AU1887392A - Air conditioning for humid climates - Google Patents

Description

AIR CONDITIONING FOR HUMID CLIMATES

This invention relates to both a method and means for air conditioning, and although generally applicable, has special value when it is associated with air conditioning for humid climates.

BACKGROUND OF THE INVENTION

Tropical and humid climates present serious design problems at various combinations of climatic conditions with room sensible and latent heat load conditions. These problems also occur in temperate climates, often at non-peak humid weather conditions when room sensible heat load is low and the room latent heat load is high. The method of this invention, and the means necessary for that method to be performed, address the dynamics of the problem arising from changing climate, ventilation requirements and changing room sensible and latent heat loads. The objectives of the invention are to achieve low running costs, good performance within comfort standards, low first costs and low space requirements for the air conditioning equipment. The main objective is to provide an improved system performance which can be achieved over a full operating range of the air conditioning system of the invention, and in particular, overcome the problems arising from humid air, inadequate ventilation and the associated health hazards (the "sick building syndrome").

The modern air conditioning system involves so many

interacting variables that any attempt to make a

scientifically valid assessment of performance over a

complete operating range for each design, let alone to undertake a full optimisation, has always been regarded as impractical. Optimisation of one particular parameter, or of the range of one particular variable, at one particular operating condition affects other variables to an extent which can render the study valueless. These other variables are not necessarily affected at that particular operating condition. For example, in a system being selected in a temperate climate the peak refrigeration requirement usually occurs when high people loads, requiring high ventilation air supply, coincide with the afternoon peak of a hot day, during which transmission is at its maximum. A dehumidifier coil can be selected which, at this peak load condition,

satisfactorily meets that design requirement while making provision for future changes and allowing a safe margin for errors in the load estimates. However the performance of this dehumidifier coil during certain critical part load conditions depends on the way in which the designer has chosen to satisfy the peak load condition. An excellent peak load selection can, at part load, result in a sick building with high humidity and outside air intake which is well below the minimum levels prescribed by the relevant Standard.

PRIOR ART

In tropical climates it is conventional practice to pre-cool the outside air however it is not usually sent on to the return air heat exchanger since it is in no position to offset the room sensible and latent heat loads properly.

The chilled water serves only a small outside air flow rate based on ventilation specifications, and a very much larger return air flow rate determined by the room sensible heat difference between the design room dry bulb temperature and the supply air. In Singapore for the design of a multistorey office building this is usually in the ratio of outside air to return air flow rate of 1 to 10.

In existing practice the chilled water flow rate would be selected to have a water temperature rise of about 8ºC. for the outside air heat exchanger resulting in a reduced mass transfer of moisture from the air. In this invention the outside air heat exchanger is served by a chilled water flow rate which is inordinately large for the relatively small outside air flow rate passing through it. This is because it is based on the requirements of the peak simultaneous demand of the return air complex to which the outside air coolant flows. Hence, quite opposite to existing practice the water temperature rise across the outside air heat exchanger is very low, often under 1ºC. The combination of a higher chilled water flow rate, a low water temperature rise, a high water velocity through the outside air heat exchanger and an outside air heat exchanger design with many circuits with relatively short paths maximise

dehumidification. This is further enhanced by the outside air coil condition being close to the saturation curve during the critical part-load condition when the room sensible heat ratio is low and the outside air condition has a high

humidity ratio. In this invention, the path of the coil condition curve through the outside air heat exchanger follows down adjacent to the saturation line of the

psychrometric chart. The path of the outside air coil condition curve has a very much steeper slope than the path of the return air coil condition curve. On the mixing of the leaving outside air with the leaving return air from their respective heat exchangers, the dew point of the supply air to the room is reduced. The deep dehumidification obtainable through the outside air heat exchanger serves as a

replacement of methods where over-cooling and reheating or equivalent is necessary to avoid excessive humidities in humid and tropical climates. Figure 8 illustrates on a psychrometric chart the paths of the outside air and return air heat exchangers at part load conditions.

A serious problem is encountered in the otherwise excellent variable air volume system performance when for example at 50% room sensible heat load the outside air flow rate in a constant population density building doubles in percentage terms relative to the return air flow rate. During humid outside air conditions this often results in excessive relative humidity. In this invention quite the opposite occurs. The outside air is now shown to help, not hinder the supply air to be at a lower dew point and yield for the same load ratio line a lower room relative humidity.

