HVAC Hydronic System with Split Buffer Tank for Zero-Mixing System Operation

Abstract

The present invention relates to HVAC-systems operating under new ZERO-MIXING (ZM) water flow condition as innovative way to promote consistent highly energy efficient performance on SOURCE-heating/cooling thermal production and BUILDING's system distribution (FIG. 1). ZM technology is applicable; but not limited, to large-residential, commercial, institutional, and industrial facilities. Current state on HVAC technology, for system hydronics loop-flow, do not provide flows temperature segregation mechanisms between heating/chiller-plants hot/cold water supply and warmer water system returns. The result, a system that continuously operates at WATER MIXING conditions that impair equipment efficiency and output, and therefore, overall system energy performance.

Description

FIELD OF INVENTION

The present invention relates generally to heating and cooling systems, and more particularly to hydronics for heating, cooling and ventilation (HVAC) systems that use a water-based working fluid to fulfill heating/cooling loads of a building and/or heating/cooling needs of an industrial process.

BACKGROUND OF THE INVENTION

Despite technology advancements in heat-power/cooling generation and distribution modern building's HVAC equipment rarely achieve lab-test high efficiencies during actual lifetime operation. As a result, even the most renowned LEED¹/ZCB²-certified buildings barely hit the 70% Annual Fuel Utilization Efficiency (AFUE³) mark. Today, modern cities' infrastructure, regardless of retrofit-upgrade opportunities, continue running under an inefficient and ineffective HVAC platform that allows for unnecessary spending of non-renewables and greater production of GHG emissions. 1 Leadership in Energy and Environmental Design. 2 Zero Carbon Building Standards. 3 Annual Fuel Utilization Efficiency. In short, the AFUE indicates, for each dollar spend on energy for heating by gas, oil, or another fuel, just how much of it shows up inside the occupied space of the building as heat.

As architects and HVAC engineers remain challenged by the existing gap between building design and actual energy performance many professional associations for science and technology, seeking to advance heating, ventilation, air conditioning and refrigeration systems design and construction, continue prototyping and experimenting with new resources to better understand system behaviour.

There is today a continuous struggle to develop the ultimate system that can effectively integrate energy-production with system-distribution, that can also perform at maximum capacity while excelling at part-loads. Yet, under existing design practices and the current state of technology for equipment connectivity (hydronics), favoring system WATER-MIXING conditions, a technological leap to close the gap is needed.

This is the case for building HVAC systems running with condensing boilers with lab-reported high efficiency ratings C≈98%) that once integrated into the system are forced by hydronics to take a life of their own with unexpected consequence on overall system performance. Before sold in the North American market highly efficient condensing boilers are tested and certified under ANSI⁴/AHRI⁵ standards. ANSI Steady State Efficiency Test (SSET) simulates artificial conditions by regulating parameters such as fuel, air-intake temperature/volume, air/gas mixture, water/brine temperatures, boiler entering flow, etc., all fixed during boiler firing to ensure that boiler efficiency performance is rated at steady state regime (FIG. 2). SSET certified efficiencies become the basis for market technical information and set the standards for product competitiveness. This is done even though lab conditions are rarely achievable and do not properly relate to actual operating conditions. 4 American National Standards Institute. 5 Air-conditioning, Heating, & Refrigeration Institute.

FIG. 3. shows typical boiler efficiency curves for a condensing boiler with condensing operation limit at water return temperature below 57° C. Condensing boilers achieve high efficiency by condensing water vapours in the exhaust gases thus recovering its vaporisation latent heat, otherwise wasted. They can provide significant energy savings due to operating efficiencies as high as 98%, as compared to a peak efficiency of 80% for conventional boilers. Condensing boilers required low Entering Water Temperatures (EWT<57° C.) to realize the advertised efficiency and AFUE performance. ANSI specifies condensing-boiler performance testing with 80° F./27° C. EWT and a ΔT=100° F./55° C. (temperature rise), a condition not generally achieved in the field with current operation practices.

Complications during design phase confront engineers with issues arising from differing boiler and building heating operating conditions. There are system conditions which arise from optimal performance of the heating system's terminal units, and there are system conditions which are optimal for efficient boiler plant performance. These two sets of conditions are rarely, and perhaps only accidentally, the same. What is optimal for the former is usually not optimal for the latter, and vice versa. Attempts by designers to accommodate the needs of the boiler when designing system can, and often do, compromise the performance of their systems. Ignoring the needs of the boiler creates short-cycling⁶, and the energy lost from it often serves to undo the gains made by state-of-the-art system designs. It is as though there are really two systems being designed, the heating system and the boiler plant, which requirements are always different, usually different enough to make differences irreconcilable. 6 Boiler short-cycling consist of a firing interval, a post-purge, and idle period, a pre-purge, and return to firing. During these events, the boiler may exhaust through the chimney between 3% to 7% its hourly boiler output

Despite host of information available on the web from boiler factory representatives, engineers, and commissioning agents, recommending practices and discussing design issues, specifying a condensing boiler does not guarantee achieving the expected savings since condensing boilers only operates in the condensing range from time-to-time. Manufacturers and engineering challenges become building owner's problem when they are sold on the idea that installing a condensing boiler will grant the building the opportunity to achieve high operating efficiencies, with highly expected savings, ignoring that WATER-MIXING is the real problem impairing boilers and system performance.

FIG. 4. shows a schematic representation of a typical HVAC-system running under WATER-MIXING conditions. Modern systems use a variety of heating/cooling SOURCES to produce the necessary heat/cold energy output for building comfort. Many options include condensing and non-condensing boilers, air/ground or water heat-pumps, solar thermal panels, Combined Generation and Power (CHP), or chillers for cold water production, among others. In any case and depending in the service configuration, for BUILDING heating or cooling ancillary equipment (such as: radian baseboard, convective baseboard, Fan-Coil (FC), Domestic Hot Water (DHW), Make-up Air units (MUA), Unit Heaters (UH) or industrial batch-process heating), they produce a thermal load that is distributed through primary-secondary piping network (hydronic loops) via water/brine media. In this process, pipe water is used as thermal-mass to store transitory loads in the system.

Thermal buffer in FIG. 4. (middle) represents the system thermal-mass accumulated into the building hydronics during reheating/cooling operation rather than an actual thermal reservoir/buffer, with temperature Tmix used to control SOURCE output.

In the case of condensing-boiler heating (SOURCE), hot water BWS is produced and pump into system hydronics/thermal-mass. Warmed water SWR, returning from BUILDING's ancillary MAU/UH/others, encounters with BSW during the reheat exchange process. A major problem with existing commercial systems resides in the fact that hydronics allow loop water to recirculate without any ability to separate/redirect the encountering of flows that have very dissimilar thermal characteristics (BWS-hot or SWR-warmer conditions). The lack of a mechanical barrier to control the encounter of hot and warm flows inside hydronics allows the WATER-MIXING phenomena to happen.

