Patent Grant
10612800
Thermodynamic systems and methods for selectively heating and/or cooling a target space, and more particularly such a thermodynamic system in which ambient air comprises the working fluid.
Heating, Ventilating, Air Conditioning and Refrigeration (HVACR) is the technology of low temperature preservation and environmental comfort within a sheltered area. Simply stated, the goal of HVACR is to provide thermal comfort within a controlled space, such as within a refrigerator/freezer, a residential structure, a hotel room, banquet and entertainment facilities, in industrial and office buildings, on board marine vessels, within land vehicles, and in air/space ships to name but a few.
A conventional HVAC system is depicted schematically on the right-hand side of
Every viable refrigeration system must have a heat source target space and a heat sink target space. The refrigeration task is to move heat from the target space of the heat source to the target space of the heat sink. The term “target space” refers broadly to any space that is served by a refrigerant, for heating, ventilating and/or air conditioning. Thus, broadly, the term “target space” includes both of the inside and outside ambient air environments which are served and/or used by the refrigerant.
Stepping through the vapor compression cycle depicted in
In order to measure this work and its results, various industry associations and standards bodies around the world define Rating Points. Rating Point protocols standardize the measurement of refrigerants including parameters for the mechanical systems within which they circulate. Outdoor temperatures range from 27° C.-55° C. while indoor temperatures range from 20° C.-27° C. Only the currently mandated replacement refrigerant, R410A, will be discussed here.
Heat can be perceived as always flowing downhill, that is from a higher temperature to a lower temperature. The amount of excess refrigerant lift needed is determined by the needed approach air temperature differential on both sides of the refrigeration task. Because this Approaching Temperature is more specifically the difference between the temperature of approaching air and the refrigerant temperature it will be identified in the following as the approaching Air to Refrigerant Temperature Differential or A-RTD. Refrigerant alone creates the needed temperature differential because the approaching ambient air temperature does not change until it comes in contact with the different temperature of the refrigerant, through the heat exchanger. Refrigerant alone creates the needed temperature differential by moving evaporator and condenser temperatures outward beyond the refrigeration task (THIGH−TLOW). Tevap is necessarily always lower than TLOW. Tcond is necessarily always higher than THIGH. The size of this approaching A-RTD controls the rate of heat transfer with the heat exchanger to and from environmental air. The excess refrigerant lift is set to transfer heat into the air flows of the target environment at speeds near the system capacity, so the air vs. refrigerant temperature differential is optimally about 20° C. for present technology. The total A-RTD on both sides then presents a total excess refrigerant lift of 40° C. beyond the refrigeration task at whatever temperatures THIGH and TLOW happen to occupy at the time.
In practice, room temperature is usually determined by the preference of the room's occupants. The occupants express their choice for personal comfort by setting the thermostat, TLOW as shown in
It may be helpful at this stage to define the terms “sensible heat” and “latent heat.” When changes in heat content cause changes in temperature, the heat is called sensible heat. When the addition or removal of heat does not change the measured temperature but instead contributes to a change of state, the change in heat content is called latent heat. A pound of liquid water changes temperature from 32° F. (its freezing point) to 212° F. (its boiling point) with the addition of a mere 180 Btu/lbm of (sensible) heat. No surprise, since the British thermal unit is actually defined by the amount of heat required to change the temperature of one pound of water by one degree Fahrenheit. Moving that same pound of water at 212° F. from the liquid state to the vapor state still at the same temperature of 212° F. however, requires an additional 970 Btu/lbm of (latent) heat. Only after 100% of the liquid water molecules have been vaporized will the temperature of the water vapor then begin to rise above 212° F. In other words, in transition to the vapor state, each molecule of water will store 5.39 times more heat than is needed to move that same molecule from 32° F. to 212° F., from freezing to boiling, and it stores all this latent heat without changing temperature.
In the USA, the internationally recognized standard room temperature of 23° C. would be stated in Fahrenheit as 73.4° F. But the internationally recognized standard room temperature is not recognized as room temperature in the USA. Commercial interests in the USA have re-defined room temperature to circumvent regulations at the expense of human comfort. The American Society of Heating, Refrigerating, and Air-Conditioning Engineers (ASHRAE) raised the “industry accepted” definition of Room Temperature to 80° F. as the industry response to (regulated) consumer demand for increased efficiency. By turning thermostats up 7° F., ASHRAE could report a sensible heat capacity improvement while leaving everything in the mechanical performance of the equipment they sold entirely unchanged. This sleight of hand allowed the HVAC industry to raise Tevap, without cutting the Approaching Temperature. The industry's claim of energy improvement was delivered in appearance only and not in fact. The same inside Approaching Temperature differential of 20° C. was maintained by turning up the heat on people, human occupants, in order to reduce the excess refrigerant lift. Instead of cooling the occupants as before, they warmed things up to cut the energy needed to cool the evaporator as well. The industry gets to look good no matter how much the occupants feel bad. Of course the occupants can still turn their thermostats down where they want them. That does not translate into any adverse consequences for the industry.
This change in the Room Temperature standard created a significant new problem where individuals choose to comply with the industry's energy stipulation of the higher thermostat setting now at 80° F. Raising the evaporator temperature also cuts the amount of humidity removed. In other words, the higher Tevap increases relative humidity in the inside target space, i.e., the controlled space occupied by people. Stated as the Sensible Heat Ratio, the fraction of total cooling capacity delivered as sensible heat was thereby increased without cost or technical advancement. Raising Tevap directly cut the amount of condensation. Smaller amounts of total cooling capacity literally ran down the drain as cold water. But higher levels of temperature and humidity have supported epidemic increases in mold, fungus, and dust mites, sick building syndrome, and even Legionnaire's Disease. Yet ASHRAE continues to advertise and rate systems based on sensible heat capacity alone.
ASHRAE also stipulates that the energy expended in moving the inside air mass is not to be included in reports of system performance. Regardless of the fact that inside mass air flow must be reported and maintained, ASHRAE Standard 27-2009 stipulates that the energy needed to move this mass flow of air is not to be recorded. Refusal to account for the cost of this inside air movement data is claimed to be justified by the wide range of home ducting air resistance. Omitting the energy cost of moving the entire mass flow of inside air makes it possible to substantially overstate the performance of all units on sale in the USA.
As shown in
The inside operating costs, which include the resistance to moving air through the unpredictable routing of building ducts, is difficult to assess with any degree of confidence. In contrast, the outside or “air side” operating cost can be more consistently estimated. Because the outside fan is more nearly comparable to blowing air through a hole in the wall after it draws the air through a fin-and-tube heat exchanger whose design is integral to the unit being rated, the cost of moving a chosen mass flow of air through the fins of the outside heat exchanger is normally included when measuring the rated performance of a residential split system at the 95° F. Rating Point. Total efficiency may be increased up to a maximum by increasing the mass flow of air, when refrigerant side mass flow is held constant.
Increasing the Approaching A-RTD, will also increase the rate of heat transfer. In the best of all possible worlds, nature provides the desired cooler outside temperatures. In any real world where air conditioning is needed both the inside and the outside ambient temperatures are given by conditions outside the control of the refrigeration engineer. The only means of increasing the Approaching A-RTD is to change the refrigerant temperature, increasing the excess refrigerant lift. The losses of increasing excess refrigerant lift (pressure ratio) always overwhelm the gains, but it is a necessary evil up to a point. The two mass air flow rates, the two Approaching Temperatures, and the pressure ratios are inter-dependent and the incremental benefits related to each are not linear.
In order to optimize the design of air-side operating efficiency, it would be necessary to manage the trade-offs among three separate subsystems: heat exchanger, refrigerant compressor, and external air blower. Observe that all three subsystems (heat exchanger, refrigerant compressor, and external air blower) are mirrored by similar components which exist in both the inside target setting and in the outside target setting as well. Optimization would further necessitate the inclusion of a real time controller to adapt as conditions change. Compressor and blower efficiencies appear to have plateaued in recent decades. The size of the heat exchanger is sometimes increased to reduce operating costs. This raises the purchase price and justifies the report of increased operating efficiency, but adding fins and tubes does not improve the underlying technology. As was the case with ASHRAE's surreptitious re-setting of the room temperature datum to a higher value, the industry claims to have increased efficiency in spite of the fact that the technology and its performance remain unimproved.
The preceding description thus reviews the basic tenants of vapor compression technology accompanied by the mandatory approach air temperature differentials required to sustain heat transfers on both sides of a closed loop system like that depicted in
All such divergent refrigeration systems lift the temperature of the refrigerant from the lowest refrigerant temperature (defined to be below TLOW) by an amount equal to the chosen Approaching A-RTD. In vapor compression systems, this temperature differential is created by setting the temperature of the refrigerant in the evaporator, Tevap, below TLOW by an amount equal to the engineered Approaching A-RTD. The refrigerant must then be lifted to the highest refrigerant temperature, Tcond, correspondingly above THIGH by an amount also equal to the Approaching A-RTD. In vapor compression systems Tcond is the temperature of the refrigerant boiling point in the condenser. For residential and commercial air conditioning, ASHRAE standards set the Approaching Air-Refrigerant Temperature Differentials near 20° C. beyond both sides of the working temperatures. The working temperatures themselves are commonly separated by less than 20° C. in most climates so the total refrigerant lift exceeds three times (3×) the difference between the working temperatures. Thermodynamically, the consequences are far more severe as mathematically demonstrated below. (NOTE: Evaporator and Condenser temperatures must be translated from Celsius into the absolute temperature Kelvin scale, where Kelvin=Celsius+273.)
The limiting value of the Coefficient of Performance (COP) is defined thermodynamically by the following equation:
COP=TLOW/(THIGH−TLOW)
Using the numbers previously established by ASHRAE (and certified by NIST) for the 95° F. Rating Point, the best possible COP attainable between the two working temperatures can be calculated as:
COP=296/(308−296)=24.6
But after accounting for the stated excess refrigerant lift, where the condenser temperature is 55° C. and the evaporator temperature is 3° C. (
COP=276/(328−276)=5.3
In the late 1990s, the EU threatened a complete ban on CFC/HCFC refrigerants. About the same time, Normalair Garrett Limited of Yeovil, Somerset, England, now a wholly owned subsidiary of Honeywell International Inc., launched a commercial closed loop air cycle refrigeration system demonstrating life cycle costs competitive with vapor compression. Still in use on some German bullet trains, this closed loop air cycle system has not enjoyed further commercial adoptions. Because the turbine pumping losses characteristic of all “reverse Brayton Cycle” refrigeration systems are substantially higher than the vane and piston pump losses used in vapor compression, air cycle operating costs are typically considered unacceptably high among those of skill in the HVAC community. The academic community uniformly describes the pumping losses in such systems as excessive.
In contrast, the open air cycle systems have some attractive attributes. Of course, harmful refrigerants are avoided when ambient air is used as the refrigerant. An open air cycle offers the possibility for eliminating excess refrigerant lift on one side of the cycle. By using ambient air as the refrigerant, the open air cycle is already in possession of all the heat at its ambient working temperature so it requires no excess refrigerant lift at the working temperature where it originates. Half of the excess refrigerant lift with its attendant penalty is thereby avoided. The air temperature must nonetheless be lifted beyond the opposite working temperature by the needed excess refrigerant lift. To accomplish this, open loop air cycle systems nonetheless routinely require pressure ratios of about 2.5 or above, in spite of the fact that they inherently cut the excess refrigerant lift in half.
