L&T EduTech · Coursera specialization · 4 courses · 9 modules
Design of Ventilation & Air Conditioning Systems
A working study guide for the whole specialization — psychrometry, heat load, equipment selection, smoke control and demand-based ventilation — with every method carried through to a number against Doha design conditions, in SI and English units side by side.
This guide was written without access to the Coursera platform. The course pages could not be reached from the environment that produced it, so nothing here is a transcript of the lectures, slides or graded quizzes. It is reconstructed from the published course and specialization descriptions, the codes those courses name, and established design method.
Two consequences you must hold onto: the practice questions are written from first principles, not copied from Coursera — treat them as revision, not as a leaked question bank; and every table value that would normally be read out of an ASHRAE handbook is tagged Verify and left editable, because it was set from established magnitudes rather than read off the page.
How to trust the numbers
Engineering documents fail in a specific way: a plausible number with no traceable origin gets copied forward until someone builds it. Every quantitative block in this guide therefore carries one of three tags, and they mean different things.
Computed inside this guide from first principles. You can follow every step and reproduce it. Errors here are mine and are traceable.
Quoted from a named clause of a code or standard, cited inline. Confirm the edition your project mandates — values move between editions.
A representative magnitude, not a read table value. Never carry one of these into a deliverable without replacing it from the handbook or your project basis of design.
CLTD, SCL, CLF, shading coefficients and climatic design data all live in copyrighted ASHRAE tables that could not be opened while writing this. Reproducing them from memory to four significant figures would look authoritative and be untrustworthy — the worst combination in a technical document. Instead the method is exact, the magnitudes are defensible, and every one of them is an input box you overwrite.
Doha design basis
These inputs drive every worked example in the guide. They open on values representative of Doha and of QCS-compliant Gulf construction. Overwrite them with your project basis of design and the calculations re-solve.
ASHRAE Fundamentals Ch. 14 lists separate design conditions for peak dry bulb and for peak dew point. Doha's two are far apart, and that gap is the single most consequential fact in this guide — Case 1 sizes the airflow, Case 2 sizes the coil.
Indoor design & ventilation Standard
Zone geometry & envelope
Internal gains & cooling load factors
CLTD & SCL table values — replace these first Verify
Read the real values from ASHRAE Fundamentals (1997) Tables 30–34 at your latitude, month and solar hour, or from the calculation basis your project has already approved. The defaults are magnitudes for a 24°N July afternoon peak, not table entries.
K = 0.65 for a light-coloured surface will cut your roof load noticeably, and it is legitimate — but only if the finish reflectance is written into the specification and survives value engineering. Claiming the credit and then letting a dark membrane onto the roof is how a plant ends up 10% short on the hottest afternoon of the year. Default here is K = 1.00.
Specialization map
Four courses, nine modules, numbered in Coursera's order. The design dependency runs differently: C1's load feeds the equipment in C3 and the fans in C4, while C2 is sized by code and stands apart. A heat load error in C1 does not stay in C1; it propagates into the chiller you select in C3 and the fan you specify in C4.
| Course | Modules | Governing method | Output it hands forward |
|---|---|---|---|
| C1 Basics of Air Conditioning & Heat Load | 3 | Psychrometry; CLTD/CLF/SCL | Zone loads, airflow, coil duty, ADP |
| C2 Ventilation, Life Safety & Smoke Extraction | 2 | NFPA 92; ASHRAE 62.1; NBC | Pressurization and extract rates |
| C3 Equipment Selection, Design & Sizing | 2 | Block load, diversity, hydraulics | Chiller, pump, tower, duct, ESP |
| C4 Cutting-edge Technology | 2 | DCV; VRF; fan laws | Part-load and energy strategy |
Master mind map
Course 1 · Module 1
Refrigeration cycle & psychrometric properties
Two instruments underpin everything that follows: the pressure–enthalpy diagram, which tells you what the machine is doing, and the psychrometric chart, which tells you what the air is doing. Learn to read both and the rest of the specialization is bookkeeping.
The vapour-compression cycle
Four components, four processes, one closed loop. The refrigerant is not consumed — it is a courier, picking up heat where you don't want it and putting it down where you don't care. Everything you pay for is the cost of moving it uphill in pressure.
What each component actually constrains
- Compressor
- lift, capacity, part-load
- Condenser
- head pressure ← ambient
- Expansion device
- superheat control
- Evaporator
- suction pressure, ADP
In Doha the condenser is the binding constraint almost every time. Ambient sets head pressure, head pressure sets compressor work, and no amount of evaporator selection recovers it.
Compressor types, and where each belongs
| Centrifugal | Large plant, >300 TR. Best full-load efficiency; surge limits turndown. |
| Screw | 100–500 TR. Tolerates high lift — a real advantage in Gulf ambient. |
| Scroll | Small packaged, VRF. Good part-load in multiples. |
| Reciprocating | Legacy and small industrial. High lift, higher maintenance. |
| Absorption | Heat-driven. Only where waste heat is genuinely free. |
A vendor quotes COP or kW/TR at AHRI 550/590 standard conditions — 29.4 °C entering condenser water. Worked example D in Course 3 arrives at 40.6 °C entering condenser water for a Doha tower on a design day. That is an 11 K gap, and chiller power rises steeply with condenser temperature. If the submittal does not carry a performance point at your condenser water temperature, the efficiency number in it is not the efficiency you will buy.
Reading the psychrometric chart
The chart is a two-property map. Fix any two independent properties and every other one is determined — that is the whole idea. Below is the real chart with this guide's Doha state points plotted on it, so the abstractions have somewhere to land.
The eight processes, as directions
Every air-handling process is a vector on the chart. Learn the compass once and you can read any air-side schematic by eye.
