REFRIGERATION

CLAIM OF PRIORITY

[0001] This application is related to, and claims priority to, the provisional utility application entitled "Closed Low Grade Heat Driven Rankine/Reverse Rankine Cycle Driving a Reverse Brayton Cycle Using an Ejector with a Liquid Desiccant Loop to Produce Shaft Power and Refrigeration by Adiabatic Expansion and Evaporation (The Sherbeck Cycle) " filed on December 22, 2015, having an application number of 62,271,134 and Attorney Docket No. 10046P002Z (Ml 6-038P), the entire contents of which are incorporated herein by reference.

COPYRIGHT NOTICE

[0002] A portion of the disclosure of this patent document contains material which is subject to copyright protection. The copyright owner has no objection to the facsimile reproduction by anyone of the patent document or the patent disclosure, as it appears in the Patent and Trademark Office patent file or records, but otherwise reserves all copyright rights whatsoever. TECHNICAL FIELD

[0003] Embodiments of the invention relate generally to the field of thermodynamics and energy conversion, and more particularly, to methods and systems for implementing a closed low grade heat driven Rankine cycle driving a reverse Brayton cycle using an ejector with a liquid desiccant loop to produce shaft power and refrigeration by adiabatic expansion and evaporation (the Sherbeck cycle).

BACKGROUND

[0004] The subject matter discussed in the background section should not be assumed to be prior art merely as a result of its mention in the background section. Similarly, a problem mentioned in the background section or associated with the subject matter of the background section should not be assumed to have been previously recognized in the prior art. The subject matter in the background section merely represents different approaches, which in and of themselves may also correspond to embodiments of the claimed inventions.

[0005] Energy efficiency is an ongoing endeavor in many areas of engineering and technology and market advantage may often be achieved through even very small gains in efficiency. Described herein are means for attaining greater efficiency in energy conversion.

[0006] Additional benefits are achieved through the described means including, for instance, a closed loop system whereas prior solutions are open loop and reuse of operating fluids.

[0007] Prior solutions which utilize an open loop cycle cause the machinery to foul with minerals from the water and dust from the air. Moreover, because such prior solutions are open loop cycles, when combined with turbo machinery, operation is excessively noisy and consumes excessive quantities of water.

[0008] The present state of the art may therefore benefit from systems and methods for implementing a closed low grade heat driven Rankine cycle driving a reverse Brayton cycle using an ejector with a liquid desiccant loop to produce shaft power and refrigeration by adiabatic expansion and evaporation as is described herein.

BRIEF DESCRIPTION OF THE DRAWINGS

[0009] Embodiments are illustrated by way of example, and not by way of limitation, and can be more fully understood with reference to the following detailed description when considered in connection with the figures in which:

[0010] Figure 1 depicts a system architecture in accordance with an embodiment of the invention;

[0011] Figure 2 depicts the four sub-cycles of a Sherbeck cycle in accordance with an embodiment of the invention;

[0012] Figure 3A depicts a waste heat driven steam Rankine cycle in accordance with an embodiment of the invention;

[0013] Figure 3B depicts a desiccant solution cycle in accordance with an embodiment of the invention;

[0014] Figure 3C depicts a reverse Brayton cycle in accordance with an embodiment of the invention;

[0015] Figure 3D depicts an evaporative cooler process in accordance with an embodiment of the invention;

[0016] Figure 4 depicts a Sherbeck cycle in accordance with an embodiment of the invention;

[0017] Figure 5 is a flow diagram illustrating a method for implementing a closed low grade heat driven Rankine cycle driving a reverse Brayton cycle using an ejector with a liquid desiccant loop to produce shaft power and refrigeration by adiabatic expansion and evaporation in accordance with described embodiments;

[0018] Figure 6 depicts an ejector refrigeration cycle which is heat driven in accordance with an embodiment of the invention;

[0019] Figure 7 depicts a Sherbeck cycle psychometric chart in accordance with described embodiments;

[0020] Figure 8 depicts another Sherbeck cycle in accordance with described embodiments;

[0021] Figure 9 depicts a T-s or temperature-entropy diagram of the reverse Brayton cycle within the Sherbeck cycle in accordance with described embodiments;

[0022] Figure 10 depicts yet another Sherbeck cycle in accordance with described embodiments; and

[0023] Figures 11A and 11B depict a complete Sherbeck cycle system in accordance with one embodiment of the invention.

DETAILED DESCRIPTION

[0024] Described herein are systems and methods for implementing a closed low grade heat driven Rankine cycle driving a reverse Brayton cycle using an ejector with a liquid desiccant loop to produce shaft power and refrigeration by adiabatic expansion and evaporation.

[0025] The 'Sherbeck' cycle as referred to and as described herein in conjunction with the various figures operates differently than prior solutions and provides many operational and efficiency advantages over such prior solutions.

[0026] A system designed to utilize exhaust heat to provide cooling and

dehumidification for air conditioning is referred to herein as the Sherbeck cycle. Such technology is composed of subsystems based on reverse Brayton cycles, steam Rankine cycles, evaporative cooling cycles and desiccant solution cycles. Such sub-cycles operate internally at atmospheric and sub-atmospheric pressures. A Sherbeck cycle system therefore utilizes only air and water as working fluids, which are non-hazardous and environmentally friendly, a key feature that distinguishes a Sherbeck cycle system from other cooling and dehumidification systems available in the marketplace today. Using a numerical model for such a Sherbeck cycle system it has been demonstrated that commercially acceptable performance for dehumidification in dry and moderately humid ambient conditions is attainable. [0027] Commercialization of the described Sherbeck cycle may provide means for off- grid refrigeration and electric power to the developing world and to remote locations which require off-grid power and/or refrigeration. The a Sherbeck cycle plant has the potential to be made at a modest cost and thus is feasible to deploy into a wide array of commercial, industrial, residential, medical, military, and various support solutions and power/refrigeration

implementations.

[0028] All known heat driven refrigeration schemes lose performance and degrade in operational efficiency when utilized in high temperature ambient conditions. However, the operational efficiency of power generation and refrigeration, especially in high temperature ambient conditions, may be boosted for such power generation plants through the use of such available water.

[0029] Moreover, unlike prior systems, because the described system is a closed loop system, a wide variety of working fluids can be employed at the discretion and preference of the operator and/or manufacturer. The system is capable of operating anywhere from sub- atmospheric, to very high atmospheric, using, for instance, super-critical carbon dioxide. The system is capable to work with low grade heat to a high heat limited only by the materials available.

[0030] At low temperatures, much of the system could be made very inexpensive and at high volumes through the use of plastics and with relatively few moving parts, the system enjoys a long useful operational life and is simple to maintain.

[0031] Further still, if the output shaft power from operating the system is then used to drive an alternator, to charge a battery for instance, then some of the converted energy could be stored for later use. Consider for instance the storing of solar power converted to shaft power through the described system to provide lighting at night.

[0032] Figure 1 depicts a system architecture 100 in accordance with an embodiment of the invention.

[0033] In particular, there is depicted an alternator 101 by which to output conversion power, for instance, converting shaft power to electrical output power for storage or immediate use.

[0034] The depicted cycle utilizes a gas, a refrigerant, and a desiccant to produce shaft power, and refrigeration from low grade heat, including heat which may be waste heat from another process or solar heat.

[0035] Although ejector refrigeration systems have been around since 1901 there is recent attention being paid to improving such systems to raise their coefficient of performance and cooling power because they can operate on low grade waste heat or solar power using natural refrigerants like water without the ozone depletion or global warming concerns associated with other refrigeration practices.