Despite the favorable attributes of the open loop air cycle, the routinely high pressure ratios (about 2.5 or above) necessarily incur unacceptably high operating losses. All devices heretofore proposed for open air cycle applications have been characterized by these prohibitively high pumping losses. A variety of alternative mechanisms have been proposed for open loop systems. But just like the turbines used in the closed cycle system of Normalair Garrett, the same problems with pumping losses have kept all proposed mechanisms from approaching commercial viability. All devices heretofore proposed for open air cycle refrigeration, as expected, fall within the category of divergent refrigeration as defined above so they necessarily all pay the same penalties for excess refrigerant lift. For example, U.S. Pat. No. 5,732,560 to Thuresson, granted Mar. 31, 1998, proposes to overcome friction with a rotary screw machine apparently made to function at pressure ratios near 2.5. In another example, U.S. Pat. No. 4,429,661 to McClure, granted Feb. 7, 1984, proposes a divergent refrigeration system that rejects heat into elevated temperatures using a single compressor. U.S. Pat. No. 6,381,973 to Bhatti, granted May 7, 2002, forthrightly relies on the production of what the Bhatti patent calls “very cold air” by turbines. Because Bhatti's ambient air is heated to a temperature well above the automobile engine compartment, as is needed to reject heat there, the exit temperature is substantially below freezing. The divergent refrigeration pressure ratio here is necessarily at or above 3.
U.S. Pat. No. 3,686,893 to Edwards, granted Aug. 29, 1972, describes yet another divergent refrigeration system based on an open air cycle. Edwards' pressure ratios correspondingly range from 2.5 to 4 and higher. Importantly, Edwards has published engineering results corresponding to his patented system (Analysis of Mechanical Friction in Rotary Vane Machines, Purdue e-Pubs, 1972). This publication acknowledged a measured COP of 0.45 with what Edwards calls a “volume ratio” of 2.5. Research indicates that after decades of development, the inventor of the aforementioned U.S. Pat. No. 3,686,893 (Edwards) shifted attention from the automotive open air cycle system (pressure ratio 2.5), toward more promising use in compressing standard refrigerants (e.g., R114) at pressure ratios near 4 and above. (The Controlled Rotary Vane Gas-Handling Machine, Purdue ePubs, 1988.) Edwards succeeded in reducing pumping losses for his device only at these higher pressure ratios. Subsequently, the published literature suggests that Edwards abandoned the open loop air cycle altogether in favor of conventional closed loop vapor compression split residential systems, a strong indicator that the open air cycle concepts embodied in U.S. Pat. No. 3,686,893 could not be successfully commercialized.
Another example is US2013/0294890 by Cepeda-Rizo, published Nov. 7, 2013. (The Applicant does not admit that Cepeda-Rizo is prior art to subject matter disclosed herein which rightfully claims the benefit of an earlier filing date.) The Cepeda-Rizo reference offers a fundamentally fresh approach to overcoming the well-defined set of deficiencies associated with open air cycle divergent refrigeration systems. Previous open air cycle divergent refrigeration systems proposed either high speed turbines characterized by leakage at low pressure ratios or multiple-vane pumps characterized by high friction loads. Cepeda-Rizo offers an adaptation of the legendary Tesla Turbine (concept, never successfully reduced to practice) asserting that its operating problems can be overcome at the pressure ratio of 2.5. If ultimately successful in overcoming the additional new challenges that Cepeda-Rizo will demand from the Tesla Turbine, Cepeda-Rizo acknowledges the best case theoretical COP of 1.5 and only an abysmal 0.4 COP overall.
The COP also provides a theoretical best case standard for comparison to actual equipment. COP, which is dimensionless, may be computed as the quotient of a relative temperature difference or as heat moved divided by work performed, heat and work being interchangeable in this context. In addition to test conditions already defined at the 95° F. Rating Point, the Energy Efficiency Ratio (EER) adds a standard for coping with differences in relative humidity. That being said, the EER is always proportional to the COP. Expressed mathematically, EER=COP*3.41. The Seasonal Energy Efficiency Ratio (SEER) applies a profile of temperature and humidity to match a range of climatological expectations. Nonetheless, it all comes back to COP which can thus be used to baseline comparisons between present known technology and proposed new solutions.
The National Institute of Standards and Technology (NIST) published a comparison of performance for refrigerants R410A and R22 across a range of temperatures. Compared to the best theoretical performance for lifting the refrigerant from 3° C. in the evaporator to 55° C. in the condenser, best case COP=5.3, NIST observed COPs as low as 3.93 (“Properties and Cycle Performance of Refrigerant Blends Operating Near and Above the Refrigerant Critical Point”, Task 2: Air Conditioner System Study Final Report by Piotr A. Domanski and W. Vance Payne, published September 2002 by National Institute of Standards and Technology Building and Fire Research Laboratory, APPENDIX B. SUMMARY OF TEST RESULTS FOR R410A SYSTEM.), dropping to 1.06 at an outside temperature of 68° C. This is the consequence of the compressor having to work harder to increase condenser pressure, hence system pressure ratios, as required to maintain the needed excess refrigerant lift for temperatures at or near the critical point of R410A or whatever refrigerant is being used. At temperatures above the critical point, a refrigerant will no longer condense. Maintaining the same Approaching Air to Refrigerant Temperature Differential as outside temperatures rise is crucial because the presumed benefits of latent heat progressively disappear as temperatures approach the R410A refrigerant critical temperature.
The contribution of latent heat disappears altogether above the critical point. For R410A the critical point is 161.83° F. or 72.13° C. Above this point the vapor will not condense. A benchmark of latent heat contribution at the 95° F. Rating Point provides an informative reference. Enthalpy numbers for the Pressure vs. Enthalpy graph of
The Pressure vs. Enthalpy graph of
The compressor discharge temperature shown in
The descending dashed line in
These measures fail to include the cost of moving the entire heat load into and out from the target environments with fans. Fans (or blowers) deliver the entire mass flow of air needed to move this heat twice, once on either side of the refrigerant loop. The energy cost of operating fans and blowers to provide the mass flow of air required on both the heat source (supplying) and heat sink (supplied) sides of the vapor compression heat exchangers is not reported in the conventional published cycle charts. The conventions of thermodynamics simply define these costs to be outside the definition of their system. Correspondingly, the numbers reported in
By restating the refrigeration problem with a wider boundary, recognizing the participation of target space air movement across the evaporator and condenser, it is possible to acknowledge the impact of several unavoidable problems. Being outside the thermodynamic boundaries of a closed loop refrigeration system, the latent heat regime is neither challenged nor charged commercially with the penalties that necessarily accrue. Correctly accounting for these inherent and unavoidable penalties can be focused into four problems: specific heat, pressure, pressure ratios, and humidity.
First problem, specific heat. Because R410A operates at or near the critical point, the contribution of latent heat is sharply reduced while contributions from sensible heat increase to take over completely as the refrigerant approaches “vapor phase only” temperatures in the condenser. The specific heat for R410A in the evaporator is less than 0.1953 Btu/lbm. The specific heat of air is 0.240 Btu/lbm. Air has a 23% higher specific heat than R410A, providing an attractive alternative to any refrigerant that fails to supply substantial contributions from latent heat.
Second problem, pressure. The higher operating pressures of R410A have troubled its introduction, compelling the replacement of the R22 systems equipment in total, rather than merely replacing their refrigerant. The R410A systems cost more and are more expensive to maintain. Indeed, far more expensive refrigerants accompanied by far more demanding mechanical systems are being introduced with barely incremental performance gains, if any at all.
Third problem, pressure ratios. Higher pressure ratios are defined by increased compression work and necessarily higher energy costs as pressure ratios increase. The relatively high Pressure Ratio for operating R410A refrigerant loops is increasingly problematic from the energy consumption point of view. At the chosen Rating Point (95° F.=35° C.) the resulting Pressure Ratio is 3.93 rising quickly above 4 with warmer outside temperatures as shown in
Pcomp/Pevap=(495.5 psia)/(126.07 psia)=3.93
To establish a reference for compression work needed in the R410A refrigerant loop,
Fourth problem, humidity. As humidity rises, performance drops precipitously due to the previously acknowledged high latent heat of water. The process of cooling air often results in cooling the air below its dew point, precipitating water which is discarded as waste, typically consuming 20%-35% of total cooling capacity. This was discussed in some detail above in relation to the inside approach air temperature. The Rating Point model calls for raising the temperature of recirculated inside air by about 10° C., a sensible heat of 18 Btu/lbm. This strategy avoids a considerable cost for removing humidity. Condensing water vapor consumes the full 970 Btu/lbm, 970/18=53.9 times more than the cost of cooling dry air by 10° C. There is no cooled air to show for this considerable expenditure of energy. Quite the opposite. The entire cooling load of condensation runs down the drain as chilled water, after having released the full 970 Btu/lbm heat of fusion directly into the air stream that is intended to be cooled.
Once the approach air differential is established, the fans on either side of the refrigerant loop become final controls for all heat transfer, limiting or enhancing efficiency. Yet fans and blowers generally operate well below half of their own announced efficiency.
Fan and blower driven systems raise pressures measured only in inches of water, as shown in
The cost of moving “inside” air is not even recorded, much less acknowledged in commercial statements of operating performance. Estimating the inside (target space) fan or blower resistance of duct work is difficult because it is said that the length and routing of ducts cannot be anticipated or averaged for a residence size matched to the unit capacity. This consideration has been used by the association and manufacturers to justify why the inside air movement cost is omitted from system performance measures. The industry's resistance to acknowledging inside air movement costs stands to fend off regulation in spite of the fact that the industry's sales engineers and jobbers must undeniably size every purchase and installation using estimates from recognized rules of thumb which are universally applied.
Unlike advertising claims which typically emphasize favorable facts and downplay or omit unfavorable details, typical energy requirements for fans and blowers can be found in repair and training manuals. These sometimes more reliable sources of information separate compressor data and air movement costs which are often otherwise unreported. Relevant factors which can be gleaned from these ancillary sources of data include a recognition that air movement energy is reliably proportional to system heating and cooling energy. No one will be surprised to learn that mass flow matches system capacity. Consequently, so-called rules of thumb appear to be reliable and widely accepted. Such rules of thumb, or benchmarks, include the following:
A) Inside mass air flow of 400 CFM is required for a ton of cooling capacity.
B) Energy usage is 1.1 kW/ton at the Department of Energy mandated COP of 3.2.
C) The outside fan uses 10% of reported energy consumption. The compressor alone draws 90%, 0.99 kW/ton. Use 1 kW/ton.
D) Inside air movement energy costs about 2.5 times the outside unit with wide variability, use 0.25 kW/ton.
E) Sensible Heat Ratios are 65 to 80 leaving latent heat losses of 20%-35%. Use 0.30 kW/ton.
Taking all of these things together, state-of-the-art entrenched beliefs favoring two-phase refrigeration solutions fail to recognize the following truths.
1) Latent heat makes no contribution to refrigeration whatsoever above the 95° F. Rating Point.
2) Consequently, all heat rejection at and above the 95° F. Rating Point is provided in the vapor phase.
3) The specific heat of air in the vapor phase is higher than refrigerants in the vapor phase.
4) All heat rejection is delivered at pressure ratios at or above 4.
5) Until recently, vapor compression had been delivered by a primitive single vane pump. Newer refrigerants have mandated a return to multiple piston devices, needed to meet their higher pressure requirements.
6) Compression of air as an alternative to environmentally unfriendly refrigerants has been largely dismissed because: a) it is assumed that the heat capacity of air cannot match the heat capacity of two-phase refrigerants and, b) the pumping losses would be too high to do it anyway.
7) Incredible improvements in COP are available as pressure ratios drop below 2, and to astonishing levels, literally skyrocketing (see
8) Commonplace pump designs ranging from 100-year-old vacuum cleaners to 150-year-old Roots Blowers will achieve adequate pumping efficiencies at pressure ratios in ranges near 1.1.
Accordingly, it will be appreciated that there exist substantial opportunities to improve the operating efficiencies of HVACR systems by the recognition and better exploitation of these factors in systems and methods that circulate ambient air from a target space across a heat exchanger and then return that same air back to the target space at a higher or lower temperature.