FIG. 5 shows a typical HVAC heating system configuration with primary-loop made up of boiler-1 and 2, and secondary system (secondary-loop) with floors' fan-coils FC1, FC2 and FC3. When heat is required by building ancillary FC1, FC2, or FC3, hot-water is produced by boiler(s) B1/2 and pumped into the hydronics system via common header. Pump P3 takes the heat-load from common header and distributes it throughout the building floors via FC1, FC2, and FC3. Warmed-water from FC1, 2, and 3 returns back to common heater for reheating. It is in common heater mixing point (FIG. 5, atop) that warmer-water from the secondary loops mix with hot-water from the boiler's primary-loop. The mixing process will continue until rising temperature in the thermal-mass reaches system temperature setpoint (@ T-Tmix=82° C.) and FC(s) stop calling for heat, in response to floor thermostat. Boiler B1/2 may reinitiate operation once FC(s) have exhausted piping thermal load with Tmix falling below Temperature sensor “T” minimum range. Boiler-Building resulting water mixing temperature “Tmix.” Can be estimated as:

$\mathrm{Tmix}=\left(BSW+S\mathrm{WR}\right)/2@\mathrm{mixing}\mathrm{point}\mathrm{in}\mathrm{FIG}.-5.$

In a typical system BSW is determined by the Outdoor Control Reset (OCR⁷) based on outdoor weather conditions, while SWR is determined by building's ancillary design. 7 Theoretically, the purpose of outdoor reset is to reduce energy use and cost without sacrificing comfort. The control system lowers the boiler water temperature when the outdoor temperature is warmer and increases it when the outdoor temperature is colder

As water mixing temperature Tmix. increases, during the reheating process, and circulates back to boilers, boilers condensing capability start diminishing, see FIG. 3. At BWR=57° C. boiler(s) begin to lose its ability to capture additional heat from combustion flue gases water-content and thermal efficiency plunges by 10.12% at 82° C. Beyond BWR=57° C. expensive condensing boilers will behave as a regular inexpensive non-condensing one. Condensing boilers required low BWR/EWT to realize the advertised efficiency and AFUE performance.

Contrary to ANSI/AHRI-SSET high efficiency test Standard⁸ for boiler optimal output, set at EWT=26.7° C. and LWT=80° C., with ΔT=LWT-EWT=53.3° C. (see FIG. 2), system WATER-MIXING conditions force building operators to run boilers at LWT-EWT=20° C. 8 ANSI Z21.13/AHRI Standard 1500-2015 for Steady State Efficiency Test [SSET].

Reset-Differential (BRD), constraining not only hot water production/delivery capacity but also impairing overall system performance. Primary and secondary pumping design capacities are also forced to oversized resulting in higher energy consumption, larger runs, with greater equipment wear and tear.

Laboratory test of Applicant's Split Buffer Tank (SBT) (previously patented examples of which are disclosed in Applicant's Canadian Patent CA2701528 and U.S. Patent U.S. Pat. No. 8,997,511, each of which is incorporated herein by reference in its entirety) proved that, under system WATER-MIXING conditions, increasing BRD beyond 20° C. only exacerbate the system performance problem due to thermal-mass reduction. And, that boiler operation below BRD=20° C. constrains boiler output with no added benefit to system performance either.

On the flip side, as water flow temperature Tmix going into building's FCs increases, from setback lower temperature condition, building's FCs efficiency and heat output ramp-up. FCs max output is reached just and point when boiler(s) is shutoff by system setpoint temperature “T”.

FCs' heat output Q can be calculated with the use of equation 1 and 2 below. See FIG. 6.

$\begin{array}{cc}Q=U\times A\times \mathrm{MLTD}.& \left(1\right)\end{array}$ $\begin{array}{cc}\mathrm{MLTD}=\left(\Delta T1-\Delta T2\right)/\ln \left(\Delta T1/\Delta T2\right)& \left(2\right)\end{array}$

Where:

Q     Rate of heat transfer in the heat exchanger between system water
      supply and building room-air.
U     Overall heat transfer coefficient,
A     Heat transfer surface area in the heat exchanger,
MLTD  is the log mean temperature difference, calculated from the inlet
      and outlet temperatures of both fluids for room-air entering/
      leaving temperature and water entering (Tmix.)/leaving water
      temperature. See FIG.-6.
ΔT    Entering/leaving temperature differential between the two
      flows (Room air/water supply).
ΔT1   twin − taout = Tmix. − taout.
twin  Fan-Coil Water Supply Temperature. Rising with Tmix.
twout Fan-Coil Water Return Temperature. Leaving heat exchanger -
      going back to boilers for reheating.
tain  Fan-Coil Air Entering Temperature. From conditioned space.
Taout Fan-Coil Air Leaving Temperature. To conditioned space.

Equation-2 temperature differentials between hot-water supply and warmer-air stream ΔT1 and ΔT2 determines the output for the FCs' heat-exchanger and therefore the efficiency of the equipment. Since system WATER-MIXING is continuously diluting the thermal-mass temperature Tmix, in an up-and-down fashion, it forces FCs output to fluctuate from optimal design point, forcing secondary-system pumps and fans to work for longer period of time at partial loads. This, causing higher electricity bills and greater equipment wear and tear.

System WATER-MIXING, resulting from hydronics evils, is the worst enemy to overall HVAC system performance. Constant changing Tmix affects the system in many ways, including:

• • Overall HVAC-system inefficient equipment operation,
  • Overall HVAC-system reduced equipment output operation,
  • Heating/Cooling plant cycling (temperature controlling),
  • Heating/Cooling plant-building system water supply/return looping upset,
  • Heating/Cooling plant, building ancillary equipment oversize design.

WATER-MIXING damaging conditions are present in every HVAC application, also including air/ground heat-pumps, solar-thermal, hot-water production (DHW), chilled water, Combined Heat and Power (CHP), and other industrial processes, involving fluid reheating and cooling. It affects overall system efficiency and equipment output in the same manner.

Applicant has invented novel hydronic HVAC systems addressing such shortcomings of conventional hydronic HVAC design.

SUMMARY OF THE INVENTION

According to one aspect of the invention, there is provided a hydronic HVAC system comprising:

• • a heating/cooling source for heating or cooling a working fluid;
  • a split-buffer tank comprising a vessel in which there is contained a separation disk that divides an interior space of the vessel into separate supply and return chambers on opposing sides of said separation disk for respective holding of differently temperatured volumes of the working fluid in isolated fashion from another within said separate supply and return chambers, said separation disk being freely movable back and forth in an axial direction of the vessel to vary the relative sizes of said supply and return chambers and thereby accommodate volumetric variation between said differently temperatured volumes of the working fluid;
  • a primary fluid loop and a primary pumping system installed therein in a manner operable to pump the working fluid through the heating/cooling source from the return chamber of the split buffer tank to the supply chamber of the split buffer tank;
  • a secondary fluid loop and a secondary pumping system installed therein in a manner operable to pump the working fluid from the supply chamber of the split buffer tank to at least one heating/cooling unit that uses the working fluid for heating/cooling purposes, said secondary fluid loop comprising at least one return line through which the working fluid is returnable from at least one heat/cooling unit to the return chamber of the split buffer tank;
  • at least one control valve installed in said at least one return line, said at least one control valve being settable into a plurality of different operating states, including at least a recirculation state operable to recirculate said working fluid back through said secondary fluid loop, and a bypass state operable to return said working fluid back to the return chamber of the split buffer tank; and
  • a control system configured to operate the secondary pumping system, thereby causing circulation of the working fluid through the secondary fluid loop, and during said circulation of the working fluid through the secondary fluid loop, operate the at least one control valve by performing at least the following steps on an ongoing basis:
    • (a) monitoring a fluid return temperature of the working fluid in said at least one return line;
    • (b) determine whether the fluid return temperature fulfills a targeted minimum temperature differential relative to an output temperature setpoint of the heating/cooling source to achieve operating efficiency thereof in an optimal range;
    • (c) when the return temperature of the working fluid fulfills said targeted minimum temperature differential, set or maintain said at least one control valve in the bypass state; and
    • (d) when the return temperature of the working fluid does not fulfill said targeted minimum temperature differential, set or maintain said at least one control valve in the recirculation state.