[0036] Different schemes attempting to raise the performance of heat driven

refrigeration are emerging, but those available to date suffer various drawbacks and challenges.

[0037] For instance, prior solutions such as the Maisotsenko cycle-based power system, the Rankine cycle, and the Brayton cycle are all well known and conventionally available solutions, but each may nevertheless be improved upon.

[0038] The depicted system 100 depicts a power plant capable of addressing the operational problems that prior known "open cycle" plants are known to encounter in practice owing to the simple fact an open cycle permits fouling from dust in the air, minerals in the makeup water, and noise pollution from the inlet and exhaust being excessively noisy. For instance, the closed cycle of the described Sherbeck system generates water from the absorbed conditioned space humidity whereas open cycle systems require a water supply to input moisture or water into such open systems. In such a way, the Sherbeck system's closed cycle provides both a more economical and a more environmentally friendly implementation.

[0039] Conversely, the depicted plant is a closed system for which cooling is accomplished with dry heat exchange coils to the air and optionally supplemented by external misting of the coils when operating in hot weather and implementations having high temperature ambient conditions.

[0040] Mineral fouling of the coils may thus be cleaned externally without necessitating disassembly the plant which therefore improves useful life and simplifies maintenance.

[0041] Alternatively, in areas where a suitable source of cold water exists, heat may be rejected to water. The heat to drive the plant may be taken directly from a hot gas stream, or the plant may alternatively be configured to run off of hot liquid, or vapor.

[0042] As depicted within system 100 representing such a plant is ejector 105 having Q _e refrigeration load Heat Exchanger (HTX) 110 leading to evaporator 1 15 which in turn is connected with low pressure air (LP AIR) 120 and then turbine or expander 125 which connects with aforementioned alternator 101. Evaporator 115 additionally is connected through e- Thermal Expansion Valve (TXV) 195 with condenser/cooler 190.

[0043] Condenser/cooler 190 is depicted as outputting HP (High Pressure) Air Saturated 136 to dryer 135 which dries the air and outputs dry 130 air output to the turbine or expander 125.

[0044] From dryer 135, weak solution 141 is depicted via the low grade heat driver Rankine cycle driving a Brayton cycle with liquid desiccant loop for power generation and refrigeration as depicted at element 145 which is connected with QIN 150 and then to QOUT 155 and then to Liquid desiccant cooler 160 and then back through strong solution 140 to dryer 135.

[0045] Second output from QIN 150 is depicted as vapor 165 leading to condenser QOUT 170 and in turn Boiler Feed Water Pump (BFWP) 175 which then connects back to

condenser/cooler 190 and ejector 105 through the condenser/cooler 190 to boiler QIN 151 connection shown. Condenser/Cooler 190 is additionally connected with ejector 105 by way of

[0046] Notably, the input to the closed loop system is heat, including waste and low grade heat and the output, as depicted, is rotational, shaft power, or electrical power.

[0047] Figure 2 depicts the four sub-cycles 201 of a Sherbeck cycle in accordance with an embodiment of the invention. As can be observed, waste heat 227 is input into the system beginning the steam Rankine cycle 226 on the left of the diagram. From there, shaft work 228 is input into the reverse Brayton cycle 231 and latent heat 229 is input into the desiccant solution cycle 232. Cold dry air 237 proceeds from the reverse Brayton cycle 231 to the evaporative cooler 234 and dehumidified room air 238 proceeds from the desiccant solution cycle 232. At the evaporative cooler 234, warm humid air 233 is returned into the reverse Brayton cycle 231 and cold water less than 20° Celsius (element 239) is provided as a second input into the cooling coil

242 (e.g., operating at 20° Celsius). From the cooling coil 242, comfortable room temperature air

243 (e.g., air at approximately 24° Celsius with a relative humidity of 30%) is then released into the air conditioned space 246, such as an indoor swimming pool area, etc. Next, warm and humid air 244 is then returned to the desiccant solution cycle 232 and the process continues.

[0048] According to described embodiments, the Sherbeck cycle utilizes four sub-cycles including the steam Rankine cycle 226, a reverse Brayton cycle 231, an evaporative cooling cycle (e.g., via evaporative cooler 234) and a desiccant solution cycle 232. Figure 2 depicts the various relationships between the 4 sub-cycles in terms of their respective inter-relationships, inputs, outputs, and how they are utilized together to achieve cooling and dehumidification. According to described embodiments, the Sherbeck cycle utilizes exhaust heat or low/middle grade heat to achieve air conditioning as an input. Such heat sources are very often considered waste heat 227 and are commonly released into the environment rather than being utilized to produce commercially viable workloads, such as air conditioning and dehumidification.

[0049] Unlike traditional air conditioner technology, the Sherbeck cycle decouples sensible cooling and latent cooling thus leaving the cooling coil responsible for only sensible cooling loads while the desiccant solution cycle 232 is utilized to remove humidity (e.g., latent load cooling) from the air conditioned space 246.

[0050] Because the sensible cooling and latent cooling loads are decoupled, there is no longer a need for the cooling coil to reach below dew point in order to remove moisture by condensation. Consequently, the cooling coil may operate at higher temperatures within the Sherbeck cycle system when compared with conventional air conditioner systems and yet attain the same comfort levels (e.g., as measured by temperature and relative humidity).

[0051] The Sherbeck system therefore requires a lesser temperature difference between outdoor air and coolant in the cooling coil, which thus allows the reverse Brayton cycle 231 which is utilized to create the cooling effect to reach much higher efficiencies when compared with conventional systems.

[0052] Figure 3A depicts a waste heat driven steam Rankine cycle 302 in accordance with an embodiment of the invention. In particular, there is shown a T-s diagram depicting multiple state points R5 to R2 to R3 to R4 which correspond to the chart provided at the bottom of Figure 3A which depicts in greater detail for each of the various state points 361 , enthalpy 362, entropy 363, pressure 364, and temperature 366.

[0053] As depicted here, the heat source is coming from exhaust heat (e.g., waste heat). Steam is assumed to be heated up to 200°C. In the steam Rankine cycle, condensation temperature is assumed to be 100°C. Therefore, the temperatures of points R4 and point R5 are 100°C. In order to simplify the steam Rankine cycle, quality of point R4 and R5 are assumed to be 1 and 0. Water 349 at point R5 will be pumped to point R2 with 90% pump isentropic efficiency. Steam at point R3 will expand to point R4 with 85% turbine isentropic efficiency.

[0054] Additionally depicted is the equation 367 for the utilization of both shaft work and latent heat according to the depicted cycle showing that total thermal efficiency of this steam Rankine cycle is around 7.6%. Nearly 92% of input heat energy is rejected again as the heat of condensation from R4 to R5. In the Sherbeck cycle systems described herein, this rejected latent heat is used for desiccant solution regeneration.

[0055] The equations below show the steps needed to solve the cycle efficiencies and states of each point where equation 1 gives the thermal efficiency of the steam Rankine cycle, as

W

follows: ψαηΐύηβ = ^q≡ ^— , where W , is the shaft power output of the auxiliary turbine q l in + W pump ,

(kW), where W _P ump3 is the power input of pump #3 (kW), and where Qin is the heat provided to the boiler (kW).

[0056] Next, equation 2 defines the isentropic efficiency of the steam turbine, as follows:

h - h

Tfcmkine, turb =— ^ — , where h _r 3 and h _r 4 are the enthalpies of points R3 and point R4

(kJ/kg), and where h _r 4a is the enthalpy of point R4a, which is the turbine outlet state achieved by an equivalent isentropic expansion.