According to a one aspect of this invention, a system and method is provided for transferring heat between a heat exchanger and a gaseous medium in a thermodynamic system, while implementing a technique referred to as Convergent Refrigeration. A plenum is provided for a gaseous heat transfer medium. The plenum is inlet gated at an upstream location with a first rotary pump. The gaseous medium has an incoming pressure and temperature entering the first rotary pump. The plenum is outlet gated at a downstream location with a second rotary pump. A heat exchanger is operatively located within the plenum in-between the first and second rotary pumps. Heat is transferred into or out of the gaseous medium with the heat exchanger. The heat exchanger has a Heat Exchanger Temperature, and the gaseous medium in the plenum upstream of the heat exchanger has an Approaching Temperature. A particular attribute of this aspect of the invention relates to the step of counter-conditioning the Approaching Temperature by reducing the Approaching Temperature below the Heat Exchanger Temperature when heat is transferred into the gaseous medium from the heat exchanger and elevating the Approaching Temperature above the Heat Exchanger Temperature when heat is transferred out of the gaseous medium to the heat exchanger. The gaseous medium is returned to the incoming pressure within the second rotary pump, and work is harvested directly from at least one of the first rotary pump and the second rotary pump in the process.
This first aspect of the present invention implements the novel technique of counter-conditioning to improve overall efficiency of the system. Counter-conditioning intentionally manipulates the Approaching Temperature, moving the air temperature toward the opposite working temperature rather than away from it as occurs in prior art (i.e., Divergent) systems. By changing the ambient air stream temperature, the Air to Refrigerant Temperature Differential (A-RTD) is increased thereby improving heat transfer with respect to the heat exchanger. The Approaching Temperature is reduced below the Heat Exchanger Temperature when heat is to be transferred into the air from the heat exchanger, and conversely the Approaching Temperature is elevated above the Heat Exchanger Temperature when heat is to be transferred out of the air to the heat exchanger.
According to another aspect of this invention, a system and method is provided for transferring heat from a heat source to a heat sink in a thermodynamic system. In this case, a supply-side sub-system is in thermal communication with ambient air in a heat source, and a delivery-side sub-system is in thermal communication with ambient air in a heat sink. A heat transfer sub-system is operatively disposed between the supply-side sub-system and the delivery-side sub-system for moving heat from the supply-side sub-system to the delivery-side sub-system. Each of the supply-side and delivery-side sub-systems, respectively, provide a plenum having an upstream air inlet and a downstream air outlet. The respective plenums are inlet gated at an upstream location with a first rotary pump. The air to each plenum has an incoming pressure and temperature as it enters the first rotary pump. The respective plenums are outlet gated at a downstream location with a second rotary pump. A heat exchanger is operatively located within the plenum in-between the first and second rotary pumps. Air is moved air across each respective heat exchanger within the plenum, and as a consequence heat is transferred into or out of the air by the heat exchanger. This transfer of heat naturally provokes a change in the volume of the air within each respective plenum. In each sub-system, the first rotary pump is asynchronously operated relative to the second rotary pump so that air exiting the respective outlet is approximately equal to the incoming pressure. And work is harvested directly from at least one of the first and second rotary pumps in response to changes in the volume of the air in the plenum.
This second aspect of the present invention implements a novel dual paired, or back-to-back, arrangement in which two independent sub-systems are located on opposite sides of a shared heat exchanger. Profoundly innovative and unexpected efficiencies are revealed when two such refrigerated air flow sub-systems are arranged back-to-back, to feed and receive heat through a common (passive or active) heat exchanger, thereby dramatically increasing COP (Coefficient of Performance) at all operating temperatures.
These and other features and advantages of the present invention will become more readily appreciated when considered in connection with the following detailed description and appended drawings, wherein:
Referring to the Figures, wherein like numerals indicate corresponding parts throughout the several views, one embodiment of the invention is shown in
The hybrid heat pump and heat engine system 20 includes a working fluid (e.g., air) flow path 24, generally indicated in
A heat exchanger 32 is disposed in the flow path 24 between the inlet 26 and the outlet 28. In the exemplary embodiment of
In the exemplary embodiment of
Two arcuately spaced transition points correspond with maximum compression and maximum expansion of the working fluid. In the particular embodiment illustrated in
Working fluid ports are provided to move the working fluid into and out of the device 36. In the embodiment illustrated in
The compression chamber inlet 52 and the expansion chamber outlet 58 are generally longitudinally aligned with one another relative to the stator housing 38 for simultaneously communicating with the same chamber 48, 50. In other words, the compression chamber inlet 52 and the expansion chamber outlet 58 may be located on opposite longitudinal ends of the stator housing 38 so as to communicate simultaneously with a common chamber or chambers 48, 50. Thus a compression chamber port (inlet 52 in this example) and an expansion chamber port (outlet 58 in this example) are continuously in communication with at least one common chamber at or near a transition point. A pump 60 may be disposed in the flow path 24 between inlet 26 and the compression chamber inlet 52 for propelling the working fluid into the stator housing 38 through the compression chamber inlet 52.
The rotor 42 is rotatably disposed within the stator housing 38 for rotating in a first direction. While the rotor 42 is rotating, the vanes 46 slide along the inner wall 44 of the stator housing 38 and simultaneously reduce the volume of the compression chambers 48 and increase the volume of the expansion chambers 50. In the exemplary embodiment, vane-type device 36 accomplishes the simultaneous compression and expansion because the cross-section of the inner wall 44 of the stator housing 38 is circular and the rotor 42 rotates about an axis A that is off-set from the center of the circular inner wall 44. Alternatively, the stator housing 38 could be elliptically shaped and the rotor 42 could rotate about the center of the elliptical stator housing 38. Other configurations are of course possible, including those described in U.S. Pat. No. 7,556,015 as well as those described in priority document U.S. Provisional Application Ser. No. 61/256,559 filed Oct. 30, 2009, the entire disclosure of which is hereby incorporated by reference and relied upon.
The embodiment of
Next, the air is pushed out of the vane-type device 36 through the expansion chamber outlet 58 by the air entering the vane-type device 36 through the compression chamber inlet 52. Finally, the air is discharged to the atmosphere 30 through the outlet 28. The difference in the pressure of the air entering the expansion chambers 50 and the atmospheric pressure represents potential energy. The expansion chambers 50 of the vane-type device 36 harness that potential energy and use it to provide power to the rotor 42.
The system includes a combustion chamber 62 in the flow path 24 between the compression chamber outlet 54 of the vane-type device 36 and the heat exchanger 32. During the standard heating/cooling mode, described above, the combustion chamber 62 remains dormant. However, during an optional high heating mode, a fuel introduced into the combustion chamber 62 is combusted, or burned, in the working fluid to greatly increase both its temperature and pressure within the flow path 24. The fuel may be any suitable type including for examples natural gas, propane, gasoline, methanol, grains, particulates or other combustible materials.
The compression chambers 48 of the vane-type device 36 compress the air by a first predetermined ratio, and the expansion chambers 50 of the vane-type device 36 expand the air by a second predetermined ratio. In the
In operation, during the high heating mode, the pump 60 propels atmospheric air into the vane-type device 36 through the compression chamber inlet 52. The temperature and pressure of the air both increase as the air is compressed in the compression chambers 48. The pressurized and warmed air then exits the vane-type device 36 through the compression chamber outlet 54 and flows into the combustion chamber 62. In the combustion chamber 62, the fuel is mixed with the air and combusted to greatly increase the pressure and temperature of the air. The air then flows through the heat exchanger 32 where it dispenses heat to warm the target space 22. Next, the valve 64 directs a predetermined amount of the air to the expansion chamber inlet 56 of the vane-type device 36 and the remaining air to the secondary expander 66. In the vane-type device 36, the pressurized air is expanded, preferably to or nearly to the atmospheric pressure, before it is discharged out of the flow path 24 and to the atmosphere 30 through the outlet 28. The air in the secondary expander 66 is also expanded, preferably to or nearly to atmospheric pressure, while powering the generator 68 to produce electricity. After the air is expanded by the secondary expander 66, it is also directed to the outlet 28 to be discharged to the atmosphere 30.
Through reconfiguration, the embodiment of
The vane-type device 36 can also work in a closed loop system 70, as generally shown in
The closed loop system 70
In the first operating mode, the rotor 42 rotates in a first direction, causing the pressure and temperature of the working fluid in the compression chambers 48 to increase as the volume of those compression chambers 48 decreases. That working fluid then flows into the high-pressure side heat exchanger 72 where it dissipates heat to either the target space 22 or the atmosphere 30. The pressurized and cooled working fluid then flows into the expansion chambers 50 through the expansion chamber inlet 56. In the expansion chambers 50, the temperature and the pressure of the working fluid decrease as the volume of the expansion chambers 50 increases. The working fluid leaves the expansion chambers 50 through the expansion chamber outlet 58 and flows to the low-pressure side heat exchanger 74. In the low-pressure side heat exchanger 74, the working fluid receives heat from either the target space 22 or the atmosphere 30 before flowing back into the compression chambers 48.
Similar to the open loop embodiment of
An open-loop air aspirated hybrid heat pump and heat engine system 20 having a compressor 76 separated from the expander 78 is generally shown in
An energy receiving device is mechanically connected to the expander 78 for harnessing potential energy from the air in the flow path 24 as will be discussed in further detail below. In the exemplary embodiment, the energy receiving device is a generator 68 for generating electricity. The electricity can then be used immediately, stored in batteries or inserted into the power grid. Alternatively, or additionally, the energy receiving device could be a mechanical connection between the expander 78 and the compressor 76 for powering the compressor 76 with the energy reclaimed from the air in the flow path 24. The energy receiving device could also be any other device for harnessing the energy produced by the expander 78.
A controller 82 is in communication with the compressor 76 and the expander 78 for controlling the hybrid heat pump and heat engine system 20. The controller 82 manipulates or switches the system 20 between different operating modes: a standard heating/cooling mode (in which the target space 22 can be either heated or cooled), and a high heating mode (in which the target space 22 is heated). The operating mode may be selected by a person, or the controller 82 can be coupled to a thermostat for automatically keeping the target space 22 at a desired temperature.
In reference to
In the standard heating mode, the controller 82 directs the compressor 76 to compress the air from the inlet to increase the pressure and the temperature of the air, as will be understood by those skilled in the art. The pressurized and warmed air then flows through the flow path 24 to the heat exchanger 32. The heat exchanger 32 dispenses heat to the target space 22 to warm the target space 22. Although the air in the flow path 24 is cooled by the heat exchanger 32, the air remains pressurized when compared to the air entering the flow path 24. This difference in pressure represents potential energy, which can be harnessed. The generator 68, which is coupled to the expander 78, harnesses this potential energy while the expander 78 expands the pressurized air to reduce the pressure of the air. Preferably, the air is expanded back to the same pressure at which it entered the flow path 24. Following the expansion, the air is discharged from the flow path 24 through the outlet 28.
In the high heating mode, the compressor 76 receives air aspirated from the inlet 26 and then compresses the air to increase its pressure and also its temperature (in compliance with relevant thermodynamic gas laws). The pressurized and high temperature air then flows through the flow path 24 to the combustion chamber 62, which mixes a suitable fuel with the air and then combusts the mixture. The combustion of the fuel and air mixture further increases both the pressure and the temperature of the air in the flow path 24. The pressurized and heated air then flows through the heat exchanger 32 and dispenses heat to the target space 22. Air leaving the heat exchanger 32 in the high heating mode remains substantially highly pressurized relative to the ambient air pressure, and therefore represents a valuable amount of potential energy. The generator 68 maybe of any suitable type that is effective to convert this potential energy into another form, such as electricity and/or mechanical energy. This potential energy may be harnessed while the expander 78 expands the air to reduce the pressure of the air, or accumulated for conversion at a later time. In other words, any residual pressure energy put into the air through the initial compression and combustion processed is subsequently re-claimed by the generator 68. Once the potential energy has been reclaimed, the low pressure air is then discharged from the flow path 24 through the outlet 28 back into the environment 30.