The existing practice of pre-cooling of the outside air also differs from the invention in important aspects. Firstly the coolant flow paths are different because in existing practice the chilled water to the central heat exchanger does not usually flow onto the return air heat exchanger as in the case of this invention. Secondly, in the existing practice chilled water flow rate through the outside air heat

exchanger is not based on the chilled water requirements of the return air heat exchanger complex during simultaneous building peak performance. Consequently the circuiting and coil design is very different from the prior art, and the flow of fluids and heat and mass transfer strategy to offset the sensible and latent heat loads over a full climatic range (including its internal variations) is essentially very different.

Still further, existing practice never includes any bypassing arrangement around a throttled return air heat exchanger so as to ensure a maximum flow at all times through the outside air heat exchanger including the critical low room sensible heat ratio, a feature which is employed in this invention when return air heat exchanger design is compacted to fit into small spaces such as ceiling plenums and where the return air heat exchangers have a fixed size.

Still further, the existing practice must service the outside air heat exchanger with a separate source of chilled water from that required by the return air heat exchanger.

In this invention there is a reduced chilled water rate to the return air heat exchanger complex in that the outside air heat exchanger carries out part of the role of the return air heat exchanger complex.

As a consequence, in buildings where the outside air heat exchanger serves a large return air heat exchanger complex, the return air heat exchanger complex is reduced in size and requires smaller chilled water risers since it has been partially relieved of offsetting some of the return air heat loads.

Unlike existing practice, in this invention the engineering design requires a knowledge of the building heat loads over its climatic range with an understanding of the inter-relationships between peak heat load and part load

performance. This data is revealed in detail in Tables 1 and 2 and in performance map Fig 3.

Existing practice for the peak return air condition would be to design for a water temperature rise of about 7ºC. or 8ºC. Excellent peak load performance may be obtained. However the chilled water temperature rise at a critical part load condition would have been well above 7.01ºC. shown below for Table 5 Standard part load condition because as the heat load decreases the water temperature rise increases when a fixed size heat exchanger is employed. At part load, the room condition would have been excessively humid. In this case, the water temperature rise at peak had to be 2.85ºC. in order to prevent the water temperature rise exceeding 7.01ºC.

indicated in Table 5. Obviously, it was the part load performance which determined the peak load water flow rate. In the prior art, these fundamental relationships do not enter into the design considerations.

The failure of existing systems in humid and tropical

climates to prevent high room humidities during critical conditions has often resulted in the supply air temperature to the rooms causing moisture to condense out within the room near the supply diffusers owing to the room dew point being too high above the supply air temperature. One solution conventionally used to raise the supply air temperatures does not resolve the fact that the room is excessively humid.

Furthermore it exacerbates the problem since it causes the return air heat exchanger to dehumidify less. An object of this invention is therefore to provide a lower room humidity which does not exceed acceptable standards and which is comfortable with the lower supply air temperatures.

Other control means which are being used in prior art have included the throttling of coolant flow, but as indicated in our US Patents 4876858 and others, the conventional method of throttling the coolant flow in both constant air volume and variable air volume systems when room heat loads decrease has only a limited range when a single dehumidifying coil is employed. It can be shown that it is thermodynamically not possible to economically control both room dry bulb

temperature and room humidity condition to always be within engineering comfort standards by throttling the coolant flow over a wide range of conditions. As the coolant flow rate decreases the dehumidification capacity of the coil also decreases. This is in sharp conflict with the fact that at low room sensible heat loads and constant or increasing room latent heat loads the dehumidification capacity of the coil is required to increase. The consequence is a high room humidity condition which may well be above engineering comfort standards, and can induce serious health hazards.

This condition arises particularly at low room sensible heat ratios (which may be defined as the ratio of room sensible to room total heat, the total heat of course including both room sensible and latent heat.)

This invention addresses the need to offset sensible and latent heat combinations over a range of climatic conditions . The dynamics of the numerous coil conditions imposed are subject to a large number of contending variables. Conventional practice does not fully consider the effect of variations in room and outside air conditions, and the main object of this invention is to improve air conditioning methods and equipment and more effectively meet demands, for good performance, for human comfort and health and for low running costs without the need for wasteful overcooling and reheating.

BRIEF SUMMARY OF THE INVENTION

In this invention, the dehumidifier of an air conditioning system is divided into two portions, one being an outside air heat exchanger and the other a return air heat exchanger. The configuration of the air conditioner is so arranged that coolant flows first through the heat exchange conduits of the outside air heat exchanger, and the outside air flows through the outside air heat exchanger. The coolant chilled water for example has only a small water temperature rise across the outside air heat exchanger before it flows through the heat exchange conduits of the return air heat exchanger, and the return air passes through the return air heat exchanger. With this arrangement the uncooled outside air is cooled by the coolant when the coolant is at its lowest mean

temperature to achieve a maximum mean partial pressure difference of moisture and a maximum dew point temperature difference between the humid outside air and the wetted cooling surfaces of the heat exchanger. The effect is to increase dehumidification of the outside air at a high ratio of mass transfer to heat transfer because of the coil

condition curve being constrained to pass down along the curvature of the saturation line of the psychrometric chart. The outside air then mixes with the treated return air after it leaves the return air coil thus assisting the

dehumidification of the treated return air.