[0057] Next, equation 3 defines the isentropic efficiency of pump #3 similar to the steam h - h turbine efficiency, with equation 3 being as follows: tfankine, pump =— — .

[0058] Next, equation 4 gives the required pump work Wpump as a function of enthalpies at points R2, R5 and the mass flow rate (kg/s), as follows: W = [h _r2 - h _r5 ] * m _rankine .

[0059] Similarly, equations 5, 6, and 7 are utilized to compute the work output, heat input, and heat rej ected, beginning with equation 5 as follows: W _auxturb = (h _r3 - h _r4 ) * m _rcmkme .

Equation 6 is as follows: Q _in = [h _r3 - h _r2 ] * m _mnkme . And equation 7 states:

1 latent ^~ (^r4 ^~~ h _r s ] * ^m _r ankme ·

[0060] In traditional air conditioning, air must first be cooled to the point where it meets an outlet humidity requirement and then heated back up. Conversely, in desiccant cooling, air will be dehumidified by direct contact with the desiccant solution. The air will then reject some heat to ambient and some to the cooling coil. The cooling coil temperature in desiccant cooling is may therefore be higher than the cooling coil temperature in traditional air conditioning systems.

[0061] A steam Rankine cycle is not efficient at converting low grade heat to work because it excessively condenses out latent heat to ambient. Conversely, the Sherbeck cycle system exploits both the shaft work and the latent heat rejected by the steam Rankine cycle. The shaft work is utilized to drive a reverse Brayton cycle, which produces cold air for sensible cooling. The latent heat then regenerates a desiccant solution which provides dehumidification. Furthermore, the steam Rankine cycle as a work producing block is an environmentally friendly altemative to organic Rankine cycles for which fluid leakage is an environmental concern further because it poses less working hazard than a sorption system with pressurized ammonia.

[0062] On the right side of Figure 3A there is depicted a process diagram of the steam Rankine cycle for which there are depicted the air 347, water vapor 348, water 349, and desiccant solution 352 phases. At the top left there is provided an auxiliary turbine 356 which receives water vapor 348 from the heat source 354 and pushes the water vapor 348 and air 347 into the solution cycle 358 (Q-in). Water 349 is then output from the HTX #2 condenser (Q-Out) 357 and input into pump #2 at element 353 thus completing the cycle.

[0063] Figure 3B depicts a desiccant solution cycle 303 in accordance with an embodiment of the invention. Desiccant solution regeneration efficiency is assumed to be 85%, as defined by equation 9, which states: rflesiccant =— ^ ^latent ' ^∞olm s— _^ where Q _{latent cooling} is the

Q condenser, heatinp t

amount of latent cooling power provided by the desiccant solution cycle to the air streams being dried (kW) and where Q _{condenser heatinpu} the heat input from steam Rankine cycle condenser (kW).

[0064] For instance, there is depicted an exemplary cycle for indoor swimming pool air 369 originating from the air conditioned space 381 in which the SPA1 (swimming pool air) 372 is input into dryer #2 373 which outputs dry air SPA2 at element 374 and also outputs reject heat 376. The key at the lower left depicts the various phases elements present within the cycle including air 347, water vapor 348, water 349, the desiccant solution 352. The reject heat 376 proceeds to be cooled by a cold storage water cooling coil at element 368 which is then released into the air conditioning space 381. The output dry air SPA2 at element 374 outputs desiccant solution 352 which continues to heat exchanger HTX #2 at element 378 where it is then looped back as reject heat 384 and returned to dryer #1 at element 382. In the middle there is the circulating air 383 which removes condensed fresh water 377 which is output from the cycle. The equations 386, 387, 388, and 389 depicted on the right define the various details of the process.

[0065] Figure 3C depicts a reverse Brayton cycle 304 in accordance with an embodiment of the invention. Such a cycle includes both heat and mass recuperation to push external heat transfer to the extreme temperature ends of the cycle. Gate valve 355 connects outdoor air to help balance any pressure differences in the reverse Brayton cycle after shutting off the system.

[0066] According to one embodiment, the reverse Brayton cycle in Sherbeck cycle systems utilize moist air as the refrigerant. In such an embodiment, the moisture content changes throughout the cycle. The cold, low pressure air from the turbine outlet is passed through evaporative media to increase the cooling capacity. Excess moisture is then removed at the high pressure side to keep the turbine inlet dry. In addition to several stages of heat rejection to ambient including indirect evaporative cooling, the compressed air is cooled via recuperation with the low pressure side in HTX#1 as shown at element 320. As shown here, HTX# 1 (element 320) assists dryer 397 by condensing out some of the moisture. From the inlet at B7, dryer 397 dehumidifies the air stream to 3% relative humidity and heats the air stream to a temperature of 46°C at point B8 from which the air then rejects heat to environment as shown at element 396. Additionally depicted is element 399 showing the water at operations B 12 and B 13 going to the evaporative cooler as well as the air at operation B9 headed to the evaporative cooler.

[0067] So as to simplify the calculation of state points in the modeling of the reverse Brayton cycle, air is treated as ideal gas, beginning with equation 10 which gives the temperature and pressure relation for isentropic compression, as follows: , where

k = C _p / C _v is the specific heat ratio. Equation 1 1 shows the isentropic efficiency of compressor, hb3a where is the ideal enthalpy of air at the compressor exit in an isentropic process, as follows: r/eomp = -^ ² — , where hb3, is the real enthalpy of state point B3, and where hb2 is the enthalpy of state point B2, representing the compressor inlet.

[0068] Equation 12 states: T _bs = T _mv + AT _reject , where is given as a reasonable assumption of temperature difference between environment temperature and exit temperature depicted given that air must be rejected as heat to the environment.

[0069] Equation 13 states: W _{ldrct evapcooling} = [h _b5 - h _b6 ]* m _b6 , where indirect evaporative cooler effectiveness is assumed to be 0.9 and where Tcoidwater represents the temperature of the cold water given the equation 14: T _b7 = T _coldwater .

[0070] Equation 15 states: h _bg = h _b7 , where dehumidification through desiccant solution is assumed to be a constant enthalpy process and heat rejection to reach point B88 is modeled with a temperature difference of 5°C according to equation 16, as follows: T _bss = T _mvir + 5 .

[0071] Equation 17 provides the COP of the reverse Brayton cycle as follows:

171 * h — 171 * h

COP _revbray = ——— ——— , where the numerator is the enthalpy difference

between turbine outlet and compressor inlet and in which denominator is the sum of auxiliary turbine power input, and pump #2 (element 345) power input, and cooling power from HTX#1 (element 320).

[0072] Figure 3D depicts an evaporative cooler process 305 in accordance with an embodiment of the invention.

[0073] The evaporative cooling process 305 depicts how an evaporative cooler is modeled on the psychometric chart depicted at element 391. Depicted by the chart psychometric chart 391, point Bl lies outside the moist air states because it represents a concurrent flow of moist air and liquid water, which becomes fully evaporated by point B2.

[0074] The purpose of the evaporative cooler is to add cooling capacity since the air coming from reverse Brayton cycle is dry and can absorb water vapor. Water flowing in the direct evaporative cooler will be cooled by both sensible cooling (heat transfer) and latent cooling (mass transfer). In practice, during operation of such an evaporative cooler, sensible cooling and latent cooling will happen simultaneously. However, in order to simplify the model, the process is separated into several steps. For instance, after air at B10 is introduced into the evaporative cooler, sprayed water is cooled down by cold air and the air is humidified and cooled to point Bl l . Second, in order to make air become saturated at point B2, more water is injected after point Bl l . Third, cold and super-saturated air in HTX#1 (element 308) will cool the counter flowing stream within the reverse Brayton cycle. [0075] Equation 18 states: T _B2 = T _elr - 5 , where the temperature of point B2 (e.g.,

HTX#1 exit) is assumed to be 5 °C higher than the cold water temperature at Elr (element 321). Use of cold water temperature is a design choice is may be varied according to implementation requirements. Equation 19 states: R _BU = 0.9 , where relative humidity at point Bl 1 is assumed to be 0.9.