Among the several embodiments presented herein, the invention may be defined in one sense as a system and method for circulating ambient air from a target space across a heat exchanger and back to the target space at a higher or lower temperature. According to still other aspects, the present invention may be defined as a system and method for transferring heat to or from a heat exchanger to a gaseous medium within the subject thermodynamic system. Before advancing further in the detailed description, it will be helpful to re-state the main components and primary elements of the invention, from which these several aspects can be better understood to accomplish the various objectives of this invention.
Within and among these various aspects, the above-described flow path 24 comprises a plenum for a gaseous heat transfer medium, which in the preferred embodiments comprises air. However, in some embodiments it is contemplated that the gaseous heat transfer medium could be a refrigerant gas other than air. The plenum 24 has an upstream inlet 26 in fluid communication with the target space 22 and a downstream outlet 28 in fluid communication with the target space 22.
Ambient air is drawn from the target space 22 into the inlet 26 of the plenum 24 at an incoming pressure and an incoming temperature. As stated above, the target space 22 may be either the inside or outside ambient air zone, depending upon which is the subject of focus with respect to the refrigerant being considered. The drawing step may include positioning a filter device at or near the inlet 26 to filter particulate from the incoming air. The plenum 24 is inlet gated at an upstream location with a first pump 76 which may comprise a rotary device like that shown in
In some embodiments, air is taken into the first rotary pump 76 using substantially atmospheric pressure from the target space 22. That is to say, the first rotary pump 76 may be configured to allow its expansion chamber 50 to fill with air using atmospheric pressure, such as by remaining open and exposed to air from the target space 22, as in
The plenum 24 is outlet gated at a downstream location with a second rotary pump 78, as shown in
A heat exchanger 72 is operatively located within the controlled pressure zone of the plenum 24, i.e., in-between the first 76 and second 78 rotary pumps. By concurrently rotating the first 76 and second 78 rotary pumps, air traveling through the plenum 24 is moved across the heat exchanger 72. The heat exchanger 72 may be viewed as always possessing an instantaneous Heat Exchanger Temperature. And the air in the plenum 24 that is upstream of the heat exchanger 72 will always have an Approaching Temperature that may be different (higher or lower) from the Heat Exchanger Temperature. When the air interacts with the heat exchanger 72, such as by flowing through fins, heat is transferred either into or out of the air. That is to say, if the Heat Exchanger Temperature is higher than the Approaching Temperature, heat will flow into the air from the heat exchanger 72. But if the Heat Exchanger Temperature is lower than the Approaching Temperature, heat will flow out of the air and into the heat exchanger 72.
Because the second rotary pump 78 gates the downstream end of the plenum 24 and prevents backflow, rotation of the second rotary pump 78 is required to discharge the air from the outlet 28 of the plenum 24. Accordingly, whenever heat is transferred, air will be discharged from the outlet 28 at a differentiated temperature relative to the incoming temperature.
Whenever the Heat Exchanger Temperature is different from the Approaching temperature, the temperature of the air within the plenum 24 downstream of the heat exchanger 72 is altered by the transfer of heat to or from the heat exchanger 72. This transferring of heat provokes a change in the volume of the air within the plenum 24. As is well-documented and generally known to those of skill in the art, because air is a gaseous medium, a temperature increase in the air will cause the volume of the air to increase when constant pressure is maintained. That is, the air expands when it is heated. And conversely, the volume of the air decreases in proportion to decreases in its temperature. Cooling air contracts. Therefore, when heat is transferred into the airstream by the heat exchanger 72, the volume of the air within the plenum 24 will increase by a mathematically determinable amount. And when heat is transferred into the heat exchanger 72 from the flowing air within the plenum 24, the volume of the air within the plenum 24 will decrease by a mathematically determinable amount.
In some embodiments of the present invention, a generally constant pressure of the air transiting the plenum 24 is maintained at the aforementioned ultra-low range notwithstanding the temperature-induced volume changes therein. Maintaining a generally constant, ultra-low pressure within the plenum 24 may be accomplished by proportionally varying the rotation speed of the first rotary pump 76 relative to the second rotary pump 78. This exercise is particularly beneficial when combined with the afore-mentioned option of controlling the first rotary pump 76 so as not to directly increase or decrease air pressure greater than about 10-20% (and most preferably in the ultra-low range of 0-10%) relative to the incoming pressure. In fact, a variety of beneficial results are to be gained when maintaining this constant low pressure, which benefits will be discussed later.
In some alternative embodiments of the present invention, a counter-conditioning step is performed to improve overall efficiency of the system. Counter-conditioning refers to an intentional manipulation of the Approaching Temperature to deliver Convergent Refrigeration, which by definition will not fall within the scope of the Fan Replacement technique. That is to say, a system configured according to the principles of this invention can be operated to achieve both Fan Replacement and Convergent Refrigeration, however not concurrently. In particular, counter-conditioning occurs when the Approaching Temperature is manipulated to increase the Air to Refrigerant Temperature Differential (A-RTD).
Conventional (prior art) refrigeration was categorized above as Divergent Refrigeration. Divergent Refrigeration offers no option for improving heat transfer except by increasing excess refrigerant lift. Refrigerant lift is increased only by moving the refrigerant temperature farther away from the working temperatures which define the refrigeration task. The prior art open air cycle methods and systems, discussed previously, all require that when using air as the refrigerant its refrigerant temperature must be changed substantially beyond the opposite working temperature. Only by providing this excess refrigerant lift is it possible for Divergent Refrigeration methods and systems to induce the requisite flow of heat. Divergence is defined by excess refrigerant lift on the opposite side of the companion working temperature.
Convergent Refrigeration delivers exponentially greater efficiencies while utilizing much smaller pressure ratios. In other words, it is not the employment of an open air cycle that defines Convergent Refrigeration; rather it is the unprecedented capability to move a comparable amount of heat with a significantly smaller amount of work.
In Divergent Refrigeration, the Approaching Temperature of the ambient air stream is always defined by one of the working temperatures THIGH or TLOW. Convergent Refrigeration changes the Approaching Temperature of the ambient air stream just prior to the heat exchanger even when the heat exchanger is of the type used by a traditional Divergent Refrigeration system. Because the temperature of the ambient air stream is otherwise defined by one of the working temperatures, Convergent Refrigeration is said to counter-condition the air stream, moving its temperature toward the opposite working temperature rather than away from it as would be required in every Divergent Refrigeration system or contrivance. Correspondingly, some embodiments of Convergent Refrigeration will be seen to be augmenting or supplementing Divergent Refrigeration systems. By changing the ambient working temperature, in other words counter-conditioning the Approaching Temperature convergently, the A-RTD is increased thereby improving heat transfer with a conventional heat exchanger. The Approaching Temperature is reduced below the Heat Exchanger Temperature when heat is to be transferred into the air from the heat exchanger 72, and conversely the Approaching Temperature is elevated above the Heat Exchanger Temperature when heat is to be transferred out of the air to the heat exchanger 72. Convergent Refrigeration can operate essentially between the working temperatures, THIGH and TLOW, rather than beyond these temperatures. No known prior art refrigeration system is capable of operate essentially between the working temperatures, THIGH and TLOW. Divergent Refrigeration can only operate outside and beyond the working temperatures, THIGH and TLOW. Moreover, even then Convergent Refrigeration provides for the reduction of excess refrigerant lift by optimization of the heat transfer temperature which cannot be practiced in any other type of open air cycle known.
Specific details pertaining to this counter-conditioning step used to deliver Convergent Refrigeration are provided below, along with supporting mathematical proofs. At this point in the description it may be valuable to note that the counter-conditioning step includes manipulating the first rotary pump 76 relative to the second rotary pump 78 to change the pressure of the air (or other gaseous medium) in the plenum 24. That is to say, the manipulating step includes reducing the pressure of the air relative to the incoming pressure when the heat exchanger 72 transfers heat into the air, and increasing the pressure of the air relative to the incoming pressure when the heat exchanger 72 transfers heat out of the air. In one embodiment, a controller, such as controller 82 in
Counter-conditioning changes the Approaching Temperature of the air stream within the plenum 24, increasing its temperature differential with respect to the Heat Exchanger Temperature. Counter-conditioning increases the rate of heat transfer to or from the air within the plenum 24. Fan Replacement, on the other hand, may leave the Approaching Temperature unchanged and in that case would not affect the rate of heat transfer except perhaps by increasing or decreasing the mass flow rate. Thus, a contrast between the concepts of Fan Replacement and Convergent Refrigeration can be clearly seen: Fan Replacement seeks to maintain a generally constant (preferably ultra-low) pressure within plenum 24, whereas Convergent Refrigeration (or counter-conditioning) seeks to intentionally manipulate the pressure within the plenum 24 to facilitate heat transfers between the air and the heat exchanger 72. The present invention makes use of substantially the same physical equipment to accomplish both Fan Replacement and Convergent Refrigeration, however both techniques are practiced mutually exclusively. The controller 82 thus regulates the system to operate either in Fan Replacement mode or in Convergent Refrigeration mode.
Accordingly, the techniques of Fan Replacement and counter-conditioning (i.e. Convergent Refrigeration) may be implemented independently from one another. That is to say, the present invention can be configured to accomplish Fan Replacement exclusively, or counter-conditioning exclusively, or both. Nevertheless, in all scenarios the air (or other gaseous medium) is returned to the incoming pressure within the second rotary pump 78 prior to discharge. Said another way, the system and methods of this invention always seek to exhaust air from the outlet 28 of the plenum 24 at very close to the incoming pressure. By this means, the invention aims to harvest work directly from at least one of the first 76 and second 78 rotary pumps in response to changes in the volume of the air in the plenum 24 due to heat transfers under constant pressure. Rather than expelling energy in the form of pressurized or de-pressurized air from the plenum 24, back into the atmosphere where it undergoes free (i.e., wasted) expansion, in all forms of this invention the work potential of volume change due to heat transfer is captured and harvested to the extent possible. Importantly, in every case, the energy spent increasing or reducing pressure in the plenum 24 is directly recovered so there are little to no energy losses due to adiabatic heating or cooling per se.
One possible way to harvest the energy is depicted in
The most powerful iterations of Fan Replacement and counter-conditioning (i.e., Convergent Refrigeration) are embodied within a dual paired, or back-to-back, arrangement in which two independent systems are located on opposite sides of a shared heat exchanger 72, like those examples depicted in
The following descriptions detail the various embodiments of
It must be emphasized that direction of flow could be reversible and pump sizes do not govern the outcome when rotation speeds can be sufficiently controlled by the controller 82. The controller 82/transmission 86 apparatus or electronics will raise or lower the pressure in the plenum 24 electively, regardless of flow direction and pump size. For example,
Although not shown in
Turning now to
For the sake of this illustration, therefore, the thermodynamic system in
Turning now to the expansion side of the thermodynamic system in the preceding example, working fluid upstream of the valve 84′ is maintained at 1.2 ATM. The valve 84′ is controlled by a regulator 96 or control system so that it remains open long enough to admit a volume of working fluid into the expansion side of the rotary device 36′ so as to achieve the desired operating conditions. The regulator 96 may be configured so as to maintain constant operating pressures, specified volumetric flow rates of the working fluid and/or desired temperature rejections from the high side heat exchanger 72. Alternatively, the regulator 96 may be coupled to rotation of the rotor 42 so that it closes the valve 84′ when the rotor 42 reaches a specified angular position. The opening and the closing of valve 84′ by the regulator 96 is based, ideally, on the amount of heat moved (in this example via the high side heat exchanger 72). Thus, considering a lobe 92 crossing the inlet 88, the retractable vane 94 will be closed against the outer surface of the rotor 42 with working fluid at the differentiated pressure (1.2 ATM) filling behind the lobe 92. This lobe 92 will be allowed to rotate sufficiently with the valve 84′ in an open condition until the desired volume of working fluid is contained in the expansion chamber.