Worked example A — state points from two properties
Given Doha's Case 1 outdoor design condition as dry bulb and wet bulb, derive every other property. This is the calculation the chart performs graphically.
Relative humidity came out at 10.8%. Does that pass a smell test for Doha in the afternoon? Yes — a 23.5 K wet-bulb depression is desert air, and the coast's humidity arrives at night and in the shoulder months, not at the 45 °C peak. If you had computed 40% RH at 45 °C you would be carrying 23 g/kg and an enthalpy near 105 kJ/kg, which is not a condition Doha reaches.
Cross-check the specific volume: 0.9105 m³/kg is 9.3% above the 0.8333 m³/kg standard-air basis. Hold onto that number — it is the reason the familiar 1.23 and 3010 shortcut coefficients drift in Module 2.
| Property | Room | OA Case 1 | OA Case 2 | Room (Eng) | Case 1 (Eng) | Case 2 (Eng) |
|---|---|---|---|---|---|---|
| Dry bulb | 24.00 °C | 45.00 °C | 31.60 °C | 75.2 °F | 113.0 °F | 88.9 °F |
| Wet bulb | 17.07 °C | 21.50 °C | 28.66 °C | 62.7 °F | 70.7 °F | 83.6 °F |
| Dew point | 12.95 °C | 7.42 °C | 27.80 °C | 55.3 °F | 45.3 °F | 82.0 °F |
| Humidity ratio | 9.299 g/kg | 6.395 g/kg | 23.825 g/kg | 65.1 gr/lb | 44.8 gr/lb | 166.8 gr/lb |
| Relative humidity | 50.0 % | 10.8 % | 80.4 % | — | — | — |
| Specific volume | 0.8544 | 0.9105 | 0.8964 | 13.69 ft³/lb | 14.59 ft³/lb | 14.36 ft³/lb |
| Enthalpy | 47.81 | 61.80 | 92.78 | 20.56 Btu/lb | 26.57 Btu/lb | 39.89 Btu/lb |
Per litre per second of ventilation air, relative to the room: Case 1 removes 11.5 g/h of moisture (a latent credit of 8.0 W per L/s), while Case 2 adds 58.3 g/h (a latent load of 40.5 W per L/s). Same building, same city, same code-required ventilation rate — and the latent duty swings by a factor of five and changes sign. Any load calculation that runs only the 45 °C condition has not seen the case that sizes the coil.
On a pressure–enthalpy diagram, the expansion process is drawn as a vertical line. Why?
A chiller submittal states 0.58 kW/TR with no qualifying condition. Which entering condenser water temperature does AHRI 550/590 use for nominal rating?
Doha Case 1 outdoor air holds 6.395 g/kg; the room is at 9.299 g/kg. What is the latent effect of ventilation air at this condition?
Which pair of properties can not be used together to fix a state on the psychrometric chart?
The apparatus dew point (ADP) is best described as:
A coil's bypass factor represents:
Why is direct evaporative cooling a poor proposal for a Doha project?
Which compressor type is most often preferred for 100–500 TR duty in Gulf ambient, on the grounds of high-lift tolerance?
Compute the enthalpy of moist air at 24.0 °C dry bulb with a humidity ratio of 9.299 g/kg. Use h = 1.006t + W(2501 + 1.86t).
Air at 45.0 °C has a vapour pressure of 1.03120 kPa. Saturation pressure at 45.0 °C is 9.59322 kPa. What is the relative humidity, in percent?
Course 1 · Module 2
Heat load calculation
The load is the deliverable everything downstream depends on. Get it wrong high and you buy plant that never modulates properly; get it wrong low and you own the problem for the life of the building. This module builds it component by component, twice — once at each Doha design condition.
Three distinctions that must stay straight
External vs internal
External gains swing with sun and ambient and are what the CLTD tables address. Internal gains are largely constant while the building is occupied. Only external gains care what time it is.
Sensible vs latent
Sensible changes temperature, latent changes moisture. They are added separately and met by different coil behaviour. A coil can be perfectly sized on total duty and still fail to hold humidity.
Room vs coil load
Ventilation air is not a room load. In a mixed-air system it never enters the room unconditioned — it hits the coil. Put it in the room load and you will oversize airflow and undersize nothing useful.
The CLTD / CLF / SCL method
A solid wall does not pass heat the instant the sun strikes it. Mass absorbs, stores and re-releases the gain hours later, and the radiant fraction of any gain warms surfaces first and the air afterwards. The CLTD family handles both lags with pre-computed tables: an equivalent temperature difference for opaque surfaces, a solar cooling load for glass, and a cooling load factor for internal gains.
Cooling load factors below 1.0 credit you for heat still stored in the structure at the calculation hour. That credit is only real if the plant runs continuously. If the system shuts down at night, CLF must be taken as 1.0 — the stored energy has nowhere to go but into the morning pull-down. A Gulf office on a 06:00 start with CLF = 0.87 in the calculation is a building that will not hold setpoint until mid-morning in August.