According to another aspect of the invention, there is provided a method of controlling hydronic heating or cooling, said method comprising:

• • having hydronic HVAC system comprising:
    • a heating/cooling source for heating or cooling a working fluid;
    • a split-buffer tank comprising a vessel (1) in which there is contained a separation disk (2) that divides an interior space of the vessel into separate supply and return chambers on opposing sides of said separation disk for respective holding of differently temperatured volumes of the working fluid in isolated fashion from another within said separate supply and return chambers, said separation disk (2) being freely movable back and forth in an axial direction of the vessel to vary the relative sizes of said supply and return chambers and thereby accommodate volumetric variation between said differently temperatured volumes of the working fluid;
    • a primary fluid loop in which the heating/cooling source is fluidly connected between the supply and return chambers of the split buffer to receive said working fluid from the return chamber and heat/cool said working fluid, before loading thereof into the supply chamber of the split buffer;
    • a secondary fluid loop in which at least one heating/cooling unit is fluidly connected between the supply and return chambers of the split buffer to receive said working fluid from the supply chamber and use said working fluid to address a heating/cooling load, before returning said working fluid to the return chamber;
  • during circulation of the working fluid through the secondary fluid loop, performing the following steps:
    • (a) monitoring a fluid return temperature of the working fluid in the secondary fluid loop at a location downstream of the at least one heating/cooling unit and upstream of the return chamber of the split buffer;
    • (b) determining whether the fluid return temperature fulfills a targeted minimum temperature differential (ΔT) relative to an output temperature setpoint of the heating/cooling source to achieve operating efficiency thereof in an optimal range;
    • (c) when the fluid return temperature fulfills said targeted minimum temperature differential (ΔT), returning the working fluid to the return chamber of the split buffer; and
    • (d) when the fluid return temperature does not fulfill said targeted minimum temperature differential (ΔT), recirculating the working fluid back through the secondary loop.

According to yet another aspect of the invention, there is provided split buffer tank for storing temperature-distinct volumes of working fluid therein as a buffer between primary and secondary fluid loops of a hydronic heating/cooling system, said split buffer tank comprising:

• • a vessel having first and second ends that are spaced apart from one another in an axial direction of the vessel, and that are situated oppositely of one another across an interior space of the vessel delimited between said first and second ends, and a circumferential wall that spans axially between the first and second ends of the vessel and closes circumferentially around the interior space thereof;
  • a separation disk that is contained in said vessel and divides the interior space thereof into separate supply and return chambers on opposing sides of said separation disk for respective holding of differently temperatured volumes of the working fluid in isolated fashion from another within said separate supply and return chambers, said separation disk being movable back and forth in the axial direction of the vessel to vary the relative sizes of said supply and return chambers and thereby accommodate volumetric variation between said differently temperatured volumes of the working fluid;
  • a shaft spanning across the interior space of the vessel in the axial direction thereof, said shaft penetrating through the first and second ends of the vessel and also through the separation disk, which is slidably disposed around said shaft for sliding movement back and forth along the shaft;
  • a supply chamber inlet installed on the first end of the vessel to receive the working fluid from one or more heating/cooling sources in a primary loop of the hydronic heating/cooling system, said supply chamber inlet opening into the supply chamber of the vessel and also receiving a first end portion of the shaft that penetrates through the first end of the vessel, thereby externally supporting said first end portion of the shaft outside the vessel;
  • a return chamber inlet installed on the second end of the vessel to receive the working fluid back from a secondary loop of the hydronic heating/cooling system, said return chamber inlet opening into the return chamber of the vessel and also receiving a second end portion of the shaft that penetrates through the second end of the vessel, thereby externally supporting said second end portion of the shaft outside the vessel;
  • a supply chamber outlet on the circumferential wall at a location situated axially near the first end to enable supply of the working fluid from the supply chamber into the secondary loop; and
  • a return chamber outlet on the circumferential wall at a location situated axially near the second end to enable output of the working fluid from the return chamber into the primary loop.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 schematically illustrates a novel HVAC boiler system of the present invention, in which a split buffer tank (SBT) is installed between primary and secondary loops of the system with zero mixing (ZM) between the hot boiler water supply (BWS) and the cool system water return (SWR), and operates in conjunction with a bypass valve in the secondar loop to control a temperature differential in the SBT and thereby optimize boiler performance.

FIG. 2 schematically illustrates parameters of a boiler Steady State Efficiency Test (SSET) in accordance with the prior art.

FIG. 3 shows typical boiler efficiency curves for a condensing boiler of the prior art.

FIG. 4 shows a prior art example of a conventional HVAC boiler system in which mixing of the BWS and SWR occurs, to the detriment of overall system efficiency.

FIG. 5 shows a typical hydronic HVAC heating system of the prior art.

FIG. 6 shows a design chart of a typical fan coil heat exchanger of the prior art.

FIG. 7 schematically illustrates a novel SBT of the present invention, in which supply and return inlet fittings installed on opposite ends of the vessel also double as external shaft mounts for supporting a central guide shaft on which an internal separation disk of the SBT is slidable back and forth.

FIG. 8 is a more detailed schematic of the novel HVAC boiler system of FIG. 1.

FIG. 9a is a detailed schematic of a variant of the HVAC boiler system of FIG. 1, showing optional use of the SBT to serve multiple heating loads of different types in a zero-mixing fashion.

FIG. 9b shows another variant of the HVAC boiler system of FIG. 9a, illustrating how recirculation of working fluid in the secondary loop may cross from one heating circuit to better reduce overall SWR temperature and thereby optimize the temperature differential in the SBT.

FIGS. 10a & 10b show thermal-mass load profiles for a conventional water-mixing HVAC system and a novel zero-mixing HVAC system of the present invention, respectively. FIG. 10c illustrates comparative thermal load fractions of those two systems.

FIG. 11 shows an example of boiler water supply and system water return temperature curves for a conventional hydronic HVAC heating system of the type shown in FIG. 5 or 14.

FIG. 12 shows an example of boiler water supply and system water return temperature plots for the novel HVAC boiler system of FIG. 8.