At this point, which may correspond to one of the phantom representations of a lobe 92 in the 4-5 o'clock positions of
In some cases, it may be desirable to over-expand the working fluid to effect additional cooling, but the working fluid will be returned again to the inlet pressure prior to discharge through the outlet 28. To deliver over-expansion, outlet 28 would be equipped with a check valve identical to 84 but set to release exhaust at the outlet pressure, in this case 1.0 ATM. Over-expansion would result from exactly the same normal process with the single exception that the inlet valve 84′ would be closed sooner. Because a smaller mass of air is admitted behind the rotating lobe 92, its pressure would be reduced below the exit pressure by the time rotating lobe 92 reaches the exit port leading to outlet 28. Therefore, the check valve set to 1.0 ATM will remain closed. In the following cycle, the lobe 92 leaving TDC will perform compression on the lower pressure over-expanded gas which was just established on its leading edge by the previous sweep of the chamber. As this lobe 92 sweeps clockwise it will perform an ordinary compression sweep. As soon as the gas is re-compressed to its exit pressure, check valve (installed on exit port leading to outlet 28) will crack open and release the gas as exhaust. This over-expansion technique returns the working fluid to the inlet pressure. Over-expansion is employed either to quick cool (self-cool) the inner walls of a chamber or to provide a pneumatic flywheel mechanism to temporarily store and balance rotating energy.
In another example of the system of
When the trailing side of a lobe 92 crosses the expansion chamber outlet, working fluid at the differentiated pressure (0.8 ATM) is emitted to the low side heat exchanger 72, where it absorbs heat in the counter-conditioning manner described above. Upon reentering the rotary device 36′ through the compression chamber inlet, the working fluid now has a higher temperature, but remains at or near the differentiated pressure of 0.8 ATM. The valve 84 associated with the compression chamber outlet 90 is again, in this example, configured as a check valve whose cracking pressure is equivalent to the pressure of the high side heat exchanger 72 which, in this example, is 1.0 ATM or ambient conditions. Thus, the working fluid in the compression chamber (i.e., on the leading edge of lobe 92) re-compresses from differentiated pressure (0.8 ATM) to the inlet pressure (1.0 ATM) until such time as the valve 84 automatically opens. Thereafter, working fluid in the compression chamber is expelled to the atmosphere in the target space 22 which is at the inlet pressure. Appropriate temperature sensors and/or pressure sensors 98 monitor the amount of heat being moved through the heat exchanger 72 and provide feedback to make appropriate corrections to close the valve 84′ at the precise moment so that heat is moved with the minimum theoretical application of work. These operations occur without decreasing the volumetric efficiency of either the compression or expansion chambers. In fact, the full volume of all chambers is fully utilized at maximum efficiency at all times.
Of course, the device illustrated in
Another novel feature of this device 36′ is that the working fluid moves through the four modes of intake, expansion, compression and exhaust modes without a change in lobe 92 direction. That is, the lobes 92 continue rotating with the rotor 42 without requiring a reversal of direction as is characteristic of piston and cylinder devices. Furthermore, it is well known that in the typical piston and cylinder device, peak and minimum pressures are generated when the piston is in its Top Dead Center and Bottom Dead Center positions which usually means that both ends of the connecting rod are aligned with crank shaft center line. In most piston/cylinder configurations, whenever both ends of the connecting rod align with crank shaft center line, the component of force able to produce or receive torque is zero. Only for those brief instants when then crank arm is offset 90 degrees is the leverage maximized so that the component of force able to produce or receive torque is at its peak value. By contrast to the typical prior art piston/cylinder arrangement, the device 36′ presents a configuration in which the peak power can be sustained for a longer percentage of the cycle. In other words, the working fluid (e.g., air) either receives mechanical energy from or imparts mechanical energy to the lobes 92 at maximum leverage for a corresponding larger portion of the rotation of the rotor 42. This results in a more efficient, powerful and smoother performance, as compared with a comparable piston/cylinder device. When operated as a combustion engine, it also invites the opportunity to function with a reduced size or weight flywheel, if indeed a flywheel is even needed. The mention of combustion in connection to the device of
In these preceding examples associated with
Moving now to
As described in U.S. Pat. No. 8,424,284, the mechanisms and methods define themselves within a refrigeration paradigm which requires excess refrigerant lift. Any such practice, method, or mechanical technology requiring the temperature of the refrigerant to be lifted by the amount of Approaching Temperatures, in addition to the difference between the working temperatures, TLOW and THIGH, is to be labeled Divergent Refrigeration.
U.S. Pat. No. 8,424,284 has outlined the use of compression or expansion to raise or lower the temperature on one side of a heat exchanger 72 by means of using the ambient air as the working fluid refrigerant. This pump-based procedure uses adiabatic compression for cooling. The temperature of ambient air is raised from TLOW to THIGH, the difference between the two working temperatures. And in addition, the temperature is further raised by an amount above THIGH equal to the outside approach air temperature differential. (See
The previously described
Recognizing that the cost of moving air alone commonly exceeds 30% of conventional air conditioning costs, it is attractive to consider simply replacing fans with pumps. Fans and blowers are notoriously inefficient. In addition to electric motor losses which range typically from 10%-25%, the fans themselves frequently waste as much as 85% of applied energy. These are the worst sorts of pumping losses. When viewed as air moving devices, pumps inherently develop the negligible pressure needed to propel a static column of air. Pumps move air as a relatively cost free byproduct that fans and blowers produce only wastefully. By simply reallocating the wasted energy of fans to very minor compression/expansion tasks it is possible to “refrigerate” many air streams without additional cost overall. These air streams already deliver the entire mass flow of air needed to perform all HVACR tasks.
The previously mentioned technique of Fan Replacement identifies the opportunity to reclaim losses from free expansion. When heat is exchanged with air inside the plenum 24, the volume of the air changes. This volume change is even defined into the coefficient of specific heat for heat transfer at constant pressure. For air at atmospheric pressure the work potential of changing volume is equal to 40% of the heat transferred. Instead of using fans, air can be moved by well-established commercial pumps proven to deliver efficiency above 95% at needed pressure ratios. At present such pumps are more expensive than fans, but lower cost options and operating cost offsets will be described.
The prevailing latent heat argument asserts that air does not provide sufficient heat capacity for refrigeration. This widely held belief falls categorically before the indisputable fact that all latent heat (vapor compression) refrigeration necessarily requires a mass flow of air sufficient to carry all the heat into and out from every vapor compression system—twice in fact. Air alone carries the entire heat load of vapor compression on both sides of every vapor compression system. This fact confirms that air possess adequate heat capacity. Furthermore, at higher temperatures as explained below, there is no contribution from latent heat in the vapor compression cycle anyway. The reality is that a vapor phase refrigerant with a lower specific heat than air can do and does do the entire refrigeration job without latent heat, and contrary to popular beliefs it does so even inside what is identified as a vapor compression refrigeration system.
According to one aspect of the present invention, referred to as the Fan Replacement technique, traditional fan blowers are replaced with pumps 76, 78 located at opposites ends of a gated plenum 24 so as to capture lost energy of free expansion during heat transfers. The bonus is a direct work dividend equal to 40% of all the heat moved. Convergent Refrigeration systems radically increase efficiency by eliminating excess refrigerant lift across the heat exchanger 72 from the nominal values of Tevap and Tcond, but the identified excess refrigerant lift barely hints at the unacknowledged and extreme energy waste of high pressure ratios, the temperature swings of superheat which are actually required to do the job of vapor compression refrigeration. Convergent refrigeration accomplishes the task with counter-conditioning as previously outlined, using a heat transfer temperature (the midpoint of any appropriate air-to-air heat exchanger or heat pipe) nominally set between the two working temperatures. The distinctive advantage of Convergent Refrigeration is improved efficiency with a reduction in total refrigerant lift for operation between any two working temperatures. The entire energy cost of running compressors to supply the extreme pressures of vapor compression refrigeration loops is zeroed out by any suitable air-to-air heat exchanger 72. This yields particular benefits when placed between two counter-conditioned Convergent Refrigeration air flows as described below.
The cooled vapor will condense and return to the liquid state. The cooled liquid then flows back to the hot end for another load of heat. This natural heat conveyor runs naturally, i.e., without requiring any additional input power. Only a single boiling point is involved and the pressure is unchanged throughout this closed two-phase refrigerant system. As is well understood for such refrigerants, the boiling point may be regulated by simply moderating the heat pipe system pressure. All the power for transporting and eliminating unwanted heat is supplied by the energy of the heat to be eliminated.
Air flows are separated in this illustration by a partition 102 which prevents mixing of the heat flows or air streams. In practice the hot and cold ends of the heat tube 100 may be some distance apart. The hot end may be in direct conductive contact with a heat source such as a component inside a computer enclosure (e.g., computer chip), a CPU cooler, any heat-emitting electronics enclosure or cell phone processing chip as mentioned previously. The liquid boiling point may be set to match precisely the temperature of the heat input by changing the pressure on the liquid (refrigerant) inside the heat pipe 100. Indeed, the liquid refrigerant may even be pumped for some distance and to new elevations at low cost because no change in pressure is required.
Those of skill in the art will understand that the specific configuration of a heat pipe 100 as illustrated in
One may ask, “What is the least costly way to change the temperature of the air in the room?”. It has always been known and always understood that, whether heating or cooling, the needed mass of air must be passed over a heat exchanger. Air has the needed heat capacity. It has long been known that heat transfers into the air faster when the approaching air temperature is farther away from the temperature of the heat exchanger. However, it is not well understood that 40% of the heat is lost in free expansion when gasses expand and contract (due to temperature changes) without harnessing the potential work available within the context of those volume changes. Heretofore, no recognition has been given to the fact that the cost of changing the air temperature before it interacts with the heat exchanger can be much less than the cost of changing the heat exchanger temperature by the same amount. The present invention explains how this behavior can be realized with significant advantages in commercial HVACR.
The present invention proposes better ways to heat and cool air, through the techniques of Fan Replacement and Convergent Refrigeration (i.e., counter-conditioning), which will be described in even greater detail below. By placing the heat exchanger 72 within a plenum 24 gated between two pumps 76, 78, it is possible to capture a 40% energy rebate provided by nature every time heat is transferred into air, which is the basis of the Fan Replacement concept. This same 40% guaranteed energy rebate is also provided in counter-conditioned air flows wherein the pressure is increased, i.e. heat source air streams intended to reject heat from TLOW. (The mechanics of reducing air pressure between two pumps unfortunately requires the initial reduction of pressure in the plenum 24 as well as its maintenance, so the opportunity to capture work from volume change exists only in positive pressure mechanical systems. This provides an argument for counter-conditioning only the heat source air stream and utilizing Fan Replacement exclusively on the heat sink side to reclaim all the benefits of work due to volume change throughout. The best theoretical heat transfer temperature is thermodynamically nonetheless still clearly the midpoint between the two working temperatures. It remains to be seen how practical mechanical considerations may influence improvements in real world settings.) To secure the most favorable temperature gradient between air and any convective heat source or sink, it costs less to change the temperature of the air (i.e., counter-condition) than to change the temperature of the heat exchanger 72 by divergent refrigeration means. This is Convergent Refrigeration.
