS Diagram), it becomes a saturated vapor that is immediately fed into the condenser 16 where it is liquefied by transferring its latent heat of vaporization to the evaporating water. The thermodynamic state parameters of saturated vapor and saturated liquid of the R32 working fluid at 277.26° K is determined from the NIST computer program. The resulting numerical values of these thermodynamic state parameters of the expanded saturated vapor at point D are:

T_D=277.26° K, P_D=9.2561 Bar, H_D=515.98 J/gm, S_D=2.1395 J/gm °K.

The corresponding numerical values of the saturated liquid at point A are:

T_A=277.26° K, P_A=9.2561 Bar, H_A=207.22 J/gm, S_A=1.0259 J/gm °K.

Thus, the latent heat of vaporization {circumflex over (Q)}_V that is extracted from the expanded saturated vapor by the evaporating water inside the condenser 16 can be calculated by:

{circumflex over (Q)}_V=H_D−H_A=308.76 J/gm (working fluid)   (1)

(To demonstrate the enormous amount of potential thermal power represented by the evaporation of water under these thermodynamic conditions inside the condenser 16 it should be pointed out that an evaporation rate of only 1 gm/sec of water will generate a heat absorbing thermal power of 2.491 KW. This fact has very important consequences that will give the invention a remarkable performance.)

Since the expansion step CD is isentropic, the values of the thermodynamic state parameters at point C. can be determined from the fact that P_C=300 Bar and S_C=S_D=2.1395 (J/gm °K). With these two parameters determined, one obtains the values of all four thermodynamic state parameters at point C by the NIST computer program:

T_C=541.07° K P_C=300.00 Bar H_C=686.57 J/gm S_C=2.1395 J/gm °K

The specific mechanical work Ŵ_E generated by the isentropic expansion step CD in the engine's thermodynamic cycle is calculated from the values of the thermodynamic state parameters:

Ŵ_E=H_C−H_D=170.59 J/gm   (2)

(Specific mechanical work refers to the work generated by 1.0 gm of working fluid that is determined from the thermodynamic state parameters and will be denoted by the symbol Ŵ.)

After the working fluid is condensed into a liquid at sub-ambient temperature T_A=277.26° K, (39.40° F.) and pressure P_A=9.2561 Bar, it is fed into the thermally insulated storage vessel 20 with the same thermodynamic state parameters. The liquified working fluid is withdrawn from the storage vessel 20 at 277.26° K (39.40° F.) and fed into the thermally insulated isentropic compressor system 26. This isentropic compression step is represented by the short vertical line segment AB on the TS Diagram (FIG. 3). In the preferred embodiment, this initial high pressure will be 300 Bar (296.1 ATM or 4,351 lbs/in²). (The isentropic compressor 26 and expander 40 systems operating at these high pressures are not uncommon in the design of prior art cryogenic engines. Thus, the basic designs of the compressor 26 and expander 40 systems will be essentially identical to those in the prior art.)

By using the NIST computer program, the compressed working fluid is withdrawn from the isentropic compressor system 26 with its thermodynamic state parameters equal to:

T_B=290.31° K, P_B=300 Bar, H_B=234.46 J/gm, S_B=1.0259 J/gm °K.

The corresponding flow point is shown as point B on the TS Diagram of FIG. 3.

The amount of specific mechanical work Ŵ_Comp consumed in the isentropic compression step Â{circumflex over (B)} is determined by

W_Comp=H_B−H_A=27.24 J/gm (working fluid)   (3)

Thus, the continuous flow, steady-state, net output work Ŵ_Net generated by one gm of working fluid circulating around the closed cycle ABCD (FIG. 3) of the engine is

Ŵ_Net=Ŵ_E−Ŵ_Comp=143.35 J/gm (working fluid)   (4)

Since the amount of heat energy {circumflex over (Q)}_V that is extracted from the expanded saturated working fluid at point D by the evaporating water inside the condenser 16 where the latent heat of vaporization of water {circumflex over (Q)}_L=2,491.12 J/gm, the ratio R defined by:

$\begin{array}{cc}R=\frac{{\hat{Q}}_L}{{\hat{Q}}_V}=\frac{2491.12}{308.76}=8.0681& \left(5\right)\end{array}$

is equal to the mass flow ratio of the engine's working fluid (R32) relative to the mass flow rate of water being evaporated inside the condenser 16 where {circumflex over (Q)}_L is equal to the latent heat of evaporation of 1.0 gm of water at the condenser's operating temperature. In the case of this example where the relative humidity is 10% and the dry-bulb temperature is 60° F. (which is equal to the assumed ambient air temperature of the natural atmosphere), Table 3 indicates that the wet-bulb temperature will be 39.40° F. (277.26° K). The corresponding latent heat of evaporating water at this temperature indicated in this Table is {circumflex over (Q)}_L=2,491.12 J/gm Hence, in view of equation (5), the engine operating at a temperature of 60° F. and a relative humidity of 10% as in this example will have a working fluid/water mass flow ratio R=8.0681. Thus, the actual rate of water consumption {dot over (m)}_W necessary to operate the engine with a continuous working fluid mass flow rate {dot over (m)} is given by the equation

$\begin{array}{cc}{\stackrel{.}{m}}_W=\frac{\stackrel{.}{m}}{R}& \left(6\right)\end{array}$

Thus, it only requires the evaporation of 1/R=0.1239 gm of water to sustain the low temperature heat sink inside the condenser 16 for each gm of working fluid circulating around the engine.

After leaving the isentropic compression system 26 at point B on the TS Diagram (FIG. 3), the compressed liquefied working fluid at 290.31° K (62.87° F.) is fed into the thermally insulated heating system 28 where it is isobarically heated to 541.07° K (514.24° F.)=T_C. After this heating step, the compressed heated working fluid at a pressure of 300 Bar is fed into the thermally insulated compressed gas storage system 34. This high pressure compressed gas storage system 34 represents a large load-leveling energy storage system for the down-stream isentropic expansion system 40. It enables the expansion system 40 to be operated using the large supply of high pressure gas stored inside the storage system 34 at very high power levels for brief time periods. (High pressure expanders such as those used in cryogenic engines weighing just a few kilograms with very small dimensions are capable of generating several hundred kilowatts of mechanical power by expanding compressed gas having pressures of several thousand psi as in the present engine. The forces are so high they can actually break the mechanical connecting rods and other parts.)

The compressed gas is withdrawn from the storage system 34 and fed into the multistage isentropic expansion system 40 which completes the cycle at point D on the TS Diagram (FIG. 3).

