CA2023468C - Air management system - Google Patents
Air management systemInfo
- Publication number
- CA2023468C CA2023468C CA002023468A CA2023468A CA2023468C CA 2023468 C CA2023468 C CA 2023468C CA 002023468 A CA002023468 A CA 002023468A CA 2023468 A CA2023468 A CA 2023468A CA 2023468 C CA2023468 C CA 2023468C
- Authority
- CA
- Canada
- Prior art keywords
- air
- fan
- orifice
- flow path
- diameter
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Expired - Fee Related
Links
Classifications
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F28—HEAT EXCHANGE IN GENERAL
- F28C—HEAT-EXCHANGE APPARATUS, NOT PROVIDED FOR IN ANOTHER SUBCLASS, IN WHICH THE HEAT-EXCHANGE MEDIA COME INTO DIRECT CONTACT WITHOUT CHEMICAL INTERACTION
- F28C3/00—Other direct-contact heat-exchange apparatus
Landscapes
- Engineering & Computer Science (AREA)
- Mechanical Engineering (AREA)
- General Engineering & Computer Science (AREA)
- Other Air-Conditioning Systems (AREA)
- Air-Conditioning Room Units, And Self-Contained Units In General (AREA)
- Structures Of Non-Positive Displacement Pumps (AREA)
- Pharmaceuticals Containing Other Organic And Inorganic Compounds (AREA)
- Acyclic And Carbocyclic Compounds In Medicinal Compositions (AREA)
- Nitrogen Condensed Heterocyclic Rings (AREA)
- Central Air Conditioning (AREA)
Abstract
SUBSTITUTE
REMPLACEMENT
SECTION is not Present Cette Section est Absente
REMPLACEMENT
SECTION is not Present Cette Section est Absente
Description
2023~
AIR MANAGEMENT SYSTE~
This invention relates generally to an air flow system, and more specifically, is directed to a flow through air management system for an outdoor unit of an air conditioning system.
Air conditioning systems, including heat pump systems, for conditioning residences and other interior spaces frequently utilize a combination of components such that the condenser unit of an air conditioning system is located outside of the residence and the evaporator unit of the system is located in communication with the interior space to be cooled. In a heat pump application, the system might have an outdoor heat exchange unit outside the residence and an interior heat exchange unit in communication with the interior space to be conditioned. These systems further utilize a compressor and appropriate expansion device and piping such that heat may be transferred either to the region to be heated or from the region to be cooled. Each outdoor unit has an electric motor and fan associated therewith such that outdoor air may be drawn through the heat exchanger of the unit. This air then is ingested by the fan and discharged through a fan guard to the outdoor environment. In these draw-through systems, because the fan is downstream of the heat exchanger, it ingests coil tube-wakes and coil turbulence.
These disturbances interact with the fan blades and exacerbate fan noise. Further, the fan discharge air exits the outdoor unit at high velocity, which equates to high kinetic energy, and is dissipated as heat when the discharge air mixes with the atmospheric environment. Thus, this direct discharge of high velocity air represents a q~
.
2023~68 significant system loss which is reflected in an increased fan shaft power.
It is an object of the present invention to provide a blow-through air management system for an outdoor heat exchange unit which overcomes the drawbacks of the prior art.
It is another object of the present invention to provide a blow-through air management system for an outdoor heat exchange unit to reduce air-side noise and reduce fan shaft power.
It is yet another object of the present invention to provide a blow-through air management system for an outdoor heat exchange unit which allows the envelope of the outdoor unit to be minimized to provide a compact unit.
These and other objects are achieved according to a preferred embodiment of the present invention by providing an outdoor heat exchanger system which comprises a heat exchange coil, a fan assembly spaced from the heat exchange coil for blowing air through said coil, and an orifice plate spaced around the fan which forms a tip gap between the fan and the orifice for the purpose of providing a very compact outdoor unit which allows for efficient and quiet.air movement through the outdoor unit.
The various features of novelty which characterize the invention are pointed out with particularity in the claims annexed to and forming a part of this specification. For a better understanding of the invention, its operating advantages, and specific objects obtained by its use, 2023~
reference should be made to the accompanying drawings and descriptive matter in which there is illustrated and described a preferred embodiment of the invention.
In the accompanying drawings, forming a part of this specification, and which reference numerals shown in the drawings designate like or corresponding parts throughout the same;
Figure 1 is a perspective view of an outdoor heat exchange unit of the prior art.
Figure 2 is a schematic sectional plan view of an outdoor heat exchange unit utilizing the present invention.
Figure 3 is a comparative diagram showing the relationship between shaft horsepower and coil frontal length/orifice plate - to - coil length ratio.
Figure 4 is a comparative diagram showing the relationship between sound power and the coil frontal length/orifice plate - to - coil length ratio.
Figure 5 is a comparative diagram showing the relationship between both sound power and shaft power, and fan diameter of the present invention for large air conditioning units.