FIG. 13 shows typical PD curves for primary-secondary or primary-only conventional HVAC heating systems in FIGS. 5 and 14. For a particular BIN temperature period it gives an indication of the additional purge loses caused by boiler short-cycling operation (during starts and stops) due to lack of thermal mass. This when compared to Zero-Mixing system.

FIG. 14 shows a typical primary-only hydronic HVAC heating system of the prior art, which lacks the secondary loop or the primary-secondary system of FIG. 5.

FIGS. 15a and 15b respectively illustrate temperature sensing configurations used in the novel HVAC systems of the present invention (FIG. 15a), and in the typical HVAC systems of the prior art (FIG. 15b).

FIGS. 16a & 16b illustrate conventional temperature sensing of condenser boiler water supply for domestic hot water or industrial purposes, both in a direct heating context (FIG. 16a) and an indirect heating context (FIG. 16b).

FIGS. 17a & 17b illustrate temperature sensing of condenser boiler water supply using an SBT in the zero-mixing context of the present invention, useful in both in the case of direct heating (FIG. 17a) and indirect heating (FIG. 17b).

FIG. 18 schematically illustrates a typical primary-only HVAC chiller system according to the prior art.

FIG. 19 schematically illustrates a novel HVAC chiller system of the present invention employing the same novel combination of the SBT and cooperating bypass valve as the HVAC boiler systems of FIGS. 8 and 9, thereby demonstrating optional use thereof in a variety of different heating/cooling systems.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS

Innovative ZM HVAC-system operating configuration in FIG. 1 offers a distinctive concept that provides intelligence to existing HVAC-system hydronics. It uses an SBT (FIG. 7) as system coupling-decoupling point for thermal exchange between system heating/cooling SOURCE and BUILDING distribution. During heating, hot-BWS flow can be stored and redirected into BUILDING secondary-loop where is need hot, as Cold-SWR is also stored and consistently supply back to the SOURCE where is needed cold. Orderly controlled BWS/SWR flows condition allow heating/cooling generation and building's distribution ancillary equipment optimal performance maximizing thermal load generation and delivery. ZM operation alone represents a dramatic change in the way in which HVAC-systems store and distribute heating/cooling loads involving flows with very dissimilar thermal characteristics. BWS and SWR flows thermal properties will never be weaken by detrimental system WATER-MIXING damaging operation. Similar exchange process takes place during system chilled water cooling operation.

FIG. 8 shows modification of the typical HVAC heating system in FIG. 5 so as to incorporate an SBT and run under the inventive ZERO-MIXING operation scheme described herein. Here, since boiler hot water BWS at 82° C. can now be produced and stored into the SBT with NO-MIXING, SWS high temperature water can be conveyed to BUILDING's FCs with no dilution/fluctuation and maximize FCs performance. Similarly, building SWR low temperature water-return going back to the boilers. Low temperature water bypassed through a control valve CV1 installed in a return line of the secondary loop is the key to boilers efficient runs described in FIG. 2. The lower the temperature T_SWR of the SWR, the closer boiler(s) B1, B2 is/are operating under SSET conditions.

Contrary to the conventional system in FIG. 5, where OCR output is used to control boiler BWS, OCR output in ZM systems may be used only to control SWR low temperature flow bypass to SBT through CV1, to maximize temperature differential between the two chambers of the SBT. At system design condition CV1 setpoint may be configured to bypass SWR below 38° C., for example as programmed into a programmable logic controller (PLC) that is responsible for control of CV1. The PLC receives the outdoor temperature from the OCR and based thereon sets an appropriate CV1 setpoint temperature. The PLC is also connected to an SWR temperature sensor at or near CV1, so that measured T_SWR can be compared against the current CV1 setpoint, above which the PLC will put CV1 into its bypass state feeding back to the SBT. If the measured T_SWR is below the current SWR setpoint, the PLC places CV1 in its recirculation state, whereby SWR flows at higher temperature are recirculated back to FCs through port-B of CV1 for additional loads release. Overall systems efficiency depends on the ability of designers to integrate SOURCE-BUILDING equipment to allow for higher ΔT (BWS-SWR) operation.

In a ZERO-MIXING multi-service system (FIG. 9a) a plurality of control valves CV1, CV2, and CV3 are installed in respective return lines of different fluid circuits in the secondary loop, each of which feed different types of ancillary equipment. Each control valve CV1, CV2, and CV3 is accompanied by a respective water temperature sensor installed in the same return line, whether at the valve or upstream thereof, but downstream from the ancillary equipment of that same circuit. Each such water temperature sensor is connected to the PLC or other controller responsible for the control of these valves CV1, CV2, and CV3. Each of the respective control valves CV1, CV2, and CV3 has its respective setpoint in the PLC finetuned to bypass low temperature water-return SWR back to the SBT without compromising unit output/room comfort. Given the cost of a condensing boiler, designer goal is to make system operate as efficient as possible with the lowest SWR.

In the first multi-service example (FIG. 9a), each control valve CV1, CV2, and CV3 feeds high temperature water-return SWR back into the same circuit from which it was received. In another multi-service system (FIG. 9b), CV1 in the building's heating panel (HP) circuit is instead forced to recirculate any high temperature return water back into the MUA circuit to reduce SWR temperature. This demonstrates how any circuit potentially subject to particularly high SWR temperatures relative to other circuits can have its respective control valve installed in a manner that, when put in recirculation mode, feeds into a different circuit of a lower SWR temperature.

FIG. 10a depicts the energy loading profile for the typical HVAC system in FIG. 5 operating under WATER-MIXING condition. Chart shows that system initiate operation recovering from standby at Tmix=27° C. ramping up BWS/BWR temperatures as water recirculate for reheating through the boiler(s), with boilers BRD=BWS-BWR=20° C., and into secondary-loop via common header. Boiler will stop operation when outlet water temperature BWS reaches 82° C. and activate internal Boiler Limit Reset Switch (BLRS). A customary boiler manufacturer recommendation. Dark grey triangular area represents energy/heat accumulated in system thermal-mass during system-cycle operation. Even though boiler(s) capability allows them to rise BWS temperature to 82° C. secondary-loop Tmix is limited to 62° C. due to constrains caused by BLRS setting of 20° C. (82° C.-62° C.). A common cause of WATER-MIXING operation in commercial systems (FIGS. 5,14, and 18).

FIG. 10b depicts the energy loading profile for the HVAC system in FIG. 8 operating under ZERO-MIXING condition. Chart shows that system initiate operation recovering from standby at BWR=27° C., with BWS ramping up to 82° C. during first 2 minutes operation. Boilers output remains steady with BSW=82° C., recirculating warm-water from a return chamber RC in the bottom half of the SBT to a supply chamber SC in the top half of the SBT, until a separation disk between those two chambers (FIG. 7) moves to the bottom of the tank, whereupon flow from the top supply chamber is bypassed to a return chamber outlet of the SBT, and back to the boiler(s). At this point, a temperature sensor T1 (FIG. 8) installed between the SBT's return chamber outlet and the boiler(s) reaches an output temperature setpoint T_sp of the boiler, as dictated by a boiler controller responsible for the boiler's operation.