The amount of heat energy {circumflex over (Q)}_H absorbed by the working fluid inside the heating system 28 in the preferred embodiment heated by burning small amounts of combustible fuel is given by:

{circumflex over (Q)}_H=H_C−H_B=4452.11 J/gm   (7)

In the preferred embodiment, Butane is used as the clean-burning fuel. The heat of combustion Ĥ_C of Butane is 64,652 J/gm. By designing the engine's thermally insulated heating system 28 using two serially connected heating chambers 30, 32 such that all of the heat generated by burning the Butane fuel is absorbed by the compressed working fluid (represented by the curve {circumflex over (B)}Ĉ on the TS Diagram) the mass flow rate {dot over (m)}_B of burning Butane is given by:

$\begin{array}{cc}{\stackrel{.}{m}}_B=\frac{{\hat{Q}}_H\stackrel{.}{m}}{{\hat{H}}_C}=\frac{{\hat{Q}}_H\stackrel{.}{m}}{64652}=0.00700\stackrel{.}{m}& \left(8\right)\end{array}$

where the required heating power P_H is given by

P_H={circumflex over (Q)}_H{dot over (m)}  (9)

The thermal efficiency η of the engine is

$\begin{array}{cc}\eta =\frac{{\hat{W}}_{\mathrm{Net}}}{{\hat{Q}}_H}=\frac{143.35}{452.11}=0.317& \left(10\right)\end{array}$

This efficiency is about three times greater than that of internal combustion engines used for propelling road vehicles which is about 0.12.

The mechanical output power P(Watts) generated by the engine corresponding to the present example can be expressed as:

P=Ŵ_Net{dot over (m)}=143.35{dot over (m)}_(Watts)   (11)

In view of equation (6), this power can also be expressed in terms of the evaporation rate of water {dot over (m)}_W (gm/sec) in the condenser as:

P=RŴ_Net{dot over (m)}_W=1,156.56{dot over (m)}_W (Watts)   (12)

At this point in the thermodynamic analysis describing the basic performance of the preferred embodiment of the invention operating under the assumption of equal mass flow rates {dot over (m)} of working fluid at all flow points, A, B, C, and D described in the TS Diagram (FIG. 3), a thermodynamic analysis of the vehicle's automatic re-compression system will now be presented that occurs when the vehicle is parked and not being driven. At a certain time determined by the control computer 24, the expansion system 40 is automatically started such that all of the mechanical output power P of expansion 40 system is used to operate the re-compression system 26 that withdraws liquefied working fluid from the storage vessel 20 and re-compresses it at a certain mass flow rate {dot over (m)}. Consequently, the required mechanical power P_Comp necessary to achieve this by the re-compression system 26 given by

P_Comp=Ŵ_Comp{dot over (m)}=P=Ŵ_E{dot over (m)}₀   (13)

where {dot over (m)}₀ is the rate of mass flow of working fluid that the expansion system 40 withdraws from the storage system 34 to expand in the expansion system 40 to generate the mechanical power P to operate the re-compressor system 26. Hence, the net mass flow rate {dot over (m)}_Net of working fluid being accumulated in the storage system 34 is given by

{dot over (m)}_Net={dot over (m)}−{dot over (m)}₀   (14)

Thus, in view of equation (13), the accumulated mass flow {dot over (m)}_Net can be expressed as

$\begin{array}{cc}{\stackrel{.}{m}}_{\mathrm{Net}}=\stackrel{.}{m}-{\stackrel{.}{m}}_0=\stackrel{.}{m}-\frac{{\hat{W}}_{\mathrm{Comp}}\stackrel{.}{m}}{{\hat{W}}_E}=\stackrel{.}{m}\left[1-\frac{27.24}{170.59}\right]=0.8403\stackrel{.}{m}& \left(15\right)\end{array}$

Hence, by making use of equation (13), we obtain

$\begin{array}{cc}{\stackrel{.}{m}}_{\mathrm{Net}}=\mathrm{.8403}\left(\frac{P}{{\hat{W}}_{\mathrm{Comp}}}\right)=\mathrm{.03085}P& \left(16\right)\end{array}$

From equation (13) we obtain

{dot over (m)}=0.03671P   (17)

Consequently, in view of equations (15), (16), and (17), the mass flow rate {dot over (m)}₀ that working fluid must be expanded through the engine's expansion system 40 to generate an amount of power P to operate the re-compression system 26 while the vehicle is parked is given by

{dot over (m)}₀={dot over (m)}−{dot over (m)}_Net=0.03671P−0.03085P=0.005862P   (18)

Since the heating power P_H required to heat a mass flow rate of working fluid {dot over (m)} passing through the heating system 28 (along the curve {dot over (B)}Ċ on the TS Diagram) is given by

P_H={dot over (m)}Q_H=Ĥ_C{dot over (m)}_B

where {dot over (m)}_B is equal to the required burning rate of Butane fuel, it follows from equation (17) that this burning rate can be expressed as

$\begin{array}{cc}{\stackrel{.}{m}}_B=\frac{\stackrel{.}{m}Q_H}{{\hat{H}}_C}=\frac{0.03671Q_HP}{{\hat{H}}_C}& \left(19\right)\end{array}$

Thus, if the expansion system 40 is running at a mechanical power P=4,000 Watts, the net amount of working fluid {dot over (m)}_Net being accumulating in the energy storage system 34 would be 123.40 gm/sec. The rate of mass flow {dot over (m)}₀ through the expander 40 obtained by equation (18) is 23.45 gm/sec. The mass flow rate {dot over (m)}_W that water would be consumed in the condensing system be

$\begin{array}{cc}{\stackrel{.}{m}}_W=\frac{{\stackrel{.}{m}}_0}{R}=0.000727P=2.906\mathrm{gm}\text{/}\sec & \left(20\right)\end{array}$

The required burning rate of Butane fuel obtained from equation (19) would be 1.0268 gm/sec.

By operating the system for only 15 minutes, a total of 111,060 gm of compressed working fluid would be added to the compressed energy storage system 34. This is enough heated compressed gas to propel the vehicle 61 miles at a speed of 50 mph without operating the re-compression system 26. This is how the compressed gas energy storage system 34 can be automatically maintained close to maximum capacity before the vehicle is used. No fuel is burned during this 61 mile distance. The only thing that is consumed in the 61 mile drive is a few kilograms of water M_W given by

$\begin{array}{cc}{M}_W=\frac{m_{\mathrm{Net}}}{R}=\frac{111060\mathrm{gm}}{8.0681}=13.8\mathrm{kg}& \left(21\right)\end{array}$

Table 9 is a summary of the above performance calculations of the engine corresponding to the assumed atmospheric conditions (relative humidity=10%, ambient air temperature=60° F.).