[0076] Optimization of steam Rankine cycle leverages the one free variable provided by the power block of the steam Rankine cycle, specifically, the low pressure level or condensation temperature. As condensation temperature is lowered, more shaft work will be delivered to the cooling cycle, but less latent heat will be provided for desiccant solution regeneration. Although reduction of condensation temperature down as far as environment temperature is possible, doing so may not yield practical results as the heat is needed to regenerate desiccant solution well above ambient temperature. Consequently, optimal performance suggests that condensation temperature should be no lower than 80°C or the efficiency of the desiccant regeneration process will be much lower.

[0077] Because condensation temperature plays an important role in influencing the steam Rankine cycle efficiency, it would be reasonable to set the temperature to 80°C or higher according to certain embodiments.

[0078] Optimization of the reverse Brayton cycle leverages two free variables that influence the COP of reverse Brayton cycle. One being the cold water temperature and the other being the pressure ratio, where decreasing each variable increases the COP. Higher cold water temperature in the evaporative cooler (E1R) results in a higher temperature at point B2 and less temperature difference between B2 and B88, where temperature at point B88 is 5°C higher than environment temperature after heat rejection.

[0079] As cold water temperature increases, the COP of the reverse Brayton cycle will increase as well, however, it will drop with increasing environment temperature. In the reverse Brayton cycle, the lower the pressure ratio, the higher COP as defined by equation 17 above. When pressure ratio reaches 101.3/70, both the reverse Brayton cycle and Sherbeck cycle will reach their maximum efficiency. If the pressure ratio exceeds this limit, the system will cease to operate properly as the reverse Brayton cycle does not generate cold air with temperatures below the cold water temperature. Consequently, such a Sherbeck cycle system imposes its own upper limit of COP.

[0080] When environment temperature is in low range (e.g., 24°C-33°C), then a pressure of 70kPa would be the theoretical optimal set point for B9 to maximize COP. As environment temperature increases, the optimal low pressure level decreases. If environment temperature exceeds 42°C, the ideal expansion pressure will be 40kPa. [0081] Exemplary optimal set points for a Sherbeck cycle system may therefore be 80°C condensation temperature and 18.5°C cold water temperature. The low pressure level at B9 may be set depending on ambient temperature such that (a) when environmental temperatures are in the range of 24°C-33°C then the pressure is set to be 70kPa, (b) when environmental temperatures are in the range of 34°C-37°C then the pressure is set at 60kPa, (c) when environmental temperatures are in the range of 37°C-42°C then pressure is set to 50kPa, and finally (d) when environmental temperatures are above 42°C then pressure is set to 40kPa. Other pressures and temperature ranges may be utilized according to specific design considerations.

[0082] According to one embodiment, the temperature of superheated steam at R3 is assumed as 200°C, however, depending on the heat source, this could be different. If superheated steam temperature increases, the COP of Sherbeck cycle system will increase as well.

[0083] Because the Sherbeck cycle system decouples the sensible cooling and latent cooling, it can operate at various ratios of latent load fraction. Also because the desiccant cooling circuit bypasses most of the losses generated in the system (by the power and cooling cycles), the system may operate more efficiently for dehumidifi cation than for cooling. However, COP will increase with increasing latent cooling load percentage and vice versa, where COP is defined to treat both latent and sensible cooling as equally important.

[0084] For the reverse Brayton cycle, because the indirect evaporative cooler utilizes outdoor air to achieve evaporative cooling, the relative humidity of the outdoor air will affect the COP of the whole system, such that the COP of the reverse Brayton cycle drops with increasing relative humidity of outdoor air and vice versa.

[0085] Consequently, implementation of Sherbeck cycle systems are especially useful where air dehumidification is much more important than air cooling. Moreover, because Sherbeck cycle systems utilize air and water as working fluids, the system is extremely environmentally friendly when compared to known conventional systems.

[0086] Figure 4 depicts a Sherbeck cycle 400 in accordance with an embodiment of the invention.

[0087] In the context of thermodynamics and energy conversion, the arrangement of the depicted Sherbeck cycle 400 provides for a closed loop system arrangement whereas prior solutions operate on an open loop basis.

[0088] For instance, another benefit and improvement over prior solutions is that the depicted Sherbeck cycle 400 produces refrigeration and produces shaft power via turbine 498 and alternator 499 as outputs. The depicted Sherbeck cycle 400 additionally reduces noise pollution while utilizing a closed loop methodology.

[0089] According to certain embodiments, waste heat from industrial processes are utilized as the heat input 401 into the depicted Sherbeck cycle 400 to produce refrigeration and shaft power.

[0090] It is common for manufacturing and industrial processes (e.g., food preparation, water desalinization, etc.) to generate large quantities of heat waste. Sometimes this heat waste is exchanged to the ambient air or exchanged with water, however, the depicted Sherbeck cycle 400 utilizes such waste heat from other processes as heat input 401 into the system.

[0091] The economics of reusing such waste heat take into account the cost of a system to convert the waste heat into useful outputs, such as the refrigeration and shaft power described with regard to the depicted Sherbeck cycle 400 and the value of the refrigeration and shaft power such that a pay-off or break even duration may be determined to offset the capital outlay to implement such a system.

[0092] Other non-monetary considerations may also be applicable, such as

environmental benefits, off-grid capability, or applicable statute, law, policy, or regulation mandating the use of such a system or mandating a reduction in expelling waste heat energy into the environment.

[0093] Increasing the conversion efficiency and lowering the cost to build, provision, and maintain such systems therefore improves these considerations and increases the likelihood that commercial, industrial, and residential operators will utilize such a system.

[0094] According to certain embodiments the depicted Sherbeck cycle 400 utilizes a turbocharger having a completely self contained bearing system capable of operating above the first critical, the second critical, and optionally the third critical.

[0095] In accordance with an optional embodiment, liquid desiccant of the air is located immediately preceding the turbine 498 entrance by which the air is cooled permitting greater quantities of refrigerant or water to be condensed and collected as a liquid rather than having to go through the desiccant loop.

[0096] In accordance with another embodiment, coefficient of performance (COP) is proportional to the quantity of heat input 401 provided into the system. Therefore, in accordance with a particular embodiment, heat required to regenerate the desiccant is taken from the condensers 425. In such a way, heat energy is utilized at least twice, for instance, once from the heat input 401 and again as waste heat energy from the condensers 425, thus making the system operate with greater conversion efficiency. Other heat re-generation sources internal to the system during operation may likewise be utilized but it is observed that a greatest quantity of such heat energy, other than the heat input 401, is provided from the condensing cooling action.

[0097] Further depicted is steam jet cooling via the spray evaporator 445 and steam ejector pump 435 which operate in conjunction with the chilled water 440 tank and boiler 430. [0098] Another potential source of heat input 401 is from industrial scale refrigeration which may be utilized as heat input 401 or alternatively as energy for supplying boiler feed water. Both are usable where there is ready access to a supply heat and water to produce steam.

[0099] Turboexpanders for both cooling and energy recovery may likewise be utilized in conjunction with the described system.