Figure 6 is a comparative diagram showing the relationship between both sound power and shaft power and fan diameter of the present invention for small air conditioning units.
. ~ . - . -. . . ,:
20~3~
Figure 7 is a comparative diagram showing the relationship between internal system losses and the ratio of the coil height and the fan diameter of the present invention.
Figure 8 is a schematic view of the velocity profile of the air flowing through the fan orifice of the present invention.
Figure 9 is a schematic view of fan and reflared orifice of the present invention.
Figure 10 is a comparative diagram showing the relationship between sound power and the ratio of the fan tip gap and the fan diameter of the present invention.
The preferred embodiment described herein refers to an air management system for an outdoor heat exchange unit of a split air conditioner. It is to be understood, however, that this invention has like applicability to the outdoor portion of room air conditioners or packaged terminal air conditioners.
Figures 1 show as an outdoor unit 10 of a draw-through design wherein the outdoor air (as indicated by arrows "a") is drawn through the coil 12 by the fan 14 and is discharged out grill 16 in the top of the unit 10. Compressor 18 circulates refrigerant to the indoor coil (not shown) of the heat pump prior art or the like, through refrigerant lines 19 and 19l and coil 12.
Figure 2 is used to illustrate the primary advantages of a blow-through design. Air is drawn from the relatively '~
t 20~34~
guiescent environment through inlet grill 26 and orifice plate 23 into the fan 24 and discharged through coil 22.
Consequently, the blades 25 of fan 24 operate in a relatively turbulent-free airstream. This is not the case for a drnw-through unit 10 of Fig. 1 where the blades of the fan 14 must operate in a more turbulent airstream, the turbulence being generated by the presence of the coil 12 upstream of the fan. Turbulence causes fluctuations in blade lift which exacerbate fan noise. Consequently, the blow-through system will tend to operate more guietly.
The blow-through system is more efficient because the air discharge 1066es are lower. In terms of pressure, the discharge loss equation from a draw-through system of Fig. 1 i~ -Pdt = 1 PVdt2 (1) which in terms of discharge area is:
Pdt 1 p ~ ~ (2) where Pdt is the draw through discharge pressure loss, whereAf is the projected discharge area of the fan, and where Q
is the fluid Slow.
Similarly, the discharge loes equation from a blow-through system or Fig. 2 is:
2 iAcDil]
~ ~.
., . . ~ .
2023~8 where Pbt is the B low through discharge pressure loss, and where ACoil is the face area of the coil (i.e. the frontal length {1~ or the coil 22 times the height ~h~ or the coil).
Combining equations (2) and (3), Pbt = Af 2 (4) Pdt Acoil Now, the projected discharge area (Af) must be less than the coil face area (ACoil)~
f Acoil (5) Therefore, [ Af ]
2c 1 (6) ~ coil ]
and using (6) in (4), Pbt < 1 (7) Pdt which means that the discharge pressure loss of a blow-through air management system (Pbt) must be less than that of an equivalent draw-through system (Pdt).
Figure 3 is a diagram showing the relationship between the ratio of the coil frontal length (1) and the orifice plate-to-coil face distance (x) given as an abscissa and the shaft power (w) of the fan in watts given as an ordinate.
The diagram is the result of an experiment in which orifice plate (23) to coil (22) distance (x) was varied both for blow-through and draw-through air management systems. The air flow rate through the system was 3660 cfm and all other . . - . .. ,.: ..
-.
.'. . , ~: .
," ... .
'; : ., ~ . . ;
.. . ~ . , 2~3~8 experimental parameters were held constant. The results are in terms of the dimensionless ratio l/x (coil frontal length/orifice-plate-to-coil length) because it is more meaningful than x alone.
As the orifice plate-to-coil face distance (x) decreases, l/x and shaft power (w) increase. The dashed line of the blow through curve is an extrapolation of actual experimental data based on the draw-through curve. It can be seen, that at a constant shaft power a blow-through system (represented by the lower curve) can be considerably more compact (i.e. smaller x for same 1) than a draw-through system (represented by the upper curve). Similarly, for the same scaled dimensions (i.e. l/x=constant), the blow-through system can be significantly more efficient. Moreover, the curves demonstrate that the blow-through system x dimension can be less than half that of a draw-through system, for constant 1 and shaft power. Or, for the same l/x, the blow-through system will require 11 percent less shaft power.
Figure 4 is a diagram showing the relationship between the ratio of the coil frontal length (1) and the orifice plate-to-coil distance (x) given as an abscissa, and the sound power (dBA) given as an ordinate. The diagram is the result of the same experiment set forth in Fig. 3. The blow-through and draw-through results are extrapolated to show the expected trends. The advantages of a blow-through system over the draw-through system are shown in Fig. 4.