TABLE 9
Performance of the Low Temperature Condensing Heat Engine Using R32 Working Fluid
When the Ambient Air Temperature is 60° F. (288.72° K) and Relative Humidity = 10%
Working Fluid                    R32 Refrigerant
Critical Temperature Of Working Fluid TC = 351.25° K (172.56° F.)
Critical Pressure of Working Fluid PC = 57.82 Bar
Ambient Air Temperature          288.72° K (60.00° F.)
Relative Humidity                10%
Condensing Temperature           277.26° K (39.40° F.)
Initial Compression Of Working Fluid (Liquid) 300 Bar (4,351 lbs/in2) at 277.26° K (39.40° F.)
Specific Compression Work        ŴComp = 27.24 J/gm (working fluid)
Specific Output Work Of Expander ŴE = 170.59 J/gm (working fluid)
Specific Heat of Evaporation of Water {circumflex over (Q)}L = 2491.12 J/gm
Specific Heat of Vaporization of Working Fluid {circumflex over (Q)}V = 308.76 J/gm
Mass Flow Ratio R (Working Fluid/Water) R = 8.0681
Net Specific Output Work Of Engine ŴNet = 143.35 J/gm (working fluid)
Heat Absorbed in Heating System  {circumflex over (Q)}H = 452.11 J/gm
Heat of Combustion of Butane fuel ĤC = 64,652 J/gm
Mechanical Output Power P(Watts) P = ŴNet{dot over (m)} = 143.35{dot over (m)} (Working Fluid) Watts
Mechanical Output Power P(Watts) P = RŴNet{dot over (m)}W = 1,156.56{dot over (m)}W (Water) Watts
Required Butane Burning Rate     m .  B  =     Q ^    m .     H ^  C   =    454.11   m .   64652  =  0.0070       m .        (  gm   /   sec  )
Thermal Efficiency of Engine:    η =    W ^   Q ^   = 0.317
Rate that working fluid is recompressed by Expander {dot over (m)}Net = .03085P gm/sec
operating at a power of P (Watts) while vehicle is parked
Fuel Burning Rate During Re-compression m .  B  =     m .    Q H     H ^  C   =    0.03671   Q H   P    H ^  C   =  0.0002567      P

Table 10 gives the net power output P of the engine generated by various water evaporation rates {dot over (m)}_W inside the condenser and the corresponding mass flow rate {dot over (m)} of the working fluid.

TABLE 10
Continuous Power Output Of The Low Temperature Condensing Heat Engine Using R32
Working Fluid When The Ambient Air Temperature is 60° F. (288.72° K.) And The Humidity
is 10% Corresponding To Various Water Consumption Rates
{dot over (m)}W (gm/sec water)	{dot over (m)} (R32 working fluid)	{dot over (m)}B (fuel burning rate)	P(Total KW)	P (HP)
1.00	8.07	0.056	1.157	1.552
2.00	10.07	0.070	1.444	1.936
3.00	24.20	0.169	3.470	4.653
4.00	32.27	0.226	4.626	6.204
5.00	40.34	0.282	5.783	7.755
6.00	48.41	0.339	6.939	9.306
7.00	56.48	0.395	8.096	10.857
8.00	64.55	0.452	9.252	12.408
9.00	72.61	0.508	10.409	13.959
10.00	80.68	0.565	11.566	15.509

Table 11 gives a listing of the power consumed by rolling friction (tires) and aerodynamic drag at various speeds corresponding to a four passenger vehicle having a gross mass of 1,230 kg (2,712 lbs), an aerodynamic drag coefficient C_d=0.19, a rolling drag coefficient C_t=0.005, and a total frontal area of 2.50 m². The data was extrapolated from the published article “Propulsion Technology: An Overview,” Automotive Engineering, Vol. 100, No. 7, July 1992, pp. 29-33.

TABLE 11
Level Road Power Requirements For Low Drag Vehicles
	Speed	Tires	Aero	Total
	(mph)	(KW)	(KW)	(KW)
	22	0.603	0.279	0.882
	25	0.670	0.382	1.052
	34	0.921	0.993	1.914
	50	1.349	2.943	4.292
	60	1.617	5.384	7.001
	67	1.808	7.524	9.332
	81	2.176	13.122	15.298

Assuming that the high-efficience low-temperature condensing heat engine described in the above thermodynamic analysis and summarized in Tables 9 and 10 is installed in a vehicle designed with these low drag coefficients, Table 11 indicates that a propulsive power of 4.3 KW (5.8 HP) will propel the vehicle along a level road at a speed of about 50 mph. This continuous power generation (4.3 KW) will be the power generated in the preferred embodiment of the invention. However, because of the design of the load-leveling, high-pressure stored energy system 34, and its operation that enables a variable mass flow rate to leaving this storage system 34, it will be possible for the engine to generate bursts of accelerating power far exceeding that of any present day automobile propelled by a conventional internal combustion engine for several seconds.

According to the above thermodynamic calculations where the engine is operating with an ambient air temperature of 60° F. and a humidity of 10% (which is close to the average atmospheric temperature of 69.5° F.), the actual net achievable continuous specific output work is Ŵ_Net=143.35 J/gm (working fluid). Consequently, in view of equation (11), in order to achieve a continuous power output of 4.3 KW, the required mass flow rate tit of the working fluid is given by

$\begin{array}{cc}\stackrel{.}{m}=\frac{4,300J\text{/}\sec }{143.35J\text{/}\mathrm{gm}}=30.00\mathrm{gm}\text{/}\sec \left(\mathrm{working}\mathrm{fluid}\right)& \left(22\right)\end{array}$

However, this working fluid is never expelled from the engine. It keeps circulating around the engine in a closed cycle described by the TS Diagram of FIG. 3. What is consumed by the engine when using the fuel burning heating system 28 is a small amount of fuel that is used to heat the compressed working fluid in the heating system 28 to a temperature of 541.07° K (514.24° F.) which is significantly lower than that of internal combustion engines, and a small amount of ordinary water that is evaporated inside the condenser 16 to condense the expanded working fluid at sub-ambient temperatures thereby enabling the engine to operate as a closed-cycle condensing heat engine with ultra high fuel efficiency.

The required rate {dot over (m)}_B that Butane has to be burned to heat the compressed working fluid in the heating system 28 to generate the 4,300 watts of propulsive power to propel the vehicle at a constant speed of 50 mph is given by equation (8). This burning rate would be

{dot over (m)}_B=0.0070{dot over (m)}=0.210 gm/sec   (23)

This is less than 1/10 the rate gasoline has to be burned in a vehicle propelled by a conventional internal combustion engine to achieve the same speed.

The specific gravity of liquified Butane at 60° F. is 0.5839. Thus, a small 5 gallon tank of Butane fuel would be sufficient to propel the vehicle moving at 50 mph a distance of 731 miles assuming that the air temperature remains constant at 60° F. and the humidity remains at 10%. The fuel cost would be about $10.00. In comparison, the fuel cost to drive a conventional automobile 731 miles propelled by an internal combustion engine would be about $200.