[00100] Previously available open-cycle and closed-cycle desiccant refrigeration systems conventionally depend upon a single working fluid and reject heat to the atmosphere through the use of evaporative coolers when the cycles are closed. Conversely, when the cycles are open on such conventional systems, they will reject heat by exhausting the working fluid which is typically water or moist air. Either method necessitates a large supply of treated fresh water which is wasteful and environmentally damaging.

[00101] The depicted Sherbeck cycle 400 negates the need to exhaust such waste heat or exhaust the working fluid while improving conversion efficiency.

[00102] According to certain embodiments, the waste heat from power generation is used to drive a gas cycle turboexpansion refrigerator while allowing waste heat to be rejected to the atmosphere in an air to air heat exchanger. In such an embodiment, the working fluid of the gas cycle (e.g., moist air 420) is kept dry by a desiccant air dryer 450 thus producing dry air 455 for the turbine 498 with the water being extracted out by the dryer, after being separated to separating tank 415 from the desiccant solution.

[00103] As the gas heads to the turbine 498 it goes through an air dryer 450 where a weak solution 460 of the sorbent is sprayed to dry the gas of refrigerant. This creates the strong solution 405 once cycled through vapor generator 402 where refrigerant vapor is released from the sorbent by heating via the heat input 401. The liquid desiccant returns to air dryer 450 by way of a liquid desiccant cooler 410.

[00104] The water extracted from the moist air is then utilized as working fluid for a boiler, driven by waste heat, which ultimately provides the pressure change needed to drive the gas cycle refrigeration. This eliminates the requirement for treated water other than that which would be used as the working fluid in the closed-cycle system.

[00105] In such a way, waste process heat is utilized to generate a cold source without requiring the traditional infrastructure of compression refrigeration; further still, there is no loss of working fluids, still further, durability and steady state operation suitable for a central plant cooling are provided, and the system may be adapted to reject heat directly to the atmosphere rather than having to cycle through a source of fresh treated water.

[00106] During operation and other than maintenance periods, once the system is filled and started, only a heat input 401 source is required to keep the heat pump operating. Energy to keep air moving across the radiators and to operate the pump is also required but is available from turboexpander with an acceptable load.

[00107] In accordance with an alternative embodiment of the described Sherbeck cycle 400 rather than outputting shaft power from the alternator 499, improved cooling/refrigeration is produced. For instance, improving cooling power without outputting shaft power increases COP (coefficient of performance) and simplifies construction. In such an embodiment, an air-to-air heat exchange as shown on the T-s diagram at Figure 9 is placed between air exiting the evaporative cooler and air before entering the liquid desiccant dryer and then the turbine. Further still, the turbine drives the compressor shown on the T-s diagram rather than driving an alternator.

[00108] In accordance with an alternative embodiment of the described Sherbeck cycle 400, the liquid desiccant air dryer is placed immediately before the entrance of the turbine which provides the advantage of condensing greater quantities of water out of the flow before entering the dryer.

[00109] In an alternative representation of the T-s diagram at Figure 9, a second temperature jog is added to the high pressure line at the bottom to show a temperature rise and pressure loss of the liquid desiccant dryer before the turbine entry. A small temperature drop corresponding to the water injection after the evaporative cooler is also present and heat transfer from the high pressure air stream extends to the coldest part of the low pressure air stream after the water injection.

[00110] In accordance with an alternative embodiment of the described Sherbeck cycle 400, the liquid desiccant loop is used to increase efficiency still further by adding a second semipermeable membrane air dryer downstream of the turbine air dryer in the conditioned space which produces temperatures which are in a comfortable range and which may be maintained without the need for the cooling loop to reach the dew point to extract out the latent cooling load.

[00111] Raising the cooling loop temperature will increase the COP (coefficient of performance) still further. Additionally, temperatures in a comfortable range may also be maintained with a chilled water supply as high as about 70-F. Having the condensate from the high pressure air flow condensation in the cooler/dryer/condenser/liquid desiccant regenerator, and air-to-air heat exchanger drain into a fill-to-spill tank further allows the moisture taken from the conditioned space to be collected as a liquid which is advantageous as the liquid water collected will be directly potable so long as corrosion within the cycle is eliminated.

Alternatively, the collected water may be utilized to assist the cold sink by evaporative cooling.

[00112] In accordance with an alternative embodiment of the described Sherbeck cycle 400, use of a float valve to control water flow from the fill-to-spill tank to the evaporative cooler, allows a constant mass of water within the cycle regardless of the latent load. In the passive limiting case, the conditioned space air dryer is an array of vertical plates on a wall. A delta T (e.g. a temperature differential) between the air temperature in the room and the liquid desiccant solution therefore drives an air flow over the dryer by buoyancy. Alternatively, such a float valve may be integrated into an air handler.

[00113] In yet another embodiment, passive cooling is achieved by use of a chilled ceiling, chilled beams, or valences without concern of condensation occurring. If the inside liquid desiccant dryer included heat transfer fluid passages, they will yield heat and humidity which is beneficial during the cold season. By providing year round comfort control the return on investment is improved thus shortening the payback period and in turn driving marketability.

[00114] In accordance with an alternative embodiment of the described Sherbeck cycle 400, the moisture content of air flow exiting the evaporative cooler is permitted. Generally, evaporative coolers and cooling towers are designed to minimize excess water in the exiting air flow as this water is wasteful, and can be damaging. However, some moisture exhaust may be desirable in certain use cases. Configurations may produce air at about 95% of saturation.

Because the air goes through an air-to-air heat exchange before entering the compressor, a metered flow of water from the cooling loop to air exiting the evaporative cooler provides additional evaporative cooling up to the compressor inlet or beyond if the compressor specifications permit.

[00115] In accordance with one embodiment, an alcohol/water mix is sprayed directly into the centrifugal compressor inlet of a turboprop engine to provide a power boost in hot weather. Such a practice may be suitable for short durations depending on the turboprop design and configuration utilized. In accordance with such an embodiment, cooling power of the air-to- air heat exchanger is improved and by lowering the compressor inlet temperature, efficiency is improved also. Because the evaporative cooler is a counter flow arrangement, water exits at a colder temperature than the air which will therefore extend the low temperature loop on the T-s diagram depicted at Figure 9.

[00116] Figure 5 is a flow diagram illustrating a method 500 for implementing a closed low grade heat driven Rankine cycle driving a reverse Brayton cycle using an ejector with a liquid desiccant loop to produce shaft power and refrigeration by adiabatic expansion and evaporation in accordance with described embodiments. Some of the blocks and/or operations listed below are optional in accordance with certain embodiments. The numbering of the blocks presented is for the sake of clarity and is not intended to prescribe an order of operations in which the various blocks must occur. Additionally, operations from flow 500 may be utilized in a variety of combinations. [00117] Method 500 converts input heat energy into output rotational shaft power as depicted at block 505.

[00118] At block 510 operation evaporates a refrigerant in a first gas.

[00119] At block 515 operation receives the input heat energy at a vapor generator having the first fluid therein.

[00120] At block 520 operation vaporizes the first fluid at the vapor generator to create high pressure vapor.

[00121] At block 525 operation ejects the high pressure vapor through an ej ector to create a low pressure by drawing the first gas through a turbine.

[00122] At block 530 operation drives an alternator through Adiabatic expansion of the first gas in the turbine to output the rotational shaft power from the alternator.

[00123] At block 535 operation condenses the first gas via condenser, wherein the first gas flows through a separator before cycling back into the turbine and further wherein the refrigerant flows returns to the boiler feedwater pump drawn by the low pressure through an expansion component and into an evaporator.