If the outside air heat exchanger is directly coupled to the return air heat exchanger the outside air heat exchanger performance in dehumidifying is partially impaired when during part load the coolant flow rate through the outside air heat exchanger is reduced due to the throttling within the return air coil complex. This is no serious problem when the system employed uses the low face velocity/high coolant velocity technology, including the staging of the return air heat exchanger size as is the case in the example of the Singapore lecture theatre. This design has three stages of return air heat exchanger size and consequently has

sufficient assistance from the outside air flow treatment to meet the design specifications in spite of the direct coupled nature of the outside air heat exchanger to the return air heat exchanger. In fact it is not necessary in this case to require the leaving condition from the outside air heat exchanger to have a dew point below that of the supply air dew points. Figures 2, 3, 4, 5 and Tables 1 and 2 below describe this performance.

On the other hand much greater assistance is required when a Standard high face velocity, fixed size return air coil is employed. The coolant flow through the outside air heat exchanger in this case is designed not to be constrained by the throttling of coolant through the return air heat

exchanger. This is accomplished by means of a bypass and a modulating valve designed to maintain the required coolant flow through the outside air heat exchanger. Table 4 below is an example of a design for a centrally located outside air heat exchanger which is serving an unconstrained coolant flow rate by means of a bypass and valve assembly. Figure 5 is a schematic diagram which includes such a bypass where a single outside air heat exchanger serves a number of return air heat exchangers from a remote position close to the chillers.

Figure 1 is a similar bypass and valve arrangement where the outside air heat exchanger is adjacent to the return air heat exchanger it serves.

More specifically the invention may be said to consist of means dividing a dehumidifier of an air conditioner into an outside air heat exchanger and return air treatment means comprising a return air heat exchanger, coolant flow conduits connecting said heat exchangers, the configuration of said conduits and dehumidifier being so arranged that said coolant flows first through a heat exchange conduit of the outside air heat exchanger and subsequently through a heat exchange conduit of the return air heat exchanger, and the coolant flow rates through the outside air heat exchanger is based on the maximum simultaneous demands of the return air heat exchanger total complex at peak load, and is dependent on the water temperature rise at peak load to be compatible with the water temperature rise occurring at all load conditions within the range and being low enough to result in a room relative humidity which does not exceed the maximum

permissible design limited, and air flow directing means which direct outside air to flow over the coil of the outside air heat exchanger, and which direct return air to flow over the coil of the return air heat exchangers, such that the maximum temperature differential between air flow and coolant occurs in the outside air heat exchanger.

This invention includes direct coupling of the coolant flows between the outside air and return air heat exchangers, but in one of its aspects also includes throttling means in said conduits of the return air heat exchanger arranged such that the coolant fed to the outside air coil section continues only in part to the return air heat exchanger, when a lower total coolant flow rate is required. By means of a bypass conduit across the return air heat exchanger, flow rate can be maintained at a high level through the outside air heat exchanger, and the outside air can be cooled to very low dew point temperatures in order to dehumidify sufficiently with only a small increase in temperature of the cooling water flowing into the return air coil. This outside air then on mixing with the relatively larger portion of treated return air is controlled so that both the supply air temperature and moisture content are at a condition to maintain the room target dry bulb temperature and the room humidity within comfort conditions. The invention may be interfaced with existing variable air volume and constant air volume

installations, and in many instances utilise standard temperature and throttling equipment otherwise employed.

This invention is enhanced if fundamental principles are used, concerning establishing design parameters taking into account the influence of these variables on the design requirements. Performance data hereunder indicates

evaluation.

This invention will be seen to be in contrast to conventional practice. In conventional practice minimum water temperature rise of say 5ºC. to 9ºC. in the outside air coil is designed for the chilled water (if that is the coolant used). In this invention the water temperature rise may be kept less than 3ºC., sometimes less than 1ºC. Furthermore, the performance achieved by this invention will meet the comfort standards, over a much greater range of conditions and at the same time provide an energy savings of typically 35% of the running costs over a full year when compared with variable air volume systems, and 50% when compared with constant air volume systems, as is disclosed hereunder.