The rate {dot over (m)}_W that the water is evaporated in the condenser 16 required to generate a continuous power output corresponding to a given mass flow rate {dot over (m)} of working fluid is given by the working fluid/water mass ratio equation (6). Thus, the rate that water has to be evaporated inside the condenser 16 to enable the engine to generate a continuous power of 4.3 KW when the ambient air temperature is 60° F. and the relative humidity is 10% is given by

$\begin{array}{cc}{\stackrel{.}{m}}_W=\frac{\stackrel{.}{m}}{R}=\frac{30.0}{8.0681}=3.72\mathrm{gm}\text{/}\sec \left(\mathrm{water}\right)& \left(24\right)\end{array}$

In comparison, the rate that gasoline is consumed in propelling a standard full-size automobile on a level highway at a speed of 50 mph is approximately 3 gm/sec. Thus, the present invention will provide an engine for propelling road vehicles by evaporating ordinary water, at zero cost at a rate not too much greater than the rate that gasoline is burned in an internal combustion engine which is now approaching $4/gal in the United States (and over twice this cost in Europe). In addition, with a water tank 10 having a capacity of 150 liters (40 gal), a vehicle propelled by the present invention moving at an average speed of 50 mph when the ambient air temperature is 60° F. and relative humidity is 10% will have a range of 560 miles which is about twice that of conventional vehicles propelled by internal combustion engines burning gasoline. When the water tank is nearly empty, the driver stops at an ordinary gas station, refills the water tank in one or two minutes without any cost, and resumes driving.

As a comparison, if the vehicle is an electric vehicle, the time to recharge the batteries would be about 8 hours which has to be repeated about every 60 miles. (See the article, “It's The Battery, Stupid,” by Stuart Brown, Popular Science, February 1995, pp. 62-64, 78.) Moreover, the weight of the storage batteries would be much more than the weight of 40 gal of water. The recharging cost would also be very high. But since less than one percent of the gas stations across the United States are equipped with recharging systems for electric vehicles, such long-distance cross-country trips are essentially impossible with electric vehicles. For a conventional vehicle using gasoline in an internal combustion engine, the cost to re-fill a 20 gallon gas tank is approaching $80.

FIG. 4 is a schematic block diagram illustrating the basic design and operating principles of the isentropic multistage expansion system 40 (FIG. 2). Since the engine is designed to generate bursts of power that can exceed 100 KW, the mass flow rates {dot over (m)} of the engine can range from just a few gm/sec to over 100 gm/sec. Therefore, in the preferred embodiment, the expansion system is designed to operate as three parallel multistage expansion systems 42, 44, 46, each comprising two individual serially connected isentropic reciprocating expanders having equal mass flow rates. When the engine is operating at 4.3 KW where {dot over (m)}=30.00 gm/sec, each group of expanders will operate with a mass flow rate of 10.00 gm/sec. In the first group 42, the expanders are 48, 50. In the second group 44, the expanders are 52, 54. And in the third group 46, the expanders are 56, 58. Thus, the system comprises 6 expanders. To achieve smooth torque on the engine's drive shaft, each group of expanders are designed to generate equal rotational forces on the drive shaft that are serially connected having equal rates of mass flow and pressure ratios. In the numerical example, the high and low pressures of each of the three parallel multistage systems 42, 44, 46, are equal to 300 Bar=P_B and 9.256 Bar=P_D, respectively. Hence, the total pressure ratios r of each group of expanders is given by:

$\begin{array}{cc}r=\frac{P_B}{P_D}=\frac{300.00}{9.256}=32.411& \left(25\right)\end{array}$

Consequently, each expander will be designed with an expansion ratio r₀ equal to

r₀=32.411^1/2=5.693   (26)

The detailed design and construction of the expanders in the expansion system 40 in the applicant's engine are well known in the prior art of cryogenic engines and is omitted from this disclosure. See for example, “Performance of An Air Expansion Engine,” by J. E. Jensen, Advances in Cryogenic Engineering, Vol. 1 1960, Plenum Press, New York, pp. 105-110. Examples of the design of multistage high pressure expanders used for propelling road vehicles with cryogenic engines using liquid nitrogen as a working fluid are described in the article: “Liquid Nitrogen As An Energy Source For An Automotive Vehicle,” by M. V. Sussman, Advances in Cryogenic Engineering, Vol. 25, 1980, Plenum Press, New York, pp. 832-837. The expanders described in this paper are very similar to the applicant's expander's.

As described above, by increasing the mass flow rate {dot over (m)} of the working fluid withdrawn from the compressed gas energy storage system 34, it will be possible to achieve significantly higher propulsive powers. Assuming that the peek power of each expander is 30 KW, the maximum accelerating power of the engine will be 180 KW=241 HP. This is higher than most conventional internal combustion engines used for propelling automobiles. However, since continuous high powers on the order of 100 KW is completely unnecessary for propelling ordinary passenger vehicles, the fact that the internal combustion engines that propel conventional automobiles require, by necessity, the generation of continuous power levels of 100 KW or more is another reason why internal combustion engines used for propelling ordinary passenger vehicles are impractical and very inefficient.

Efficiencies of high pressure expanders often exceed 90%. In addition, high pressure expanders operating in road vehicles can be easily designed to decelerate the vehicle by re-compressing the working fluid using the vehicle's kinetic energy as an energy storage system. Thus, by merely reversing the input and output gas flow into these expanders, they can be made to operate as isentropic re-compressors at very high efficiency by re-compressing the discharged gas, and feeding it back into the upstream expander (operating as a compressor) and back into the compressed gas storage system 34. Thus, in the preferred embodiment, the expansion system 40 will be designed such that each isentropic expander can operate in reverse as isentropic compressors such that the expansion system 40 will provide both positive propulsive thrust and regenerative negative decelerating thrust at an efficiency of about 90%. The power that generates the decelerating breaking comes from the vehicle's rotating drive shaft thereby decelerating the vehicle.

FIG. 5 is a horizontal cross section through the thermally insulated compressed gas storage system 34 that operates as a load-leveling compressed gas energy storage system before the heated compressed gas contained therein is fed into the expansion system 40. In order to withstand the very high pressures of the heated compressed gaseous working fluid, the system 34 is designed as a plurality of 4 serially connected, thermally insulated cylindrical high pressure gas storage vessels 60 mounted parallel to each other on the bottom of the vehicle. They are constructed with ultra high strength composite material 62. The design and construction of ultra high strength compressed gas storage vessels are well known in the prior art and the detailed construction is omitted. (See, for example, “Strength Design Criteria for Carbon/Epoxy Pressure Vessels,” by Stephen R. Swanson, Journal of Spacecraft & Rockets, Vol. 27, No. 5, 1990; and “On The Design Of The Φ800 Wire-Wound Superhigh Pressure Vessel,” by Dazeng Su, Pressure Vessel Technology, Vol. 1, Design & Analysis, 1989, Pergamon Press, pp. 341-347.) In order to eliminate heat loss, the external walls of the high pressure gas storage vessels 60 are enclosed within a thick blanket of multilayer cryogenic thermal insulation 64. (See Chapter 7, Cryogenic Fluid Storage and Transfer Systems, pp. 447-535 in Cryogenic Systems by Randall Barron, McGraw-Hill Book Company, New York, 1966.)

In order to enable the engine to operate at power levels above the continuous 4.3 KW of continuous power for certain time periods, the mass flow rate of compressed working fluid leaving the compressed gas storage system 34 has to be greater than the mass flow rate entering the system without any appreciable loss of pressure. To achieve this, the total amount of compressed gas stored in the 4 compressed gas storage cylinders 60 has to be fairly large.