The following analysis separates the cost of moving air with pumps from the cost of compression mirrored by a complimentary expansion in the same air stream by using a pair of Dresser Roots® Blowers. As illustratively depicted in
Dresser URAI® blower performance is specified for the whole family of blowers in the available literature. (Dresser, Universal RAI and Roots are registered trademarks of Dresser, Inc. Data provided in URAI Spec Sheet S-12K84 rev. 0608 provides the basis for conclusions which follow.) Mass flows are suitable as stated because air flows in refrigeration systems are normally driven by fans. The desired changes in pressure (temperature) maintain the same mass flow. Dresser URAI® specifies inlet pressure of 14.7 psia at 68° F., specific gravity 1.0. Vacuum discharge is 30″ Hg as well as all relevant performance data for commercial purchase. It can be seen in the published literature that at 1 psig and 6 psig, the energy cost to both move and compress a cubic foot of air increases roughly linearly across the range of flows and pressures regardless of the device actually chosen. Because the proposed air flows of convergent refrigeration systems will operate primarily near atmospheric pressure ±10%, rarely exceeding 20% differences, only the published data associated with 1 psi governs the relevant conclusions. Others provide confirming data beyond this range.
Rather than simply moving the air, the objective of the counter-conditioning utilized by Convergent Refrigeration is to move a comparable mass flow of ambient air through a pressure differential sufficient to change its Approaching Temperature to a desired level in relation to the heat exchanger 72. In conventional systems, the ambient (target environmental) mass flow is passively fed across a heat exchanger 72 whose temperature is separately engineered to provide the desired rate and direction of heat flow. Contrast this to Convergent Refrigeration systems of this present invention where the ambient (target environmental) mass flow is used as the refrigerant. The temperature of CR mass flows is engineered to provide the desired rate and direction of heat flow now being exchanged with a passive heat exchanger 72 whose source or sink is thermodynamically considered to be outside the thermodynamic system under consideration.
Correspondingly, in order to compare the energy that would otherwise be required simply to move the air, it is necessary to identify the cost of compressing the air and subtract that compression cost from the reported cost of compressing and moving the air. The reported cost of compressing air as reported inherently includes the cost of moving the air, so the thermodynamic work assignable to compression alone is easily computed and subtracted from the reported total to reveal the cost of moving air alone in these Roots Blowers.
For the case where no heat is transferred following compression, a follow-on expansion process might recover the entire energy cost of compression directly by complimentary mechanical means. The Roots® Blower offers such a mechanism, as one example of a suitable mechanism implementing the pumps 76 and 78. Other types of rotary pumps 76, 78 are also possible as described herein. Notably this energy recovery mode during expansion is different from both the compression operation and the vacuum pump for which data is available. But a free-wheeling exit pump 78 would not sustain the plenum 24 pressure as needed for heat transfer under constant pressure. An electrical load would be provided to the motor/generator 68 (
For the case where heat is acquired within the plenum 24 (i.e., heat is moved from a higher temperature heat exchanger 72 into lower Approaching Temperature air flowing through the plenum 24), the resulting increase in volume of the air in the plenum 24 will directly increase the energy recovered at exit, in the fashion of a heat engine. Thermodynamically, the addition of heat yields work. The introduction of heat between the two pumps 76, 78, as in
For the case where heat is rejected within the plenum 24 (i.e., heat is moved from the higher temperature air flowing through the plenum 24 into a lower temperature heat exchanger 72) the resulting decrease in volume will directly decrease the energy recovered at exit. In this case the departure of heat from the air mass within the plenum 24 reduces the volume of the air (but not its mass) by 40%. Strikingly, this reduction of volume also affects the system and its net energy consumption in a manner analogous to the heat engine behavior described above because work can be extracted from the larger volume of air entering the plenum. Because the plenum 24 pressure must be maintained in Fan Replacement, the exit pump 78 energy expenditure is offset by the greater volume of air drawn through the entry and energy is recovered there.
When all is accounted for, the transfer of heat makes a 40% contribution to offset the losses related to compressing and expanding the air within the plenum 24. This net contribution may substantially offset pumping losses depending on the capability of the pumps 76, 78 as well as on the compression ratios and the heat finally transferred. Because this exercise is limited to published pump performance at a pressure of 1 psig, a pressure ratio of 1.068, it can be confidently assumed that compression costs will be offset by expansion gains and vice-versa. Looking at the operating energy requirements reported by Dresser®, the full value of compression/expansion energy may be subtracted from the operating energy cost, leaving all losses chargeable to air movement alone.
Any pump actually designed and developed for these low pressure ratios may be expected to meet or exceed all currently reported performance specification. Because the Roots® Blower was intended for much higher pressure ratios, it is reasonable to benchmark compression performance at 90%, knowing that the entire cost of compression and expansion will be directly offset, i.e. zeroed out. For example, Dresser® Frame #718 delivers 1590/0.81 CFM/BHP total or 2628 CFM/Kw for air movement alone, after the cost of compression has been removed. Compared to residential HVAC air flows (2,000 CFM/Kw inside and 4,000 CFM/Kw outside), any such 2628 CFM/Kw unit will deliver heating and cooling comfortably within the energy budget of present fan systems alone.
The analysis has identified several factors which control the energy needed to change the pressure of a mass flow of air within a gated plenum 24 between two pumps 76, 78. Whether the temperature between the pumps 76, 78 is changed or not, and whether heat is transferred or not, the complimentary compression/expansion energy can be definitively identified. Subtracting this fully recovered compression/expansion energy component from the total pumping energy reveals the cost of moving air through the system, nominally through the connected system where the follow-on pressure is measured only in inches of water. The cost of moving air through the dual pump system is well below the cost of moving the same mass flow of air with fans. This simple reality confirms that the two-pump and plenum air moving system can confidently be accurately labeled as Fan Replacement.
The common Roots® Blower was initially developed more than a century ago for high compression applications. It is machined from cast metals. Even when adapted for supercharging high performance automotive vehicles, the lighter weight versions of the Roots® Blower still rely on machined castings. In U.S. Pat. No. 7,621,167 to Staffend, issued Nov. 24, 2009, a method is taught for replacing such castings with light weight roll-formed products that inexpensively deliver three orders of magnitude better surface finish than the best attainable machined casting. The results displayed above can be mass produced with dramatic cost reductions. Much more importantly, the combination of inexpensive mass production with the disruptive market opportunity presented by Convergent Refrigeration invites a vast new wave of innovation for related HVAC products as well as many other pumps and engines throughout the Pressure v. Volume product space.
All traditional fans waste the work component of cp, the coefficient of specific heat under constant pressure. This is the energy saving opportunity that is currently unrecognized, even denied, in academic and industry teachings on heat transfer. The present invention identifies and takes advantage of this phenomenon, resulting in the equivalent of a 40% instant energy rebate.
Using a pair of Roots® Blowers for the two rotary pumps 76, 78 operating at pressure ratios within 20% of atmospheric, and more preferably within 10%, the efficiency of each blower or positive displacement pump is near 0.9. Combined efficiency is thus characterized as 0.9*0.9=0.81. Utilizing a typical 3-ton household air flow of 1250 CFM through the HPT heat exchanger 72 HRM 3040 calls for the following power.
kW=CFM/(11674*Motor Eff*Fan Eff)
kW=1250/(11675*(0.9*(0.9*0.9)))=0.147 kW
As expected, using a pair of positive displacement pumps 76, 78 will move air more efficiently than the traditional fan they replace. When heat is exchanged with the transient air column moving through the plenum 24, the bonus harvest of ˜40% of the heat exchanged will be reduced by pumping losses. Nonetheless, Fan Replacement at or near the ultra-low pressure ratio of 1.0 still yields a net gain quite close to this goal.
The Fan Replacement technique of this present invention corrects for the widespread, perhaps universal failure to comprehend the work lost as free expansion in common situations involving cp. The premier academic authority (incorrectly) defines convection with the stipulation that the density of the gas does not change during heat transfer. In spite of the fact that the amount of heat exhausted by both automotive and jet aircraft engines is correctly computed with cp, textbooks uniformly fail to mention that the work component of heat engine exhaust is necessarily never captured in convective heat transfer in the same manner as it is in combustion contexts. The work component of cp is wasted as free expansion in the exhaust of every heat engine. The same failure to recognize the work component of cp is pervasive throughout the literature on refrigeration as well.
Fan Replacement means quite literally to replace the traditional fans in forced air convection systems with a plenum 24 gated at each end with a rotary pump 76, 78. Traditional fans will blow the same mass flow of air into heat exchangers regardless of changing heat demands, mindlessly intent on driving out the air that was previously heated. In contrast, the Fan Replacement technique meters in fresh ambient air at the full value of its Approaching Temperature as needed to attain the greatest efficiency in managing optimum mass air flow and temperature differential in contact with the heat exchanger 72. As costly as it may be to run two rotary pumps 76, 78 in a forced air convection system, the benefit in accelerating heat transfer has justified the expense. Traditional fans are energy inefficient; the opportunity to claim an instant energy rebate of 40% is presently wasted as free expansion whenever traditional fans are used. Fan Replacement collects the 40% guaranteed energy rebate by simply enclosing the heat exchanger 72 in a plenum 24 gated by two pumps 76, 78.
Consider a simple prior art space-heater, such as a 1000 Watt tungsten space heater equipped with a built-in 100 Watt fan. In this example, the 1000 Watt tungsten heating element corresponds to the heat exchanger. The 100 Watt fan moves a definable mass flow of air. Using principles of this invention, the same mass air flow can be moved across the tungsten filament using a pair of pumps 76, 78, consuming the same 100 Watts that would otherwise run the fan. An honest 400 Watt rebate is achieved on the Kilowatt space-heater when the principles of Fan Replacement are applied. The Kilowatt of heat costs a net 600 Watts. Of course the same price must be paid for moving the same air over the same heat exchanger. This example illustrates a simplified case of the Fan Replacement technique, in which the heating element (i.e., the heat exchanger 72) is located within a plenum 24, and the built-in blower fan is replaced with the pumps 76, 78 gating opposite ends of the plenum. Beyond the suggested repackaging of any tungsten filament space heater, Fan Replacement will harvest otherwise wasted energy from a myriad of similar devices and circumstances. Consider, for example, the notorious cost of running (cooling) computers especially in computer centers. Instead of paying twice (once for the cost of running the computer and once again for the refrigeration to cool it) Fan Replacement can cut the cost of running the computer by 40% while cooling it at the same time. The operating costs for the average Data Center are cut by 70% with Fan Replacement.
The configuration, processes, and uses of the Fan Replacement technique will next be described in relation to the heating and cooling requirements of a target space 22 in which the heat exchanger 72 is supplied by water. For heating only, water-supplied room heat exchangers have been prominent in buildings as well as in homes. The oldest configurations utilize hot water or steam for heating. Updates have transformed the old fashioned radiator into stylish baseboard units. Modern building systems integrate cooling water and heating water into the same circulated water systems. Modern building systems are supplied by cooling towers as well as boilers. In the most energy efficient of all new configurations, the year around water supply will utilize geothermal water sourcing. Because the Approaching Temperatures presented by cooling towers are so much smaller than the Approaching Temperatures presented by water heated in boilers or steam, fans will be present in all cases where cooling is to be incorporated. Fans are needed to accelerate heat transfer in cooling, due to the much smaller Approaching Temperatures supplied by either cooling towers or geothermal sources.
The potential for replacing fans in other configurations where air is blown over a heat exchanger supplied by other refrigerant types, in particular air, CO2, CFC's, HCFC's, etc., are as varied as are the other refrigerant types and the circumstances in which they are used. Different configurations, processes, and uses, can be engineered to each refrigerant type.