Worked example B — full zone heat load
A top-floor south-west corner office module in Doha: 96 m² (12 m × 8 m), 3.0 m ceiling, exposed roof, two exposed façades at 40% glazing, ten occupants. QCS-compliant envelope. Computed at both design conditions.
| Element | Value | English | Basis |
|---|---|---|---|
| Floor area | 96.00 m² | 1 033 ft² | 12 m × 8 m |
| Room volume | 288.0 m³ | 10 171 ft³ | 96 m² × 3.0 m ceiling |
| Exposed roof | 96.00 m² | 1 033 ft² | top floor |
| Glazing — south | 17.28 m² | 186.0 ft² | 12 m × 3.6 m × 0.40 |
| Glazing — west | 11.52 m² | 124.0 ft² | 8 m × 3.6 m × 0.40 |
| Opaque wall — south | 25.92 m² | 279.0 ft² | gross less glazing |
| Opaque wall — west | 17.28 m² | 186.0 ft² | gross less glazing |
| Shading coefficient | 0.3218 | — | SHGC 0.28 ÷ 0.87 |
| Ventilation — 62.1 VRP | 53.8 L/s | 114.0 CFM | 2.5×10 + 0.3×96, Ez = 1.0 |
| Infiltration | 12.00 L/s | 25.4 CFM | 0.15 ACH × 288 m³ |
| Surface | CLTD table | LM | K | Case 1 corrected | Case 2 corrected |
|---|---|---|---|---|---|
| Roof | 35 K | 0 | 1.00 | 45.35 K | 34.20 K |
| Wall — south | 16 K | −2 | 1.00 | 24.35 K | 13.20 K |
| Wall — west | 20 K | 0 | 1.00 | 30.35 K | 19.20 K |
| Glass — conduction | 13 K | — | — | 23.35 K | 12.20 K |
| Mean outdoor temperature | — | — | — | 38.25 °C | 27.10 °C |
Worked for the roof at Case 1: (35 + 0) × 1.00 + (25.5 − 24.0) + (38.25 − 29.4) = 35 + 1.5 + 8.85 = 45.35 K. The +8.85 K term is Doha doing the work — it is the gap between the local mean outdoor temperature and the 29.4 °C the base tables assume.
| Component | Case 1 · W | % RSH | Case 2 · W | % RSH | Case 1 · Btu/h |
|---|---|---|---|---|---|
| Roof | 1 306 | 15.9% | 985 | 14.3% | 4 457 |
| Glass solar — west | 1 594 | 19.4% | 1 594 | 23.2% | 5 440 |
| Glass solar — south | 806 | 9.8% | 806 | 11.7% | 2 752 |
| Glass conduction — south | 767 | 9.3% | 401 | 5.8% | 2 616 |
| Glass conduction — west | 511 | 6.2% | 267 | 3.9% | 1 744 |
| Opaque wall — south | 316 | 3.8% | 171 | 2.5% | 1 077 |
| Opaque wall — west | 262 | 3.2% | 166 | 2.4% | 895 |
| Equipment | 1 037 | 12.6% | 1 037 | 15.1% | 3 538 |
| People — sensible | 675 | 8.2% | 675 | 9.8% | 2 303 |
| Lighting | 668 | 8.1% | 668 | 9.7% | 2 280 |
| Infiltration — sensible | 283 | 3.4% | 106 | 1.5% | 964 |
| Room sensible heat · RSH | 8 225 | 100% | 6 876 | 100% | 28 064 |
| People — latent | 550 | — | 550 | — | 1 877 |
| Infiltration — latent | −98 | — | +496 | — | −335 |
| Room latent heat · RLH | 452 | — | 1 046 | — | 1 542 |
| Outdoor-air load (at the coil) | 826 | — | 2 699 | — | 2 819 |
| Grand total heat · GTH | 9 503 | 2.70 TR | 10 621 | 3.02 TR | 32 425 |
31.8 m² per ton (342 ft²/TR) for a top-floor corner office with 40% glazing in Doha. Gulf offices with a QCS-compliant envelope generally land between 25 and 35 m²/TR, and worst-case corner zones sit at the tight end. This is inside the band and toward the efficient side — consistent with the low-E glazing (SHGC 0.28) and insulated roof (U = 0.30) specified.
If your own calculation returns 15 m²/TR for a similar space, you have almost certainly double-counted ventilation into the room load, used CLF = 1.0 on everything, or left the glazing at clear-single SHGC. If it returns 55 m²/TR, check whether the roof was omitted or the solar orientation swapped.
Case 1 produces the larger room sensible load — 8 225 W against 6 876 W — because the 45 °C ambient drives conduction hard. Case 2 produces the larger grand total — 10 621 W against 9 503 W — because outdoor air at 23.8 g/kg carries a latent load of 2 699 W where Case 1's dry air carried only 826 W total.
So the two cases size different things: Case 1 sizes the airflow, the ductwork and the fan. Case 2 sizes the coil and the chiller. Running only the 0.4% dry-bulb condition — the habit most load calculations fall into — undersizes the cooling coil by 11.8% and leaves the building unable to hold humidity on exactly the days occupants complain about it.
Why RTS replaced CLTD, and when to still use CLTD
ASHRAE withdrew the CLTD/CLF/SCL method from the Fundamentals handbook and replaced it with the Radiant Time Series. The reason is structural, not cosmetic.
| Aspect | CLTD / CLF / SCL | Radiant Time Series (RTS) | Heat Balance (HB) |
|---|---|---|---|
| Status | Withdrawn (1997 Fundamentals) | Current recommended simplified method | Rigorous reference method |
| How lag is handled | Baked into fixed tables for representative constructions | Explicit conduction time series + radiant time series coefficients | Simultaneous surface energy balance, solved hourly |
| By hand? | Yes — this is its whole advantage | Painful — 24-hour convolution per surface | No |
| Your construction not in the table | Pick the nearest group and accept the error | Generate coefficients for the actual assembly | Modelled directly |
| Where you meet it | Exams, hand checks, legacy Carrier E20 | Current ASHRAE practice, spreadsheets | Inside HAP, TRACE, IES |
Use CLTD as a hand check, not as the submitted basis. Its enduring value is that one engineer with a calculator can reproduce it and catch an order-of-magnitude error in a software output in twenty minutes. Its weakness is that its tables encode specific wall constructions and a specific base condition, and the correction terms stretch further the more your project departs from them — a Doha project already carries a +8.85 K correction before anything else. If a subcontractor submits CLTD as the primary basis for a large plant, that is a legitimate technical comment: ask for RTS or a validated software output, and keep your CLTD as the independent check.