The control of humidity within the air conditioned space can be readily effected in this invention by varying the flow of coolant through the return air heat exchanger, and the return air heat exchanger mainly serves to offset loads of sensible heat, and usually latent heat, which have been generated within that space. Separate humidity control means are not generally necessary, since free range of humidity is achieved through design to be contained within limits of comfort requirements.

The invention makes possible full ventilation requirements for part-load if used with the more common variable air volume systems. Although the invention can be used

independently with great advantage, if combined with the high coolant velocity/low face velocity multi-stage systems of the aforesaid US Patent 4876858, very high energy savings can be achieved.

BRIEF SUMMARY OF THE DRAWINGS AND TABLES

An embodiment of the invention is described hereunder in some detail with reference to and as illustrated in the

accompanying drawings in which:

Fig 1 is a basic diagrammatic representation of an air conditioning system embodying the split system of this invention, for a single zone system where the outside air heat exchanger is adjacent to the return air heat exchanger. It includes a bypass with modulating pressure control to maintain full coolant flow rate through the outside air heat exchanger when return air heat exchanger flow is being throttled during part load conditions.

Fig 2 is a schematic view showing the detailed circuiting of the split system,

Fig 3 shows a "map" illustrating the conditions between three stages of air conditioning which are determined by load, and utilise the valve arrangement of Fig 2,

Fig 4 is a psychrometric chart which illustrates the

embodiment described hereunder with respect to peak

conditions when this invention is interfaced with a system which has a return air heat exchanger with multi-stages. The outside air heat exchanger is not required to have a leaving off coil condition below the supply air DPT. The system has the OA coolant flow rate coupled directly to the RA coolant flow rate as indicated in Fig 2. The performance specifications are fully satisfied.

Fig 5 diagrammatically illustrates a multistorey installation illustrating the split system of this invention with a single outside air heat exchanger remotely located from separate RA heat exchangers and including a bypass with regulating valve and separate RA treatment means,

Fig 6 indicates the performance of the three stage coil of Fig 2 and compares it with conventional VAV and CAV

performance with respect to relative humidity over the full climatic range when the lecture theatre has full occupancy,

Fig 7 is a psychrometric chart which illustrates the

embodiment described hereunder with respect to peak

conditions, for the multistorey split system of this

invention with a single outside air heat exchanger remotely located from a fixed size RA heat exchanger complex and including a bypass with regulating valve,

Fig 8 is a psychrometric chart which illustrates the

embodiment described hereunder with respect to part load humid conditions for the same system of Fig 7,

Table 1 sets out change over values in three stage example in Figs 3, 4 and 5, for falling thermal loads,

Table 2 sets out change over values in three stage example of Table 1 for rising thermal loads,

Table 3 shows some of the values for the arrangement of

Tables 1 and 2, comparing however the performance when an LFV/HCV split system of the invention is compared with

corresponding conventional VAV and CAV systems,

Table 4 indicates details of the outside air heat exchanger including performance at peak and part loads, and is relevant to Figs 7 and 8 ,

Table 5 compares a "prestige" system utilising two stages with a single stage "Standard" system which is of lower cost, and

Table 6 indicates annual running costs of the invention compared with a conventional VAV installation and a

conventional CAV installation.

The following standard abbreviations are used in the

specification, Tables and drawings:

OA : Outside air

RA : Return air

DBT : Dry bulb temperature

WBT : Wet bulb temperature

DPT : Dew point temperature

WTR : Water temperature rise

CHWS : Chilled water supply

WPD : Water pressure drop in kilo Pascals

Rm SHR : Room sensible heat ratio

C : Temperature ºCelsius

LPS : Litres/second

RH : Relative humidity %

WV : Water vel. m/sec.

KW : Kilowatt energy

B1 : Bottom of stage 1

T2 : Top of stage 2

VAV : Variable air volume system

CAV : Constant air volume system

CCC : Coil condition curve

LFV/HCV : Low face velocity (air)/High coolant

velocity (water)

SA : Supply air temperature to the room

HE : Heat exchanger, air conditioning

dehumidifier, cooling coil While, as said above, the invention can be utilised

independently with some advantage, the invention can achieve the highest efficiency if it is used in conjunction with the low face velocity/high coolant velocity disclosure in our aforesaid US Patent 4876858. (Hereinafter the subject matter of that patent will be abbreviated to "LFV/HCV".)

Referring to the drawings, Fig 1 shows the principle of the invention in a very much simplified embodiment, wherein an outside air heat exchanger 10 is arranged to receive chilled water from a chiller 11, and all of the chilled water flows through the outside air heat exchanger 10 rising in

temperature only about 1ºC., and then through a return air heat exchanger 12, the water passing through to lines, the lower line being illustrated to contain the regulating valve 13 which is a throttling valve for controlling the amount of water which flows through the heat exchanger 12, a pump 14 and back through the chiller to be recirculated.