A first purpose of this Fan Replacement configuration, as described above, is simply to capture the work otherwise lost in free expansion. By replacing the traditional fan as the air moving device with a pair of pumps 76, 78 gating opposite ends of a plenum 24, it becomes possible to contain the heat exchanger 72 in the plenum 24 wherein the pressure may be maintained as a constant while heat is transferred to or from the moving column of air. Because any heat exchange necessarily provokes a change in the volume of the air inside the plenum 24, the very process of maintaining a relatively constant pressure (ultra-low differential) assures that the work associated with free expansion will be recovered. To reiterate, the pressure inside the plenum 24 is maintained generally constant by controlling the relative speeds of the rotary pumps 76, 78 via their respective motor/generator units 68 (
In another example, heat is being transferred into the heat exchanger 72 from the transient air column within the plenum 24, in an air-conditioning mode of operation. In this case, the volume of air inside the plenum 24 will be induced to shrink, such that the pumps 76, 78 must be controlled to maintain a generally constant static pressure inside the plenum 24 (i.e., less than 20% relative to ambient atmospheric pressure, and more preferably within the ultra-low 0-10% range). In this case, the motor/generator unit 68 associated with the second rotary pump 78 may be used to slow the rotating speed of the second rotary pump 78 (relative to the first pump 76) thereby reducing the net energy consumption required to move air through the plenum 24. The energy reduction in this case is also calculated to be about 40%.
Acceptable performance from commercially available Roots® Blowers has been validated for pressure ratios up to 1.06. At a pressure ratio between 1 and 1.2, and even more preferably between 1 and 1.1, these devices may move a mass flow of air more efficiently than common fans. At a pressure ratio between 1 and 1.2, and even more preferably between 1 and 1.1, these devices can move the same mass flow of air within the energy budget of the fans they replace and at the same time capture the energy otherwise lost through free expansion. Note that in this case the energy rebate of about 40% has been captured only in relation to the transfer of heat for which the subject HVAC system is already specifically in service to achieve. That is to say, the HVAC is being operated—at cost—to change the temperature of ambient air. Rather than neglecting the energy inherent in the free expansion of the air due to its changing temperature, the concept of Fan Replacement will supplement even established HVAC systems by harvesting a 40% energy rebate (otherwise lost to free expansion) wherever forced air convection is now used. The full value of the so-called rebate is thus captured here. Nonetheless, once the heated (or cooled) ambient air exits the Fan Replacement system, that air will return to room temperature within the target space 22 under circumstances of free expansion, i.e. without yielding work.
The advantage of replacing fans in every forced air convection application is clear, depending only on the relative offset cost of the replacement pumps 76, 78 and plenum 24 arrangement. In perhaps every configuration where traditional fans blow air over heat exchangers, those fans can be replaced to advantage using the techniques of Fan Replacement. Recognizing that well over 30% of conventional (prior art) air conditioning energy goes to moving the air through heat exchangers, it is attractive to consider the replacement of fans with pumps 76, 78 configured within a gated plenum 24 as described herein. Fans and blowers are notoriously inefficient, commonly wasting as much as 85% of applied energy. These losses result primarily from the wasteful way that fans and blowers attempt to propel air into the resistance of a static column of air. The technique of Fan Replacement capitalizes on the opportunity to reclaim ˜40% of the heat energy exchanged while moving air with well-established commercial pumps 76, 78 proven to deliver efficiency above 95% at the needed pressure ratios of between 1 and 1.2, and more preferably in the ultra-low range between 1 and 1.1.
The core concept of a plenum 24 gated on each end with a rotary pump 76, 78 used to implement the Fan Replacement configuration described above, can be further modified to improve the Approaching Temperature relative to the refrigerant. The efficiency of forced air convection depends on both the speed of air flow and the Approaching Temperature differential. The Approaching Temperature differential can be defined as the difference between the approaching air temperature and the refrigerant temperature. Fan Replacement naturally provides for speed control of the mass air flow entering the heat exchanger 72 by increasing or decreasing the rotating speeds of the first 76 and second 78 pumps. However, in a completely novel fashion the aforementioned system used to implement the Fan Replacement concept has the inherent capability to actively/intentionally alter the Approaching Temperature, thereby refrigerating the transient air flow within the plenum 24. This novel application of the core concept of a plenum 24 gated on each end with a rotary pump 76, 78 provides the mechanism and the procedure to implement an entirely new refrigeration practice which can be differentiated authentically from conventional refrigeration practices. The modification of Fan Replacement as described is necessarily the activity which Convergent Refrigeration defines to be counter-conditioning. In other words, the same apparatus may be used to replace fans by simply admitting ambient air at its unaltered Approaching Temperature, Fan Replacement, or the entering air stream may be counter conditioned, which is then Convergent Refrigeration.
All known refrigeration techniques documented in thermodynamic and HVAC industry literature are readily and consistently classed as Divergent Refrigeration. As stated above in connection with
Convergent Refrigeration uses counter-conditioning to dramatically reduce the needed refrigerant lift, raising thermodynamic efficiency to unprecedented levels in common refrigeration tasks. Counter-conditioning alters Approaching air flow Temperatures. Convergent Refrigeration mechanisms can substantially alter the economics of whatever is going on “on the other side” of the heat exchanger 72. In that sense, Convergent Refrigeration can be said to “reach through” the heat exchanger 72.
For example, in systems where the heat exchanger 72 is fed by water (for either heating or cooling), building energy professionals agree that for every degree they can reduce the energy spent changing the temperature of the refrigerant supply water they can cut the cost of delivering that refrigerant water temperature by 1.5%. In other words, for every degree of improvement in the Approaching Temperature (convergently reducing the excess refrigerant lift), the operating cost of the underlying HVAC plant is reduced by 1.5%. These are far greater cost reductions and energy efficiency gains than are delivered just by the acceleration of convection in local heat transfer. By temporarily (convergently) raising TLOW, the low temperature ambient air stream, toward its opposite working temperature, by even a degree or two, significant savings can be realized. The same relationships are commonly found when Convergent Refrigeration is used to (convergently) lower THIGH, the high temperature ambient air stream, toward its opposite working temperature. More notably, large gains can be delivered in refrigeration efficiency as measured by the COP. A single degree of counter-conditioning temperature change yields a huge change in COP. COP=273/(274−273)=273
Convergent Refrigeration increases (i.e., counter-conditions) the Approaching Temperature, simultaneously accelerating heat transfer and in some cases increasing the aforementioned energy rebate described by application of the Fan Replacement concept. And in most if not all cases, the underlying cost of improving the refrigerant supply temperature will be found to be large relative to the cost of increasing the Approaching Temperature according to these principles of Convergent Refrigeration. In other words, refrigerant lift (as seen by the underlying refrigerant supply system) may be cut with large and favorable consequences because the Approaching Temperature can be maintained within air movement costs covered by Convergent Refrigeration. In addition to delivering very large benefits overall, the economics of dramatically slashing background heating and cooling plant costs skyrocket when focus is placed on room by room heating and cooling. The optimum reductions rapidly cut total costs in half or better, especially when occupancy may be less than one or two shifts for five days out of seven rather than 24×7. The most attractive gains come from geothermal where year-around heating and cooling can be accomplished by Fan Replacement mechanisms that are also capable of delivering convergently counter-conditioned Convergent Refrigeration air flows, completely eliminating the costs of the vapor compression apparatus and refrigerants which still accompany geothermal use. Consider a geothermally supplied heat exchanger 72 in
The company Heat Pipe Technology, Inc. (HPT) provides the following formula to compute power required to drive an air flow through a heat pipe 100 like that depicted illustratively in
kW=CFM/(11674*Motor Eff*Fan Eff)
The common Roots® Blower provides exceptional efficiencies at the pressure ratios needed for Fan Replacement (as described above) and also for Convergent Refrigeration flows. Not only is volumetric efficiency exceptional at all but the lowest air flows, the compression efficiency is so well matched by expansion efficiency that the Roots® device is often selected as a vacuum pump for other applications.
As mentioned above, when using a pair of Roots® Blowers for the two rotary pumps 76, 78 operating at pressure ratios near 1.1, the efficiency of each blower or positive displacement pump is near 0.9. Thus, the efficiency of two rotary pumps 76, 78 operating in the Convergent Refrigeration context is 0.9*0.9=0.81. The Convergent Refrigeration context is more generally between about 1.2 and 1, however pressure ratios closer to 1.1 and below provide the most favorable efficiencies as can be readily confirmed by
The drawing shown as
As previously stated, all prior art vapor compression refrigeration schemes can be characterized as Divergent Refrigeration because the required excess refrigerant lift diverges from the refrigeration task. (
The large temperature change characteristic for every prior art Divergent Refrigeration system including vapor-compression systems can be diagrammed as the Brayton Cycle on a Ts diagram taking into account the required excess refrigerant lift, i.e., between Tevap and Tcond. Convergent Refrigeration, on the other hand, is performed between THIGH and TLOW. That is to say, Convergent Refrigeration can be diagrammed as a Brayton cycle on a Ts diagram operating within the confines of the refrigeration task as shown to scale in
When the expansion work can be used to directly offset the compression work, as with a turbine, the net Work that must be added from an external source is reduced by the amount of energy recaptured during expansion. The resulting COP increases exponentially as pressure ratios fall within the aforementioned range between 1.2 and 1. No such possibility exists for prior art vapor compression systems because the low pressure region is constantly under suction from the compressor. As documented in the discussion of
Anticipating a total ban on CFC and HCFC refrigerants in Europe, competitive life cycle costs for air cycle systems were certified in the late 1990s. The closed loop air cycle systems developed for trains at that time are still viable and continue to be re-adopted for Germany's most advanced bullet trains. When the expected ban on CFC/HCFC refrigerants was overwhelmed by political pressure, the wider adoption of air cycle refrigeration was blocked before the 20th century drew to a close. The less effective HCFC refrigerants, now widely mandated, still fail to provide life cycle cost competition against proven air cycle alternatives. Without question, HCFC refrigerants are more expensive than air used as a refrigerant. Of far greater consequence however, the newer refrigerants require higher pressures with resultant high rates of leaks and resultant uncontrolled maintenance discharge of harmful gases. More consequentially the new refrigerants dictate more expensive mechanical systems all delivering barely negligible increases in performance, if any at all. An unbiased review of HCFC's lesser capabilities will reveal them to be vulnerable to direct displacement in today's market by the environmentally friendly, less mechanically complex and more cost-effective air cycle refrigeration concepts described herein.
As mentioned earlier,
The high pressure ratios required by new refrigerants are easily out-performed by low pressure Convergent Air Flows. Likewise, the high pressure ratios of closed loop air cycle refrigeration can be out-performed by the much lower pressure open-loop air cycle principles of this invention. Once it is recognized that present refrigeration systems already expend the energy needed to move the entire mass flow of air required for refrigeration and they already move the needed mass flow of air without exception necessarily on both sides of each and every single vapor compression refrigeration loop, there can be no reasonable argument against using air as the refrigerant, certainly no argument based on the heat capacity of air. That said, there is no justification for retaining the vapor compression refrigerant loop. In prior art configurations, fans and blowers move the needed mass flow of air on both sides (Zone 1/TLOW/Heat Source and Zone 2/THIGH/Heat Sink) of the refrigeration paradigm. However, as has been demonstrated, fans and blowers of the prior art move all the needed air expensively, inefficiently, wastefully, without energy recovery and ignoring free expansion. Not only is the air that is already being moved through existing HVACR systems sufficient to refrigerate all the ambient air, that same air can be moved for a lower cost and refrigerated at the same time within the Convergent Refrigeration mechanisms described herein. In companionship with embodiments of the more basic Fan Replacement mechanisms, this set of Convergent Refrigeration tools will fundamentally disrupt all prior understandings of practical refrigeration.
When the Expansion Work is subtracted from the Compression Work, the COP as traced in
As a mnemonic device it is convenient to anchor Convergent Refrigeration performance in
Compared to the COP of 3.93 NIST reported (above) for cooling only an 8.3° C. (15° F.) refrigeration task at the 95° F. Rating Point, a compelling case for disruptive technology can be made. Convergent Refrigeration therefore has the potential to usher in an entirely new order of energy efficiency within the HVACR industry.