In a mixed-air system, where does the ventilation (outdoor air) load belong?
Correct the roof CLTD for Case 1. Table value 35 K, LM = 0, K = 1.00, indoor 24.0 °C, mean outdoor 38.25 °C. Use CLTDcorr = (CLTD + LM)·K + (25.5 − ti) + (to,mean − 29.4).
West glazing contributes 138 W/m² of solar load against south glazing's 47 W/m² in this July calculation. Why?
When must cooling load factors (CLF) be taken as 1.0?
Solar gain through the west glazing: area 11.52 m², SHGC 0.28, SCL 430 W/m². Remember the CLTD method uses shading coefficient, where SC = SHGC / 0.87.
Which design condition sizes the cooling coil for this Doha zone, and by how much would the other one miss it?
Should a cooling load factor be applied to the latent gain from occupants?
ASHRAE 62.1 ventilation rate procedure: office space, 10 occupants, 96 m², Rp = 2.5 L/s·person, Ra = 0.3 L/s·m², Ez = 1.0. Find Voz.
Why did ASHRAE withdraw CLTD/CLF/SCL in favour of the Radiant Time Series?
A subcontractor's load calculation for a similar Doha office returns 15 m²/TR. What is the most likely cause?
Course 1 · Module 3
Psychrometry × heat load — selecting the coil
Module 2 produced watts. Watts do not buy equipment. This module converts the load into the four numbers a supplier actually needs: airflow, supply air condition, apparatus dew point and total coil duty — and shows why the two Doha design cases hand you two different answers.
The three sensible heat factors
Three ratios, three different jobs. Confusing them is the most common error in this part of the syllabus, so anchor each one to the question it answers.
Worked example C — coil selection at the governing case
Carrying Example B's Case 2 loads forward: RSH 6 876 W, RLH 1 046 W, outdoor air 2 699 W, grand total 10 621 W. Supply air off the coil is taken at 13.0 °C.
Drawing the ESHF = 0.8390 line from the room condition to saturation gives ADP = 11.47 °C. Solving the actual mass and energy balance and extending the real process line gives 11.27 °C. The classical graphical method is within 0.20 K here — comfortably good enough for selection, and a fair reason to trust the chart construction the course teaches. The rigorous route is used above only because it closes the energy balance exactly, which lets every number in this guide be checked against every other.
The same zone at Case 1 — a dry coil
Run the identical procedure at the 45 °C design condition and the coil behaves like a different machine.
| Quantity | Case 1 · 45.0 °C DB | Case 2 · 27.8 °C DP | What it means |
|---|---|---|---|
| Room sensible RSH | 8 225 W | 6 876 W | Case 1 sizes airflow |
| Room latent RLH | 452 W | 1 046 W | occupants dominate in Case 1 |
| Outdoor-air sensible | 1 266 W | 473 W | 21 K vs 7.6 K temperature difference |
| Outdoor-air latent | −440 W | +2 226 W | changes sign |
| GSHF | 0.999 | 0.692 | Case 1 is essentially all sensible |
| Coil SHR | 0.998 | 0.693 | Case 1 coil runs dry |
| Supply airflow | 601 L/s | 502 L/s | 16% turndown at Case 2 |
| Grand total | 9 503 W · 2.70 TR | 10 621 W · 3.02 TR | Case 2 governs the coil |
At Case 1 the coil sensible heat ratio is 0.998 — the outdoor air is dry enough to absorb the entire occupant latent gain on its own, and almost no condensate forms. That is comfortable, but it has three consequences worth writing down:
- The apparatus dew point construction degenerates. With the process line nearly horizontal, ADP and bypass factor stop being meaningful selection quantities at this condition. Select on Case 2.
- Humidity control is unavailable when you might want it. A dry coil cannot dehumidify. Any space needing tight humidity (archives, laboratories, some clean areas) needs a separate strategy, not a bigger coil.
- Condensate drainage is still not optional. The plant runs wet for most of the year — the coil is only dry at this specific design extreme.
Fan, duct and VAV box: 601 L/s (1 274 CFM) from Case 1. Coil and chiller: 3.02 TR (10.62 kW) from Case 2, at ADP 11.27 °C and coil SHR 0.693. Specify both duties on the equipment schedule. A supplier given only the total tonnage will select a coil with a sensible heat ratio near 0.75–0.80 by default, and it will not hold the room at 50% RH on a September morning in Doha.
Which sensible heat factor is used to locate the apparatus dew point?
Mixed air enters a coil at 24.76 °C, leaves at 13.00 °C, and the apparatus dew point is 11.27 °C. Compute the bypass factor.
At Case 1 the coil sensible heat ratio is 0.998. What follows?
Dry-air mass flow is 0.61120 kg/s and the supply air specific volume is 0.8219 m³/kg. What is the supply airflow in L/s?
A supplier is given only "3.02 TR" for this coil. What is the most likely failure mode?
Why does this guide solve the supply state from a mass and energy balance rather than reading it off the ESHF construction?
Between Case 1 and Case 2 the outdoor-air latent load moves from −440 W to +2 226 W. What must the air system be able to do?
Room sensible heat is 6 876 W and room latent heat is 1 046 W. Compute RSHF.
Course 3 · Modules 1–2
Equipment selection, design & sizing
Course 1 produced a zone duty. This course turns it into hardware you can put on a schedule and defend in a bid evaluation — a chiller, a tower, two pump circuits, a duct and a fan — and shows where the numbers a vendor quotes stop meaning what you think they mean.
The plant, end to end
Two water loops and one air loop, in series thermally. Heat leaves the room in air, crosses into chilled water at the coil, crosses into refrigerant at the evaporator, is pumped uphill in pressure by the compressor, crosses into condenser water, and finally leaves into the Doha air at the tower. Every crossing costs a temperature difference, and every temperature difference costs compressor power.