For the same sound power (dBA), which in effect is noise, the blow-through orifice-plate-to-coil (x) is about 70 percent less than an equivalent draw-through. Moreover, for :, 2~3~fi3 the same orifice-plate-to-coil (l/x=constant), the blow-through system will be about 1.2 dBA quieter.
In residential air conditioning systems, it is desirable to maintain the sound power level or the air management system below 74 dBA. Thus, Figure 4 indicates that an l/x > 5.5 would be unacceptable in that sound power would approach and excsed 74dBA. Consequently, the maximum value of l/x for noise control is 5.5. From Figure 3, it can be seen that most of the efficiency benefit is achieved for l/x >2.5. As a result, the range of 1/x for favorable efficiency and noise control while maintaining compactness of the air management system is:
5.5 > l/x > 2.5.
Figure 5 is a diagram showing the relationship between the diameter (D) of fan 24 and the sound power (DBA) and shaft power (w) of the fan, wherein the diameter (D) of the fan is given as an abscissa and the sound power (dBA) is given as an ordinate and the shaft power (w) is given as another ordinate. The analysis of fan sound power and shaft power as a function of fan diameter is based on the method of Wright, T., from "A Velocity Parameter for the Correlation of Axial Fan Noise", Noise Control Engineering, July-August 1982, Vol 19/Number 1, pg 17-25. The relative shapes of the curves are not a strong function of fan static pressure rise (Ps)~ The analysis was performed on a large chassis unit or a four (4) to five (S) ton air conditioning system, with an outdoor fan speed of 856 RPM, with an air flow of 3660 cfm, and a static pressure rise (Ps) of 0.26 inches of water.
Therefore, the conclusions drawn from this curve are not materially affected by the choice of Ps.
.
:
.
2 0 ~
As apparent from Fig. 5, it is found that most of the shaft power reduction benefit of increasing diameter (D) of a four (4) to five (5) ton unit is achieved at about D=450mm. Any diameter larger than this would be generally acceptable from an efficiency viewpoint. However, the sound power reaches 74 dBA at D=650 mm, which would be the maximum acceptable limit compatible with the sound leadership objective.
Consequently, the diameter range for acceptable efficiency and sound is:
450 mm < D < 650 mm.
A similar analysis to the above was used for a small chassis unit, i.e. one and one-half (1 1/2~ to three (3) ton units.
Figure 6 is a diagram showing the relationship between the diameter (D) of fan 24 and the sound power (DBA) and shaft power (w) of the fan, wherein the diameter (D) of the fan is given as an abscissa and the sound power (dBA) is given as an ordinate and the shaft power (w) is given as another ordinate. Figure 6 presents the sound and shaft power as a function of fan diameter for this smaller system. Again, Wright's method was used with an outdoor fan speed of 856 rpm, an alr flow of 1800 cfm, and a static pressure rise of 0.2 inches of water.
The optimum diameter is defined as the one which manifests minimum sound, therefore, from Figure 5, the optimum diameter is 520 mm. Similarly, the optimum diameter for the small chassis unit is 415mm. To establish the maximum ~DmaX) and minimum (Dmin) geometry for the small chassis unit, the ratios Dmax/DOptimum and Dmin/Doptimum large chassis (l.c.) are multiplied by Doptimum for the small chassis (s.c.).
~, .
, 2023~3 That is:
(Dmax) x (Doptimum)s.c. (Dmax)s.c. ( 8) ( optimum)l.c.
x 415 = 519 mm ( 9) Therefore, (Dmax)s.c. 519 mm (lo) ( min) x (Doptimum)s.c. (Dmin.)s.c (11) ( optimum) x 415 - 359 mm (12) Therefore, ( min)s.c. 359 mm (13) Consequently, the diameter range for optimum efficiency and sound for the small chassis unit is:
359mm < D < 519mm (14) This range is apparent from Figure 6.
Because the small chassis diameter range has been scaled from the large chassis analysis, the same values for the h/D
range i.e. 1.1 < h/D ~ 1.6 apply.
Figure 7 is a diagram showing the relationship between the ratio of the coil height ~h) and the fan diameter (D) given as an abscissa, and system internal losses (K), a dimensionless loss factor given as an ordinate. This analysis of system internal losses (K) does not include coil losses. As apparent from Fig. 7, losses are minimized .
.
'. ' : ' ' ,.
, ;" ' ' ' ' ~ ':
. - . :
2023~fi~
(efficiency maximized) as h/D decreases. Sound would be minimized as well because the fan is called upon to do less work, hence it would make less noise. The height of the coil (h) must be larger than the fan diameter (D) because of the space required by the orifice 23. Therefore, this into account, the minimum h/D ratio is about 1.1. There is a preferred h/D ratio range because of the preferred range of fan diameters. Therefore, the maximum h/D ratio is equivalent to the maximum-to-minimum diameter ratio, from Figure 5, times the minimum preferred h/D ratio. This relationship is:
(h)maX Dmax (H)min = x (15) (D)Max Dmin (D)min or (h)max (650) = x(}.1) (16) ( )max 450 (h)max = 1.6 (17) (D)max Summarizing, the optimum coil height-to-fan diameter ratio range is:
1.1 ~ h/D 1.6 (18) The radius of curvature of the orifice is especially critical to sound performance as a result of over-speed .