The amount of mechanical output work W generated by isentropically expanding compressed gas from an initial pressure P₁, and initial volume V₁, to a greater volume V₂, is given by the well known equation:

$\begin{array}{cc}W=\frac{P_1V_1}{\gamma -1}\left[1-{\left(\frac{V_1}{V_2}\right)}^{\gamma -1}\right]& \left(27\right)\end{array}$

where γ is a constant equal to the specific heat ratio c_P/c_V of the gas at constant pressure and constant volume. Consequently, the maximum output work W_max obtained in the limit V₂→∞ is given by

$\begin{array}{cc}{W}_{\mathrm{Max}}=\frac{P_1V_1}{\gamma -1}& \left(28\right)\end{array}$

For the R32 working fluid, the constant γ=1.816. In the preferred embodiment, the inside length and diameter of each storage cylinder 60 is 3 m and 0.3 m, respectively. Hence, the total interior volume V₁ of the 4 cylinders 60 is 0.820 m³. Since the design pressure is 300 Bar (3×10⁷ N/m²), the total stored energy in the system is 3.02×10⁷ J. This is enough stored energy to generate 4.3 KW of power continuously for about 1.0 hours without operating the re-compression system 26. The condensed working fluid generated in the condenser 16 is simply withdrawn from the condenser 16 and fed into the thermally insulated liquified gas holding vessel 20 where it is accumulated (FIG. 2). The engine is designed to enable all the sub-systems (condenser 16, compressor 26, heating system 28, compressed gas holding system 34, and expansion system 40) to operate independently of each other with different mass flow rates {dot over (m)} and at different times. This enables the mass flow rate {dot over (m)} of working fluid passing through the condensing system 16 to be significantly different from the rate of mass flow through the various expanders for fairly long time periods.

As described above, the engine is designed to operate automatically even when the vehicle is parked and not being used. This is an important design feature of the invention. The purpose is to automatically restore all the energy storage vessels 60 to full capacity before the engine is turned off. This will enable the engine to instantly operate at very high power levels even after not running for several days. In addition, when the engine is operating to propel the vehicle, all of the various sub-systems (e.g., the condenser 16, the compressor 26, and the expansion system 40), are automatically controlled via servo control actuators 65 mounted at various control points that control the mass flow rates of the working fluid and the evaporating water to achieve maximum efficiency. This is achieved by using many electronic sensors (transducers) 38 mounted at all the flow points described in FIGS. 2 and 3. Other sensors 68 continuously monitor the local air temperature, humidity, and barometric pressure, and continuously feeds this information into the engine's control computer 24 that determines the best time to operate all the automatic subsystems when the vehicle is parked and not being driven by the driver. When the vehicle is being driven, sensors 70 connected to the driver's control actuators (accelerator and breaking control pedal etc.) are sent to the control computer 24 and processed along with all the other incoming sensor signals. The computer 24 then generates optimum control signals which are fed into the various electronic actuators mounted near the various flow points to operate all the subsystems of the engine to achieve the driver's desired accelerating power and decelerating regenerative breaking with maximum efficiency.

Since all the subsystems of the engine are designed to operate independently of each and at different times, and since the holding capacities of the liquified gas storage vessel 20, the water vessel 10, and the compressed gas storage system 34 are very large, there are several ways that the engine could be operated to obtain more efficiency. For example, when the compressed gas energy storage system 34 is full to maximum capacity which will occur when the liquified gas vessel 20 is nearly empty, the vehicle could be driven without the re-compression system 26 turned on. In this case the engine's specific output work Ŵ_Net=Ŵ_E=170.59 J/gm. Consequently, the propulsive power P of the engine would be given by:

P=Ŵ_Net{dot over (m)}=170.59{dot over (m)}  (29)

Therefore, to propel the vehicle at 50 mph, (which would require a propulsive power of 4,300 Watts) the required mass flow rate {dot over (m)} of the working fluid would only be

$\stackrel{.}{m}=\frac{P}{{\hat{W}}_{\mathrm{Net}}}=25.21\mathrm{gm}\text{/}\sec$

instead of 30.00 gm/sec. At this rate, the vehicle could be driven 50 miles during one hour and the amount of working fluid used would be 91 kg which would be liquified and fed into holding vessel 20. The amount of water consumed would only be 13.8 kg (13.8 liters=3.5 gal). Thus, the only “fuel” consumed during the trip would be 3.5 gal of water at zero cost.

Although the design and construction of high-pressure expanders and heat exchangers are well known in the prior art of cryogenic engines (see Cryogenic Systems, by R. Barron, McGraw-Hill Book Company, New York, 1966) the design and construction of the low temperature condensing system using evaporating water as the heat sink for a closed-cycle condensing heat engine designed for propelling road vehicles operating according to the principles set forth in the present invention is new and requires a detailed description. It is one of the most important operating systems of the engine.

Before presenting the design and operating principles of the preferred embodiment of the sub-ambient condensing system 16 that is artificially created by evaporating water, it is important to understand the basic theoretical aspects of heat flow through a metal conductor due to a temperature gradient. For a material of constant cross section and constant thermal conductivity K, in a steady state, the magnitude of the heat current {dot over (Q)} flowing through the material is given by the following well-known Fourier heat conduction equation:

$\begin{array}{cc}\stackrel{.}{Q}=\mathrm{KA}\frac{\Delta T}{\tau }& \left(30\right)\end{array}$

where A is equal to the cross sectional area of the material, ΔT is equal to the difference in temperature between the two ends, and τ is equal to the distance between the two ends (thickness).

For example, for copper K=0.92 cal/sec-cm-°C. Hence, a thin sheet of copper 100 cm wide and 10 cm long having a thickness τ=0.20 cm, and a temperature difference ΔT=0.10° C. between the surfaces will generate a heat flow {dot over (Q)} of:

$\stackrel{.}{Q}=\mathrm{KA}\frac{\Delta T}{\tau }=0.92\times 100\times 10\left(\frac{0.1}{0.2}\right)=460.0\mathrm{cal}\text{/}\sec =1,923J\text{/}\sec$

In order to achieve heat transfer inside the condenser between the evaporating water and the condensing working fluid, there must be a certain temperature difference to obtain the heat transfer effect. However, in the above thermodynamic analysis of the engine it was assumed that the vapor entering the condenser had exactly the same temperature as the evaporating water. This is a common assumption used in the thermodynamic analysis of all heat engines. However, in the real situation, the actual temperatures will never be equal to each other because in this case there would be no heat transfer. Thus, the actual performance of any heat engine will always be somewhat less than that computed by the theoretical thermodynamic analysis. However, if the temperature differences are small, the actual performance will be fairly close to the calculated theoretical performance. Thus, the condenser described herein will be designed on the assumption that the temperature difference between the working fluid leaving the condenser will be within 0.2° K of the evaporating water temperature. It follows from equation (30) that in order to achieve this, the area of the adjacent heat transfer surfaces must be fairly large, and the thickness must be small.