Cooling towers set the condenser temperature, and the condenser sets the bill
A cooling tower cannot cool water below the ambient wet bulb. What it achieves is the wet bulb plus an approach — the margin the tower is selected for. Range is the temperature drop across the tower, fixed by flow and heat rejection, not by tower size.
AHRI 550/590 rates chillers at 29.4 °C entering condenser water. This plant, on a Doha design day, runs at 40.6 °C. Reversed-Carnot scaling across that gap, at a fixed evaporator leaving temperature of 6.7 °C:
| Condition | Condenser in → out | Carnot COP | Quoted 0.58 kW/TR becomes |
|---|---|---|---|
| AHRI 550/590 nominal | 29.4 → 34.9 °C | 9.92 | 0.58 kW/TR |
| Doha design day | 40.6 → 46.1 °C | 7.10 | ≈ 0.81 kW/TR |
| Power multiplier | Δ 11.2 K | ×1.40 | ≈ +40% |
On this 98 TR plant that is roughly 57 kW rising to 80 kW at design. Second-law efficiency is not actually constant across the range, so treat +40% as an argument, not an answer. The argument is this: a kW/TR figure with no stated condenser condition is not a number you can evaluate a bid on. Ask for the performance point at your condenser water temperature, and ask for it in writing.
Pumping arrangements — where the operating cost hides
Constant primary
One fixed-speed circuit through evaporator and coils, three-way valves at the loads. Simple and robust. Pumping energy is constant at full flow regardless of load — the expensive option over a year.
Primary–secondary
Constant flow through the chillers, variable flow to the loads, decoupled by a bypass. The conventional arrangement. Watch the decoupler: reverse flow through it is the classic low delta-T syndrome symptom.
Variable primary
One variable-speed circuit, two-way valves, minimum-flow bypass to protect the evaporator. Lowest pumping energy and fewest pumps, but requires a chiller that tolerates variable evaporator flow — confirm that in the submittal, do not assume it.
Designing at a wide ΔT saves pumping energy (below: 8 K needs 25% less flow than 6 K, and 25% less pump power). But if coils are selected on optimistic entering conditions, or three-way valves leak, or control valves are oversized, the plant never achieves the design return temperature. Flow rises to compensate, pumps run out, and chillers load up on flow rather than tonnage. The wider your design ΔT, the less margin you have before this bites — so a wide ΔT must be matched by two-way valves, correctly sized control valves, and coil selections you have actually checked.
Worked example D — from zone load to plant
One zone does not buy a chiller. Scaling to a plant needs two judgements that are easy to get wrong in opposite directions: how much of the building looks like the worst zone, and how much of the peak is coincident.
- 2.22 GPM/TR chilled water at ΔT 6 K. The rule is 24/ΔT(°F); 6 K = 10.8 °F gives 2.22. Exact.
- 3.12 GPM/TR condenser water against the familiar 3 GPM/ton. Slightly high because Doha's condenser runs hot, which is the right direction.
- 5.76 m/s duct velocity, inside the 5–7 m/s band for a main. Above 8 m/s you are buying noise complaints.
- 1.37 W/(L/s) specific fan power, comfortably under typical 1.6–2.0 limits for a system of this type.
The 30% perimeter fraction and the 0.90 diversity factor together moved the answer from a naive 5 100 × 110.6 = 564 kW down to 345 kW — a factor of 1.63. Both must be stated explicitly in the basis of design and agreed, not buried in a spreadsheet. If a reviewer disagrees with either, the plant size changes materially, and that is a conversation to have at design stage rather than after the chillers are ordered.
Technical bid evaluation — what to demand and what to reject
A bid evaluation that compares price against nameplate capacity compares almost nothing. These are the points where offers actually differ.
| Equipment | Demand in the offer | Reject or query if |
|---|---|---|
| Chiller | Performance at site condenser condition (40.6 °C entering), not only AHRI; IPLV and full-load kW/TR; evaporator minimum flow if variable primary; fouling factor used; sound power by octave band | Only AHRI-condition performance quoted; fouling factor of zero; IPLV cited alone as the efficiency claim |
| Cooling tower | Thermal performance certified at the specified wet bulb, range and approach; drift rate; fan power; water consumption; materials for Gulf water chemistry | Performance quoted at a wet bulb below the project design; approach quoted without the corresponding flow |
| AHU | Coil entering and leaving conditions, total and sensible duty, face velocity, rows and fin spacing; fan curve with the operating point marked; casing leakage and thermal-bridging class | Total duty given without SHR; face velocity above ~2.5 m/s with no drift eliminator; fan curve absent |
| Pump | Certified curve with duty point, efficiency and NPSH required at duty; motor margin; operation away from best efficiency point | Duty point sits far from BEP; NPSH margin not demonstrated; curve is a catalogue family rather than the selected impeller |
| Fan | Curve with system resistance overlaid; specific fan power; sound power; motor and drive losses stated as included or not | Operating point on the unstable left of the curve; SFP quoted excluding drive losses |
Ask every bidder to quote performance at the project's stated conditions on a single common data sheet you issue, rather than accepting each vendor's own format. Offers quoted at differing conditions cannot be compared, and the difference is routinely larger than the price spread you are agonising over. This one requirement converts a bid evaluation from a price comparison into an engineering comparison.
A cooling tower is selected for a 5 K approach at a 30.6 °C design wet bulb. What is the leaving water temperature, and can a larger tower beat it?
A 345.1 kW plant operates at COP 5.8. Compute the condenser water flow at a 5 K range. Use Qrej = Q(1 + 1/COP) and ṁ = Qrej/(4.187 × range).
