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Publication numberUS3686893 A
Publication typeGrant
Publication dateAug 29, 1972
Filing dateDec 22, 1969
Priority dateDec 22, 1969
Publication numberUS 3686893 A, US 3686893A, US-A-3686893, US3686893 A, US3686893A
InventorsEdwards Thomas C
Original AssigneePurdue Research Foundation
Export CitationBiBTeX, EndNote, RefMan
External Links: USPTO, USPTO Assignment, Espacenet
Air refrigeration device
US 3686893 A
Abstract
A single fluid (air to air) refrigeration device which includes compressor and expander sections with a single heat exchanger disposed therebetween employing an open reversed Brayton cycle in operation.
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Claims  available in
Description  (OCR text may contain errors)

United States Patent Edwards Aug. 29, 1972 [22] Filed:

[54] AIR REFRIGERATION DEVICE [72] Inventor: Thomas C. Edwards,

Lafayette, Ind.

[73] Assignee: Purdue Research Foundation Dec. 22, 1969 [21] Appl. No.: 886,979

West

[52] U.S. Cl. ..62/402, 417/348, 417/349,

[51] Int. Cl ..F25d 9/00 [58] Field of Search ..62/402; 418/86, 259, 266, 267; 417/348, 349, 392, 393, 405, 406

[56] References Cited UNITED STATES PATENTS 2,733,663 2/ 1956 Marshall ..417/406 2,924,178 2/1960 Hogan ..417/348 3,301,471 1/1967 Clarke ..417/392 1,726,462 8/1929 Wittig ..4 1 8/ 86 2,073,833 3/ 1937 De Bothezat ..62/402 2,754,660 6/ 1956 Morrison ..62/402 FOREIGN PATENTS OR APPLICATIONS 554,525 3/ 1923 France...... ..62/402 Primary Examiner-William J. Wye

Attomey-John R. Nesbitt [57] ABSTRACT A single fluid (air to air) refrigeration device which includes compressor and expander sections with a single heat exchanger disposed therebetween employing an open reversed Brayton cycle in operation.

11 Claims, 14 Drawing Figures This invention relates to an improved air conditioning system, particularly suited for medium capacity cooling requirements, such as in automobile air conditioning.

The air conditioning systems currently in use in automobiles employ a two phase refrigeration system. The components are complicated and also expensive. For example, the vapor-liquid compressor in the conventional system comes in several variations (swashplate, slider-crank), all of which contain many close tolerance parts and require high pressure seals against contamination and refrigerant leakage. Such systems also require two-liquid-vapor-to-air heat exchangers with similar problems. The conventional systems also require an expansion valve and a number of high pressure refrigerant lines and suitable fittings. Independent means must be provided on the evaporator side to conduct the air from the refrigerated space surrounding the evaporator coils to the area to be cooled.

Single fluid refrigeration devices are not new. For example, the so-called dense air cycle (a closed reversed Brayton cycle) has been known for years and was formerly used mainly with large reciprocating piston machinery. This system has long been abandoned for commercial purposes because of the large size required to provide even a moderate amount of cooling.

SUMMARY It is one object of this invention to provide a single fluid refrigeration device of compact. size to fulfill moderate cooling requirements as they occur for automotive uses, as well as for other applications.

Another object is to provide a compact single fluid refrigeration device of relatively inexpensive construction compared to conventional two phase systems now in use.

Another object of this invention is to provide means for use in any rotary pump or compressor to increase effective vane travel to make possible increased fluid handling capacities.

FIG. 1 is an exploded perspective view of one embodiment of the present invention.

FIG. 2 is a cross-sectional view through the rotor and stator of the preferred embodiment of the present invention.

, FIG. 3 is a graphic representation of the P-V curve of the preferred embodiment shown in FIG. 2.

FIG. 4 is a graphic representation of the T4 curve of the preferred embodiment shown in FIG. 2.

FIG. 5 is a diagrammatic view through the rotor and stator of another embodiment of the present invention diagrammatically showing a movable heat exchanger outlet port.

FIG. 6 is a diagrammatic view through the rotor and stator of another embodiment of the present invention including an adjustable slider means inside the rotor core to supply cooled fluid to the expansion side of the device.

FIG. 7 is a diagrammatic view through the rotor and stator of another embodiment of the present invention wherein the rotor axis is eccentrically shifted as another means of compensating the VHXE volume.

FIG. 8 is a diagrammatic view through the rotor and stator of another embodiment of the present invention with the rotor being located inside a nonsymmetrical stator, whereby still another means of VHXE compensation is provided.

FIG. 9 is a graphic representation of the P-V curve of a non-VI-IXE compensated design.

FIG. 10 is a graphic representation of the T-S curve of a non-VHXE compensated design.

FIG. 1 1 is a diagrammatic view through the rotor and stator of the preferred embodiment of the present invention and associated intake and exhaust ducts through the rotor and stator indicating the travel of a I given air mass through 45 of rotation of the rotor.

FIG. 14 is a cross-sectional view showing a movable exhaust port.