A feature of this invention is that the coolant flow rates through the outside air heat exchanger is based on maximum simultaneous demands of the return air heat exchanger (which can be a bank of heat exchangers) at peak load, but peak load seldom exists and when single fixed size RA heat exchangers are installed there is provided a bypass conduit valve assembly 15 which comprises a bypass conduit 16 and a

regulating valve 17 which, in effect bypasses a combination of heat exchanger 12 and regulating valve 13. It is also important to note that the outside air passing through heat exchanger 10 and return air passing through heat exchanger 12 mix after having passed through the heat exchangers and is driven by a supply air blower or fan 18 to supply air to a space (room) which is to be air conditioned.

In Fig 2, the outside air heat exchanger sits adjacent, in the same cross-section as the return air heat exchanger perpendicular to the air flow. It illustrates a return air heat exchanger which permits three different stages

representing three different active heat exchanger sizes. When Valve A and Valve B are open the total return air heat exchanger is active in Stage one. When Valve B is closed the heat exchanger has an intermediate size, when both Valves A and B are closed the heat exchanger is at its smallest active size. In addition the conventional modulating valve is indicated in the return chilled water conduit leaving the return manifold. The above staging is described in the aforesaid US Patent 4876858. The changeover points between staging are indicated on the Performance Table 1 and 2 for both falling and rising room sensible heat loads.

Fig 2 illustrates a simpler form which can be termed a "split adjacent" system wherein chilled water supply goes into a manifold 21 and is directed through that manifold into the outside air heat exchanger 10 in a series of parallel

circuits (with respect to coolant flow), there being

illustrated 18 circuits each with four passes. The

downstream manifold 26 receives the water from the return air heat exchanger 12, and redirects it to the chiller but through a modulating valve 27.

The physical arrangement which is in common use however is not limited to a "split adjacent" system as shown in Fig 1 and Fig 2, but normally utilises a single outside air heat exchanger 10 which is relatively large and this delivers air to a series of levels of a multistorey building through a duct 30 as shown in Fig 5, from which the air is taken through side ducts 31 to be mixed in outlet ducts 32 from air which is passed from the air conditioned space through the return air heat exchangers 12. Between 10% and 30% of return air is usually spilt or leaked from the air conditioned room and is replaced by outside air which comes through the outside air heat exchanger 10. There is thus a complex of return air heat exchangers, and the coolant flow rates from the chillers (three in number in Fig 5) is necessarily based on the maximum simultaneous demands of all the return air heat exchangers 12 at peak load, and should not be less than those demands but preferably a little more. This in turn dependent on water temperature rise at peak load which must be compatible with the water temperature rise occurring at all load conditions within the range, and must be low enough to result in a room relative humidity which does not exceed the maximum permissible design limit. The range of the air conditioned space is usually between 22ºC. and 27ºC.

temperature and between 30% and 60% relative humidity.

Our aforesaid US Patent 4876858 (or any one of its equivalent other family patents) is included in this specification by way of reference, and the control means of the stages which are clearly set forth in that patent are utilised with only minor description herein.

Figs 3 and 4 illustrate the three stages of air conditioning in a lecture theatre in Singapore wherein difficulty is encountered in reducing humidity, and Table 1 (which

comprises Tables 1A and 1B). Table 1B should be read in association with the falling loads of Fig 3. Figs 7 and 8 are also relevant to this description, Fig 7 showing the peak condition of outside air having a temperature of 32ºC. and outside air wet bulb temperature of 27 corresponding

approximately to a humidity ratio of .020 (about 68%). The room condition is shown in Fig 7 as having a temperature of about 24ºC. and the relative humidity of a little more than 50%.

A feature of the LFV/HCV design methodology is that a range of design conditions, which is representative of the complete operating range envisaged for the plant, is considered during the selection of a dehumidifier coil. A global "Performance Map", covering the performance of the coil over the range of design conditions, is generated, and is illustrated in Fig 3. Showing the three stages of this embodiment, Fig 3 has eight lines which slope upwardly to the left and which identify the number of occupants (the population) of the conditioned room, ranging from 150 maximum to five minimum. The change over control shown on the map of Fig 3 is substantially the same as in our aforesaid U.S. Patent No 4876858, and is therefore not repeated herein. It will be noted however that the main control is coolant velocity regulated by throttle valves 13, or variable delivery pump 35, or both.