FIG. 6 is a schematic transverse cross section through the preferred embodiment of the condenser 16 that is 120 cm long, 60 cm wide and 16 cm thick mounted vertically under the hood near the front of the vehicle such that a constant stream of unsaturated air 72 flows through it over wetted pads 74 with a fairly high rate of mass flow. The wetted pads 74 are mounted along the external surfaces 76 of a vapor conduit 78 (condensing conduit) containing condensing working fluid 80. FIG. 7 is a schematic transverse cross section through the vapor conduit 78 illustrating its rectangular (channel-like) geometry. In the preferred embodiment, the inside dimensions of the vapor conduit 78 are 1.0 cm high and 12.0 cm wide that is divided into 12 separate parallel ducts (condensing tubes) 82 by internal walls 84 as shown in FIG. 7. In order to achieve high heat transfer through the external walls of the condensing conduit 78, the conduit 78 is made of copper with an external wall thickness of 0.20 cm.

FIG. 8 is a longitudinal cross section through the rectangular condensing conduit 78 illustrating the design and construction of the water evaporation pads 74 that are mounted on both sides of the condensing conduit 78. In the preferred embodiment, the water pads 74 are made of thin layers of specially made water-absorbing padding material of the type used in commercial evaporators operated by evaporating water on wetted surfaces. (See Evaporative Air-Conditioning by G. Bom et al., World Bank Technical Paper No. 421, The World Bank, Washington D.C., 1999, pp. 23-25.) The pads 74 are wetted by a network of small diameter capillary tubes 86 containing pressurized water from the water pump 14. (See FIG. 2.) The water is discharged from these capillary tubes 86 via several hundred spaced apart holes 88 having a very small diameter.

The condensing conduit 78 is mounted inside the condenser 16 (transverse to the air stream 72 flowing through it) such that it traverses back and fourth 12 times inside the condenser 16 as illustrated in FIG. 6 such that the total mean length of the condensing conduit 78 is 1,464 cm (14.6 m or 48 ft). Hence, the total wetted surface area A_W of the condensing conduit 78 is 2×[1464 cm×12 cm]=35,136 cm². Since the air stream 72 is blowing at a fairly high speed in a transverse direction through the condenser 16 over the wetted surfaces 74 which only travels a distance of 12 cm (transversely) over the wetted surfaces 74, each gram of air in the air stream 72 picks up very little evaporating water. In addition, the external lateral walls of the condenser 16 are fitted with multilayer cryogenic thermal insulation 90 to prevent any heat from entering the condenser 16 aside from that carried by the incoming air stream 72. However, since the specific heat of the air in the airstream 72 is very low relative to the evaporating water, and since the air passes through the condenser 16 in such a brief time period, essentially all of the heat absorbed by the evaporating water comes from the heat flow through the copper walls of the condensing conduit 78 carrying the condensing working fluid 80. This enables essentially all of the latent heat of the evaporating water to be used as a heat sink for absorbing the heat of vaporization from the saturated vapor 92 flowing through the condensing tubes 82.

Since the saturated vapor 92 must enter the condensing tubes 82 at a temperature slightly higher than the temperature of the evaporating water so that the latent {circumflex over (Q)}_L of the evaporating water can be used for absorbing the latent heat of vaporization {circumflex over (Q)}_V of the working fluid, the system is designed such that this temperature difference ΔT=0.2° C. Consequently, it follows from equation (30) that the total heat flow from the vapor flowing through the condensing conduit 78 and absorbed by the evaporating water is:

$\begin{array}{c}\stackrel{.}{Q}={\mathrm{KA}}_W\frac{\Delta T}{\tau }\\ =0.92\times 35136\times \frac{0.2}{0.2}\\ =33524\mathrm{cal}\text{/}\sec \\ =135152J\text{/}\sec \end{array}$

In the above thermodynamic analysis where the engine is operating at an ambient air temperature of 60° F. and a relative humidity of 10%, the latent heat {circumflex over (Q)}_L of evaporating water is 2,491.12 J/gm. Consequently, since the preferred embodiment of the condenser 16 will give a maximum heat transfer rate {dot over (Q)}=135,152 J/sec, it will be possible to operate the engine at a maximum water mass flow rate {dot over (m)}_W given by:

$\begin{array}{cc}{\stackrel{.}{m}}_W=\frac{\stackrel{.}{Q}}{{\hat{Q}}_L}=\frac{135152J/\sec }{2491.12J/\mathrm{gm}}=54.25\mathrm{gm}/\sec & \left(31\right)\end{array}$

Since the engine's specific output power without operating the recompressor system 26 is given by

P_W=RŴ_E{dot over (m)}_W=1376.34{dot over (m)}_W (Watts)

this condensing system will be able to run the engine continuously at a maximum power of 75 KW (100 HP). This is 17 times greater than the preferred continuous power output of 4.3 KW which, according to equation (24), only requires a water evaporation mass flow rate of 3.72 gm/sec.

Since the design of the high pressure isentropic compressor 26 (FIG. 2) is well known in the prior art of cryogenic systems, a detailed description is omitted from this disclosure. (See the book Cryogenic Systems by Randall Barron, McGraw-Hill, 1966, New York.)

FIGS. 9 and 10 are transverse and longitudinal schematic cross sections through the preferred embodiment of the main compressed gas heating chamber 32 of the heating system 28 which comprises two serially connected heating chambers 30, 32 where the first heating chamber 30 is a pre-heating chamber utilizing the combustible gases discharged from the main heating chamber 32 as the heat source for the first pre-heating chamber 30. Both of the heating chambers 30, 32 have essentially the same designs. (See FIG. 2.) The heat source for the main heating chamber 32 is generated by burning small amounts of a combustible, clean-burning fuel, such as Butane. After the combustible gases are discharged from the second heating chamber 32 at a temperature of 541.07° K, they are fed into the first heating chamber 30 via a thermally insulated conduit 94 (FIG. 2) and used as the heat source for the first heating chamber 30. Since the preferred mass flow rate tit of working fluid required to generate 4.3 KW of continuous propulsive power is 30.00 gm/sec, which has to be fed through and heated inside the heating system 28 which can be much higher when generating bursts of power at 50 KW or more, the inlet conduit 96, is sub-divided in the chambers 30 and 32 into six separate sub-conduits 98 (FIG. 10) containing equal mass flow rates. When the engine is operating at a power of 4.3 KW, each of these sub-conduits 98 has a mass flow rate of 30/6=5 gm/sec. As shown in FIG. 10, the sub-conduits 98 are attached to each other and mounted inside the heating cylinder 32 in a spiraling mounting structure 100 to achieve maximum heat transfer with the hot combustion gases 102 passing through the heating chamber 32. The heating chamber 32 is designed as a fairly long thermally insulated cylinder 104 that is constructed with material having very low thermal conductivity such as fibreglass. A thick jacket of multilayer thermal insulation 106 is mounted around the outer walls of the cylinder 104. The entire inside surfaces 108 of the cylinder 104 are coated with a mirror-like material 110 having very high reflectivity such that the heat generated inside the cylinder 104 by the burning fuel is absorbed by the working fluid passing through the heating coils 98. In the preferred embodiment, both heating chambers 30, 32 are mounted under the vehicle's chassis.