In the illustrated example T_sp=82° C., and so when sensor T1 reaches setpoint condition T1=82° C., the boiler controller shuts off the boiler(s) while the SBT remains fully loaded with hot water at 82° C. Arrival of the separation disk at the bottom of the tank marks the end of the cycle (100% load). At any given time, if system calls for heat, SWS water is pumped from the top supply chamber SC of the tank and supplied to secondary-loop's FCs via hydronics system. Additional reference may be made to Applicant's aforementioned prior patents for further details on the SBT's sequence of operation). System low temperature water return SWR is then bypassed through CV1 and stored at the bottom of SBT, pushing upward the separation disk. As system SWR flow accumulate at SBT bottom, separation disk is pushed top the top of tank where it is bypass into valve T2, at SWS outlet, and low temperature SWR flow turn-on boilers. Pumps P1/P2/P3 can operate concurrently. System-OCR output may be used to control T2 setpoint or to reduce BWS. Dark grey rectangular area represents the energy accumulated in system thermal-mass during system cycle operation.

The ZERO-MIXING Plant Advantage.

FIG. 11 depicts Boiler-BWS and Building-SWR characterization curves for System WATER-MIXING operation in FIG. 5, 14. Here the BWS curve is shaped down by system's OCR while SWR curve underlines water temperature back to the boiler. Along supply curve BWS and at any given outdoor weather condition (depicted by BIN temperature histogram in the background, the city of Calgary used as example), boiler capacity is subject to output limitations due to the interaction between BRD (20° C.) and BLRS (82° C.). At maximum design conditions (−33° C. DB), boilers-1 and 2, coming online from standby/setback mode with system water Tmix at 27° C., will only be able to raise the system's water temperature to 62° C. (82° C.-20° C.) for a boiler output constrain of =36% [(62° C.-27° C.)/(82° C.-27° C.)] (see also FIG. 10a). If non-condensing efficiency factor is also considered, since SWR>57° C. during BIN 20° C./−30° C. and boilers condensation stops at point C, additional boiler output drops by ˜10.12%, for real boiler maximum output limited to ˜54%. A cutdown capacity of 46% from claimed manufacturer output due to mixing operation. Not to be confused, manufacturers are right on boiler's lab-output from the SSET-test. It is actually operation at mixing conditions what change boilers output.

Boilers' output drawbacks are the result of system hydronics favouring WATER-MIXING in the reheating/delivery process and the unavoidable boiler/system-running scheme with BRD=20° C., reducing gradually to zero at BIN 20° C./10° C. To overcome such deficiencies engineers are forced to make uneconomical decisions and overdesign plant-capacity to makeup for the lost of it. As weather conditions improve, system rides down the BWS-SWR curves, with boiler output capacity points AB reducing even lower to A′B′. At BIN 10° C./0° C. Boilers may find themselves running for longer time at intermittent operation (short-cycling). New ZERO-MIXING HVAC-system (FIG. 8) runs under a quite different BWS/SWR flat patter (FIG. 12) maintaining a constant BWS=82° C. while returning SWR=30° C. and superior boiler output for any given outdoor temperature (BINS −30° C./−40° C. to 20° C./10° C.). Since SWR is always maintained below 57° C., boiler's performance emulates SSET parametric conditions leading to much higher annual system AFUE values. Estimated boiler's output for any BIN can be calculated with the help of equation below as:

$\begin{array}{cc}{Q}_b=4,200\times \mathrm{FLOW}\times \left(\mathrm{BWS}-\mathrm{SWR}\right)& \left(3\right)\end{array}$

Qb   Boiler Output Rate (kWatts)
FLOW Boiler(s) Flow (m3/sec)
BWS  Boiler Water Supply (° C.)
SWR  Building/System Water return (° C.)

At systems design condition C−33° C. DB, FIGS. 11 and 12, typical condensing boiler output can be calculated as:

$\begin{array}{c}{Q}_M=4,200\times \mathrm{FLOW}\times \left(20°C.\right)\left(\mathrm{WATER}-\mathrm{MIXING}\mathrm{Operation}\right)\\ Q_Z=4,200\times \mathrm{FLOW}\times \left(44°C.\right)\left(\mathrm{ZERO}-\mathrm{MIXING}\mathrm{Operation}\right)\end{array}$

Based on boiler output comparison Q_M/Q_z=0.5 it can be concluded that, boiler output Q_M gets reduced by 50% due to mixing operation. As building weather conditions move into milder summer periods Q_M output reduces progressively, with Q_M=25% Q_z at BIN 10° C./0° C. Boilers continuous run at near 100% capacity, under ZERO-MIXING operation, reduces time operation with longer standby periods at milder weather conditions. Since return water is always maintained below 57.2° C., boiler's performance emulates SSET parametric conditions leading to much higher annual system AFUE values, improving building OPEX from fuel savings, carbon tax levy reduction, maintenance and early replacement cost from continuous wear and tear.

The SBT/Building Thermal-Mass Advantage.

Today, HVAC system are designed for loads that rarely occurred (˜2% of seasonal operation), with boilers spending nearly 7,400 hrs (98%) running at partial loads (between bins 20° C./10° C. to −10° C./−20° C.) to overcome smaller building seasonal heat loses due to milder weather conditions. Since in the absence of thermal buffers or larger boiler's thermal-mass, stored piping water thermal-mass is the last resource for a system starving from heat, remaining heat in hydronics is rapidly withdrawn until water temperature condition falls below the system's control system setpoint “T”, forcing boilers to fire. At milder outdoor temperatures (bin 10° C./0° C., see FIG. 11) boilers fire at a fraction of the design load indicated by the temperature differential between points A′B′. Such a small output, in comparison with boiler output at ZERO-MIXING conditions (FIG. 12) determined by point C′D′, along with the low piping water thermal-mass, causes boilers to run in an on-off fashion (short-cycling) in their effort to serve building loads. Each short-cycling is accompanied by a small pre-purge⁹ and a post-purge chimney discharge, with boiler wasting between 3% to 7% of hourly heat output. 9 Pre-purge/Post-purge. Boiler's burner cycle before it lights-off to prove system safety and purge out the combustion chamber. post-purge comes at the end of the burner cycle to purge out the combustion chamber before the fan shuts off.

FIG. 10c shows WATER-MIXING/ZERO-MIXING energy loading operation, both compared at system design condition (−33° C. DB). Experimental data results showed that ZERO-MIXING system took about ⅓ of the WATER-MIXING system hour-operation to reach 100% load, while a WATER-MIXING conventional system took an hour fraction to barely achieve 54%. In comparison, ZERO-MIXING complete 100% load in one hour cycle while conventional system had an additional run to complete 54% of the hourly load, with two purge differential and additional heat loses estimated at 10%. See calculation below:

Purge Differential-PD=4 Purge (WM)−2 Purge (ZM)=2×5% average heat loss=10% Purge Differential (PD) dashed-line in FIG. 13 depict incremental purging events for system in FIG. 5 when compared to more efficient ZERO-MIXING system in FIG. 8. Note that for the low-mass Primary-Only system (see grey dashed line in FIG. 13), more boiler on-off runs with additional pre/post-purges will be necessary to deliver the same required building loads.