A chiller is quoted at 0.58 kW/TR with no condition stated. Your plant runs at 40.6 °C entering condenser water. What is the most defensible response?
Widening design ΔT from 6 K to 8 K cuts chilled water flow 25% and pump power with it. What is the corresponding risk?
Build the external static pressure: filters 150 Pa, cooling coil 250 Pa, supply duct and fittings 300 Pa, diffusers 50 Pa, return path 80 Pa, sound attenuator 60 Pa.
An AHU offer states "cooling capacity 3.02 TR" and nothing else. What is the first thing you ask for?
The block load calculation blended 110.6 W/m² perimeter with 60 W/m² interior at 30/70, then applied 0.90 diversity — moving 564 kW down to 345 kW. What does this demand?
Which pumping arrangement gives the lowest pumping energy, and what must be confirmed before specifying it?
Course 2 · Modules 1–2
Ventilation, life safety & smoke extraction
The one branch of the specialization that is not sized by thermal load. These systems are sized by code, geometry and fire size, they are tested by a witnessed procedure rather than a performance curve, and they are the systems most likely to fail at handover because nobody checked the criteria until the authority arrived.
Stack effect — and why Doha reverses it
Warm air is less dense. In a tall building a temperature difference across the façade drives a vertical pressure gradient, and somewhere up the building sits a neutral pressure plane where inside and outside pressures are equal. Above it, flow is one way; below it, the other.
Pressurization — two criteria, and the one that governs
A stairwell pressurization system must satisfy two requirements at once, and they pull in opposite directions. Enough pressure to keep smoke out; not so much that a person cannot open the door.
Worked example E — life safety systems
E1 · Car park ventilation and smoke extraction
An ACH rate takes no account of how many vehicles run, for how long, or where. It is a code compliance floor, not an engineering answer. Where the code permits, CO-based demand ventilation sized on actual vehicle movements is both more defensible and much cheaper to run — but the smoke-extraction duty still has to be met at full rate on demand, so the fans do not get smaller. Confirm which rate your authority mandates before sizing anything.
E2 · Stairwell pressurization
The fan is sized on the open-door case at 3 780 L/s. With every door closed, that same fan drives the stairwell far past 50 Pa — and the calculation above says the door force fails at about 100 Pa. A fixed-speed fan with no relief will therefore trap occupants behind doors they cannot open, which is the precise opposite of the system's purpose. A barometric relief damper or a variable-speed fan on pressure control is not an enhancement here; it is what makes the system safe.
E3 · Atrium smoke extraction
96.8 m³/s is nearly ten times the entire normal ventilation rate of the 2 000 m² car park above. Smoke control is not a variation on comfort ventilation; it is a different order of magnitude, and it drives shaft space, structural openings, generator sizing and façade make-up air.
Note also that most of the plume mass is entrained room air, not combustion products — which is why the smoke layer is only 53 °C, and why the extract rate grows with the 5/3 power of clear height. Lowering the design smoke-layer interface by a few metres reduces the extract rate dramatically. That is a coordination conversation with the architect, and it is worth having early.
- The design fire size, and who approved it. Everything scales from it. A 5 MW assumption presented without authority agreement is an unaccepted risk, not a design.
- Make-up air. You cannot extract 96.8 m³/s from a sealed atrium. Make-up must be provided at low level below 1.0 m/s so it does not disturb the smoke layer — and it must be shown on the drawings, not assumed.
- Door-force calculation at the maximum pressure, not just at the design pressure. See E2 above.
- Fan and damper response times, and the cause-and-effect matrix that drives them, cross-checked against the fire alarm submittal.
- The witnessed test procedure, agreed before installation. Smoke-control commissioning is a pass/fail demonstration in front of the authority; discovering the criteria at that point is far too late.
In Doha during the cooling season, which way does the stack effect drive infiltration in a tall building?
Open-door criterion: 2 doors open, each 0.9 m × 2.1 m, minimum velocity 1.0 m/s. What airflow is required?
A stairwell fan is sized on the open-door case with no pressure relief. What happens when all doors close?
Door force at 50 Pa: closer force 30 N, door 0.9 m wide × 2.1 m high, knob 0.075 m from the edge. Use F = Fdc + Kd·W·A·Δp / [2(W − d)], Kd = 1.0.
The atrium plume calculation gives 104.65 kg/s of smoke at only 53.4 °C for a 5 MW fire. Why is it so cool?
An atrium submittal specifies 96.8 m³/s of extract and shows no make-up air provision. Is this acceptable?
A 2 000 m² car park with a 3.0 m ceiling requires 10 air changes per hour for smoke extraction. What is the rate in L/s?
Which item, missing from a smoke-control submittal, invalidates every other number in it?
Course 4 · Modules 1–2
Cutting-edge technology
Design load happens on a handful of afternoons a year. Everything else is part load — which is where the energy actually goes, and where the technologies in this course either earn their premium or quietly fail to.
Demand control ventilation — derive the setpoint, never copy it
CO₂ is not a pollutant to be controlled at some universal threshold. It is a proxy for occupancy. At steady state the difference between indoor and outdoor concentration tells you how much outdoor air each person is receiving — and the setpoint that corresponds to your design ventilation rate is specific to your design.
| Occupants | Vbz required | Per person | Steady-state rise | Setpoint at 420 ppm outdoor |
|---|---|---|---|---|
| 10 (design) | 53.8 L/s | 5.38 L/s | 967 ppm | 1 387 ppm |
| 8 | 48.8 L/s | 6.10 L/s | 852 ppm | 1 272 ppm |
| 5 | 41.3 L/s | 8.26 L/s | 630 ppm | 1 050 ppm |
| 3 | 36.3 L/s | 12.10 L/s | 430 ppm | 850 ppm |
First: this zone's design condition corresponds to 1 387 ppm, not the 1 000 ppm that gets copied into specifications as though it were a code limit. Setting the controller to 1 000 ppm here would force more outdoor air than ASHRAE 62.1 requires, all day, every day — a demand control system that increases energy consumption.