OPERATION OF THE DEVICE FIG. 1 shows an exploded view of the basic device of this invention and its related components. The stator 1 of the device is machined such that the interior walls are closely parallel. The cross-section of the interior of the stator is substantially elliptical in shape. The stator 1 is constructed with a series of ports which are specially located for the purpose of allowing air to flow through them at various times in the proper sequence. The rotor 4 is equipped with a series of slots 5 disposed axially on the rotor and which receive vanes 6. The slots are placed symmetrically around the circumference on the rotor, and into these slots fit the sliding vanes 6. These vanes are maintained in contact during rotation of the rotor by a pressure acting outwardly from the core 7 of the rotor. This pressure may be established by means of an external pressure source (a common method in the art), by an internal air cushion means described in US Pat. No. 3,434,655, or by means of an outwardly expansible elastic band 101 as shown in FIG. 11. Band 101 is thought to be novel and will be described in detail hereafter. Not shown in FIG. I are the two endplates which are equipped with bearings and shaft seals and said plates firmly hold the rotor-vane assembly in place inside the stator in a conventional manner. The finned tubular heat exchanger 8 is connected directly (or through a simple manifold which is not shown) to the heat exchanger inlet 9 and outlet 10 ports which are machined into the stator.

The air duct 1 1 shown schematically in FIG. 1, serves to separate the warm incoming air indicated by arrow marked 12, from the cooled air indicated by arrow marked 13 passing to the space being cooled. As shown I f in FIG. 11, the outlet leg 102 of the air duct may be 15. As the rotor turns a total of 360/2, where Z is the number of vanes, (45 for the 8-vane machine shown),

from VNLT to Volume of Compression, denoted VOC a maximum volume is reached. The trailing edge of the vane volume segment VOC defined by vane 14 then just passes the last inlet port 3.

During another rotation of 45, the vane volume segment is compressed to the Heat Exchanger Volume, Compression side (denoted VHXC). At this point the leading vane 16 that defines VHXC passes the heat exchanger inlet port 9 and the heated and compressed volume of air is pumped out of VI-IXC and into the heat exchanger 8 as the rotor continues to rotate another 45. Some of the compressed air is carried over through the Clearance Volume (denoted V9 C) located between the inlet ports 9 and outlet ports of the heat exchanger. The heat of the compressed air is then partially rejected as it is pumped through the heat exchanger.

During the compression process on the left (as viewed in FIG. 2), the vane volume segment on the right now defined by vane 17 and vane 18 and called the Heat Exchanger Volume Expansion side (denoted VHXE), has been accepting the relatively cooled air from the heat exchanger. The cooled air contained in VI-IXE is then expanded and thereby greatly cooled as VHXE is expanded during the further 45 rotation of the vane volume segment to the position called the Volume of Expansion (denoted VOE) and defined by leading vane 19 and trailing vane 18. This expansion process recovers a large fraction of the initial work of compression that occurred on the left side of the device.

The next 90 of rotor rotation moves the volume segment under consideration from VOE to the Volume of Exhaust (denoted V9OE) and this forces the cooled air out through exhaust ports 20 into the space being cooled.

One full revolution of the rotor 4 takes a charge of air through a complete open cooling cycle, as indicated by path a-b-c-d-e-f-g-a in FIGS. 3 and 4. During one 360 revolution the deviceperforms the following functions:

1. draws in ambient air,

2. compresses and heats the air,

3. pumps the hot air through the heat exchanger for the purpose of cooling it,

4. expands the relatively high pressure cooler air to atmospheric pressure, lowering its temperature considerably and recovering compression work, and

5. pushes and circulates the cooled air into the space being cooled.

The foregoing description illustrates on a qualitative basis the advantages of this new and novel system. It can be demonstrated on an analytic basis that the device is capable of supplying substantial quantities of cooling power in small sizes. The following analysis, which has been substantiated by experiment, analytically proves the utility of one embodiment of the device described.

THEORETICAL APPRAISAL OF DEVICE PERFORMANCE The following analysis contains these physical assumptions and restrictions:

l Air behaves as in ideal gas,

2. Air flows are considered as essentially steady,

3. Minor pressure losses and mechanical friction are neglected,

4. Kinetic energy changes are negligible.

(Note: While this analysis considers air as the working fluid, it is not limited thereto.)

The performance (cooling capacity and coefficient of performance) of a given refrigeration device as described can be predicted by a system analysis based on the first law of thermodynamics. The steady-state thermodynamics of the device depends primarily on the contour of the inside of the stator 1 with respect to the rotor axis 21, the physical location of the various ports, and the number of vanes used. Other parameters such as ambient conditions are also influential.

The cooling capacity is defined here as the product of the total specific enthalpy change achieved between the inlet and outlet air streams passing through the device and the mass flow rate of the air stream. That is,

QCAP=m i o) (1) where QCAP is the cooling capacity given on a rate basis, n'i is the mass flow rate of air through the system, T, is the incoming air stream temperature, and T is the outlet air stream temperature. C is the constant pressure specific heat of the air.