- . ~
2023~68 phenomenon. Figure 8 is a schematic view of the orifice plate 23 of Fig. 2 with velocity vectors of the air at the orifice. As air enters the orifice plate, the air nearest the orifice tends to accelerate to a higher velocity relative to the core flow through orifice opening 27.
Therefore, the ratio Vp/Vu is greater than 1.0, where Vp denotes the peak velocity (over-speed) and Vu denotes the core fluid velocity. Because the fluid with velocity Vp enters the fan tip, it can, and does, have a dramatic influence on fan noise. One would prefer to have no over-speed (i.e., Vp=Vu). Over-speed exacerbates fan noise because noise is proportional to inlet velocity. Since Vp is greater than Vu, the fan blades will make more noise than if exposed to a uniform inlet velocity of Vu. Vp is inversely proportional to the orifice radius of curvature (rO). Thus, rO becomes smaller, Vp becomes larger relative to Vu, and fan noise increases. Consequently, the larger the radius of curvature, the smaller Vp becomes with a concomitant decrease in noise. Therefore, the range of orifice radius of curvature is within the range from:
.05 rO .15 < _ < (19) Dfan Which can be called the preferable range, since unit compactness suffers for rO/dfan values much greater than .15, and noise suffers for values much lower than .05.
Figure 9 is a schematic view of the orifice plate 23 and fan 24 according to the present invention. Prior art orifices are terminated at 90 degrees and are thin plate orifices with a minimum thickness. The present orifice has a ': ' ' .
`
:: -20~3~
termination angle of 30 degrees and i8 referred to as a reflared orifice. The orifice plate 23 and fan 24 have a gap 30 (~ ) there between. The reflared orifice provides superior diffusion relative to a simple thin plate orifice.
This reflare of 30 degrees improves efficiency and sound performance.
Figure 10 is a diagram showing the relationship between the ratio of the tip gap 30 ( ~) and the fan 24 diameter (D) given as an abscissa, and the sound power (dBA) given as an ordinate. Apparent from Fig. 10 is the influence of tip gap (~) on fan noise. Tip gaps much greater than 1.5 percent exact an increasingly severe penalty on noise.
Consequently, the tip gap according to the present invention is less than 1.5 percent.
The invention has been described with reference to a particular embodiment however, it is to be understood by these skilled in the art, that variations and modifications can be made within the spirit and scope or the invention.
For example, the horizontal discharge of the unit may be changed to a vertical discharge.
- .
. ~ , . . .
AIR MANAGEMENT SYSTE~
This invention relates generally to an air flow system, and more specifically, is directed to a flow through air management system for an outdoor unit of an air conditioning system.
Air conditioning systems, including heat pump systems, for conditioning residences and other interior spaces frequently utilize a combination of components such that the condenser unit of an air conditioning system is located outside of the residence and the evaporator unit of the system is located in communication with the interior space to be cooled. In a heat pump application, the system might have an outdoor heat exchange unit outside the residence and an interior heat exchange unit in communication with the interior space to be conditioned. These systems further utilize a compressor and appropriate expansion device and piping such that heat may be transferred either to the region to be heated or from the region to be cooled. Each outdoor unit has an electric motor and fan associated therewith such that outdoor air may be drawn through the heat exchanger of the unit. This air then is ingested by the fan and discharged through a fan guard to the outdoor environment. In these draw-through systems, because the fan is downstream of the heat exchanger, it ingests coil tube-wakes and coil turbulence.
These disturbances interact with the fan blades and exacerbate fan noise. Further, the fan discharge air exits the outdoor unit at high velocity, which equates to high kinetic energy, and is dissipated as heat when the discharge air mixes with the atmospheric environment. Thus, this direct discharge of high velocity air represents a q~
.
2023~68 significant system loss which is reflected in an increased fan shaft power.
It is an object of the present invention to provide a blow-through air management system for an outdoor heat exchange unit which overcomes the drawbacks of the prior art.
It is another object of the present invention to provide a blow-through air management system for an outdoor heat exchange unit to reduce air-side noise and reduce fan shaft power.
It is yet another object of the present invention to provide a blow-through air management system for an outdoor heat exchange unit which allows the envelope of the outdoor unit to be minimized to provide a compact unit.
These and other objects are achieved according to a preferred embodiment of the present invention by providing an outdoor heat exchanger system which comprises a heat exchange coil, a fan assembly spaced from the heat exchange coil for blowing air through said coil, and an orifice plate spaced around the fan which forms a tip gap between the fan and the orifice for the purpose of providing a very compact outdoor unit which allows for efficient and quiet.air movement through the outdoor unit.