At the upper right hand side of Fig 3, there is a comparison between the conventional constant air volume (CAV) system and the conventional variable air volume (VAV) system with the split system with adjacent outside air to return air heat exchangers coupled to each other and with three stages of return air coil size. No bypass is required to meet the specifications.

The data from which the performance map of Fig 3 and Table 1 is generated are also summarised in Table 1. The map (Fig 3) can include any or all of the variables and parameters listed in the tables. For clarity it is herein illustrated by the plots of room relative humidity, which is representative of the level of comfort achieved, versus the chilled water velocity in the active coil circuits, which is representative of the mechanical operation of the plant, for range of outside air conditions and occupancy numbers. In one example shown these latter parameters are representative respectively of the externally imposed and the internally generated loads. For another example the population of the room may not vary and thus the room latent load may be reasonably constant, but the room sensible load may vary through changes in the transmission load and the equipment being operated. As a general rule, the variables and parameters depicted visually on the map are chosen to be the most appropriate for

conveying information about the plant and its operation in relation to the imposed loads. The drawings and tables relate to conditions for falling loads, but are almost the same for rising loads, the latter being separately illustrated in Fig 4 and Table 2.

The described embodiment relates to a system designed to service a lecture theatre in a tropical port such as

Singapore. The variation in the people loads and the accompanying ventilation loads are extreme in the example. The design challenge is to maintain both thermal comfort and the necessary ventilation air quantities for student numbers which range from the capacity seating of 150 down to tutorial groups of five to ten students.

The performance is achieved with Fig 2 return air

dehumidifier coils with 18 circuits. All 18 circuits of the ventilation air segment have 4 passes per circuit and are active at all times. All are active when in the high load range. In "Stage 1" (valves A and B open), only 14 circuits are active at intermediate loads with valve A open and B closed, ("Stage 2"), and the number of active circuits reduces to 12 at low loads with valves A and B closed,

("Stage 3"). The process of changing from one stage to another is referred to as "change over". There is only one chilled water feed. The water passes first through the outside air coil 10 and then through the return air coils 12 of the complex as shown in Fig 5. The symbol T in the first column of Table 1 refers to the top of a Stage, that is, where the coolant velocity in the active circuits of that Stage is at its maximum. The symbol B refers to the bottom of a Stage where the coolant velocity is at its lowest value. For example, the symbol B2 refers to the bottom of Stage 2. The total cooling requirements for each of the design

conditions defining the operating range are shown in the column headed Ref Cap.

The set points for change over from one stage to another are different for falling loads from those for rising loads to avoid hunting of the control system.

The effectiveness of the staging of the LFV/HCV coil in accommodating changes in room sensible heat load is clearly shown in Figs 2 and 3.

In Fig 4 several critical operating conditions of the system are compared with conventional 4 row deep VAV and CAV systems which were selected to satisfy peak load specifications. The energy requirements to achieve this same performance are presented in Table 2, which also provides comparative annual running costs.

The supply air dry bulb temperature is chosen in each case to minimise the risk of condensate forming on the supply air diffusers. The lower supply air temperature possible with the LFV/HCV system reflects the lower room dew point which can be achieved.

Difficulties are encountered however in circumstances of relatively low temperature say 25ºC. and relatively high humidity say 95% which is required to meet satisfactory conditions for the room being air conditioned, and Fig 8 illustrates how the invention can be used to achieve this very important result.

Fig 7 graphically illustrates the advantage of the invention. The psychrometric chart shows how the large difference between the outside and the almost constant chilled water temperature 32ºC. causes the outside air coil condition curve to follow a steep gradient downwardly to the left bringing it alongside the saturation line and reducing the humidity in this instance from 0.020 kg per kilogram of dry air down to 0.007. Return air coil condition curve has a much flatter slope, and therefore the return air heat exchanger depends on the assisted dehumidification from the outside air heat exchanger in this case where single fixed size return air heat exchangers with high face velocities were employed to fit into small spaces.

Table 4 indicates control of an outside air heat exchanger serving ten storeys totalling 33330 litres of chilled water per second, having a return air treatment means comprising at least ten return air heat exchangers each handling 10%, that is 3333 litres per second, serving an office building in Singapore.

Table 4 indicates details of the outside air heat exchanger including its performance at peak and part loads. The psychrometric charts of Fig 7 and Fig 8 indicate the coil condition curve path for peak and part load conditions respectively for the outside air heat exchanger of Fig 4.

The part load contribution of Fig 8 relates to a system where there is a bypass and risers valve assembly around the return air heat exchangers as illustrated in Fig 5.

Table 5 identifies performance citing four of the most relevant values of air conditioning under the same peak and part loads, comparing on the one hand the invention applied to a two stage return air heat exchanger low face velocity VAV system with a standard high face velocity VAV system.