Assuming the inlet air temperature is known, the problem is the calculation of the outlet temperature,

T and the air mass flow rate, m. The mass flow rate can be deduced in terms of ambient inlet conditions, maximum inlet volume, and operating speed. The outlet temperature, T,,, can be found by considering the changes of state through which the air occupying some representative volume segment passes as the rotor rotates through a typical complete cycle after steady operation has been reached. The first process to change the state of the inlet air is the compression process occurring as the maximum inlet volume, VOC, diminishes to the compression side heat exchanger inlet volume, VHXC, as the rotor rotates. The temperature of the air occupying this volume can be computed from the reversible polytropic relations as THXC=TNLT (VOC/VHXCY'" 2 where THXC is the temperature of the air in Vl-IXC, and n is the polytropic index of the process.

Immediately after VI-IXC is reached, as seen in FIG. 12, the leading vane passes the heat exchanger inlet port 9. What happens now to the mass of air 1 l2 occupying VI-IXC depends upon what has been happening to the previous generation of air charges as they passed through the heat exchanger and on to Vl-IXE, the expansion side intake volume, as shown in FIG. 13. If the volume VHXE is reduced by just the precise amount to compensate for the decreased specific volume of the air on the expansion side of the device (because of the cooling of the air as it passes through the heat exchanger), a substantially constant-pressure process will occur as Vl-IXC rotates first to V9OC and then over to VHXE. FIGS. 12 and 13 show these two successive positions of the rotor-vane assembly as the rotor rotates through 45 (as it would for an eight vane configuration). Since V9OC is very small compared with VHXC and Vl-IXE, the finite width of the heat exchanger inlet ports 9 and outlet ports has little influence upon the analysis because of the small amount of mass contained in V9OC. It is important to note, however, that the heat exchanger inlet and outlet ports are never in direct communication through V9OC, as shown in FIGS. 12 and 13, because hot air could shortcircuit the heat exchanger through this passage.

Upon rotating another 45 after the inlet process some of the mass originally occupying vane segment VHXC passed into the heat exchanger, and some of the mass in the heat exchanger passed into V9OC as it expanded to VHXE. Thus a closed system process is achieved. Applying the first law of thermodynamics to this process,

a b a b+( b a) where a is the amount of heat transferred from the system during the process from a to b and a W,, is the amount of work done by the system during the process. U, and U are internal energies of the air in the system.

Assuming that the heat transfer takes place primarily in the heat exchanger, we can write, from the steadystate form of the first law (neglecting potential and kinetic energies) "Q11 (mass through HX) C, (THXI THXO) VHXE J V900 POW vaxo pdV+ where p PHXC. Since it was assumed at the outset that the pumping and heat transfer processes occurred at substantially constant pressure, the work term is reduced to W, PHXC (VI-[XE VHXC) 5 Finally, the internal energy terms, U, and U can be written as where C, is the constant volume specific heat of the air, MHXT is the total mass in heat exchanger and THX is the bulk temperature of the air in the heat exchanger.

The substitution of Equations (4), (5), and (6) into (3 yields C (MHXC M9OC) (THXO THXI) PHXC (VHXE VHXC) (THXC) Under steady-state conditions, no mass will accumulate or deplete from the heat exchanger. That is, MHXE will be equal to MHXC. Also, it is quite reasonable to assume that THXI is equal to THXC. So Equation (7) becomes, upon solving for THXE,

THXE THXC 1.0 M9OCIMHXC) (THXO THXC) 8 Since the process has been assumed to occur at constant pressure, and by using the definition of an ideal gas, Equation (8) can be simplified further to THXE =THXC +(1.0 VQOCIVHXC) (THXO THXC) 9 Equations (8) and (9) are dimensionally homogeneous, and can be further checked by considering two known limiting conditions:

1. When the carry-over volume, V9OC, is zero,

Equation (9) gives:

, THXE=THXO,

which is the expected result because the temperature of the air passing into VHXE would be only a function of the heat exchanger outlet tempera ture, THXO. 2. When THXO is equal to THXC (no heat transfer),

Equation (9) yields:

THXE THXC,

TOLT THXE (VHXE VATMY" 10 where VATM is the correct volume to which the original inlet air mass volume segment (VOC) must be expanded to just reach atmospheric pressure, and thereby to minimize exhaust port noise.

Now the air mass flow rate, m, is derived as m (inlet air density).(Maximum volume flow rate) p, VOC 1 1 where p v lRT from the ideal gas law (where R is the gas constant), and VOC NZ(VOC V9OE). Here N is the operating speed (rev/min) and Z is the number of vane segments. V9OE is the minimum exhaust carryover volume. Thus Equation 10) can be written as Cooling efieot QCAP Work Required WORK Here the work required, WORK, is considered to be the p-v work of the machine, neglecting friction. So we can write WORKC (n/(n-l ))(PNLT) (VOC) (l (VOC/VHXCY") WORKE= (n/(n-l ))(PNLT)(VOE) (l VOE/VHXEY") 14 Thus, all the relations required to predict the operating characteristics of a given cooled-volume (VHXE) compensated device are now available. These relations have been programmed on a digital computer and the results of one example run are presented in Table I.

