JPH03168543A - Heat exchange unit for air condition system - Google Patents

Abstract

PURPOSE: To reduce air side noise and further reduce a fan axial output, by forming a tip end gap between a fan and an orifice, and disposing an orifice plate, separated around the fan. CONSTITUTION: There is accommodated in a container a heat exchanger 22 including a front surface having a length 1 between a flow passage through which air passes and a first side opposing to the flow passage, traversing the flow passage. Then, an axial fan 24 is disposed on the flow passage located between the first side and the heat exchanger 22, and an orifice plate 23 is mounted in the flow passage substantially coaxially with the axial flow fan 24 in order to guide air into the axial fan 24. Thereupon, the flow passage includes a first wall having an air intake first opening 26 and a second wall located downstream the first wall and having an exhaust second opening, in which the orifice plate 23 is placed at a position separated by a predetermined distance from the front surface with a ratio of length 1 with respect to a longitudinal distance x being 2.5 to 5.5.

Description

DETAILED DESCRIPTION OF THE INVENTION [Industrial Application Field] The present invention relates generally to air flow systems, and more particularly to heat exchange systems that optimize air flow flowing through an air management system for an outdoor unit 1 of an air conditioning system. Regarding l1.

[Conventional technology]

In many air conditioning systems that use a heat pump bunsdem for outdoor air such as in a house, the condenser unit of the air conditioning system is placed outside the house, and the evaporator unit of the system is placed in communication with the indoor space to be cooled. It uses a device coupling method that allows for In heat pump applications, the system includes an outdoor heat exchange unit outside the home and an indoor heat exchange unit in communication with the indoor space to be conditioned. These systems also employ piping, compressors, and appropriate expansion devices such that heat is transferred to or from the area to be heated or cooled. Each outdoor unit has a motor and a fan associated with the motor so that outdoor air is drawn through the unit's heat exchanger. This air is drawn in by the fan and released through the fan guard to the outdoor environment.

[Problem to be solved by the invention]

In these draw-through systems, the fan is downstream of the heat exchanger and therefore sucks in coil tube wake and coil turbulence. These disturbances interact with the fan blades, making the fan noise even louder. Furthermore, the discharge air of the fan is blown out with high kinetic energy, ie at high velocity. This emitted air is wasted as heat when mixed with the atmosphere. Therefore, this direct transmission of high-velocity air causes large system losses, which are reflected in an increase in export power.

It is an object of the present invention to provide a blow-through air management system for an outdoor heat exchange unit that overcomes the drawbacks of the prior art.

Another object of the present invention is to provide a blow-through air management system for an outdoor heat exchange unit that reduces air side noise and reduces fan shaft output.

Yet another object of the present invention is to provide a blow-through air management system for an outdoor heat exchange unit that minimizes the bulk of the outdoor unit and provides a compact unit.

[Means to solve the problem]

The above objects are achieved by preferred embodiments of the present invention. The outdoor heat exchanger system according to this preferred embodiment comprises a heat exchanger coil, a fan assembly located remotely from the heat exchanger coil for blowing air through the heat exchanger coil, and an outdoor unit. In order to provide a very compact outdoor unit that provides large and quiet air movement, it consists of orifice plates spaced apart around the fan so as to form a tip ganop between the fan and the orifice.

〔Example〕

The preferred embodiment described herein refers to an air management system for an outdoor heat exchange unit of a separate air conditioner. However, the present invention is equally applicable to the outdoor portion of a room air conditioner, that is, a package type terminal air conditioner.

FIG. 1 shows an outdoor unit with a draw-through design. Here, outdoor air is drawn through coil 12 by fan 14 (as indicated by arrow A) and discharged through grille l6 at the top of unit 10. The compressor t8 is the same as the conventional one, and circulates the refrigerant to the indoor coil (not shown) of the heat pump via the refrigerant pipes 19 and 19' and the coil t2.

The main advantages of the blow-through design will be explained with reference to FIG. Air is drawn from the relatively still environment through the artificial grill 26 and orifice plate 23 to the fan 2.
Pulled by 4, coil? released through 2. Blades 25 of fan 24 thus operate with a relatively unturbulent airflow. This is different from the draw-through unit 10 of FIG. In the picture below, the fan ■
4 blades must be operated in high turbulence generated by coil t2 upstream of the fan. Turbulent flow causes fluctuations in blade lift, which further increases fan noise. Accordingly, the blow-through system may operate more quietly than a draw-through system such as that shown in FIG.