The various features of novelty which characterize the invention are pointed out with particularity in the claims annexed to and forming a part of this specification. For a better understanding of the invention, its operating advantages, and specific objects obtained by its use, 2023~
reference should be made to the accompanying drawings and descriptive matter in which there is illustrated and described a preferred embodiment of the invention.
In the accompanying drawings, forming a part of this specification, and which reference numerals shown in the drawings designate like or corresponding parts throughout the same;
Figure 1 is a perspective view of an outdoor heat exchange unit of the prior art.
Figure 2 is a schematic sectional plan view of an outdoor heat exchange unit utilizing the present invention.
Figure 3 is a comparative diagram showing the relationship between shaft horsepower and coil frontal length/orifice plate - to - coil length ratio.
Figure 4 is a comparative diagram showing the relationship between sound power and the coil frontal length/orifice plate - to - coil length ratio.
Figure 5 is a comparative diagram showing the relationship between both sound power and shaft power, and fan diameter of the present invention for large air conditioning units.
Figure 6 is a comparative diagram showing the relationship between both sound power and shaft power and fan diameter of the present invention for small air conditioning units.
. ~ . - . -. . . ,:
20~3~
Figure 7 is a comparative diagram showing the relationship between internal system losses and the ratio of the coil height and the fan diameter of the present invention.
Figure 8 is a schematic view of the velocity profile of the air flowing through the fan orifice of the present invention.
Figure 9 is a schematic view of fan and reflared orifice of the present invention.
Figure 10 is a comparative diagram showing the relationship between sound power and the ratio of the fan tip gap and the fan diameter of the present invention.
The preferred embodiment described herein refers to an air management system for an outdoor heat exchange unit of a split air conditioner. It is to be understood, however, that this invention has like applicability to the outdoor portion of room air conditioners or packaged terminal air conditioners.
Figures 1 show as an outdoor unit 10 of a draw-through design wherein the outdoor air (as indicated by arrows "a") is drawn through the coil 12 by the fan 14 and is discharged out grill 16 in the top of the unit 10. Compressor 18 circulates refrigerant to the indoor coil (not shown) of the heat pump prior art or the like, through refrigerant lines 19 and 19l and coil 12.
Figure 2 is used to illustrate the primary advantages of a blow-through design. Air is drawn from the relatively '~
t 20~34~
guiescent environment through inlet grill 26 and orifice plate 23 into the fan 24 and discharged through coil 22.
Consequently, the blades 25 of fan 24 operate in a relatively turbulent-free airstream. This is not the case for a drnw-through unit 10 of Fig. 1 where the blades of the fan 14 must operate in a more turbulent airstream, the turbulence being generated by the presence of the coil 12 upstream of the fan. Turbulence causes fluctuations in blade lift which exacerbate fan noise. Consequently, the blow-through system will tend to operate more guietly.
The blow-through system is more efficient because the air discharge 1066es are lower. In terms of pressure, the discharge loss equation from a draw-through system of Fig. 1 i~ -Pdt = 1 PVdt2 (1) which in terms of discharge area is:
Pdt 1 p ~ ~ (2) where Pdt is the draw through discharge pressure loss, whereAf is the projected discharge area of the fan, and where Q
is the fluid Slow.
Similarly, the discharge loes equation from a blow-through system or Fig. 2 is:
2 iAcDil]
~ ~.
., . . ~ .
2023~8 where Pbt is the B low through discharge pressure loss, and where ACoil is the face area of the coil (i.e. the frontal length {1~ or the coil 22 times the height ~h~ or the coil).
Combining equations (2) and (3), Pbt = Af 2 (4) Pdt Acoil Now, the projected discharge area (Af) must be less than the coil face area (ACoil)~
f Acoil (5) Therefore, [ Af ]
2c 1 (6) ~ coil ]
and using (6) in (4), Pbt < 1 (7) Pdt which means that the discharge pressure loss of a blow-through air management system (Pbt) must be less than that of an equivalent draw-through system (Pdt).
Figure 3 is a diagram showing the relationship between the ratio of the coil frontal length (1) and the orifice plate-to-coil face distance (x) given as an abscissa and the shaft power (w) of the fan in watts given as an ordinate.
The diagram is the result of an experiment in which orifice plate (23) to coil (22) distance (x) was varied both for blow-through and draw-through air management systems. The air flow rate through the system was 3660 cfm and all other . . - . .. ,.: ..
-.
.'. . , ~: .
," ... .
'; : ., ~ . . ;
.. . ~ . , 2~3~8 experimental parameters were held constant. The results are in terms of the dimensionless ratio l/x (coil frontal length/orifice-plate-to-coil length) because it is more meaningful than x alone.