TABLE I Computer Results VHXE-Compensated Refrigeration Device Rotor Diameter: 4.92

From Table I, it is seen that quite a small unit can produce substantial theoretical cooling effects while operating at a modest speed. A ten vane model has been built and tested and these theoretical results are closely approximated in actual performance of the device, excluding the effects of mechanical friction.

At this point is must be re-emphasized that, in the foregoing analysis, it was assumed that there was no pressure drop from the compression side at VHXC to the expansion side at VHXE. That is, VHXE was adjusted to make up for the contraction of the relatively cooler gas on the expansion side. In order to obtain maximum performance from the device described it is imperative that the actual pressure drop be held to a minimum.

The amount of expansion work that is recoverable depends directly upon the pressure from which expansion begins (see Equation (13)). If PHXC is decreased for any reason from the value obtained on the compression side, something less than the maximum work recovery will be achieved and the minimum low temperature achievable (TOLT) will be increased, thereby directly limiting the cooling capacity QCAP, as shown by Equations and (l). The next portion of this analysis presents in a quantitative manner, the effects of not properly compensating for the specific volume decrease occurring on the expansion side.

Theoretical Appraisal of Device Performance Without VHXE Compensation Consider for simplicity a symmetrical stator interior with the axis of the rotor coinciding directly with the central axis of the stator, with the inlet heat exchanger port located symmetrically opposite the outlet heat exchanger port (not shown in the drawings).

To begin this analysis, consider the conservation of mass: At steady-state operating conditions, the mean amount of mass occupying the heat exchanger will not change. Therefore, the amount of mass on the compression side of the device must be equal to the amount of mass on the expansion side. This can be stated as MHXC MHXE, and in terms of a perfect gas, we have (PHXC') (VHXC) (PHXE) (VHXE) R(THXC') R(THXE) Solving for PHXE yields (VHXC) (THXE) E? (PHXC) (VHXE) (THXC) For the symmetrical geometry under consideration, VHXC VHXE, for Equation 14) reduces to PHXE PHXC (THXE/THXC) (15) So as a result of cooling, THXE will be less than THXC, and therefore, PHXE will be depressed with respect to PHXC. Since this pressure is lowered, the amount of available work of expansion is decreased (as discussed earlier), reducing the total capacity and the coefficient of performance.

In order to predict the adverse effect on performance resulting from this pressure loss, apply the First Law of Thermodynamics again.

Equations (1), (2), (3), and (6) still hold from the first law. However, the work term W is different. Assume that the pressure drop across the heat exchanger is negligible. Therefore, the pressure inside of the heat exchanger must be very nearly equal to the pressure inside VHXE, which is PHXE. However, just at the moment the leading vane of VHXC passes the heat exchanger inlet port, a surge of air will flow from VHXC to the heat exchanger thereby rapidly equalizing the pressure in the three volumes VHXC, V9OC, and VHXE, and also the heat exchanger. The pressure equalization takes place in a very short time compared with the time required for the volume VHXC to diminish to V9OC. Therefore, a substantially constant pressure pumping process from the compression side of the device to the expansion side will take place, but at a lower pressure, PHXE.

For this case the work ,w, is reduced to zero because the volumes VHXC and VHXE are equal and the process occurs at constant pressure. Therefore, Equation (5) of the previous analysis becomes C, (MHXCM9OC) (THXO-Tl-IXT) C, (MHXC) (THXE)-C,. (MHXC) (THXC) 5% Note here that THXI (l-lX inlet temperature) is used because the temperature of the air entering the heat exchanger will be somewhat different from THXC because of the expansion from PHXC to PHXE. Simplifying Equation (5 and solving for THXE yields THXE THXC k( lgoc/ xc (THXO-TI-IXI) (6) 31 m Unfortunately, Equation (6') cannot be written directly in terms of the volumes as was done previously in Equation (7) because MHXC and M9OC both depend on the temperatures and pressures in the respective volumes.

the same size unit as in previous work (See Table I for comparison).

TABLE II Computer Results Non-VHXE Compensated Refrigeration Device Physical Geometry (size) Performance Length: 6" Major Diameter: 6" QCAP: 3612 BTU/hr Ellipse Angle: 35 COP: L24 Rotor Diameter: 4.92

Not only is the cooling capacity 20 percent less than the VHXE compensated device, but the COP has been cut in half. (Although it will require more than twice as much power to operate the non-VHXE compensated device, a COP of 1.24 is considerably better than that achieved by auto air conditioning units now in service.)

FIGS. 9 and 10 are the P-V and T-S diagrams, respectively, of the noncompensated design, and FIGS. 3 and 4, respectively, are the P-V and T-S diagrams of a VHXE compensated device. The efiects of the pressure loss can be seen by comparison of these figures. In FIG. 9, for instance, the amount of recovered work of expansion is obviously less in the case when heat exchanger pressure loss arises from lack of VI-IXE compensation. Likewise, in FIG. 10, the amount of heat transferred is notably less in the case of the noncompensated model.

Now that the importance of VHXE compensation has been demonstrated, it remains to show how this compensation is accomplished in the actual device.