Blow-through systems are very efficient because air evacuation losses are lower. In terms of pressure, the equation for the discharge loss from the draw-through system of Figure t is P, . 1 (1/2)V. ．． ” (1)
This is converted into a discharge area of P,. 1 (1/2) [Q] 2 (A')" [AI) 2 (2) Here, P is the draw-through discharge pressure loss, Af is the protruding discharge area of the fan, and Q is the fluid flow rate. Similarly, the emission loss from the blow-through lens dem in Figure 2 is: P1-(1/2)(Q] 2 [A""]2 (A,.i.:]2 (3) Here Te P1
is the blow-through discharge pressure drop, A,. 1 is the area of the coil surface (i.e., the length (l) of the coil 22 x height (h
)). By combining equations (2) and (3), P. ．． /P,,-A. ”/Ac....”
(.4) Now, the protruding area (Af) must be smaller than the area of the coil (Acoil).

1<A,. 11 (5)
Therefore, [A,:l ”/Ci~...]2 ku■ (
6) Applying (6) to (4), P. / Pd = <1 (7) This means that the discharge pressure drop of the blow-through air management system (P,.) must be smaller than that of the equivalent draw-through system (P..) There is.

Figure 3 shows the coil front length (l) given as the horizontal axis.
and the distance from the orifice plate to the coil iii(x
) and the fan shaft output wattage (ω) given as the vertical axis. This figure is the experimental result and shows the distance (x) from the orifice plate to the coil surface.
have also been modified to blow-through and draw-through air management systems. The airflow velocity through the system is
3660 cfm and all other experimental parameters were kept constant. The result is converted into a dimensionless ratio e/χ (coil front length/length from orifice plate to coil). This is because e/χ has a more important meaning than χ alone.

As the distance (x) from the orifice plate to the coil surface decreases, 1/χ and the shaft output (ω) increase. The dashed line of the blow-through curve is an extrapolation of actual experimental data based on the draw-through curve. At constant axial power, the blow-through system (represented by the lower curve) is significantly more compact (i.e., χ is smaller for the same curve) than the draw-through system (represented by the upper curve). I know what I'm getting.

Similarly, it can be seen that for the same ratio magnitude (ie, constant Q/χ), a blow-through system can be much more efficient. Furthermore, the curve shows that when l and shaft power are constant, the magnitude of χ for a blow-through system can be less than half that of a draw-through system. Furthermore, compared to the same 1/χ, the blow-through system requires ↓}% less shaft output.

Figure 4 shows the surface length of the coil 111 given as the horizontal axis (
{) and the ratio of the distance (x) from the orifice plate to the coil and the sound output (dBA) given as the vertical axis. This figure is a duplicate of the same experiment as described in FIG.

Blow-through and draw-through results are calculated by extrapolation to 1
There is a tendency for people to wait for a while. In Figure 4: Well, the advantages of the blow-through system over the draw-through system are shown. The same sound output (d
BA), the distance (x) between the orifice plate and the coil for blow-through is approximately 70% larger than that for the equivalent draw-through.
small. Furthermore, the same orifice plate-coil distance M (
1'/χ constant), the blow-through system is
It is about 1.2dBA quiet.

In residential air conditioning systems, it is desirable to maintain sound output levels, or air management systems, below 74 dBA. Therefore, when e/χ≧5.5, the sound output is 74
The unacceptable nature of approaching and exceeding dBA is shown in FIG. Therefore, for noise control, the maximum value of 1/χ is 5.5. From FIG. 3, it can be seen that good efficiency can be obtained when e/χ≧2,5.

As a result, 1/χ for good efficiency and noise control while maintaining the compactness of the air management system.
The range of is 5.5≧l/χ≧2.5 Figure 5 shows the diameter (D) of the fan 24 and the fan sound output (
dBA) and the shaft output (ω). Here, the horizontal axis represents the diameter (D) of the fan, and the vertical axis represents the sound output (dBA) on one side and the shaft output (ω) on the other. Analysis of sound power and shaft power as a function of fan diameter is performed in noise control engineering.
earing) 1982 July-August, Mol
．． 1 9/No. 1. Based on the method of Wright (T.) from "Correlated Velocity Parameters for Axial Fan Noise" on pages 17-25. The relative shape of the curves is not a strong function of fan static pressure rise (Ps). The analysis is based on a large Noyashi unit,
i.e. outdoor fan speed of 856 rpm, 3 6 6
It was conducted on a 4 to 5 ton air conditioning system with an air flow of 0 cfm and a head static pressure rise (Ps) of 0.26 inches. Therefore, the result derived from this curve is P
is not significantly affected by the choice of s.