Second: halving occupancy from 10 to 5 does not halve the outdoor air. It falls from 53.8 to 41.3 L/s — a 23% reduction, not 50% — because the area component Ra × Az = 28.8 L/s does not depend on people and must keep flowing. In this zone the area term is the larger half of the requirement. DCV vendors who quote savings proportional to occupancy have not read the ventilation rate procedure.
Variable refrigerant flow — and where it does not belong
What VRF genuinely offers
- Excellent part-load efficiency through inverter compressors
- Individual zone control without terminal reheat
- Heat recovery between simultaneously heating and cooling zones
- No plant room, no chilled water distribution, small risers
What it does not solve, and must be checked
- Ventilation. VRF conditions recirculated room air. Outdoor air still needs a dedicated unit — and in Doha that unit carries the entire 2 226 W latent load computed in Course 1.
- Dehumidification at part load. A VRF unit modulating on sensible demand can leave a Gulf space cool and humid.
- Refrigerant volume limits for occupied spaces (ISO 5149 / EN 378 concentration limits).
- Condenser ambient rating. Confirm capacity at 45 °C+, not at a 35 °C nominal rating.
"State the net cooling capacity and input power at 46 °C ambient with the specified refrigerant pipe run and lift, and state the latent capacity at the space design condition." Nominal VRF capacities are typically published at 35 °C ambient with a short, level pipe run. Doha's design ambient, a long riser and a real lift can each take a bite out of capacity, and they compound. A system selected on nominal figures can be materially short on the day it is needed.
Fans and the fan laws
The fan laws are the strongest argument in building services, because the third one is cubic. Everything about variable-speed control follows from it.
Worked example F — DCV saving and fan turndown
In this zone, DCV avoids 627 W at design and fan turndown avoids 402 W at 80% speed — and the fan saving is available whenever load is below design, which is nearly always, while the DCV saving requires the space to actually be under-occupied. On a single office zone with a modest ventilation rate, variable speed on the fan is the stronger investment. DCV earns its keep in spaces with high and genuinely variable occupant density — auditoria, training rooms, prayer halls, conference centres — where the people component dominates the area component. Applying it to a lightly-occupied cellular office is a specification habit, not an analysis.
Derive the CO₂ setpoint for this zone at design: 10 occupants, Vbz = 53.8 L/s, generation 0.0052 L/s per person, outdoor 420 ppm.
Occupancy halves from 10 to 5. By how much may the outdoor air rate fall under ASHRAE 62.1?
A variable-speed fan runs at 80% of design speed with an unchanged system curve. What percentage of design power does it draw?
An engineer proposes throttling a damper to reduce airflow 20%, citing the fan laws to claim a 49% power saving. What is wrong?
A VRF system is proposed for a Doha office. Which of these does VRF not address?
A fan delivers 601 L/s and draws 823 W. Compute the specific fan power.
A VRF vendor quotes nominal capacity at 35 °C ambient. What must you require for a Doha project?
In this 96 m² office zone, which measure delivers the more reliable saving — and why?
Reference
Checklists, formulae & codes
Five review checklists, a printable formula sheet and a codes register. Ticks are stored in this browser, so the checklists survive a reload and can be worked through across several sittings.
Formula sheet
Everything used in this guide, in one place. This section prints cleanly.
| Quantity | Relation | Units |
|---|---|---|
| Humidity ratio | W = 0.621945 · p_w / (p − p_w) | kg/kg dry air |
| From wet bulb | W = [(2501 − 2.326·t_wb)·W_s,wb − 1.006(t − t_wb)] / [2501 + 1.86t − 4.186·t_wb] | kg/kg |
| Relative humidity | φ = p_w / p_ws(t) | — |
| Enthalpy | h = 1.006t + W(2501 + 1.86t) | kJ/kg dry air |
| Specific volume | v = 0.287042(t + 273.15)(1 + 1.607858W) / p | m³/kg dry air |
| Mixing | W_mix = (ṁ₁W₁ + ṁ₂W₂)/(ṁ₁+ṁ₂) — mass-weighted, likewise for h | — |
| Quantity | SI shortcut | English shortcut | Exact |
|---|---|---|---|
| Sensible | q = 1.23 · V · Δt | q = 1.10 · CFM · Δt | q = ṁ · c_p · Δt |
| Latent | q = 3010 · V · ΔW | q = 4840 · CFM · ΔW | q = ṁ · h_fg · ΔW |
| Total | q = 1.20 · V · Δh | q = 4.5 · CFM · Δh | q = ṁ · Δh |
1.23, 3010 and 1.20 assume 1.2 kg/m³ (v = 0.8333 m³/kg). Doha outdoor air at the peak dry-bulb condition has v = 0.9105 m³/kg — 9.3% lighter — so the shortcut overstates the outdoor-air load by 11.3% at Case 1 and 5.8% at Case 2, and produced a physically impossible grand sensible heat factor of 1.0026 before the rigorous method was applied. Use ṁ · Δh whenever air is far from standard density: hot ambient, high altitude, or a cold supply duct.