[00124] At block 540 operation cycles the first gas which is saturated with the refrigerant from the condenser through a dryer via a sorbent loop to dry the first gas.

[00125] Thus, it is in accordance with one embodiment that input heat energy is converted into output rotational shaft power, by way of a method or other means which include at least means for evaporating a refrigerant in a first gas; means for receiving the input heat energy at a vapor generator having the first fluid therein; means for vaporizing the first fluid at the vapor generator to create high pressure vapor; means for ejecting the high pressure vapor through an ejector to create a low pressure by drawing the first gas through a turbine; means for driving an alternator through Adiabatic expansion of the first gas in the turbine to output the rotational shaft power from the alternator; means for condensing the first gas via condenser, wherein the first gas flows through a separator before cycling back into the turbine and further wherein the refrigerant flows returns to the boiler feedwater pump drawn by the low pressure through an expansion component and into an evaporator; and means for cycling the first gas which is saturated with the refrigerant from the condenser through a dryer via a sorbent loop to dry the first gas.

[00126] In accordance with optional and alternative embodiments of method 500, the first gas is air.

[00127] In another embodiment of method 500, the refrigerant is water.

[00128] In another embodiment of method 500, the refrigerant is a refrigerant mixture.

[00129] In another embodiment of method 500, the refrigerant is an immiscible fluid. [00130] In another embodiment of method 500, the refrigerant is not water and produces sub-OC temperatures in an evaporator through one of water/ethanol or water/acetone refrigerant mixtures.

[00131] In another embodiment of method 500, evaporating a refrigerant in a first gas includes using the gas to drive a turbine in a reverse Brayton cycle to evaporate the refrigerant in the first gas.

[00132] In another embodiment of method 500, the vapor generator is a boiler.

[00133] In another embodiment of method 500, the high pressure exists between a boiler feedwater pump and the ejector nozzle where the fluid is vaporized in the boiler.

[00134] In another embodiment of method 500, the adiabatic expansion in the turbine cools the first gas.

[00135] In another embodiment of method 500, further cooling of the first gas is attained by re-evaporating the refrigerant in the gas which is the drawn by the ejector through a load heat exchanger.

[00136] In another embodiment of method 500, the expansion component comprises float valve.

[00137] In another embodiment of method 500, the expansion component comprises thermostatic expansion valve to hold a fixed superheat at the load heat exchanger exit.

[00138] In another embodiment of method 500, the expansion component comprises a thermoacoustic or a Hilsch tube expansion device to permit heat to be rejected from the cycle.

[00139] In another embodiment of method 500, cycling the first gas through a dryer via a sorbent loop comprises spraying a weak solution of sorbent to dry the first gas and remove the refrigerant from the first gas.

[00140] In another embodiment of method 500, a strong solution is produced from spraying the weak solution of sorbent to dry the first gas and remove the refrigerant from the first gas.

[00141] In another embodiment of method 500, the method further includes pumping the strong solution into a generator where refrigerant vapor is released from the sorbent by heating.

[00142] In another embodiment of method 500, the sorbent is a salt solution.

[00143] In another embodiment of method 500, the sorbent is one of LiBr or calcium chloride.

[00144] In another embodiment of method 500, the sorbent is an ionic fluid.

[00145] In another embodiment of method 500, the method further includes cycling the strong solution through a second condenser the strong solution is liquefied and returned to the separator. [00146] In another embodiment of method 500, the method further includes cooling the weak solution and pumping the weak solution back to the dryer.

[00147] Figure 6 depicts an ejector refrigeration cycle which is heat driven in accordance with an embodiment of the invention.

[00148] The Sherbeck cycle air loop involves three pressures. First, Atmospheric (high) pressure where steam from the Rankine loop, mixed with saturated air from the cooling loop in the ejector are condensed and cooled. The heat is delivered to the liquid desiccant regenerator, rejected to the cold sink, and a heat exchange with air returning from the evaporative cooler takes place. Much of the remaining water in the air is removed by the desiccant loop before entry into the turbine. Second, a medium pressure at the compressor discharge/ejector inlet. Third, a low pressure between the turbine exit, and compressor inlet, where water in the cooling loop is evaporated, and heat is exchanged with air going to the liquid desiccant dryer.

[00149] Specifically depicted at Figure 6 is state-space representation of ejector cycle with water as both motive fluid and refrigerant. Also indicated with dashed lines are the external flows, re-scaled to match on horizontal axis, thus showing the pinch points for heat transfer. In particular, the generator (transferring heat from HTF, 22- 23- 24- 25, to motive stream, 1 ->2- 3 - 4) has a pinch point where 24 heats point 2.

[00150] An ejector system model thus exhibits many results an example of which is depicted at Figure 6. In addition to the state point constraints and given values for the three coefficients, mass and energy balance equations are imposed for each of the components, and for each heat exchanging component, a heat transfer equation is used to couple the heat flow and m.KE.

temperatures of the internal and external streams where: rjentrainment

m ₅ KE ₅

[00151] This is distinct from the entrainment ratio of mass flow rates,

' which although a useful indicator of pump operation, is not an input in this model. The diffuser efficiency is also a so-called isentropic efficiency, where 7' is an artificial state that would result from isentropic expansion from state 6, where: r/diffuser

h, - h ₆

[00152] As a result, lower diffuser efficiency indicates a need for higher kinetic energy at state 6, etc.

[00153] Figure 7 depicts a Sherbeck cycle psychometric chart 700 in accordance with described embodiments. In particular, there are depicted points 1-4 at high pressure, point 8 at medium pressure, and points 5-7 at low pressure. Point 1 represented by element 710 is off the scale of the chart as depicted.

[00154] Figure 8 depicts another Sherbeck cycle 800 in accordance with described embodiments. As depicted, by taking latent load out of the air conditioned space 808 with the liquid desiccant dryer 838 via the SLDS phase from the turbine dryer 801 through the permeable membrane air handler 818, the cooling load temperature can be raised while maintaining comfort conditions. Using a fill to spill tank to feed the evaporative cooler 804 allows the collected condensed moisture from the air conditioned space to spill out of the cycle as depicted at element 828. In particular, condensate 806 captured from the re-gen 802 phase through the air to air HTX 803 exits the cycle as a liquid at atmospheric pressure. Raising cooling loop temperatures further increases the coefficient of performance (COP).

[00155] Figure 9 depicts a T-s or temperature-entropy diagram of the reverse Brayton cycle within the Sherbeck cycle in accordance with described embodiments. In particular, there is depicted the liquid desiccant regenerator 920 at point 1 heading to cooler 921 at point 2 and the liquid desiccant dryer 938 at point 3 which via QA to A HTX proceeds to point 5 and on to evaporative cooler 904 at point 5. Water injection 905 at point 6A results in 3Δ super saturated 911 air at point 6B. Additionally shown here are the compressor 919 taking point 7 at 21.5 KPA to the ej ector 922 of point 8 at 101 KPA.

[00156] With reference also to both Figure 1 and Figure 4 from above, there are three fluids, and three pressures are used to run the depicted cycle; the fluids being: 1) a gas used to drive a turbine 125 in the reverse Brayton portion, which may be air, or any other gas in which the refrigerant will evaporate; 2) a refrigerant, which may be water or may include refrigerant mixtures, and/or immiscible fluids; and 3) a desiccant, which may be as simple as a salt solution such as LiBr, calcium chloride, or an ionic fluid.

[00157] The high pressure exists between the Boiler Feed Water Pump (BFWP) 175, and the ejector nozzle where that fluid is vaporized in the boiler. The high pressure vapor is used in the ejector 105 to create the system low pressure (LP AIR 120) by drawing the gas through the turbine 125. Adiabatic expansion in the turbine 125 extracts work from the gas to drive an alternator 101 or other device with the resulting shaft power, and results in cooling of the gas.