It is a characteristic of the aforesaid LFV/HCV method that humidity does not need to be controlled, and with the

combination of the present invention, there is still no need for separate control of humidity, that being inherent in the design of the air conditioning system. However with this invention, the humidity variation ranges within the

conditioned space would be less when with the LFV/HCV systems and with multi-stages. In Table 5 such a configuration has been called "Prestige" design. In the Prestige design the chilled water flow rate is considerably reduced. As

indicated for a 100kW room sensible heat system it only required 5.08 litres per second at peak whereas the Standard design required for a 77.4 kW application is higher but still a reasonably low chilled water flow rate of 7.96 litres per second is to be further reduced because the simultaneous chilled water flow rate is less than the sum of the

individual return air heat exchanger flow rates and to be further reduced by the latent heat contribution from the outside air heat exchanger to help offset latent heat loads. When these reductions are considered the return air heat exchangers are indeed smaller and requiring smaller uses than existing systems. This contrasts when outside air and return air is pre-mixed prior to entering a common heat exchanger. For a 77.4 kW application it would require 13.1 litres per second of chilled water at peak load conditions.

Claims (24)

The claims defining the invention are as follows:

1. Air conditioning means comprising means dividing a dehumidifier of an air conditioner into an outside air heat exchanger, and return air treatment means comprising at least one return air heat exchanger; coolant flow conduits

connecting said heat exchangers, the configuration of said conduits and dehumidifier being so arranged that said coolant flows first through a heat exchange conduit of the outside air heat exchanger and subsequently through a heat exchange conduit of said at least one return air heat exchanger, and pump means having capacity to pump said coolant at a flow rate through the outside air heat exchanger sufficient to meet maximum return air heat exchanger simultaneous demands in heat and mass transfer from the moist air passing through the return air treatment means at peak load, control means controlling said flow rate through each said return air heat exchanger to be responsive to water temperature rise which is maintained low enough to result in a room relative humidity which does not exceed maximum permissible design limits over full operating range, air flow directing means which direct outside air to flow over the heat exchange conduit of the outside air heat exchanger, and which direct return air to flow over the heat exchange conduit of the return air heat exchangers and subsequently mix said air flows, such that maximum temperature differential between air flow and coolant occurs in the outside air heat exchanger.

2. Air conditioning means according to claim 1 wherein said conduits directly couple said outside air heat exchanger to said return air heat exchanger assembly such that all coolant flowing through said outside air heat exchanger also flows through said return air treatment means.

3. Air conditioning means according to claim 1 wherein said return air treatment means comprises a plurality of said heat exchangers each separately but directly coupled by said conduits to said outside air heat exchanger and coolant throttling means separately throttling coolant flow to each return air heat exchanger in response to separate temperature control means, said coolant flowing through said outside air heat exchanger being divided to flow through said return air heat exchangers.

4. Air conditioning means according to any one of claims 1 to 3 wherein said return air treatment means comprises control means sensitive to leaving temperature of mixture of air cooled separately by said outside air and. return air heat exchangers, there being no control means directly coupled to vary coolant temperature of the outside air heat exchanger in response to load variations.

5. Air conditioning means according to any preceding claim wherein said outside air heat exchanger comprises a bypass conduit and valve assembly which bypasses coolant not

demanded by said return air treatment means during part load conditions thereby allowing the dehumidification of said outside air to be maintained at a level required to achieve the desired humidity ratio in the conditioned space while maintaining the design sensible cooling required in said return air heat exchangers.

6. Air conditioning means according. to any preceding claim wherein said return air treatment means comprises a plurality of said return air heat exchangers in parallel array with respect to coolant flow and further comprising respective control means controlling coolant flow rate through each said return air heat exchanger independently in response to temperature variation.

7. Air conditioning means according to any preceding claim wherein said return air treatment means comprises a plurality of separate return air heat exchangers, coolant flow conduits joining the return air heat exchangers in an array functionally parallel with respect thereto, and also joining the outside air heat exchanger in a configuration wherein all coolant flow to the return air heat exchangers is from a downstream side of the outside air heat exchanger and all coolant flow from all air handling units is through at least one chiller to an upstream side of the outside air heat exchanger,

pumps in a circuit of said conduits effective in causing said flow,

a bypass conduit and valve assembly between said

downstream and upstream sides of the outside air heat

exchanger, said assembly comprising a pressure responsive valve operative in said bypass conduit,

and a variable flow pump operative between said outside air heat exchanger and the return air heat exchangers.