WATER-MIXING/ZERO-MIXING System Purge-Differential can be estimated for some BIN as:

BIN 10° C./0° C.	C″′D″′/A″′B″′=	5.2 PD 3,215 hr.,
		36% heating operation.
BIN 0° C./−10° C.	C″D″/A″B″=	3.5 PD 1,925 hr.,
		22% heating operation.
BIN −10° C./−20° C.	C′D′/A′B′=	2.7 PD 685 hr.,
		8% heating operation.
BIN −20° C./−30° C.		2.2 PD 158 hr.,
		1.8% heating operation.
BIN −30° C./−40° C.	CD/AB=	2.0 PD 4 hr.,
		0.04% heating operation.

Looking at year AFUE analysis for the WATER-MIXING system in FIG. 5, boiler hourly non-condensing loses of 10%, during 162 hr./1.84% seasonal operation (BINs −30° C./−40° C. and −20° C./−30° C.), are overshadowed by small but repetitive 5% hourly heat loses due to boiler short-cycling, at BINs 10° C./0° C. to −10° C./−20° C. In BIN 10° C./0° C. alone, hourly loses can be estimated at 26% (5.2 PD×5%) for 36% (3,215 hr) of the seasonal period. Analysis above supports the argument that boiler short-cycling is the major threat to system performance and that primary objective on HVAC design should focus on system thermal-mass addition, rather than creating favorable conditions for boiler condensation.

In the Primary-only system in FIG. 14 the non-condensing efficiency problem associated to high SWR>57° C. has been eliminated by matching boiler(s) output to system loads. Here, system uses variable speed pumps to supply just the required flow to FC1/FC2 or FC3 to satisfy floor heat requirements. In this system, boiler output still controlled by the OCR and BRD −20° C./BLRS −82° C. (82° C.-62° C.), as explained by FIG. 10a and BWS-SWR curves in FIG. 11. This system characterizes by the lack of thermal-mass and boiler minimum flow requirements through control valve CV1.

When compared to ZERO-MIXING system, problem in the Primary-only arises with the system need to maintain minimum flow return to boiler(s)/system. To satisfy building minimum loads, most likely happening at BIN 10° C./0° C., point A′B′ in FIG. 11, the lack of additional thermal-mass to respond to small incremental system demands makes boiler(s) prompt to short-cycle. FIG. 13 shows system PD-Primary-only curve with PD=8 at BIN 10° C./0° C. Hourly loses in BIN 10° C./0° C. alone are estimated at 40% (8 PD×5%) for 36% (3,215 hr) of the seasonal period. The biggest threat to performance in this type of systems is posed by the lack of thermal-mass rather than system SWR high water temperature, resolved by the staging of system pumps VFDs and flow-control valves CV2 and CV3.

Short-cycling is also the result of system oversized heating plant with poor boiler(s) turndown ratio or an erratic multi-boiler system water temperature control system. In any case, either a low-mass system, boiler oversize, or water temperature controlling issues, loses small in size take a huge toll on annual boiler AFUE efficiency due to their reoccurring nature. This is usually neglected when analyzing the gap between building design and actual energy performance. No matter what the case is, short-cycling can be eliminated at once with the integration of the ZERO MIXING concept into any HVAC boiler system.

The Boiler/Building System Low-Flow Advantage.

Current HVAC design fundamentals on condensing boiler-systems (FIGS. 5 and 14) reinforce the practice of output settings that call for BRD=20° C., or lower. Despite manufacturer and trade suggested advantages lab test demonstrated that BRD limits hot-water production capacities while forcing the system to operate at unnecessary pumping higher flow volumes, on both SOURCE and BUILDING loops ΔT<20° C. limitation forces secondary system terminal units not only to work harder but to run for longer periods to provide building comfort; or in the case of industrial setting, limiting optimum process setpoint conditions, sacrificing production output and quality.

A WATER-MIXING/ZERO-MIXING flow comparison can be analysed through the formula below for a particular boiler output as:

$\begin{array}{c}{Q}_Z=Q_M\\ 4,200\times {\mathrm{FLOW}}_Z\times \left(44°C.\right)=4,200\times {\mathrm{FLOW}}_M\times \left(20°C.\right)\\ {\mathrm{FLOW}}_Z=50\%{\mathrm{FLOW}}_M\end{array}$

ZERO-MIXING system high temperature differential operation reduce by 50% pumping and air-handling flows without sacrificing boiler or ancillary equipment output. Savings are not limit to operational costs but also to capital investment for new facilities with reduction of otherwise oversize equipment.

The Boiler/Building System Control Advantage.

There are many ways to control a boiler and the boiler controls can be layered. A boiler's own controls can be set for standalone operation to maintain desired boiler setpoints, and also can be linked to a more complex Direct Digital Control (DDC) system for a multi-boiler multi-stage with more complex controlling operation. Customary use of DDC-OCR controls to manage boiler output, coupled with temperature controlled variable flow pumps to produce low temperature water return for boiler condensing opportunities, always result in water mixing in the common heater (FIG. 5). No matter what strategy is used, multi-stage boiler lead/lag, boiler lead/lag with demand or boiler multi-pump rotation/operational sequencing, boiler(s) at some point during operation are forced to overfiring (short-cycling) due to mixing.

HVAC ZERO-MIXING systems use SBT as the main point for system-loops coupling/decoupling. SBT on-stream sensing at the inlets/outlets of the tank (TS1/TS2/TS3/TS4, FIG. 15a) allows accurate temperature measurements and signaling back to the boilers and building's pumps. Since SOURCE supply and BUILDING return flows are contained within the SBT common-header temperature misreading/signaling, observe in conventional systems (FIG. 15b), will never occur.

FIG. 16a/16b shows a condensing boiler direct/indirect hot water operation for domestic hot water or industrial purpose. In the figure, typical sensor TSX location induces reading error due to instantaneous signaling on tank's isotherm rather than on the average temperature of the tank, understating or overstating the actual system load. In either case, boiler will always overfire trying to satisfy system demand. Proper positioning is a determining factor in attaining maximum buffer thermal capacity due to boiler dependency on sensor signaling for firing control. A sensor location too high may force boilers to overfire in order to reach temperature setting. Too deep into the tank, may reduce buffer volumetric/thermal capacity, increasing boiler cycling for particular loads. Finding the most appropriate spot is an impossible task, especially in systems where heat demand moves through a wide range of seasonal temperature variations, causing the buffer to operate from 26° C. during summer up to 62° C. during winter.

The SBT on-stream sensing system in FIG. 17a/17b allows instantaneous accurate temperature measurements and eliminates problems with TSX sensing lagged readings, favouring better temperature control, and storing greater thermal mass, for more stable energy production and distribution. On the flow dynamic SBT separation disk helps isolate primary and secondary loop streams from mixing inside the tank, doubling DHW capacity while reducing the use of costly non-renewables and associated CO2 emissions. SBT water storing temperatures can reach up to 82° C. compared to 62° C. for commercial buffers.