METHODS OF VHXE COMPENSATION The present invention may be realized by any one of several methods by which expansion-side volume compensation is achieved.

In practice the only VHXE compensation really necessary to provide a practical and highly efiicient device using the construction previously described and shown in FIGS. 1 and 2 is to locate the heat exchanger outlet ports 10 at an average design point while operating upon air at normal ambient temperature conditions. However, more precise methods of achieving such VHXE compensation may also be employed.

One such method is to reduce the major axis 120 of the expansion side of the stator by the correct amount to make VHXE take on the correct value. FIG. 8 depicts this approach.

Another method of compensating for the increased density of the air after passing through the heat exchanger is to shift the axis of the rotor from the center of the stator to a position 121 which will diminish VHXE by the required amount. This approach is shown in FIG. 7.

It can be seen that the VHXE compensation methods shown in FIGS. 7 and 8 are simple to achieve, but they are not easily made to be continuously variable, and therefore flexible to adjust for varying ambient inlet conditions (TNLT, PNLT, etc.). Therefore, some single mean design value of rotor axis offset or expansion side major axis decrement must be chosen which offers the best over-all performance. It is conceivable that the device could be built such that the rotor axis offset and/or expansion side major axis decrement could be continuously variable.

Another method which achieves VHXE compensation, on a continuously variable basis, with simple control characteristics is depicted in FIG. 6. Using this method, air passes from the heat exchanger inlet ports 9 through the heat exchanger, and then into a passage 123 which leads to a port in communication with a hollow chamber 124 inside the rotor. The air then passes through aslit 126 defined by two thin-walled semicylinders, movable semicylinde=r 127 and stationary semi-cylinder 128, and finally through the radial slots 129 in the rotor between each vane segment to VHXE. One of these thin-walled mating Ihalf-cylinders is capable of sliding circumferentially with respect to the other, thereby controlling the azimuthal angle through which the radial rotor slots 129 communicate with the air passing inside the rotor. Therefore, the size of VHXE is effectively controlled.

This method of VHXE compensation can be controlled by a simple calibrated bellows-lever arrangement (not shown) which senses any pressure change inside the heat exchanger and shifts semi-cylinder I27 accordingly.

As: pressure would tend to build up in the heat exchanger due to less mass leaving the heat exchanger than entering it, the sensing bellows would expand slightly. This would cause a lever (not shown) to rotate movable semi-cylinder 127 counter-clockwise to increase VHXE, thereby accepting more mass from the heat exchanger and relieving the pressure build-up. Obviously the same control process would function in a reverse fashion as changing ambient conditions caused a pressure drop in the heat exchanger. This construction continuously compensates for changing ambient conditions (as well as changing operating speeds) and permits maintaining maximum performance over a wide range of operating conditions.

Another method by which this continuously variable VHXE compensation can be achieved is depicted in FIG. 5. In this version of the device, the heat exchanger outlet port 200 is constructed such that it can slide back and forth, thereby compensating VHXE for density differences and maintaining maximum performance. The outlet ports 200 can be constructed either on the end-plates (not shown) or on the stator wall 201, in which case a substantially flat section 202 is machined on the interior of the stator wall in the vicinity of the heat exchanger outlet port to facilitate easy sliding motion and sealing. This sliding relocatable port can be controlled by a simple bellows arrangement 203 (or by any other means known by the state of the art) much like that discussed previously. The same result may be obtained in the same manner as described above as shown in FIG. 6 by making the heat exchanger inlet port 198 movable or stator wall flat portion 196. Actual cooling capacity can be changed in this manner because the compression ratio is changed by moving inlet port 198.

Still another method of obtaining VHXE compensation can be made so that not only is VHXE adjusted in order to compensate for air density increases on the expansion side due to cooling, but the final outlet volume, VATM, can also be adjusted to compensate for this density increase. Adjusting VOE to the correct value is very desirable because exhaust port noise can be minimized thereby. That is, if the final outlet volume,

VATM, is adjusted so that if at just the moment the leading vane of the final outlet vane chamber passes onto the final outlet ports, the pressure of the air contained within this chamber has just reached the value equal to the surrounding pressure, no severe acoustic disturbances will occur.

FIG. 14 shows this combination of continuously variable VHXE and VOE compensation. Here it is seen that the final outlet volume, VOE, is controlled by the location of the slidable final outlet port wall 205 which can be controlled by any means known by the state of the art. Of course it is quite possible to choose some average design value for the permanent location of the final outlet port wall.

Because the refrigeration device of this invention is easily capable of cooling 80 F. intake air to -30 F the water vapor in the humid intake air will be condensed or frozen during cooling. Some means must be pro vided to get rid of the cold water (ice) particles in the exhaust air or an icing condition will be created in the cooled space into which this air is introduced. FIG. 11 shows an intake duct 98 and an adjacent outlet duct 102. This is a desirable arrangement for several reasons that are described below.