As the orifice plate-to-coil face distance (x) decreases, l/x and shaft power (w) increase. The dashed line of the blow through curve is an extrapolation of actual experimental data based on the draw-through curve. It can be seen, that at a constant shaft power a blow-through system (represented by the lower curve) can be considerably more compact (i.e. smaller x for same 1) than a draw-through system (represented by the upper curve). Similarly, for the same scaled dimensions (i.e. l/x=constant), the blow-through system can be significantly more efficient. Moreover, the curves demonstrate that the blow-through system x dimension can be less than half that of a draw-through system, for constant 1 and shaft power. Or, for the same l/x, the blow-through system will require 11 percent less shaft power.
Figure 4 is a diagram showing the relationship between the ratio of the coil frontal length (1) and the orifice plate-to-coil distance (x) given as an abscissa, and the sound power (dBA) given as an ordinate. The diagram is the result of the same experiment set forth in Fig. 3. The blow-through and draw-through results are extrapolated to show the expected trends. The advantages of a blow-through system over the draw-through system are shown in Fig. 4.
For the same sound power (dBA), which in effect is noise, the blow-through orifice-plate-to-coil (x) is about 70 percent less than an equivalent draw-through. Moreover, for :, 2~3~fi3 the same orifice-plate-to-coil (l/x=constant), the blow-through system will be about 1.2 dBA quieter.
In residential air conditioning systems, it is desirable to maintain the sound power level or the air management system below 74 dBA. Thus, Figure 4 indicates that an l/x > 5.5 would be unacceptable in that sound power would approach and excsed 74dBA. Consequently, the maximum value of l/x for noise control is 5.5. From Figure 3, it can be seen that most of the efficiency benefit is achieved for l/x >2.5. As a result, the range of 1/x for favorable efficiency and noise control while maintaining compactness of the air management system is:
5.5 > l/x > 2.5.
Figure 5 is a diagram showing the relationship between the diameter (D) of fan 24 and the sound power (DBA) and shaft power (w) of the fan, wherein the diameter (D) of the fan is given as an abscissa and the sound power (dBA) is given as an ordinate and the shaft power (w) is given as another ordinate. The analysis of fan sound power and shaft power as a function of fan diameter is based on the method of Wright, T., from "A Velocity Parameter for the Correlation of Axial Fan Noise", Noise Control Engineering, July-August 1982, Vol 19/Number 1, pg 17-25. The relative shapes of the curves are not a strong function of fan static pressure rise (Ps)~ The analysis was performed on a large chassis unit or a four (4) to five (S) ton air conditioning system, with an outdoor fan speed of 856 RPM, with an air flow of 3660 cfm, and a static pressure rise (Ps) of 0.26 inches of water.
Therefore, the conclusions drawn from this curve are not materially affected by the choice of Ps.
.
:
.
2 0 ~
As apparent from Fig. 5, it is found that most of the shaft power reduction benefit of increasing diameter (D) of a four (4) to five (5) ton unit is achieved at about D=450mm. Any diameter larger than this would be generally acceptable from an efficiency viewpoint. However, the sound power reaches 74 dBA at D=650 mm, which would be the maximum acceptable limit compatible with the sound leadership objective.
Consequently, the diameter range for acceptable efficiency and sound is:
450 mm < D < 650 mm.
A similar analysis to the above was used for a small chassis unit, i.e. one and one-half (1 1/2~ to three (3) ton units.
Figure 6 is a diagram showing the relationship between the diameter (D) of fan 24 and the sound power (DBA) and shaft power (w) of the fan, wherein the diameter (D) of the fan is given as an abscissa and the sound power (dBA) is given as an ordinate and the shaft power (w) is given as another ordinate. Figure 6 presents the sound and shaft power as a function of fan diameter for this smaller system. Again, Wright's method was used with an outdoor fan speed of 856 rpm, an alr flow of 1800 cfm, and a static pressure rise of 0.2 inches of water.
The optimum diameter is defined as the one which manifests minimum sound, therefore, from Figure 5, the optimum diameter is 520 mm. Similarly, the optimum diameter for the small chassis unit is 415mm. To establish the maximum ~DmaX) and minimum (Dmin) geometry for the small chassis unit, the ratios Dmax/DOptimum and Dmin/Doptimum large chassis (l.c.) are multiplied by Doptimum for the small chassis (s.c.).
~, .
, 2023~3 That is:
(Dmax) x (Doptimum)s.c. (Dmax)s.c. ( 8) ( optimum)l.c.
x 415 = 519 mm ( 9) Therefore, (Dmax)s.c. 519 mm (lo) ( min) x (Doptimum)s.c. (Dmin.)s.c (11) ( optimum) x 415 - 359 mm (12) Therefore, ( min)s.c. 359 mm (13) Consequently, the diameter range for optimum efficiency and sound for the small chassis unit is:
359mm < D < 519mm (14) This range is apparent from Figure 6.