As described above, the interior 112 of the cylinder 104 contains the 6 heat-absorbing tubular coils 98 containing the compressed working fluid spiraling around the cylinder's longitudinal axis 114 in a group of six tubes forming layers 116 that spiral around the cylinder's longitudinal axis 114.

As shown in FIGS. 9 and 10, a spiraling region 118 is constructed inside the cylinder 104 with spiraling walls 120 (made of copper having high thermal conductivity) such that each of the spiraling layers of 6 tubes 116 are centrally mounted inside the spiraling region 118 and separated from each other by the spiraling copper walls 120. This spiraling region 118 contains hot gasses 102 generated from the combustion of the Butane fuel. The design is such that the heating coils 98 are completely immersed within the hot gases thereby heating the compressed working fluid flowing inside the heating coils 98 to achieve maximum heat transfer. The external surfaces of the heating coils 98 are painted with black paint specially designed to maximize heat absorption from the hot gases in contact with it. The Butane fuel is burned at a maximum temperature of 3,870° F. (2,405° K) in a small combustion region 122 located at the forward end of cylinder 104. The fuel is pumped into the burning region 122 via a conduit 124 from a 50 liter (13.2 gallon) Butane fuel tank 126 via a fuel pump 128 and mixed with ambient air 130 that is also pumped into the burning region 122 by another pump 132 via another conduit 134. The mixture is ignited by a small spark plug 136.

The burning process generates the high temperature combustion gases 102 that spirals through the cylinder 104 (through the spiraling region 118) thereby heating the heating coils 98 immersed within it This high temperature gas 102 leaves the combustion region 122 and enters the spiraling region 118 containing the six spiraling 116 layers of the heating tubes 98 as it moves through the cylinder 104 towards the opposite end.

In the preferred embodiment, the heating cylinder 104 will be designed to be 1.0 m (3.28 ft) long with an inside diameter of 40 cm (1.31 ft). The high pressure heating coils 98 will be made of stainless steal with an inside diameter of 1.5 cm and a wall thickness of 0.3 cm. With these dimensions the cross sectional area of the heating coils 98 will be 1.77 cm². Since the mass flow rate of the compressed working fluid passing through each tube 98 when the engine is operating at 4.3 KW is 5 gm/sec and since the average density is 1.4 gm/cm³, the flow velocity through each tube will be 2.02 cm/sec. The pitch of the spiraling walls 120 (and layers 116 of the heating coils 98) are designed to be one revolution per 3.5 cm of longitudinal distance along the central axis 114. Consequently, the average total length of each heating coil 98 will be 2,692 cm. Thus, the average amount of time that the compressed working fluid spends in spiraling through the heating cylinder will be 1296 sec or 21 min. This is more then enough time to heat the working fluid from the intermediate temperature T₁ to the design temperature of 541.07° K (514.24° F.) described in the thermodynamic analysis of the engine. When the engine is operating at higher power levels, the heating time will just be a few minutes.

When the combustion gases 102 spiraling through the cylinder 104 reaches the end 138 of the cylinder 104 it is discharged through a small exhaust conduit 140. The end of this exhaust conduit 140 is fitted with a pressure activated relief valve 142 that only opens if a certain threshold pressure is reached. This enables the hot gases 102 to have a relatively high pressure and hence a high specific heat as it passes through the spiraling region 118 containing the six heating coils 98 on its way out of the cylinder 102.

Since the temperature inside the spiraling region 118 is highest at the burning region 122 near the beginning of the cylinder 104, the heat transfer to the compressed liquefied working fluid spiraling through the heating coils 98 is maximum at the beginning of the cylinder 104 Thus, the rate of heat absorption by the compressed working fluid flowing through the heating conduits 98 is initially very high. But since the heating coils 98 are relatively long as it spirals around inside the heating cylinder 104, the compressed gaseous working fluid 144 inside the coils 98 continues to be heated as it spirals through the heating cylinder 104 immersed in the hot combustion gases 102. Consequently, the heat transfer from the hot gases 102 to the compressed working fluid 144 will be very efficient as the working fluid spirals through the heating cylinder 104. By controlling the rate at which the fuel is burned, the fuel and air injection pressures, and the activation pressure of the exhaust gas relief valve 142, it will be possible to control the heating process such that both the compressed working fluid 144 and the hot combustion gases 102 leave the heating cylinder 104 at about the same temperature.

According to the numerical example, this exhaust temperature will be 541.07° K (514.24° F.). Thus, this hot exhaust gas 102 is fed into a thermally insulated conduit 94 which feeds this hot exhaust gas 102 from the second heating chamber 32 into the first heating chamber 30 where it is utilized as the heating source for pre-heating the compressed working fluid to some intermediate temperature T₁ before it is fed into the second (main) heating chamber 32. (See FIG. 2.) The design and construction of the first heating chamber 30 is essentially identical to that of the second heating chamber 32 described above. (In place of the burning region 122, the first heating chamber 30 uses the conduit 94 to feed the hot gases into a similar region in the first heating chamber 30.) After the hot gases is circulated through the first heating chamber 30, its temperature falls to the intermediate temperature T₁ which is not very far above the natural ambient temperature of atmosphere. Thus, with this heating system that is thermodynamically represented by the upward curve BC on the TS Diagram (FIG. 3), essentially all of the heat of combustion of the fuel is used to heat the working fluid. In internal combustion engines used for propelling road vehicles, the temperature of the exhaust gases is about 400° K (260° F.) which is another reason why internal combustion engines are so inefficient.

FIG. 13 is a schematic longitudinal cross-section of a vehicle 146 propelled by the invention showing the locations of the water tank 10, sub-ambient condenser 16, liquified working fluid holding vessel 20, isentropic compression system 26, heating system 28, compressed gas storage system 34, and isentropic expansion system 40.

The above preferred embodiment of the invention was specifically designed for propelling road vehicles. However, the engine could also be used for propelling other vehicles such as small boats and large ocean going ships of all types. It could also be used for powering trains and propeller driven aircraft. Another useful application would be for generating low-cost bulk electric power. Smaller engines operating at about 5 KW could be used for generating electricity for private homes.

As various other embodiments, changes, and modifications, can be made in the above method and apparatus for generating silent and clean mechanical power at very high efficiencies and high power densities by harnessing the thermal potential difference between an artificial sub-ambient heat sink generated by evaporating water and a high temperature heat reservoir maintained at relatively low temperatures by burning small amounts of a clean burning fuel by using a phase-changing working fluid having a critical temperature above the temperature of evaporating water without departing from the spirit or scope of the invention, it is intended that all subject matter contained in the above description or shown in the accompanying drawings should be interpreted as illustrative and not in a limiting sense.