As is clear from FIG. 5, most of the effect of reducing the shaft output by increasing the diameter (D) of the 4- to 5-ton unit was achieved when D=450 mm.

Any diameter larger than this is generally acceptable for effectiveness. However, the high power is D=650m
74.m. Reach dBA. The volume is the maximum allowable limit that is compatible with the sound priority objective.

Therefore, the range of diameters for acceptable efficiency and sound is: 450mm≦D≦650mm Similar analyzes as above were also used for small chassis units, ie 1.5 to 3 ton units. FIG. 6 shows the relationship between the diameter (D) of the fan 24 and the sound output (dBA) and shaft output (ω) of the fan.

In this figure, the fan diameter (D) is shown on the horizontal axis, and the sound output (dBA) is shown on one side of the vertical axis, and the axial output (ω) is shown on the other side.
is shown. FIG. 6 depicts the sound output as a function of fan diameter for this relatively small system. Also,
Wright's method is 8 5 6 rpm outdoor speed, 18
00 cfm airflow and 0.2 inches of head static pressure rise were used.

The optimal diameter was determined as the one that achieves the minimum sound volume. Therefore, from Figure 5, the optimal diameter is 52? It is mm.

Similarly, the optimal diameter for a small nyasi unit is
It is 415mm. To establish the maximum (D...) and minimum (D.i,) l shapes for that small chassis unit, D for the large shear (Lc.)
man/D 09 1 1au+m and D. ．． ．． /D
. ■IMUS ratio is D for a small chassis. 2. ，
■． be multiplied.

That is, [(D−−−)/(D−.+−−−) 1. j
．． x (D.p+lsusL-c-=CD...)
．． ．． 、． (8) (650/
520) x 415 = 519 Ujm
(9) Therefore, (D...). ．． 、． = 519 +n+++
(10) ((D- + -)
/ (D.pl+musH!.c.]X(D.2,1.
. ．． ),. 、． = (D.+.). ．． 、．
(h) (450/ 520) X 415-3
59 mn+ (12) Therefore, (D..). ．． 、． - 359 mm
(13) Therefore, the optimal efficiency and diameter range for the neck for small Nyashnits 1 is: 359 questions ≦ D ≦ 519 o+u+
(14) This range is clear from FIG.

The small yarn diameter range was measured from the large chassis analysis, so the h/D range, i.e. 1.1≦h/D
≦Top. The same value applies for 6.

In Figure 7, the horizontal axis shows the coil height (h) and the fan diameter (
D) and the system internal loss (k), that is, dimensionless loss, is shown on the vertical axis to express the relationship between them. This analysis of system internal losses (k) does not include coil losses. As is clear from FIG. 7, the loss is minimized (efficiency is maximized) as h/D decreases. Sound is also minimized as the fan has less work to do and generates less noise. To ensure space for the orifice 23, the coil height (h) must be greater than the fan diameter (D). Therefore, taking this into account, the minimum h/D ratio is approximately Ll. The preferred range of fan diameters determines the preferred range of h/D ratios. Therefore, the maximum h/D? is rod 1 - 2, which is equal to the small diameter ratio for the smallest dog multiplied by the smallest suitable h/D ratio. This relationship is (h). ．． ．． /CD). ．． ．． One (D,.../D,...

)X ((h).../CD). ．． , ]
(15) That is, (h.), . ．． ．． /CD). ．． ．． [(650)/450) X (1.1)
(1(i)(}1).../CD). ．． ．． = 1
．． 6 (17) In summary, the optimal coil height to fan diameter ratio range is: 1.1 ≦ }1/D ≦ 1 . 6
(18) The radius of curvature of the orifice is particularly critical to sound performance as a result of overspeed phenomena.

FIG. 8 is a diagram showing the orifice plate 28 together with the air velocity vector at the orifice position. Compared to the central flow of air through the orifice opening 27, the velocity is increased. Therefore, the V p/Vu ratio is greater than 1.0. Here, Vp indicates peak speed (overspeed), and Vu
indicates the central fluid velocity. Since the fluid at the velocity Vp hits the tip of the fan, it has a large effect on fan noise. It is preferable to have no overspeed (ie, Vp2Vu). Since noise is proportional to inlet speed, overspeed will make fan noise louder. Since Vp is larger than Vu,
The fan blades generate more noise than if they were exposed to a constant population velocity Vu. Vp is inversely proportional to the orifice radius of curvature (γ.). Therefore, γ. As V becomes smaller, Vp becomes larger compared to Vu and fan noise increases. Therefore, the larger the radius of curvature, the smaller Vp becomes, and the noise decreases accordingly. Therefore, the range of the radius of curvature of the orifice is within the range of the following equation.