HVAC Cooling Primary-only System and WATER-MIXING Operation.

Conversion to ZERO-MIXING of a conventional HVAC-Cooling Primary-only system (FIG. 18) favours operation performance by eliminating compressor cycling, due to system minimum loads, and mixing in the bypass line thermal-mass. As shown in FIG. 19, SBT and cooperating control valve CV1 can be similarly installed and put to similar purpose as they are in the HVAC boiler systems of the earlier figures.

In operation, the difference is that the supply chamber SC holds cold water received from the chiller CH, and the return chamber RC holds warmer water received back from the ancillary equipment (e.g. FCs) in the secondary loop responsible for cooling the building. Accordingly, in this cooling application, instead of bypassing the SWR to the SBT when the SWR temperature T_SWR is below the CV1 setpoint stored in the PLC, the PLC instead sets CV1 to perform such bypass when T_SWR is above the CV1 setpoint. This is because the temperature differential in this cooling system context is one where a large temperature differential requires a higher T_SWR value, since SWR is warmer than SWS. Here, compressor runs happen at equipment stationary conditions sustaining equipment max capacity and efficiency. Utilization of SBT thermal-mass also allow system to take advantage of Free-cooling options.

As a novel alternative to either the conventional Primary-secondary loop system in FIG. 5 or the conventional Primary-only system in FIG. 14/18, the novel ZERO-MIXING/SBT operation of the present invention promotes:

• • Maximize overall HVAC-System AFUE,
  • Continuous SOURCE max output operation,
  • Continuous SOURCE max efficiency operation,
  • SOURCE short-cycling elimination,
  • System Thermal-mass addition,
  • Improved temperature sensing/controlling,
  • Improve ancillary equipment operation,
  • Seamless coupling/decoupling loops hydronic,
  • New approach to system design (50% lower flow operation).

While the HVAC systems disclosed herein may use the SBT disclosed in Applicant's aforementioned prior patents, FIG. 7 illustrates a novel design of the SBT that may alternatively be used in the presently disclosed HVAC systems. The SBT features a vessel 1 having first (e.g. top) and second (e.g. bottom) ends that are spaced apart from one another in an axial direction (e.g. vertical height) of the vessel, and that are situated oppositely of one another across an interior space of the vessel that is delimited between these first and second ends. A cylindrical circumferential wall spans axially between the first and second ends of the vessel and closes circumferentially around the interior space thereof.

A separation disk 2 is contained within the vessel and divides the interior space thereof into separate supply and return chambers (SC, RC) on opposing sides of said separation disk for respective holding the differently temperatured volumes of the working fluid in isolated fashion from another within these chambers. The disk is being movable back and forth (e.g. up and down) in the axial direction of the vessel along a central longitudinal axis thereof. Such disk movement varies the relative sizes of said supply and return chambers and thereby accommodates volumetric variation between the fluid pumped into the supply chamber from the boiler, chiller or other heating/cooling source, and the fluid returned from the secondary loop.

To help guide water-driven movement of the disk 2 and maintain the disk in perpendicular orientation to the vessel axis on which it moves back and forth, a guide shaft 7 spans across the interior space of the vessel on the central longitudinal axis thereof. The shaft penetrates centrally through both ends of the vessel, and also through the separation disk that is slidable back and forth along the shaft. An outer periphery of the disc has a seal that engages the interior surface of the vessel's circumferential wall to maintain a fluid tight, but slidable, relation therewith. This allows movement of the disk, while maintaining fluid isolation between the two chambers of the vessel's interior space to prevent water from leaking across the disk from one chamber to the other.

A supply chamber inlet 5 is installed on the first end of the vessel to receive the working fluid from the one or more heating/cooling sources in the primary loop of the hydronic heating/cooling system. This supply chamber inlet opens into the supply chamber of the vessel, and also receives a first end portion of the shaft that penetrates through the first end of the vessel. The supply chamber inlet 5 thus doubles as an external support for holding the first end portion of the shaft at a position outside the vessel. The fitting installed on the vessel to form this supply chamber inlet 5 thus makes use of a singular hole at a singular location in the vessel walls to both introduce the working fluid to the supply chamber, and support a respective end of the shaft. This also simplifies the tank design by avoiding the need to somehow install shaft-holding components inside the vessel.

Likewise, a return chamber inlet 6 is installed on the second end of the vessel to receive the working fluid back from the secondary loop of the hydronic heating/cooling system. This return chamber inlet opens into the return chamber of the vessel, and also receives a second end portion of the shaft that penetrates through the second end of the vessel. Like the supply chamber inlet, the return chamber inlet 6 thus doubles as an external support for holding the second end portion of the shaft at a position outside the vessel, once again reducing the overall quantity of holes in the tank walls, and simplifying the construction by avoiding internally mounted shaft supports.

Meanwhile, a supply chamber outlet is provided on the circumferential wall at a location situated axially near the first end of the vessel, and enables supply of the working fluid from the supply chamber into the secondary loop to feed the ancillary equipment (e.g. FC1, FC2, FC3, HP, MUA, and/or DHW) installed therein. Similarly, a return chamber outlet is also on the circumferential wall at a location situated axially near the second end of the vessel, and enables output of the working fluid from the return chamber into the primary loop to feed the heating/cooling source (e.g. B-1, B-2, or CH) during operation of the primary loop's pumping system (e.g. P1, P2).

Though the SBT may use the stops and disk bypasses described in Applicant's aforementioned prior patents, the illustrated example lacks such features, and relies on movement of the disk fully past either chamber's outlet to allow water from the other chamber to pass therethrough when that other chamber is filled to its maximum capacity.

While the detailed embodiments above make reference to a PLC or other controller connected to the control valve(s) in the return line(s) of the secondary loop to control operation of these valves based on a control valve setpoint, which may be a floating setpoint that varies with outdoor temperature read by a connected OCR, and use of a separate heating/cooling source controller (e.g. boiler controller) to control operation of the heating/cooling source based on temperature readings from sensors T1, T2 that measure fluid temperature from the supply and return chamber outlets of the SBT, it will be appreciated that these and any other control functions mentioned herein may be consolidated into a singular controller, or distributed among a greater number of discrete controllers, without affecting the overall function of the inventive system. Accordingly, the term control system is used herein to denote any quantity of electronic controllers programmed to carry out any and all operations described herein on an automated basis.