An air cleaner 96 is placed in the inlet duct to keep impurities out of the system. Baffles 103 are disposed in the outlet duct causing the air (indicated by arrows) to follow a serpentine path whereby the water (or ice) particles are physically thrown outward against the duct walls 94 which in turn catch such particles so that they may be channeled through drain holes 92 into a catching trough 90. The cold moisture particles thus collected may be drained off to drain hole 88 or, if needed, may be circulated in jacket 86 past the heat exchanger to further cool the air flowing therewithin. These baffles may also be arranged in harmonic opposition to muffle the outlet port noise.

Another advantage of locating the inlet and exhause ducts adjacent one another is that a butterfly mixing valve 84 may be conveniently interposed in communication therebetween for the purpose of using the humid inlet air to both re-humidify and partially warm the cold outlet air before it is discharged into the space to be cooled.

Another novel feature that is used in conjunction with the refrigerated device described hereinbefore, is an expansible elastic or pneumatic resilient ring band 101 as shown in FIG. 1 1. The said spring band exerts an outward pressure on the reciprocatable vanes 82 as they continuously maintain sealing engagement with the elliptical walls 80 of the stator during rotation of the rotor 78 and permits a wider range of reciprocal radial motion than previously known in the art without the direct use of internal pressurization.

Other pressure systems for maintaining vane contact presently used in connection with all manner of rotary compressors and pumps are limited as to the range of radial travel that is permitted. This limiting factor in turn limits the capacity of the pump or compressor of the rotary type because the interior walls of the stator cannot be located very far away from the exterior surface of the rotor and hence the volume of fluid upon which the work is performed is limited in turn. The utility of the spring band may be further enhanced by introducing an external booster pressure (from the high pressure side of the device) through port 76 to the band interior. The advantage of using this combination of band spring action and booster pressure is that a lighter band material may be used with a resultant reduction in the frictional forces generated between vane ends 74 and the stator wall. Common means of biasing the reciprocatable vanes of a rotary pump or compressor are shown in US. Pat. No. 2,739,539 (internal air pressure) and US. Pat. No. 3,434,655 (internal air cushion). The spring band described above (with or without booster pressure), or any of the common external air biasing systems may, however, be used with good results in connection with the refrigeration device described hereinabove.

It should also be realized that the stator of the device described may be considered as a rotor, and vice versa.

I claim:

1. A refrigeration device comprising:

unitary compressor and expander means,

a heat exchanger having an inlet port connected to one side of said unitary means and an outlet port connected to another side of said unitary means, said unitary means including an intake port and an exhaust port, both of said ports being open to ambient pressure whereby a fluid enters said unitary means and is compressed, and then is passed through said heat exchanger at a substantially constant pressure, and then reintroduced into said unitary means and is expanded and exhausted to ambient pressure, such fluid passage comprising one cycle for said refrigeration device.

2. A refrigeration device comprising:

a stator including a chamber having a noncircular cross-section,

end members closing said chamber,

a rotor rotatably mounted in said chamber,

said rotor having a plurality of vane slots therein,

an inlet port and an outlet port communicating between ambient pressure and said stator chamber,

a heat exchanger having an inlet port and an outlet port both communicating with said stator chamber and both located between the inlet and outlet ports of said device which communicate with ambient pressure,

a multiplicity of vanes, at least one being at least substantially radially slidable in each of said slots,

vane biasing means which urges said vanes in at least substantially a radially outward direction into slidable but continuous contact with the chamber wall whereby a multiplicity of separate chambers are formed inside said stator chamber, said chambers being adapted to successively increase in capacity and decrease in capacity upon rotation of the rotor,

one of said separate chambers communicating with the stator inlet port to ambient pressure,

another of said separate chambers being sealed and adapted to compress fluid therewithin upon rotation of the rotor,

another of said separate chambers communicating with the heat exchanger inlet port,

another of said separate chambers communicating with the heat exchanger outlet port,

another of said chambers being sealed and adapted to expand fluid therewithin upon rotation of the rotor,

another of said separate chambers communicating with the stator outlet port to ambient pressure,

a 360 rotation of the rotor causes a given fluid segment between vanes to pass from an intake stage at ambient pressure, through a compression stage, through a heat exchanger stage, through an expansion stage, and into an exhaust stage to ambient pressure whereby substantially all the heat of compression is dissipated through the heat exchanger and substantially all the cooling effect of expansion is imparted to the fluid being exhausted at the stator exhaust port.

3. The refrigeration device of claim 2 wherein the heat exchanger outlet port is located so that the fluid volume defined by the vanes immediately on either side of said outlet port at a given rotational point has a value adapted to accept the same amount of fluid mass as entered the heat exchanger inlet port at said given rotational point.

4. The refrigeration device of claim 3 wherein said heat exchanger outlet port is at a fixed location calculated for a normal range of ambient conditions whereby an approximately VHXE compensated device is provided.

5. The refrigeration device of claim 3 including VHXE compensation means comprised of:

a fluid inlet from the heat exchanger to a chamber inside the rotor,

a pair of thin-walled semi-cylinders at least one of which is circumferentially in said rotor chamber,

fluid passages from said rotor chamber to the fluid segment of the heat exchanger exhaust port, one of said cylinders being shiftable to open and close said passages,

control means for circumferentially shifting at least one of said semi-cylinders, said means being responsive to pressure changes inside the heat exchanger whereby heat exchanger fluid is supplied to the fluid segment at the heat exchanger exhaust port in sufficient mass to provide VHXE compensation.