Because the small chassis diameter range has been scaled from the large chassis analysis, the same values for the h/D
range i.e. 1.1 < h/D ~ 1.6 apply.
Figure 7 is a diagram showing the relationship between the ratio of the coil height ~h) and the fan diameter (D) given as an abscissa, and system internal losses (K), a dimensionless loss factor given as an ordinate. This analysis of system internal losses (K) does not include coil losses. As apparent from Fig. 7, losses are minimized .
.
'. ' : ' ' ,.
, ;" ' ' ' ' ~ ':
. - . :
2023~fi~
(efficiency maximized) as h/D decreases. Sound would be minimized as well because the fan is called upon to do less work, hence it would make less noise. The height of the coil (h) must be larger than the fan diameter (D) because of the space required by the orifice 23. Therefore, this into account, the minimum h/D ratio is about 1.1. There is a preferred h/D ratio range because of the preferred range of fan diameters. Therefore, the maximum h/D ratio is equivalent to the maximum-to-minimum diameter ratio, from Figure 5, times the minimum preferred h/D ratio. This relationship is:
(h)maX Dmax (H)min = x (15) (D)Max Dmin (D)min or (h)max (650) = x(}.1) (16) ( )max 450 (h)max = 1.6 (17) (D)max Summarizing, the optimum coil height-to-fan diameter ratio range is:
1.1 ~ h/D 1.6 (18) The radius of curvature of the orifice is especially critical to sound performance as a result of over-speed .
- . ~
2023~68 phenomenon. Figure 8 is a schematic view of the orifice plate 23 of Fig. 2 with velocity vectors of the air at the orifice. As air enters the orifice plate, the air nearest the orifice tends to accelerate to a higher velocity relative to the core flow through orifice opening 27.
Therefore, the ratio Vp/Vu is greater than 1.0, where Vp denotes the peak velocity (over-speed) and Vu denotes the core fluid velocity. Because the fluid with velocity Vp enters the fan tip, it can, and does, have a dramatic influence on fan noise. One would prefer to have no over-speed (i.e., Vp=Vu). Over-speed exacerbates fan noise because noise is proportional to inlet velocity. Since Vp is greater than Vu, the fan blades will make more noise than if exposed to a uniform inlet velocity of Vu. Vp is inversely proportional to the orifice radius of curvature (rO). Thus, rO becomes smaller, Vp becomes larger relative to Vu, and fan noise increases. Consequently, the larger the radius of curvature, the smaller Vp becomes with a concomitant decrease in noise. Therefore, the range of orifice radius of curvature is within the range from:
.05 rO .15 < _ < (19) Dfan Which can be called the preferable range, since unit compactness suffers for rO/dfan values much greater than .15, and noise suffers for values much lower than .05.
Figure 9 is a schematic view of the orifice plate 23 and fan 24 according to the present invention. Prior art orifices are terminated at 90 degrees and are thin plate orifices with a minimum thickness. The present orifice has a ': ' ' .
`
:: -20~3~
termination angle of 30 degrees and i8 referred to as a reflared orifice. The orifice plate 23 and fan 24 have a gap 30 (~ ) there between. The reflared orifice provides superior diffusion relative to a simple thin plate orifice.
This reflare of 30 degrees improves efficiency and sound performance.
Figure 10 is a diagram showing the relationship between the ratio of the tip gap 30 ( ~) and the fan 24 diameter (D) given as an abscissa, and the sound power (dBA) given as an ordinate. Apparent from Fig. 10 is the influence of tip gap (~) on fan noise. Tip gaps much greater than 1.5 percent exact an increasingly severe penalty on noise.
Consequently, the tip gap according to the present invention is less than 1.5 percent.
The invention has been described with reference to a particular embodiment however, it is to be understood by these skilled in the art, that variations and modifications can be made within the spirit and scope or the invention.
For example, the horizontal discharge of the unit may be changed to a vertical discharge.
- .
. ~ , . . .
Claims (4)
EXCLUSIVE PROPERTY OR PRIVILEGE IS CLAIMED ARE
DEFINED AS FOLLOWS:
1. A heat exchange unit for use in a system for conditioning air comprising:
an enclosure defining a flow path for air to flow therethrough;
said flow path having a first wall with a first opening for ingress of the air therethrough and a second wall downstream of said first wall with a second opening for the discharge of the air therethrough;
a heat exchanger within said enclosure, including a frontal face having a length (l) between opposite first sides of said flow path and a height (h) between opposite second sides, generally transverse to said flow path which the air passes therethrough;
an axial fan positioned in said flow path between said first wall and said heat exchanger; and an orifice plate adapted to be mounted in said flow path and substantially coaxial with said axial fan for guiding the air into said axial fan, a portion of said orifice plate generally parallel to said frontal face of said heat exchanger positioned a predetermine distance (x) from said heat exchanger wherein a ratio of said length (l) to said distance (x) is in the range between 2.5 and 5.5 to reduce the air-side noise.