Claims

1. A method for generating mechanical power at high efficiency comprising the steps of:

creating an artificial low temperature heat sink at a temperature below ambient temperature by evaporating water;

condensing a vaporized phase-changing working fluid at sub-ambient temperature having a critical temperature close to said ambient temperature and above said low temperature heat sink by extracting its latent heat of condensation by said low temperature heat sink by evaporating said water;

compressing said condensed working fluid;

heating said compressed working fluid into a compressed gas by absorbing heat energy generated by burning a combustible fuel;

expanding said vaporized compressed working fluid inside an expander means thereby converting a portion of said absorbed heat energy into mechanical work; and

repeating said condensing, compressing, vaporizing, and expanding steps in a cyclic process by evaporating additional water and burning additional fuel.

2. A method as set forth in claim 1 wherein said working fluid is a refrigerant.

3. A method as set forth in claim 1 wherein said working fluid is R32 refrigerant.

4. A method as set fourth in claim 1 wherein said step of creating said artificial low temperature heat sink comprises at least one condensing tube and the steps of:

mounting water absorbing padding means in thermal contact with said condensing tube; and

dispensing water on said padding means such that the evaporation of said water generates said low temperature heat sink inside said condensing tube such that when vaporized working fluid enters said condensing tube its latent heat of condensation is absorbed by said evaporating water and condenses into a liquid at sub-ambient temperature inside said condensing tube.

5. A method as set forth in claim 1 further comprising the step of interposing compressed gas storage vessel means between said heating step and said expansion step as a load-leveling system for varying the amount of power generated by said expander means by varying the mass flow rate at which said compressed heated working fluid is fed into said expander means.

6. A method as set forth in claim 1 wherein said mechanical power is used to propel a road vehicle.

7. A method as set forth in claim 1 wherein said condensing, compressing, heating, and expanding steps can be performed independently of each other and at different times.

8. A method for operating a condensing heat engine for generating mechanical power at high efficiency comprising the steps of:

utilizing a phase-changing working fluid having a critical temperature close to ambient temperature and above the temperature of evaporating water;

creating a high temperature heat reservoir not far above ambient temperature by burning small amounts of a combustible fuel;

creating a low temperature heat reservoir below ambient temperature by evaporating small amounts of water; and

operating said condensing heat engine between said high temperature heat reservoir and said low temperature heat reservoir such that relatively high mechanical power can be generated by burning relatively small amounts of said combustible fuel.

9. A method for generating mechanical power comprising the steps of:

creating an artificial low temperature heat sink at a temperature below ambient temperature by evaporating water;

generating a high temperature heat reservoir above said ambient temperature; and

converting a portion of the thermal potential difference between said low temperature heat sink and said high temperature heat reservoir into mechanical power.

10. An apparatus for generating mechanical power at high efficiency comprising:

a phase-changing working fluid having a critical temperature close to ambient temperature and above the temperature of evaporating water;

condenser means maintained at sub-ambient temperature by evaporating water in thermal contact with said condenser means;

heating means by burning a combustible fuel;

means for feeding vaporized working fluid into said condenser means thereby condensing said vapor by extracting latent heat of condensation by said evaporating water at sub-ambient temperature;

means for compressing said condensed working fluid;

means for heating said compressed working fluid into a heated compressed gas by feeding said compressed working fluid into said heating means thereby absorbing heat energy from said combustible fuel;

means for feeding said heated compressed gas into an expander means thereby converting a portion of said heat energy absorbed from said burning fuel into mechanical work; and

means for repeating said condensing, compressing, heating, and expanding steps in a cyclic process for generating more mechanical work.

11. An apparatus as set forth in claim 10 wherein said working fluid is a refrigerant.

12. An apparatus as set forth in claim 10 wherein said working fluid is refrigerant R32.

13. An apparatus as set forth in claim 10 wherein said condenser means comprises:

at least one condensing tube;

water absorbing padding means mounted on the external surfaces of said condensing tube in thermal contact with said condensing tube;

means for dispensing water on said padding means that evaporates on said padding means thereby reducing the temperature of said condensing tube below said ambient temperature; and

means for introducing vaporized working fluid into said condensing tube which condenses into a liquid at sub-ambient temperature by extracting its latent heat of condensation by said evaporating water.

14. An apparatus as set forth in claim 10 further comprising:

compressed gas storage vessel means interposed between said heating means and said expander means as a load-leveling system; and

means for varying the amount of mechanical power generated by said expander means by varying the mass flow rate at which said compressed working fluid is fed into said expander means.

15. An apparatus as set forth in claim 10 wherein said mechanical power is used to propel road vehicles.

16. An apparatus as set forth in claim 10 wherein said condensing means, compressing means, heating means, and expanding means can occur independently of each other and at different times.

17. An apparatus for generating mechanical power at high efficiency comprising:

a phase-changing working fluid having a critical temperature close to ambient temperature and above the temperature of evaporating water;

condenser means maintained at sub-ambient temperature by evaporating water in thermal contact with said condenser means;

means for feeding vaporized working fluid into said condenser means thereby condensing said vapor by extracting latent heat of condensation by said evaporating water at sub-ambient temperature;

means for compressing said condensed working fluid;

means for heating said compressed working at a temperature above said critical temperature by absorbing heat energy from a combustible fuel;

means for feeding said compressed working fluid into said heating means;

means for feeding said heated compressed gas into an expander means thereby converting a portion of said absorbed heat energy into mechanical work; and

means for repeating said condensing, compression, heating, and expanding steps in a cyclic process for generating more mechanical work.

18. An apparatus for generating mechanical power at high efficiency comprising:

a phase-changing working fluid having a critical temperature close to ambient temperature and above the temperature of evaporating water;

condenser means maintained at sub-ambient temperature by evaporating water in thermal contact with said condenser means;

means for feeding vaporized working fluid into said condenser means thereby condensing said vapor by extracting latent heat of condensation by said evaporating water at sub-ambient temperature;

means for compressing said condensed working fluid;

means for heating said compressed working at a temperature above said critical temperature by absorbing heat energy from a combustible fuel;

means expanding said heated compressed gas thereby generating mechanical power; and

means for repeating said condensing, compression, heating, and expanding steps in a cyclic process for generating more mechanical work.

19. A condensing heat engine for generating mechanical power at high efficiency comprising:

a phase-changing working fluid having a critical temperature close to ambient temperature and above the temperature of evaporating water;

means for creating a high temperature heat reservoir not far above ambient temperature by burning small amounts of a combustible fuel;

means for creating a low temperature heat reservoir below ambient temperature by evaporating small amounts of water; and

means for operating said condensing heat engine between said high temperature heat reservoir and said low temperature heat reservoir such that relatively high mechanical power can be generated by burning relatively small amounts of said combustible fuel.