6. The refrigeration device of claim 3 including VHXE compensation means comprised of:

a heat exchanger outlet port movable with respect to its circumferential position on the chamber wall of the stator,

heat exchanger outlet port location control means responsive to pressure changes in the heat exc er d c ib ated to move said on s th t the iii lame tletlned by the vanes ifrrmediately on either side of said heat exchanger outlet port at a given rotational point has a value adapted to accept the same amount of fluid mass as entered the heat exchanger inlet port at said given rotational point.

7. A refrigeration device according to claim 2 having inlet and outlet ducts communicating, respectively, with the stator inlet port and stator outlet port,

moisture or ice particle catching means disposed in said outlet duct whereby exhaust air is dehumidified.

8. The refrigeration device according to claim 7 wherein said moisture catching means is connected to cooling means adapted to circulate said moisture in heat exchanging relationship with the fluid in said heat exchanger whereby said heat exchanger fluid is cooled.

9. The refrigeration device according to claim 2 having adjacently disposed inlet and outlet ducts communicating, respectively, with the stator inlet port and stator outlet port,

and variable mixing means communicating between said inlet and outlet ducts whereby incoming warm moist fluid may be mixed in the desired proportions with the cool, dry exhaust fluid to provide temperature and humidity adjustment.

10. The refrigeration device of claim 3 including VHXE compensation means comprised of:

a heat exchanger inlet port movable with respect to its circumferential position on the chamber wall of the stator,

heat exchanger inlet port control means responsive to pressure changes in the heat exchanger and calibrated to move said port so that the fluid volume defined by the vanes immediately on either side of said heat exchanger inlet port at a given rotational point has a value adapted to inject the same amount of fluid mass as exited the heat exchanger outlet port at said given rotational point.

11. The refrigeration device of claim 3, including VOE compensation means comprised of:

A device outlet port movable with respect to its circumferential position on the chamber wall of the stator,

device outlet port control means responsive to changes in VOE and calibrated to move said port so that the fluid pressure in the space denoted VOE at a given rotational point has a value substantially equal to the fluid pressure in the space denoted VOLT at said given rotational point.

" UNITED STATES PATENT OFFICE CERTIFlCATE OF CORRECTION- Patent. No. 893 v Dated ugust 29, 1972 Inventor(s) Thomas C. Edwards It is certified that error appears in the above-identified patent and that said Letters Patent are hereby corrected as shown below:

Amend Claim 1 to read as follows:

i, A refrigeration device comprising:

unitary compressor and expander means,

a heat; exchanger having an inlet port connected to one side of said unitary means and an outlet port connected to another side of said unitary means,

said unitary means including an intake port and an exhaust port, both of said ports being open to ambient pressure,

said unitary means including a rotor and stator, one of which has anon-circular section, and a single set of vanes extending between said rotor and said stator forming substantially sealed chambers therebetween,

and wherein said rotor and stator and single set of vanes accomplish intake pumping, compression, heat exchanger inlet pumping, heat exchanger outlet pumping, expansion and outlet pumping,

whereby a fluid enters said unitary means and is compressed, and

passed through said heat exchanger at a substantially constant 22 3 3 UNITED STATES PATENT OFFICE CERTIFICATE OF CURRECTION PAGE 2 Patimro. 3, 586,893 Dated August 29, 1972 Invntofls) Thomas C. Edwards It is certified that error appears in the above-identified patent and that said Letters Patent are hereby corrected as shcwn below:

r pressure, and then reintroduced intc said unitary means and is expanded and exhausted to ambient pressure, such fluid passage comprising one cycle for said refrigeration device.

Signed and sealed this 18th day of February 1975.

(SEAL) Attest:

- C MARSHALL DANN RUTH C. MASON I Commissioner of Patents Attesting Officer and Trademarks