．． 05≦γ. /D. ．． ,≦． 1 5 (
19) γ. /D. ．． ,but. If it is larger than 15, it will affect the compactness of unit 1, and the value will be . If it is smaller than 05, the noise will be large, so the above equation can be said to be a suitable range.

FIG. The conventional orifice is a thin plate orifice that terminates at 90 degrees and has a minimum thickness. The orifice of the present invention is a tail flared orifice terminating at a 30° position. Orifice plate 2
3 and the fan 24 there is a gap 30 (ε).

Its wide base orifice provides a superior single-plate expansion effect compared to a single plate orifice. This 30 degree skirt flare improves efficiency and sound performance.

FIG. 10 shows the tip gap 30 (
ε) and the ratio of the diameter of the fan 24 and the sound output edge (dBA) given as a vertical sleeve. The influence of the tip gas sob (ε) on noise is due to this first effect.
It is clear from Figure 0. Tip gap is L5%1
・If it is larger than , it will have a very large effect on the noise. Therefore, in this defense, the tip gas is smaller than the upper 5 percent.

As described above, it is clear that the present invention can also be applied to devices that emit horizontally as examples.

〔Effect of the invention〕

According to the present invention, air side noise is reduced and fan shaft output is reduced. Furthermore, since the envelopment of the outdoor unit is minimized, a compact heat exchange unit is obtained.

[Brief explanation of the drawing]

FIG. 1 is a perspective view of a conventional outdoor heat exchange unit, and FIG. 2 is a schematic plan view of an outdoor heat exchange unit using the present invention.
Figure 3 is a comparison diagram showing the relationship between the axis and the ratio of the coil front length/orifice plate to coil length, and Figure 4 is a comparison diagram showing the relationship between the sound output and the coil front length/orifice plate to coil length ratio. FIG. 5 is a comparison diagram showing the relationship between the ratio of the length to the coil and the ratio of the length to the coil; FIG. , 6th
Figure 2 is a comparative diagram showing the relationship between both the sound output and the shaft output and the fan diameter of the present invention for a small air conditioner unit;
FIG. 7 is a comparative diagram showing the relationship between internal system losses and the ratio of coil height and fan diameter of the present invention; FIG. 8 is a diagram of the velocity profile of air flowing through the fan orifice of the present invention; FIG. 9 is a schematic diagram of the flared orifice and fan of the present invention; FIG. 10 is a comparative diagram showing the relationship between sound output and fan tip gap and fan diameter ratio of the present invention. be. [Explanation of symbols] 10...Outdoor heat exchange unit, l2...Coil 14
．． 24... Fan, 16... Grill 18... Compressor, {9... Refrigerant piping 23... Orifice plate, 30
...Gap FIG. 3 FIG. 4 FIG. 5 FIG. 7

Claims (5)

[Claims] (1) A length ( a heat exchanger (22) comprising a front surface having a front surface with a front surface having a heat exchanger (22); an axial fan (24) disposed in the flow path between the first side and the heat exchanger; an orifice plate (23) mounted in the flow path substantially coaxially with the axial fan for guiding into the fan; and a second wall downstream of the first wall for exhaust.
a second wall having an opening; the orifice plate is located at a predetermined distance (x) from the front surface; the ratio of the length (l) to the distance (x) is from 2.5 to 5;
．． A heat exchange unit for an air conditioning system, characterized in that the heat exchange unit is within the range of 5. (2) In the heat exchange unit according to claim (1),
Heat exchange unit for an air conditioning system, characterized in that the axial fan has a diameter (D) in the range from 359 mm to 650 mm. (3) In the heat exchange unit according to claim (2),
the front surface has a height (h) between two sides perpendicular to the opposing first side of the flow path, the ratio of the height (h) to the diameter of the axial fan; However, from 1.1 to 1.
A heat exchange unit for an air conditioning system, characterized in that the heat exchange unit is within the range of 6. (4) In the heat exchange unit according to claim (2),
The orifice plate further includes a base-widening orifice having a radius of curvature (γ.), and the orifice plate further includes a base-widening orifice having a radius of curvature (γ.
A heat exchange unit for an air conditioning system, characterized in that the ratio of the radius of curvature (γ.) to D) is in the range of 0.05 to 0.15. (5) In the heat exchange unit according to claim (2),
The above orifice has a tip gap distance (
ε) and the diameter (D) of the axial fan
A heat exchange unit for an air conditioning system, characterized in that the ratio of the tip gap distance (ε) to the tip gap distance (ε) does not exceed 1.5%.