| Quantity | Relation |
|---|---|
| Opaque surface | q = U·A·CLTD_corr, CLTD_corr = (CLTD+LM)·K + (25.5 − t_i) + (t_o,mean − 29.4) |
| Glass solar | q = A·SC·SCL, SC = SHGC / 0.87 |
| Ventilation (62.1) | V_bz = R_p·P_z + R_a·A_z, V_oz = V_bz / E_z |
| Sensible heat factors | RSHF = RSH/(RSH+RLH); GSHF = (RSH+OA_s)/GTH; ESHF = (RSH+BF·OA_s)/(RSH+BF·OA_s+RLH+BF·OA_l) |
| Bypass factor | BF = (t_leaving − ADP)/(t_entering − ADP) |
| Chilled water flow | ṁ = Q/(c_p·ΔT), c_p = 4.187 kJ/kg·K · rule: GPM/TR = 24/ΔT(°F) |
| Heat rejection | Q_rej = Q(1 + 1/COP) |
| Tower | leaving = WB + approach; range = entering − leaving |
| Pump power | P = ρgQH/η |
| Fan laws | Q ∝ N; p ∝ N²; W ∝ N³ — fixed system curve only |
| Specific fan power | SFP = W_fan / V [W per L/s], including motor and drive losses |
| Quantity | Relation |
|---|---|
| Stack effect | Δp = 3463·h·(1/T_o − 1/T_i) — negative in the Gulf cooling season |
| Leakage flow | Q = 0.827·A_leak·Δp^0.5 |
| Open-door criterion | Q = N·A_door·v_min, v_min ≈ 0.75–1.0 m/s |
| Door force | F = F_dc + K_d·W·A·Δp / [2(W − d)] ≤ 133 N |
| Plume mass (z > z_l) | ṁ_p = 0.071·Q_c^(1/3)·z^(5/3) + 0.0018·Q_c |
| Flame height | z_l = 0.166·Q_c^0.4, Q_c = 0.70·Q |
| Smoke temperature | T_s = T_a + Q_c/(ṁ_p·c_p), ρ_s = 353/T_s |
| DCV setpoint | C = C_o + 10⁶·N·G/V_oz, G ≈ 0.0052 L/s per person |
| From | To | Multiply by |
|---|---|---|
| kW | TR (tons refrigeration) | 0.284345 |
| kW | Btu/h | 3 412.14 |
| L/s | CFM | 2.11888 |
| L/s | US GPM | 15.85032 |
| kJ/kg | Btu/lb | 0.429923 |
| kg/kg | grains/lb | 7 000 |
| Pa | in. w.g. | 0.0040146 |
| m/s | fpm | 196.850 |
| m² | ft² | 10.7639 |
| °C | °F | ×9/5 + 32 |
Codes & standards register
| Document | Governs | Where it bites in this guide |
|---|---|---|
| ASHRAE Fundamentals | Psychrometrics (Ch. 1), climatic design data (Ch. 14), ventilation and infiltration (Ch. 16), load calculation (Ch. 18) | Every state point; both Doha design conditions; RTS method |
| ASHRAE 55 | Thermal environmental conditions for human occupancy | Indoor design condition and acceptable humidity range |
| ASHRAE 62.1 | Ventilation for acceptable indoor air quality | Rp/Ra/Ez; the area component DCV may not turn down |
| ASHRAE 90.1 | Energy standard — envelope, lighting power density, equipment efficiency, fan power | LPD, SFP limits, minimum chiller efficiency |
| AHRI 550/590 | Water-chilling package performance rating | The 29.4 °C condenser rating condition, and why it is not yours |
| NFPA 92 | Smoke control systems | Pressurization criteria, plume equations, extract rates |
| NFPA 101 / IBC | Life safety — means of egress | The 133 N door-opening force limit |
| NFPA 88A | Parking structures | Car park ventilation and CO control |
| ISO 5149 / EN 378 | Refrigerating systems — safety and environmental requirements | VRF refrigerant concentration limits in occupied spaces |
| ISO 16890 | Air filter classification | Filter selection and the mid-life pressure drop used in ESP |
| QCS (Qatar Construction Specifications) | The governing national specification; adopts and amends international standards | Envelope U-values, glazing SHGC, ventilation and life-safety requirements |
Where QCS and an international standard disagree, the prime contract decides which prevails, and it usually names the more onerous. Confirm the edition of every document the contract invokes: ventilation rates, efficiency floors and design climatic data have all moved between editions, and a calculation to the wrong edition is non-compliant even when the arithmetic is perfect.
Sources & limits of this guide
It was written without access to the Coursera platform — the course pages could not be reached from the environment that produced it. Nothing here is a transcript of the lectures, slides or graded quizzes. The structure follows the published course and specialization descriptions; the engineering follows the codes those courses name and established design method.
Practical consequences: the practice questions are authored from first principles, not reproduced from Coursera — they are revision, not a question bank, and reproducing the real ones would breach the honor code in any case. Every value that would normally be read from a copyrighted ASHRAE table is tagged Verify and left editable. Replace those before any figure here reaches a deliverable.
Verification performed
- Every worked example was computed independently in Python and the page's own calculations reconciled against it — agreement better than 0.04% on all published quantities.
- The coil energy balance (coil duty against room load plus outdoor-air load) closes to 0.000%, which is what allows an error anywhere in the chain to be detected.
- Psychrometric routines were checked against ASHRAE reference values: pws(25 °C) = 3.16922 kPa against a published 3.1698 kPa.
- Results were cross-checked against independent rules of thumb — m²/TR, GPM/TR, duct velocity, specific fan power — each stated in the relevant sanity-check block.
- The classical ESHF/ADP graphical construction was checked against the rigorous mass-and-energy solution and agrees within 0.20 K.
Course pages this guide covers
- Design of Ventilation & Air Conditioning System for Buildings — specialization
- C1 — Basics of Air Conditioning & Heat Load Calculation
- C2 — Ventilation, Life Safety, and Smoke Extraction in Building
- C3 — Air Conditioning Equipment Selection, Design and Sizing
- C4 — Cutting-edge Technology in Air Conditioning System
Course structure was corroborated from public listings; the pages themselves were unreachable from the build environment. This guide is not affiliated with, endorsed by, or derived from Coursera or L&T EduTech course material.