As described in the Background section, ASHRAE has raised the standard for “room temperature” from 23° C. to 27° C. This allows the increase of evaporator temperature from 3° C. to 7° C. while maintaining the desired approach Air to Refrigerant Temperature Differential of 20° C. This artifice increases human discomfort while allowing the manufacturers to claim substantial improvements in performance By claiming that customers now suddenly tolerate the ASHRAE-stated higher room temperature, the manufacturers cut excess refrigerant lift to advertise increased performance But conventional wisdom suggests that the average person is ignorant of the manufacturer's surreptitious specification changes, and simply turns their thermostat down to a comfortable lower temperature thus negating the manufacturer's claimed efficiency improvements. The point is that the industry's efficiency claims are dubious. But a 1-Sided Convergent Refrigeration flow device like that depicted in
Another Convergent Refrigeration flow can be grafted onto the condenser to deliver 2-Sided Convergent Refrigeration flow, like that schematically illustrated in
Thus, not only can the expansion work of cooling be used to directly offset the compression work of heating, the energy spent creating excess refrigerant lift as well as temperature overshoot can be essentially eliminated.
The right side of
The engineering specifications of a heat pipe 100 type of heat exchanger 72 (
The optimum “end to end” temperature differential for a heat pipe 100 may be as low as 9° C. This is the total Approaching Temperature needed to secure heat transfer from one end of the heat pipe 100 to the other. At the same time, the cost of running the (prior art) compressor is eliminated altogether and the total refrigerant lift (20° C.+12° C.+20° C.=52° C., COP=5.3) needed to transfer heat on both sides of the working temperatures is reduced. Instead the ambient air temperature is moved only 4.5° C. beyond the midpoint between the two working temperatures. (6° C.+4.5° C.=10.5° C., COP=28.5) The ambient air temperature is moved twice in this example, but the COP is nonetheless dramatically reduced. The work on both sides is fully recognized in the embodiments detailed later.
The heat pipe 100 uses the energy of the heat to be moved to move the heat without any added cost of work. But more relevant to its speedy adoption, the heat pipe 100 can be tailored to match exactly the physical dimensions of a vapor compression fin-and-tube heat exchanger that it might replace. There is no cost for running the compressor and the refrigerants are inexpensive and benign. The heat pipe 100 directly replaces the (prior art) vapor compression loop while counter-conditioned convergent air flows will deliver exactly the same mass flow of environmental air for cooling and heating at common temperatures for less than the cost of running only the fans in a traditional vapor compression system. Thus, utilizing heat pipes 100 in combination with the heat exchanger 72 in a back-to-back arrangement like that shown in
Because the physical implementation of counter-conditioned convergent air flows invites a wide variety of physical dimensions and engineering interpretations, the simple schematic of two Convergent Refrigeration flows arranged in back-to-back relationship is presented in
For consistency in the schematics which comprise
The common heat exchanger 272/372 shown in
Commercial air-to-air heat exchangers 72 of this heat pipe 100 class use typical refrigerants like R134a circulating through the same fin-and-tube heat exchangers 72 employed by vapor compression systems. Such two phase refrigerants may even be pumped at very low cost while in the liquid phase. Not dependent on gravity, heat pipes 100 overcome limitations of elevation and distance. The direction of flow may be reversed easily to change over from Air Conditioning to Heat Pump operation, meeting day-night and/or seasonal demands Their boiling points may be controlled with specific pressure regulation, exactly as in vapor compression systems. But a crucial performance distinction for heat pipes 100 remains that heat is acquired at a higher temperature source and rejected into a lower temperature sink. No external energy is required to compress the vapor so that it will condense at a higher temperature. As graphically depicted in
In the summer, for example, the warmer outside air is made cooler between the pumps 276, 278 surrounding the heat exchanger 272 while the cooler inside air is made warmer. Heat will naturally migrate into the outside counter-conditioned convergent air flow through any air-to-air heat exchanger 272, which may be a heat pipe 100 or any other suitable device. Reversing these relationships transforms the system from an air conditioner into a heat pump, moving heat from the colder outside air into the building in winter. Just as the relative pump speeds will be tuned for best efficiency as inside and outside temperature and humidity changes, the boiling point of the heat pump working fluid may be moved to the optimum temperature between counter-conditioned convergent air flows to follow both the size and the direction of the refrigeration task, reversing the direction of vapor and liquid flows to meet seasonal or even daily needs.
The heat demands of very cold temperatures have been addressed and satisfied by configurations like that of
It is contemplated that the outside 224 and inside 324 plenums will represent permanent ducting that remains fixed in place while the changeover from air conditioning to heating seasonal needs is delivered simply by changing relative pump or turbine speeds. That is to say, the transition from the inside space being Zone 1 (Heat Source) in the summer to Zone 2 (Heat Sink) in the winter may be accomplished without physical relocation of the outside 224 and inside 324 plenums. In this manner, daytime cooling is readily complimented with heating on cold nights.
Depending on proximity and climate variables, the driving pumps 276/378 and 278/376 may optionally share a common shaft. That is to say, in some contemplated configurations, inlet pump 276 is mechanically coupled with outlet pump 378. And likewise, inlet pump 376 and outlet pump 278 are mechanically coupled through a common drive shaft or other power transmission device. More typically, however, each pump will be separately powered and precisely controlled using DC motor-generators, like those depicted schematically at 68 in
Throughout
For illustration, the temperature values shown in examples which follow have been taken from the commercially available engineering statements of Heat Pipe Technology, Inc. (HPT). Often demonstrating greater heat flux, the heat pipe 100 type of heat exchanger 72 can deliver temperature changes often exceeding 90% of the approach compared to 60% with prior art vapor compression. Not only will typical heat transfers be substantially higher with the same mass air flow, the total heat content will be greater because 1) the inside air flow is always “non-condensing” and 2) condensation in the outside flow will rarely occur due to significantly narrower A-RTD. In fact, there is considerable latitude to avoid condensation in the outside air stream altogether by simply moving the heat transfer temperature above the dew point of the outside air. The temperature of the inside air stream can be counter-conditioned to compensate accordingly. With the Sensible Heat Ratios of present (prior art) HVAC air conditioners running from 65% to 80%, latent heat losses due to cold water running down the drain amount to 0.30 Kw/ton. Except for the dehumidification of make-up air, this charge will be entirely avoidable in a Convergent Refrigeration system. Condensation in the outside air stream is totally avoidable, as is condensation in the inside air stream after accounting for the dehumidification of make-up air. This capability further improves the efficiency gained by evaporative cooling in the outside air stream. In fact, due to the high latent heat of water, it is certain that the best Convergent Refrigeration performance will be obtained by saturating the outer air stream to a dew point just above its cooler counter-conditioned target temperature.
This configuration reduces ˜90% of the acknowledged vapor compression energy cost. Convergent Refrigeration flows eliminate the vapor compression system altogether. Of course the Convergent Refrigeration energy budget would include the previously unreported cost of moving air through the inside heat exchanger 72. Even including these additional energy consumption parameters, however, the entire cost of refrigeration using two counter-conditioned convergent air flows back-to-back sharing a common heat exchanger 272 may fall below what the prior art would have incurred just to move the mass flows of air using fans or blowers.
As previously shown for temperatures at and above this rating point, the only usable portion of the R410A vapor compression cycle is vapor, not latent heat. And the energy needed to raise the vapor pressure to ratios of 4 and above causes extreme temperature overshoot. Vapor compression may have benefited from temperature overshoot by accelerating heat transfer, but temperature overshoot can be eliminated altogether by sustaining a precisely tuned Approaching Temperature. Accordingly, Convergent Refrigeration may be delivered within the energy budget previously required just for moving air.
Rather than use the above-mentioned 20:20:20 rule with both flows in the simple back-to-back air conditioning illustration of
As previously stated, the temperature relationships are chosen purposefully to meet the requirements of the 95 F Rating Point under the heat movement measurements published by HPT. The counter-conditioned convergent air flows here follow behaviors incidental to the choices made by HPT rather than the optimized values readily preferred in a working system. The stated values have also been validated computationally. These HPT numbers provide commercial certification of temperature relationships and deliverable technology capable of displacing vapor compression with counter-conditioned convergent air flows. Their physical dimensions provide a plug and play replacement for vapor compression heat exchangers 72 used around the world and their track record of performance and reliability is acknowledged by ASHRAE to be “second to none”!
As stated above, at any given time in a system utilizing two back-to-back counter-conditioned convergent air flows sharing a common heat exchanger 100, one half or sub-system operates in heat pump mode (the supplier of heat, the heat source) while its partner operates as the heat sink. The air conditioning example shown in
Returning again to
It can be seen, therefore, that heating (
Both inside and outside air flows are refrigerated by the same temperature change, 35.55° F.=19.75° C., somewhat less than needed to fit the 20:20:20 rule introduced above. It is noteworthy that refrigeration can be delivered under these extreme circumstances by increasing the pressure ratio to only 1.25 from the 1.15 needed at the 95° F. Rating point described in
Performance will be increased by provisions for dehumidification, make-up air, and exhaust when compared to the standard operating mode of Convergent Refrigeration. These three new capabilities detailed below far exceed the best possibilities of vapor compression alternatives.
In
In
The process for providing and dehumidifying make-up air is understood and adequately documented in the engineering of wrap-around heat pipes 100 by HPT. Although it is not detailed here, the effusive endorsement of heat pipes by ASHRAE was previously noted. The anticipated blending of outside makeup air to be dehumidified, as indicated by the direction arrow containing the “?” symbol in the upper left corner, will increase energy use. The heat exchanger 272 target temperature of 51.35° F. is below the best evaporator inlet temperatures recorded by NIST in the Domanski and Payne (2002) study previously mentioned. Clearly this target temperature meets the ASHRAE specifications for testing at the 95° F. Rating Point. Cooling work must be done in the upper path sufficient to assure that the target temperature chosen for the desired exit humidity level has been met. Because no external heat rejection occurs in the process as depicted, heat will accumulate from the latent heat of condensation.
In summary, Convergent Refrigeration (also referred to herein as counter-conditioned convergent air flow) provides an entirely new set of mechanisms and methods for minimizing heat transfer in refrigeration, delivering unprecedented high COPs with unprecedented low pressure air cycle refrigeration. In both cooling and heating applications, Convergent Refrigeration replaces the energy intensive and environmentally harmful vapor compression technology of the 20th Century with a clean, low-cost alternative. The prior art's performance mnemonic 4:400:8 becomes the new and substantially more attractive mnemonic 20:20:20 (COP:PR:ΔT)
By incorporating proven passive heat pipe 100 technology, Convergent Refrigeration uses as its refrigerant exactly the same mass flow of air required by vapor compression technology. Of the most profound importance to certify the feasibility of counter-conditioned convergent air flows, vapor compression systems demand much more than just the same mass air flow. The necessary heat capacity of circulated air has been demonstrated by vapor compression systems to provide adequate mass flow to hold and move requisite heat to and from the source (Zone 1) to the sink (Zone 2). Convergent Refrigeration uses the same mass flow of air as circulated by prior art vapor compression systems, but uses that air as its refrigerant. By moving the air across the heat exchanger 272, 372 within a plenum 224, 234 that is gated at each end with a rotary pump 276, 278, 376, 378, the air can be transformed for use as a refrigerant and thereby accomplish the purposes of this invention.
The foregoing invention has been described in accordance with the relevant legal standards, thus the description is exemplary rather than limiting in nature. Variations and modifications to the disclosed embodiment may become apparent to those skilled in the art and fall within the scope of the invention.
This application claims priority to Provisional Patent Application No. 61/256,559 filed Oct. 30, 2009.
| Filing Document | Filing Date | Country | Kind |
|---|---|---|---|
| PCT/US2016/047778 | 8/19/2016 | WO | 00 |
| Publishing Document | Publishing Date | Country | Kind |
|---|---|---|---|
| WO2017/031428 | 2/23/2017 | WO | A |
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