8. Air conditioning means according to claim 7 wherein said pressure responsive valve maintains pressure between the upstream and downstream sides of the return air heat

exchanger assembly, and the speed of said variable output pump is controlled to be effective in controlling coolant flow rate through said return air heat exchangers in response to demand thereof.

9. Air conditioning means according to any preceding claim wherein said outside air heat exchanger comprises a coolant coil not more than six rows deep, and further comprising a coolant flow pump which pumps sufficient coolant through said coil that temperature rise of said coolant during its

traverse through its said coolant coil does not exceed 3ºC. when the air conditioner operates in the design range of temperature and humidity of the air conditioner.

10. Air conditioning means according to claim 9 wherein said outside air heat exchanger coolant coil comprises a plurality of circuits which are in parallel array with respect to coolant flow; and no said circuit comprises more than five passes of coolant flow conduit connecting the upstream to the downstream side of said coil.

11. A method of air conditioning a room comprising

(a) pumping coolant through heat exchange conduits of an outside air heat exchanger,

(b) passing outside air through the outside air heat exchanger and over said heat exchange conduits thereof to both cool and dehumidify the air and impart a temperature rise to the coolant which does not exceed 3ºC.,

(c) pumping said coolant from a downstream side of said outside air heat exchanger through heat exchange conduits of at least one return air heat exchanger comprised in return air treatment means and

separate from the outside air heat exchanger,

(d) recirculating some but not all of the air of said room as return air through said return air heat exchanger to at least cool said return air, and

(e) mixing the outside air and return air after both have passed over their respective said heat exchange conduits, and air conditioning said room with said mixed air.

12. A method according to claim 11 wherein said coolant is pumped at a temperature and flow rate which cools said air passing through the outside air heat exchanger below the dew point of the supply air to the room.

13. A method according to claim 11 or claim 12 comprising controlling the coolant flow rate in step (c) by thermostat means to be at a leaving temperature from each said return air heat exchanger to obtain a design supply temperature of air mixed in step (e).

14. A method according to any one of claims 11, 12 or 13 comprising controlling the coolant flow rate in step (c) by effecting said coolant pumping with a variable delivery pump.

15. A method according to any one of claims 11 to 14

comprising controlling coolant flow rate through each said return air heat exchanger in step (c) by throttling the flow rate with a throttling valve, and controlling the throttling valve with a thermostat located in said mixed air.

16. A method according to any one of claims 11 to 15

comprising pumping coolant in accordance with step (a) simultaneously through a plurality of heat exchange conduits in said outside air heat exchanger in circuits which do not comprise more than five passes of coolant.

17. A method according to any one of claims 11 to 16

comprising controlling entry temperature of said coolant in step (a) to be not more than 9ºC., and controlling coolant flow rate in step (c) such that coolant temperature rise does not exceed 10ºC. in any one heat exchanger comprised in said return air treatment means.

18. A method according to claim 11 further comprising controlling the return air heat exchanger to a plurality of stages to vary with dehumidification requirements, and pumping said coolant at a temperature and rate which does not necessarily cool the air passing through the outside air heat exchanger below the dew point of the supply air to the room.

19. A method according to any one of claims 11 to 17 wherein said coolant is pumped through the outside air heat exchanger at a temperature and a flow rate which cools said outside air passing therethrough below the dew point temperature of the room supply air, and is pumped through the return air heat exchanger at a temperature and a flow rate which cools and dehumidifies the return air to offset the room load thereof sufficiently that said mixing of outside air and return air fully offsets the room load.

20. A method according to any one of claims 11 to 19 wherein said coolant is pumped at a sufficiently low temperature and sufficiently high flow rate through said outside air heat exchanger to constrain said outside air coil condition curve to closely follow the saturation line of a psychrometric chart to a dew point temperature below the dew point of supply air to the room,

and said coolant is pumped through said return air heat exchanger at a sufficiently low temperature and sufficiently high flow rate to terminate the coil condition curve at a condition wherein said mixing of outside air and return air fully offsets the room load.

21. A method according to any one of claims 11 to 20

comprising bypassing surplus coolant not demanded by said return air treatment means from the downstream side of the outside air heat exchanger, through a chiller, and back to the upstream side of the outside air heat exchanger during part load conditions to thereby maintain a high

dehumidification level of said outside air heat exchanger over system full operating range.

22. A method according to claim 21 comprising effecting said bypassing through a bypass conduit and valve assembly which includes a pressure responsive bypass valve to maintain a high coolant flow rate through the outside air heat

exchanger.

23. A means substantially as hereinbefore described with, reference to and as illustrated in the accompanying drawings.

24. A method substantially as hereinbefore described with reference to and as illustrated in the accompanying drawings .