[00158] Further cooling is produced by evaporating the refrigerant in the gas which is the drawn by the ejector 105 through the refrigeration load Heat Exchanger (HTX) 110. Choice of a refrigerant other than water may produce temperatures in the evaporator below freezing. For instance, water/ethanol, or water/acetone mixtures may be utilized to permit sub-OC temperatures in the evaporator 115, and add temperature glide to the condenser/cooler 190 to improve operational conditions.

[00159] The medium pressure is set by the condensing/gas cooling temperature. From the condenser/cooler 190, the flow enters a separator (dryer 135) from which the gas heads back to the turbine 125, and the refrigerant returns to the Boiler Feed Water Pump (BFWP) 175, and is drawn by the low pressure through an expansion device into the evaporator. The expansion device may be a float valve, a thermostatic expansion valve, which holds a fixed superheat at the load heat exchanger exit, a thermoacoustic, or Hilsch tube expansion device which permits another location for heat to be rejected from the cycle, etc.

[00160] Because the gas leaving the condenser/gas cooler is saturated (HP air sat. 136) with refrigerant, a sorbent loop is added to dry the gas. As the gas heads to the turbine it goes through a dryer 135 where a weak solution 141 of the sorbent is sprayed to dry the gas of refrigerant. This creates the strong solution 140 which is pumped into a generator where refrigerant vapor 165 is released from the sorbent by heating, and enters a second condenser 170 where it is liquefied and returned to the separator. The hot strong solution 140 is cooled and pumped back to the dryer 135. Commercially available heat exchangers may be utilized in the depicted cycle. For instance, the cold refrigerant flow exiting the load Heat Exchanger (HTX) 110 (Refrigeration Load HTX 110), could be used to pre-chill the liquid refrigerant before it enters the evaporator, or the dry air before it enters the turbine 125 to improve the cycles' efficiency.

[00161] Thus, as depicted, the cycle uses a gas, a refrigerant, and a desiccant to produce the shaft power as well as refrigeration from low grade heat that might be waste heat, or solar heat or other available sources.

[00162] Figure 10 depicts yet another Sherbeck cycle 1000 in accordance with described embodiments. As depicted, the air conditioned space 1055 shown on the lower left is where a second dryer 1050 is added to the desiccant loop to take the latent load from the air conditioned space and in particular, condensed moisture from the air conditioned space' spills out of the cycle as depicted at H2O spill element 1035. The mass of water in the cycle is held constant despite the humidity being removed from the space by adding a fill to spill 1034 provision to the air/water separator at the lower right side.

[00163] Air to air H2O HTX 1005 and solution to solution heat exchangers (SDL to SDL HTX 1080) are introduced in this variant to create additional temperature differences from the left side of the diagram where refrigeration is produced, and the right side where the heat source 1010 is used to power the cycle.

[00164] The optional use of a steam turbine 1094 is further depicted in this embodiment at the top right. Where the product of the steam turbine and the compressor efficiencies exceed those of the ejector, it may be preferential and economical to implement such a steam turbine. The pressure ratio of the steam loop in the initially targeted is high enough that a multi-stage turbine may be desirable. Use of an impulse turbine permits such a pressure ratio to be accomplished in a single rotor by having the flow reverse and go through the turbine blades two or more iterations in opposing directions. Such an approach additionally negates rotor thrust issues.

[00165] At one-third from the bottom right, the cooler/condenser 1040 is depicted which is utilized to heat the desiccant loop (re)generator, where moisture 1025 is driven from the weak solution to create the strong solution 1030. This is possible because the solution regenerates at temperatures as low as 60C, while at ambient pressure, the steam in the boiler ejector 1045 discharge will condense at l OOC. The boiler ejector 1045 discharge may enter the

cooler/condenser/liquid desiccant regenerator at a higher temperature which may lead to crystallization in the regenerator, but which is addressed via a shut down delay to the solution pump such that any crystallization is dissolved after every operation. The use of the

cooler/condenser to drive the regenerator eliminates a separate use of the heat from the heat source and improves the cycle coefficient of performance (COP).

[00166] In another embodiment, the rejection of heat to ambient in and/or between the two dryers (1050 and 1065) in the desiccant loop may be preferred depending on the particular implementation and use case.

[00167] In another embodiment, the addition of water injection into air flow as depicted by element 1005 leaving the evaporative cooler 1070 before entering the air to air heat exchanger is utilized which allows evaporative cooling to continue to the compressor inlet or beyond if the compressor tolerates liquid in the inlet flow. Such an approach increases the cooling power of the air to air heat exchanger, and the COP.

[00168] Figures 11A and 11B depict a complete Sherbeck cycle system 1100 in accordance with one embodiment of the invention. In particular, an evaporative cooler is depicted at the upper left having a cooling load 1 150 in which incoming air at point B10 1 125 is introduced into the evaporative cooler 1 101 , sprayed water is cooled down by cold air and air is humidified and cooled to point Bi l l 130. Point Bl 1 105 lies outside the moist air states because it represents a concurrent flow of moist air and liquid water, which becomes fully evaporated by point B2 1 180 via HTX # 1 at element 11 10. To make the air saturated at point B2 1180, more water is injected after point B l 1 1 130 via injector 11 15 at point E2r in which the water is pushed from evaporative cooler 1101 from point E2 1145 through pump #1 1 120 into the injector 11 15 as shown. Water is additionally passed from point E2 1 145 to point El 1140 at the top of the evaporative cooler 1101 where the water is released as sprayed water and cooled down to point Bl 1 1130 as noted above.

[00169] The low pressure level at point B9 1 175 is set depending on ambient temperature for the system 1100. Higher cold water temperature in the evaporative cooler 1101 at point Elr 1135 results in a higher temperature at point B2 1 180 and less temperature difference between point B2 1180 and point B88 1196.

[00170] From equation 11 above, hb3, is the real enthalpy of state point B3 1195 and hb2 is the enthalpy of state point B2 1180, representing the compressor inlet 1190.

[00171] At point B5 1106 it may be observed that heat is rejected to the environment as depicted by element 1111 as the air proceeds to the water separator #1 at element 1113.

Additionally shown is the gate valve 1114 permitting outdoor air 1116 into the system 1100 to balance any pressure differences in the reverse Brayton cycle after shut off the system 1100.

[00172] Also depicted here are point S9 1117 where water from the room and dryer #1 passes through water separator #1 1113 to point A2 1123 where pump #2 1118 (e.g., such as a circulating fan) pushes outdoor air at point AO 1119 through an extra air to air indirect evaporative cooling phase at point A3 1127 and releasing the air at point A4 1121 into the environment 1170. Water from room and dryer #1 at element 1117 is additionally passed through point Al 1122 back to the evaporative cooler 1101 through point B12 1160 where the water is cooled before entering or returning to the evaporative cooler 1101 as depicted by element 1155.

[00173] As shown, air also moves from the extra air to air indirect evaporative cooling phase at point A3 1127 through point B6 1126 into heat exchanger HTX #1 1124 where it is then passed to water separator #2 1112. Water is additionally passed from water separator #2 1112 back to the evaporative cooler 1101 through point B13 1165.

[00174] Circle connectors (a) and (b) and (d) and (e) on the right side of Figure 11A connect with the corresponding circle connectors (a) and (b) and (d) and (e) on the left side of Figure 11B to form one continuous system 1100 which is split over the two images for the sake of clarity.