Claims

1. A hydronic HVAC system comprising: a heating/cooling source (B1-B2/CH) for heating or cooling a working fluid;a buffer for holding a thermal mass warmed or cooled by said heating/cooling source;a primary fluid loop and a primary pumping system (P1, P2) installed therein in a manner operable to pump a working fluid through the heating/cooling source from a return outlet of the buffer to a supply inlet of the buffer;a secondary fluid loop and a secondary pumping system (P3) installed therein in a manner operable to pump the working fluid from a supply outlet of the buffer to at least one heating/cooling unit (FC1, FC2, FC3, HP, MUA, DHW) that uses the working fluid for heating/cooling purposes, said secondary fluid loop comprising at least one return line through which the working fluid is returnable from at least one heat/cooling unit to a return inlet of the buffer;at least one control valve (CV1, CV2, CV3) installed in said at least one return line, said at least one control valve being settable into a plurality of different operating states, including at least a recirculation state operable to recirculate said working fluid back through said secondary fluid loop, and a bypass state operable to return said working fluid to the return inlet of the buffer; anda control system configured to perform monitoring of temperature conditions of the working fluid and controlled operation of the at least one control valve between the recirculation state and the bypass state depending on said temperature conditions of the working fluid.

2. The system of claim 1 wherein said monitoring of the temperature conditions and controlled operation of the at least one control valve comprises: (a) monitoring a fluid return temperature (TsWR) of the working fluid in said at least one return line;(b) determiningwhether the fluid return temperature (TSWR) fulfills a targeted minimum temperature differential (ΔT) relative to an output temperature setpoint (Tsp) of the heating/cooling source to achieve operating efficiency thereof in an optimal range;(c) when the return temperature of the working fluid fulfills said targeted minimum temperature differential, set or maintain said at least one control valve in the bypass state; and(d) when the return temperature of the working fluid does not fulfill said targeted minimum temperature differential, set or maintain said at least one control valve in the recirculation state.

3. The system of claim 2 wherein said heat/cooling source is a heating source (B1-B2), step (b) comprises determining whether the fluid return temperature fulfills the targeted temperature differential by checking whether the fluid return temperature is less than a control valve setpoint value, step (c) comprises setting or maintaining said at least one control valve in the bypass state when the fluid return temperature is less than the control valve setpoint value, and step (d) comprises setting or maintaining said at least one control valve in the recirculation state when the fluid return temperature is greater than the control valve setpoint value.

4. The system of claim 2 wherein said heat/cooling source is a cooling source (CH), step (b) comprises determining whether the fluid return temperature fulfills the targeted temperature differential by checking whether the fluid return temperature is greater than a control valve setpoint value, step (c) comprises setting or maintaining said at least one control valve in the bypass state when the fluid return temperature is greater than the control valve setpoint value, and step (d) comprises setting or maintaining said at least one control valve in the recirculation state when the fluid return temperature is less than the control valve setpoint value.

5. The system of any claim 2 wherein the control system is configured to variably adjust the targeted minimum temperature differential.

6. The system of claim 5 wherein the control system is configured to variably adjust the targeted minimum temperature differential based at least partly on an outdoor temperature.

7. The system of claim 3 wherein the control system is configured to variably adjust the control valve setpoint value.

8. The system of claim 7 wherein the control system is configured to variably adjust the control valve setpoint value based at least partly on an outdoor temperature.

9. The system claim 1 wherein the secondary fluid loop comprises a plurality of fluid circuits through which the working fluid is respectively delivered to a plurality of different heating/cooling equipment types (HP, MUA, DHW), the at least one return line comprises a plurality of respective return lines each belonging to a respective one of the fluid circuits, the at least one control valve comprises a plurality of respective control valves (CV1/CV1′, CV2, CV3) each installed in a respective one of said plurality of return lines, and the control system is configured to monitor a respective fluid return temperature in each of said return lines, and to control the respective control valve of each return line based on the respective fluid return temperature in said return line.

10. The system of claim 9 wherein each respective control valve (CV1, CV2, CV3), in the recirculation state thereof, is operable to recirculate the working fluid back through a same one of the fluid circuits in which said respective control valve is installed.

11. The system of claim 9 wherein at least one of the respective control valves (CV1′), in the recirculation state thereof, is operable to recirculate the working fluid through a different one of the fluid circuits that that in which said respective control valve is installed.

12. The system of claim 11 wherein the heating/cooling source is a heating source (B1-B2), and said at least one of the respective control valves is a first control valve (CV1′) installed in the respective return line of a first fluid circuit whose respective fluid return temperature is greater than a second fluid circuit into which the first control valve is operable, in the recirculation state thereof, to recirculate the working fluid.

13. A method of controlling hydronic heating or cooling, said method comprising: having a hydronic HVAC system comprising: a heating/cooling source (B1-B2/CH) for heating or cooling a working fluid;a buffer for holding a thermal mass warmed or cooled by said heating/cooling source;a primary fluid loop in which the heating/cooling source is fluidly connected between a supply inlet and a return outlet of the buffer to receive said working fluid from the return outlet and heat/cool said working fluid, before loading thereof into the buffer through the supply inlet;a secondary fluid loop in which at least one heating/cooling unit (FC1, FC2, FC3, HP, MUA, DHW) is fluidly connected between a supply outlet and return inlet of the buffer to receive said working fluid from the supply outlet and use said working fluid to address a heating/cooling load, before returning said working fluid to the buffer through the return inlet; andperform monitoring of temperature conditions of the working fluid and controlled operation of the at least one control valve between the recirculation state and the bypass state depending on said temperature conditions of the working fluid.

14. The method of claim 13 wherein said monitoring of the temperature conditions and controlled operation of the at least one control valve comprises: (a) monitoring a fluid return temperature (TSWR) of the working fluid in said at least one return line;(b) determining whether the fluid return temperature (TSWR) fulfills a targeted minimum temperature differential (ΔT) relative to an output temperature setpoint (Tsp) of the heating/cooling source to achieve operating efficiency thereof in an optimal range;(c) when the return temperature of the working fluid fulfills said targeted minimum temperature differential, set or maintain said at least one control valve in the bypass state; and(d) when the return temperature of the working fluid does not fulfill said targeted minimum temperature differential, set or maintain said at least one control valve in the recirculation state.

15. The method of claim 14 wherein said heat/cooling source is a heating source (B1-B2), step (b) comprises determining whether the fluid return temperature fulfills the targeted temperature differential by checking whether the fluid return temperature is less than a control valve setpoint value, step (c) comprises returning the working fluid to the return chamber of the split buffer when the fluid return temperature is less than the control valve setpoint value, and step (d) comprises recirculating the working fluid back through the secondary loop when the fluid return temperature is greater than the control valve setpoint value.

16. The method of claim 14 wherein said heat/cooling source is a cooling source (CH), step (b) comprises determining whether the value of the fluid return temperature fulfills the targeted temperature differential by checking whether the fluid return temperature is greater than a control valve setpoint value, step (c) comprises returning the working fluid to the return chamber of the split buffer when the fluid return temperature is greater than the control valve setpoint value, and step (d) comprises recirculating the working fluid back through the secondary loop when the fluid return temperature is less than the control valve setpoint value.

17. The method of claim 14 wherein comprising, over time, variably adjusting the targeted minimum temperature differential.

18. The method of claim 14 comprising, over time, variably adjusting the targeted minimum temperature differential based at least partly on an outdoor temperature.

19. The method of claim 15 comprising, over time, variably adjusting the control valve setpoint value.

20. The method of claim 19 comprising, over time, variably adjusting the control valve setpoint value based at least partly on an outdoor temperature.