an enclosure defining a flow path for air to flow therethrough;
said flow path having a first wall with a first opening for ingress of the air therethrough and a second wall downstream of said first wall with a second opening for the discharge of the air therethrough;
a heat exchanger within said enclosure, including a frontal face having a length (l) between opposite first sides of said flow path and a height (h) between opposite second sides, generally transverse to said flow path which the air passes therethrough;
an axial fan positioned in said flow path between said first wall and said heat exchanger; and an orifice plate adapted to be mounted in said flow path and substantially coaxial with said axial fan for guiding the air into said axial fan, a portion of said orifice plate generally parallel to said frontal face of said heat exchanger positioned a predetermine distance (x) from said heat exchanger wherein a ratio of said length (l) to said distance (x) is in the range between 2.5 and 5.5 to reduce the air-side noise.
2. A heat exchange unit as set forth in claim 1 wherein said axial fan has a diameter (D) in the range between 359mm and 650mm and said orifice plate further includes a reflared orifice with a radius of curvature (ro) wherein a ratio of said radius of curvature (ro) to said axial fan diameter (D) is in the range between 0.05 and 0.15.
3. A heat exchange unit as set forth in claim 1 wherein said axial fan has a diameter (D) in the range between 359mm and 650mm and said orifice is spaced from said fan by a tip gap distance (E) wherein a ratio of said tip gap distance (E) to said axial fan diameter (D) is no more than 1.5%.
4. A heat exchange unit as set forth in claim 2 wherein said reflared orifice has a termination angle of approximately 30 degrees relative to said first sides of said flow path.
Applications Claiming Priority (2)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| US42740889A | 1989-10-27 | 1989-10-27 | |
| US427,408 | 1989-10-27 |
Publications (2)
| Publication Number | Publication Date |
|---|---|
| CA2023468A1 CA2023468A1 (en) | 1991-04-28 |
| CA2023468C true CA2023468C (en) | 1993-01-19 |
Family
ID=23694748
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| CA002023468A Expired - Fee Related CA2023468C (en) | 1989-10-27 | 1990-08-16 | Air management system |
Country Status (8)
| Country | Link |
|---|---|
| JP (1) | JPH03168543A (en) |
| KR (1) | KR910008365A (en) |
| AR (1) | AR247623A1 (en) |
| BR (1) | BR9005431A (en) |
| CA (1) | CA2023468C (en) |
| ES (1) | ES2026011A6 (en) |
| IT (1) | IT1244051B (en) |
| NZ (1) | NZ234942A (en) |
Families Citing this family (4)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| KR100550962B1 (en) * | 1998-10-27 | 2006-06-21 | 삼성종합화학주식회사 | Catalyst for Hydrogenation Purification of Terephthalic Acid |
| JP2014111998A (en) * | 2011-03-28 | 2014-06-19 | Toshiba Carrier Corp | Heat source unit |
| CN105805852A (en) * | 2016-03-28 | 2016-07-27 | 朱虹斐 | Energy saving air conditioning outer unit |
| CN106762827A (en) * | 2016-12-16 | 2017-05-31 | 上海置信节能环保有限公司 | A kind of asymmetric S types airfoil fan and its design and application process |
Family Cites Families (1)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPS5454169A (en) * | 1977-10-08 | 1979-04-28 | Asahi Chem Ind Co Ltd | Compound |
-
1990
- 1990-08-16 NZ NZ234942A patent/NZ234942A/en unknown
- 1990-08-16 CA CA002023468A patent/CA2023468C/en not_active Expired - Fee Related
- 1990-09-28 ES ES9002476A patent/ES2026011A6/en not_active Expired - Lifetime
- 1990-10-12 IT IT02172990A patent/IT1244051B/en active IP Right Grant
- 1990-10-24 AR AR90318173A patent/AR247623A1/en active
- 1990-10-26 KR KR1019900017218A patent/KR910008365A/en not_active Application Discontinuation
- 1990-10-26 BR BR909005431A patent/BR9005431A/en not_active IP Right Cessation
- 1990-10-29 JP JP2291688A patent/JPH03168543A/en active Pending
Also Published As
| Publication number | Publication date |
|---|---|
| CA2023468A1 (en) | 1991-04-28 |
| IT9021729A1 (en) | 1992-04-12 |
| KR910008365A (en) | 1991-05-31 |
| AR247623A1 (en) | 1995-01-31 |
| JPH03168543A (en) | 1991-07-22 |
| IT1244051B (en) | 1994-07-01 |
| IT9021729A0 (en) | 1990-10-12 |
| ES2026011A6 (en) | 1992-04-01 |
| NZ234942A (en) | 1992-06-25 |
| BR9005431A (en) | 1991-09-17 |
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Legal Events
| Date | Code | Title | Description |
|---|---|---|---|
| EEER | Examination request | ||
| MKLA | Lapsed |