[00175] Nevertheless, turning now to Figure 11B, it may be observed that circle connector (a) leads to the auxiliary turbine 1172 where water vapor originating from heat source 1151 is pulled through point R3 1164 through the auxiliary turbine 1172 and pushed through point R4 1171 to heat exchanger HTX #2 1156 on the condenser side condensing the water vapor to water and pushing the water through point R5 1154 through pump #3 1153 and looped back into the heat source 1151 through point R2 1152. According to such an embodiment, the heat source 1151 is waste heat, low grade heat, or exhaust heat. Steam is assumed to be heated up to 200°C. In the steam Rankine cycle, condensation temperature at HTX #2 1156 is assumed to be 100°C and temperatures of water vapor at point R4 1171 and therefore the water condensed at point R5 1154 is likewise assumed to be 100°C according to the depicted embodiment.

[00176] Water at point R5 1154 will be pumped via pump #3 1153 through point R2 1152 with 90% pump isentropic efficiency. Steam water vapor returning to the auxiliary turbine 1172 through point R3 1164 will expand to point R4 with 85% turbine isentropic efficiency. [00177] From circle connector (d) air proceeds into dryer #1 1167 via the inlet at B7 1166 where dryer #1 1167 dehumidifies and heats the air stream from which the air is then rejected to the environment at element 1159 through point B8 1162.

[00178] Further depicted is return air SPA 1 (e.g., swimming pool air) 1177 from air conditioning space 1191 being returned to dryer #2 11169 which receives desiccant solution S I 1163 from dryer #1 1167 returning dehumidified air from dryer #2 1169 as SPA2 1173 through heat exchanger HTX #3 1174 through the water cooling coil 1178 which cools the dehumidified air SPA2 by cold storage water in the water cooling coil 1178 before the air is then returned into the air conditioning space at element 1191.

[00179] Desiccant solution from dryer #2 1169 is pushed through point S2 1168 to HTX

#2 1156 on the condenser side where it is then looped through circulating air 1157 through heat exchanger HTX #4 1158 before being returned to dryer #1 1167 via point S4 1161. Condensed fresh water is additionally returned from heat exchanger HTX #4 1158 through point S9 1176 to circle connector (e) at the left side of Figure 11B to circle connector (e) at the right side of Figure 11A where the condensed fresh water then proceeds to the extra air to air indirect evaporative cooling phase at point A3 1127 of Figure 11A.

[00180] Additionally of note are the four circled heat rejection points within the complete Sherbeck cycle system 1100 including heat rejection point 1111 at the circled heat rejection point 1 on Figure 11A and each of heat rejection points 1159, 1158, and 1174 at circled heat rejection points 2, 3, and 4 respectively, each shown on Figure 11B.

[00181] It is therefore in accordance with one embodiment that there is a system to dehumidify air and cool air using an input heat source, in which such a system includes at least: an auxiliary turbine to receive input steam from a heat source, in which the input steam is to expand within the auxiliary turbine to output shaft power and release low pressure air cooled to a first temperature from an outlet of the auxiliary turbine; a heat exchanger to pass the low pressure air from the outlet of the auxiliary turbine through an evaporative media with moist air in an air to air indirect evaporative cooling process to further cool the low pressure air to a second temperature; a dryer to dehumidify the low pressure air via a desiccant solution in the dryer; an evaporative cooler to further cool the low pressure air to a third temperature via absorption of water vapor within the evaporative cooler, in which the cool low pressure air is to be sprayed with water within the evaporative cooler to humidify the low pressure air and to cool the low pressure air to the third temperature; and a cool air output to release the cool air from the system.

[00182] In accordance with another embodiment of the system, the auxiliary turbine to receive the steam input from the heat source is to implement a steam Rankine cycle for the system.

[00183] In accordance with another embodiment of the system, the heat exchanger to pass the low pressure air from the outlet of the auxiliary turbine through the evaporative media is to implement a reverse Brayton cycle for the system.

[00184] In accordance with another embodiment, the system further includes: a gate valve to connect outdoor air with the auxiliary turbine to balance pressure differences between the reverse Brayton cycle implemented by the auxiliary turbine and the outdoor air after shutting off the system.

[00185] In accordance with another embodiment of the system, the auxiliary turbine is to further output shaft power; in which the shaft power is an input to the reverse Brayton cycle for the system; and in which the shaft power is to drive the low pressure air from through the evaporative media of the heat exchanger to produce the cool air as an output from the system.

[00186] In accordance with another embodiment of the system, the dryer to dehumidify the low pressure air via a desiccant solution is to implement desiccant solution cycle for the system.

[00187] In accordance with another embodiment of the system, the auxiliary turbine is to further reject latent heat; and in which the rejected latent heat from the auxiliary turbine is used to recharge the desiccant solution.

[00188] In accordance with another embodiment of the system, the evaporative cooler is to implement an evaporative cooling cycle for the system.

[00189] In accordance with another embodiment, the system further includes: a cooling coil, in which the cool air released from the system via the cool air output passes through the cooling coil prior to system egress; and in which the cooling coil operates at a temperature above dew point.

[00190] In accordance with another embodiment of the system, the cooling coil reduces a temperature of the cool air to produce the cool air from the system and further in which the desiccant solution is to dehumidify the air; and in which the dehumidified and cooled air is released into an air conditioned space via the cool air output of the system.

[00191] In accordance with another particular embodiment, there is a Sherbeck cycle system which utilizes ocean water as a heat sink to extract heat from the system or into which heat may be rejected by the Sherbeck cycle system. Alternatively, river or lake water or another body of water may be utilized. For instance, consider an exemplary multi-megawatt (MW) steam plant adjacent to the ocean such that the ocean water is available for use as a heat sink to the Sherbeck cycle system. For instance, a highest surface temperature for such ocean water may be 83 °F and even cooler several meters deeper. Utilizing such a low heat sink temperature, the Sherbeck cycle system may be utilized to produce both desalinized water and refrigeration from low grade or waste heat. For instance, such a variation of the Sherbeck cycle system utilizes a Rankine cycle to drive the reverse-Brayton cycle, while eliminating the desiccant loop and evaporative cooling components. In such an alternative embodiment, the condenser heat rejection would instead heat sea water under a vacuum to produce desalinized water as an output from the Sherbeck cycle system. For instance, desalinization of ocean water occurs in an elevated position such that a column of brine returning to the ocean (as a bi-product of the desalinization process) produces the necessary vacuum and thus reduces the requisite work by the pump needed to bring the source sea water to a desalinization chamber. According to such an embodiment, water may be vaporized in the vacuum and condense on a plate cooled by sea water to produce the desalinized fresh water. The brine then flows through a counter-flow heat exchanger to preheat the source sea water. Such a configuration makes both the reverse-Brayton cycle and the Rankine cycle closed such that the working fluid choices may be selected from among a very wide array of choices. For instance, super-critical carbon dioxide or helium may be utilized in the reverse Brayton cycle.

[00192] While the subject matter disclosed herein has been described by way of example and in terms of the specific embodiments, it is to be understood that the claimed embodiments are not limited to the explicitly enumerated embodiments disclosed. To the contrary, the disclosure is intended to cover various modifications and similar arrangements as would be apparent to those skilled in the art. Therefore, the scope of the appended claims should be accorded the broadest interpretation so as to encompass all such modifications and similar arrangements. It is to be understood that the above description is intended to be illustrative, and not restrictive. Many other embodiments will be apparent to those of skill in the art upon reading and understanding the above description. The scope of the disclosed subject matter is therefore to be determined in reference to the appended claims, along with the full scope of equivalents to which such claims are entitled.