Patent Citations
Cited PatentFiling datePublication dateApplicantTitle
US1726462 *Oct 9, 1925Aug 27, 1929Wittig KarlCompressed-air engine
US2073833 *Aug 29, 1935Mar 16, 1937Bothezat George DeAir conditioner
US2733663 *Sep 20, 1954Feb 7, 1956 Deep well pumping apparatus
US2754660 *May 12, 1955Jul 17, 1956Union Stock Yard And Transit CApparatus for refrigerating air
US2924178 *Jan 28, 1955Feb 9, 1960John X HoganFluid proportioning pump
US3301471 *Aug 19, 1965Jan 31, 1967British Oxygen Co LtdCombined compressor and expansion engine
FR554525A * Title not available
Referenced by
Citing PatentFiling datePublication dateApplicantTitle
US3793848 *Nov 27, 1972Feb 26, 1974Eskeli MGas compressor
US3850208 *Mar 3, 1972Nov 26, 1974C HamiltonPositive displacement vapor control apparatus for fluid transfer
US3884664 *Apr 23, 1974May 20, 1975Rovac CorpThrottle valve arrangement for noise control in compressor-expander
US3886763 *Feb 26, 1974Jun 3, 1975Rovac CorpSelf-driven refrigerator
US3886764 *Jul 29, 1974Jun 3, 1975Rovac CorpCompressor-expander having tilting vanes for use in air conditioning
US3886765 *Jul 29, 1974Jun 3, 1975Rovac CorpCompressor-expander having thermal isolation and adjustment features
US3905204 *Apr 15, 1974Sep 16, 1975Rovac CorpAuxiliary porting arrangement for noise control in compressor-expander
US3913351 *May 1, 1974Oct 21, 1975Rovac CorpAir conditioning system having reduced driving requirement
US3956904 *Feb 3, 1975May 18, 1976The Rovac CorporationCompressor-expander for refrigeration having dual rotor assembly
US3965697 *Nov 8, 1974Jun 29, 1976Beierwaltes Richard RCompressor and air cooling system employing same
US3967466 *Mar 17, 1975Jul 6, 1976The Rovac CorporationAir conditioning system having super-saturation for reduced driving requirement
US3968649 *Apr 15, 1974Jul 13, 1976The Rovac CorporationExhaust emission control system
US3977852 *Apr 2, 1975Aug 31, 1976The Rovac CorporationCompressor-expander with volume compensation
US4015441 *Mar 10, 1976Apr 5, 1977Robinet Sylvia JRefrigeration apparatus
US4017285 *Oct 30, 1975Apr 12, 1977The Rovac CorporationHeat pump-refrigeration system with water injection and regenerative heat exchanger
US4064705 *Oct 26, 1976Dec 27, 1977The Rovac CorporationAir conditioning system having compressor-expander in pressurized closed loop system with solar assist and thermal storage
US4088426 *May 17, 1976May 9, 1978The Rovac CorporationSliding vane type of compressor-expander having differential eccentricity feature
US4106304 *Sep 17, 1976Aug 15, 1978Michael EskeliThermodynamic compressor
US4187693 *Jun 15, 1978Feb 12, 1980Smolinski Ronald EClosed chamber rotary vane gas cycle cooling system
US4208885 *Jun 12, 1978Jun 24, 1980Schmerzler Lawrence JExpander-compressor transducer
US4295282 *Nov 15, 1978Oct 20, 1981Minnesota Mining And Manufacturing CompanyHeat and liquid recovery using open cycle heat pump system
US5388428 *Jun 23, 1993Feb 14, 1995Harper; Murry D.Gas expansion refrigeration system
US5595067 *Dec 9, 1994Jan 21, 1997Maness; James E.Energy pump
US5642629 *Oct 11, 1995Jul 1, 1997Svenska Rotor Maskiner AbCooled air cycle system and method for operating such a system
US5732560 *Nov 17, 1995Mar 31, 1998Svenska Rotor Maskiner AbSystem and method for performing cooling
US6589033 *Sep 29, 2000Jul 8, 2003Phoenix Analysis And Design Technologies, Inc.Unitary sliding vane compressor-expander and electrical generation system
US6821099Jul 1, 2003Nov 23, 2004Tilia International, Inc.Rotary pump
US6966198Dec 12, 2003Nov 22, 2005Visteon Global Technologies, Inc.Air-cycle air conditioning system for commercial refrigeration
US7114931 *Aug 12, 2004Oct 3, 2006Tilia International, Inc.Rotary pump
US8123506 *May 29, 2008Feb 28, 2012Flsmidth A/SRotary sliding vane compressor with a secondary compressed fluid inlet
US8166775Oct 9, 2003May 1, 2012Ford Global Technologies, LlcNoise attenuating device for a heating-ventilation-cooling system of a motor vehicle
US8596068 *Nov 1, 2010Dec 3, 2013Gilbert StaffendHigh efficiency thermodynamic system
US20110100011 *Nov 1, 2010May 5, 2011Gilbert StaffendHigh efficiency positive displacement thermodynamic system
DE3544445A1 *Dec 16, 1985Jun 25, 1987Bosch Siemens HausgeraeteKuehl- und gefriergeraet
DE3714727A1 *May 2, 1987Nov 10, 1988Seifert Electronic RAir conditioning equipment
WO1996016302A1 *Nov 17, 1995May 30, 1996Henrik OehmanSystem and method for performing cooling
WO1996026398A1 *Feb 19, 1996Aug 29, 1996Svenska Rotor Maskiner AbCooled air cycle system and method for operating such a system
WO2000077404A1 *Jun 15, 2000Dec 21, 2000Diro Gmbh & Co KgAir liquefaction machine
WO2004005709A1 *Jul 2, 2003Jan 15, 2004Tilia Int IncRotary pump
Classifications
U.S. Classification62/402, 417/406, 417/349, 417/392, 417/348, 418/86
International ClassificationF04C23/00, B60H1/32, F25B9/00, F24F5/00
Cooperative ClassificationF24F5/0085, F04C23/003, F25B9/004, B60H1/32
European ClassificationF04C23/00B2, F25B9/00B2, F24F5/00L, B60H1/32